JPWO2009147826A1 - Refrigeration cycle equipment - Google Patents

Abstract

The refrigeration cycle apparatus 100 includes a low-pressure stage compressor 113, a high-pressure stage compressor 101, a radiator 103, a gas-liquid separator 107, an expansion valve 109, an expander 105, and an evaporator 111. The low-pressure compressor 113 and the expander 105 are connected by a shaft 116, and the low-pressure compressor 113 is driven by the power recovered by the expander 105 from the refrigerant. The low-pressure compressor 113 and the high-pressure compressor 101 are connected in series by an intermediate pressure channel 114. The gas-liquid separator 107 and the intermediate pressure flow path 114 are connected by a forward return path 115. The forward recirculation path 115 is configured so that the refrigerant can flow in both directions. By controlling the opening degree of the expansion valve 109, the refrigerant flow rate in the forward recirculation path 115 is adjusted.

Description

  The present invention relates to a refrigeration cycle apparatus.

  As a refrigeration cycle apparatus used for an air conditioner or a hot water heater, as shown in FIG. 13, one having a first compressor 801a, a radiator 802, an expander 803, a heat absorber 804, and a second compressor 801b is known. (Patent Document 1). The second compressor 801b and the expander 803 are connected by a rotating shaft 806, and the driving force of the second compressor 801b is covered by the power accompanying the expansion of the refrigerant in the expander 803. Therefore, it is possible to reduce the power of the first compressor 801a that is consumed to increase the refrigerant pressure to a predetermined pressure.

  According to the refrigeration cycle apparatus shown in FIG. 13, the rotational speed of the expander 803 and the rotational speed of the second compressor 801b coincide. Further, since the refrigerant expanded by the expander 803 is compressed by the second compressor 801b through the heat absorber 804, the mass flow rate of the refrigerant passing through the expander 803 and the mass flow rate of the refrigerant passing through the second compressor 801b. Matches. Furthermore, since the suction volume of the expander 803 and the suction volume of the second compressor 801b are determined at each design time point, restrictions are imposed on the suction refrigerant density of the expander 803 and the suction refrigerant density of the second compressor 801b.

That is, the product of the suction refrigerant density [rho _exi the suction volume V _exi and the expander 803 of the expander 803, the product of the suction refrigerant density [rho _c2i of the suction volume V _C2i of the second compressor 801b and the second compressor 801b Is always equal to Between the suction refrigerant density of the expander 803 and the suction refrigerant density of the second compressor _{801b, (ρ exi / ρ c2i} ) = relationship always holds the (V _C2i / V _exi). This relationship is called a constraint with a constant density ratio. To operate the refrigeration cycle apparatus at optimum efficiency, it is essential to be able to freely adjust the density [rho _exi and the density [rho _c2i depending on external conditions such as seasonal and weather. However, it can not be freely adjusted certain the density [rho _exi and the density [rho _c2i is constraint of constant density ratio, realization of efficient operation becomes difficult.

  In order to solve these problems, a refrigeration cycle apparatus shown in FIG. 14 has been proposed (Patent Document 2). The refrigeration cycle apparatus shown in FIG. 14 includes a two-stage compressor 903, a radiator 904, an expander 905, a gas-liquid separator 906, an evaporator 908, a gas injection circuit 910, and a bypass circuit 911. The two-stage compressor 903 includes a low-pressure stage compressor 901 and a high-pressure stage compressor 902. The low pressure compressor 901 and the expander 905 are connected by a rotating shaft. The bypass circuit 911 is provided with a flow rate control valve 913. By appropriately controlling the opening degree of the flow control valve 913 and allowing a part of the refrigerant to flow through the bypass circuit 911, it is possible to avoid the restriction of a constant density ratio.

JP 2003-307358 A JP 2006-71257 A

However, when a part of the refrigerant flows through the bypass circuit 911, the amount of refrigerant contributing to power recovery by the expander 905 decreases, and there is a problem that power recovery efficiency deteriorates. This problem becomes remarkable when, for example, the refrigeration cycle apparatus having the same design is applied to each of the heat pump hot water floor heater and the air conditioner. Respect CO ₂ refrigeration cycle device of one design, the density ratio can realize optimum efficiency where (ρ _exi / ρ _c2i) The present inventors have calculated, at rated conditions of floor heating 7.13, 3 in rated conditions of cooling It was 2.98 in the rated condition of heating. If the low-pressure compressor 901 and the expander 905 are designed in accordance with floor heating, 49.6% refrigerant flows through the bypass circuit 911 during cooling, and 58.2% refrigerant flows through the bypass circuit 911 during heating. In other words, the power that can be recovered is reduced to about half that of floor heating.

  An object of the present invention is to provide a refrigeration cycle apparatus capable of efficient power recovery while avoiding the restriction of a constant density ratio.

That is, the present invention
A positive displacement low-pressure stage compressor for pre-compressing the refrigerant;
A high-pressure stage compressor for further compressing the refrigerant pre-compressed by the low-pressure stage compressor;
An intermediate pressure flow path connecting the low pressure stage compressor and the high pressure stage compressor in series so that the refrigerant pre-compressed by the low pressure stage compressor is sent to the high pressure stage compressor;
A radiator for cooling the refrigerant compressed by the high-pressure stage compressor;
It is coaxially connected to the low-pressure stage compressor so that power transmission is performed, and is configured so that the entire amount of the refrigerant cooled by the radiator passes, and for recovering power by expanding the refrigerant A positive displacement expander;
A gas-liquid separator for separating the refrigerant expanded by the expander into a gas refrigerant and a liquid refrigerant;
An evaporator for evaporating the liquid refrigerant separated by the gas-liquid separator;
An expansion valve with variable opening provided on a flow path between the liquid refrigerant outlet of the gas-liquid separator and the inlet of the evaporator;
The refrigerant stored in the gas-liquid separator is preliminarily compressed by the low-pressure stage compressor in a first flow state where the refrigerant is led to the inlet of the high-pressure stage compressor without passing through the evaporator and the low-pressure stage compressor. A recirculation path connecting the intermediate pressure flow path and the gas-liquid separator so as to be able to switch between a second flow state in which a part of the refrigerant circulates back to the gas-liquid separator;
A controller for adjusting the flow rate of the refrigerant in the forward / return path in each of the first circulation state and the second circulation state by controlling the opening of the expansion valve;
A refrigeration cycle apparatus is provided.

  According to the present invention, when the suction volume of the low-pressure compressor is insufficient compared to the suction volume of the expander, the gas refrigerant is sent from the gas-liquid separator to the intermediate pressure channel through the forward recirculation path. And sucked into the high-pressure stage compressor. Thereby, the flow volume balance of a refrigerating cycle is materialized. On the other hand, when the suction volume of the expander is insufficient compared to the suction volume of the low-pressure stage compressor, part of the gas refrigerant pre-compressed by the low-pressure stage compressor passes through the intermediate pressure flow path and the forward recirculation path. It is sent to the gas-liquid separator. Thereby, the flow volume balance of a refrigerating cycle is materialized. Regardless of the value of the suction volume of the low-pressure compressor and the suction volume of the expander (design value), the flow rate balance of the refrigeration cycle is established by the action of the forward recirculation path.

  On the other hand, the pressure in the gas-liquid separator can be freely adjusted by the expansion valve. The refrigerant pressure on the radiator side can be arbitrarily adjusted by adjusting the pressure in the gas-liquid separator. For example, when the expansion valve is fully opened under an arbitrary operating condition, the pressure in the gas-liquid separator approaches the evaporating pressure of the refrigerant in the evaporator as much as possible. Then, since the gas-liquid separator and the intermediate pressure flow path are connected by the forward recirculation path, the pressure difference between the inlet and the outlet of the low-pressure stage compressor approaches zero as much as possible. That is, the amount of compression work of the low-pressure stage compressor is reduced. On the other hand, since the pressure difference between the inlet and the outlet of the expander increases, the power recovery amount of the expander increases. Based on this relationship (power recovery amount)> (compression work amount), the rotation speeds of the expander and the low-pressure compressor are increased. As a result, since the discharge refrigerant flow rate of the expander becomes excessive with respect to the discharge refrigerant flow rate of the high-pressure compressor, the refrigerant pressure on the radiator side decreases. As a result, the power recovery amount of the expander is reduced, and is balanced with the compression work amount of the low-pressure compressor, so that the refrigeration cycle is stabilized. That is, when the expansion valve is opened, the refrigerant pressure on the radiator side can be reduced.

  Conversely, as the expansion valve is gradually throttled, the pressure in the gas-liquid separator gradually increases. Since the gas-liquid separator and the intermediate pressure flow path are connected by the forward recirculation path, the pressure difference between the inlet and the outlet of the low-pressure compressor is increased. That is, the compression work of the low-pressure stage compressor increases. On the other hand, since the pressure difference between the inlet and outlet of the expander decreases, the power recovery amount of the expander decreases. Based on this relationship (power recovery amount) <(compression work amount), the rotation speeds of the expander and the low-pressure compressor are reduced. As a result, since the discharge refrigerant flow rate of the expander is insufficient with respect to the discharge refrigerant flow rate of the high-pressure compressor, the refrigerant pressure on the radiator side increases. As a result, the power recovery amount of the expander is increased to balance the amount of compression work of the low-pressure compressor, and the refrigeration cycle is stabilized. That is, if the expansion valve is throttled, the refrigerant pressure on the radiator side can be increased.

  Thus, if the opening degree of the expansion valve is appropriately controlled and thereby the rotational speeds of the expander and the low-pressure compressor are controlled, the refrigerant pressure on the radiator side can always be optimally adjusted. In addition, since the entire amount of refrigerant passes through the expander, efficient power recovery can be performed. Even if the refrigerant recirculates in the forward recirculation path (second circulation state) and a part of the recovered power is consumed for the circulation of the refrigerant, according to the present invention, the conventional example of flowing the refrigerant through the bypass circuit (FIG. 14)) in terms of energy balance. Therefore, it is possible to operate a refrigeration cycle apparatus including an expander having a volume ratio suitable for an application and a low-pressure compressor at a pressure / temperature condition desirable from the viewpoint of energy efficiency.

  The theory described above can be established regardless of the value of the volume ratio between the low-pressure stage compressor and the expander. Therefore, according to the present invention, the low-pressure stage compressor and the expander can be designed to have an arbitrary volume ratio that can theoretically minimize the annual power consumption. That is, according to the present invention, the degree of freedom in designing the refrigeration cycle apparatus is also increased.

The block diagram of the refrigerating-cycle apparatus concerning 1st Embodiment of this invention. Configuration diagram of multi-function heat pump system Control flowchart of the first embodiment Mollier diagram showing intermediate pressure control Mollier diagram showing refrigeration cycle under floor heating cycle conditions Mollier diagram showing refrigeration cycle under cooling cycle conditions Mollier diagram showing refrigeration cycle under heating cycle conditions Graph showing changes in various cycle characteristics relative to changes in volume ratio under floor heating cycle conditions The graph which shows the change of the discharge refrigerant temperature of the high-pressure stage compressor to the change of the volume ratio under the floor heating cycle condition Graph showing changes in various cycle characteristics relative to changes in volume ratio under cooling cycle conditions The graph which shows the change of the discharge refrigerant temperature of the high-pressure stage compressor to the change of the volume ratio under the cooling cycle condition Graph showing changes in various cycle characteristics with respect to changes in volume ratio under heating cycle conditions The graph which shows the change of the discharge refrigerant temperature of the high-pressure stage compressor to the change of the volume ratio under the heating cycle condition The block diagram of the refrigerating-cycle apparatus concerning 2nd Embodiment of this invention. Control flowchart of the second embodiment The block diagram of the refrigerating-cycle apparatus concerning 3rd Embodiment of this invention. The block diagram of the refrigerating-cycle apparatus concerning 4th Embodiment of this invention. Partial enlarged view of FIG. 12A showing the detailed structure of the two-stage rotary expander Configuration diagram of conventional refrigeration cycle equipment Configuration diagram of another conventional refrigeration cycle apparatus

  Hereinafter, embodiments of the present invention will be described with reference to the accompanying drawings.

(First embodiment)
As shown in FIG. 1, the refrigeration cycle apparatus 100 includes a high-pressure stage compressor 101, a radiator 103, an expander 105, a gas-liquid separator 107, an expansion valve 109, an evaporator 111, and a low-pressure stage compressor 113. .

  The low-pressure stage compressor 113 pre-compresses the gas refrigerant evaporated in the evaporator 111. The high-pressure compressor 101 further compresses the refrigerant (working fluid) preliminarily compressed by the low-pressure compressor 113. The expander 105 performs power recovery by expanding the refrigerant cooled by the radiator 103. The expander 105 is configured so that the entire amount of the refrigerant cooled by the radiator 103 passes through. That is, no bypass circuit is provided for bypassing the expander 105 and flowing the refrigerant. Since the total amount of refrigerant contributes to power recovery, the effect of improving COP (coefficient of performance) based on power recovery is high. Note that, in a normal operation that exhibits refrigeration or heating capacity, the entire amount of refrigerant passes through the expander 105, but in a special operation such as defrosting, the refrigerant may not pass through the expander 105.

  The low-pressure stage compressor 113 and the expander 105 are both constituted by positive displacement fluid machines. The low-pressure stage compressor 113 and the expander 105 are connected by a shaft 116 so that the power recovered from the refrigerant in the expander 105 is transmitted to the low-pressure stage compressor 113, and are housed in a common hermetic container 117. Yes. The cylinder volume of the low-pressure stage compressor 113 and the cylinder volume of the expander 105 are constant. Specifically, in the present embodiment, the suction volume of the low-pressure compressor 113 and the suction volume of the expander 105 are constant. The suction volume of the low-pressure stage compressor 113 is larger than the suction volume of the expander 105. Conventionally, it is known that a constant density ratio can be avoided by using a variable volume fluid machine. However, the structure of the variable volume fluid machine is complicated and causes an increase in cost. Therefore, as in the present embodiment, it is preferable to use a fluid machine having a constant suction capacity for the compressor and the expander.

  The “cylinder volume” means the volume of the working chamber (expansion chamber or compression chamber) when the suction stroke is completed, and is often referred to as “confined volume”. “Suction volume” means the volume of refrigerant sucked into the compressor or expander in one cycle (suction → compression or expansion → discharge) of the compressor or expander. In the present embodiment, the second embodiment, and the third embodiment, the “cylinder volume” is equal to the “suction volume”. However, as will be described later in the fourth embodiment, the high-pressure refrigerant may be injected into the expansion chamber while the refrigerant is expanding. In this case, the volume (intake volume) of the refrigerant sucked into the expander in one cycle of the expander exceeds the cylinder volume.

  In this embodiment, a rotary compressor is used as the high-pressure stage compressor 101, but the type of the high-pressure stage compressor 101 is not limited at all, and other positive displacement compressors such as a scroll compressor or a turbo compressor. A centrifugal compressor such as may be used. The type of the low-pressure compressor 113 and the type of the expander 105 are not limited to the rotary type as long as they can be connected to each other by the shaft 116 so that power can be transmitted. For the low-pressure compressor 113 and the expander 105, other positive displacement fluid machines such as a scroll fluid machine may be used.

  The radiator 103 is a device for cooling the refrigerant compressed by the high-pressure compressor 101, and typically includes a water-refrigerant heat exchanger or an air-refrigerant heat exchanger. The gas-liquid separator 107 is a device for separating the refrigerant expanded by the expander 105 into a gas refrigerant and a liquid refrigerant. The gas-liquid separator 107 has a liquid refrigerant outlet at the bottom, a refrigerant inlet / outlet at the top, and a refrigerant inlet at the side. The evaporator 111 is a device for evaporating the liquid refrigerant separated by the gas-liquid separator 107, and typically includes an air-refrigerant heat exchanger. The expansion valve 109 is a valve having a variable opening, for example, an electric expansion valve, and is provided on a flow path between the liquid refrigerant outlet of the gas-liquid separator 107 and the inlet of the evaporator 111.

  The above devices are connected to each other by a refrigerant pipe so that a refrigerant circuit is formed. Specifically, the outlet of the high-pressure compressor 101 and the inlet of the radiator 103 are connected by the refrigerant pipe 102. The outlet of the radiator 103 and the inlet of the expander 105 are connected by a refrigerant pipe 104. The outlet of the expander 105 and the inlet of the gas-liquid separator 107 are connected by a refrigerant pipe 106. The liquid refrigerant outlet of the gas-liquid separator 107 and the inlet of the expansion valve 109 are connected by a refrigerant pipe 108. The outlet of the expansion valve 109 and the inlet of the evaporator 111 are connected by a refrigerant pipe 110. The outlet of the evaporator 111 and the inlet of the low-pressure compressor 113 are connected by a refrigerant pipe 112. The outlet of the low-pressure stage compressor 113 and the inlet of the high-pressure stage compressor 101 are connected by a refrigerant pipe 114.

  Further, the refrigerant inlet / outlet at the upper part of the gas-liquid separator 107 and the refrigerant pipe 114 are connected by the refrigerant pipe 115. Hereinafter, in this specification, the flow path formed by the refrigerant pipe 114 is referred to as an intermediate pressure flow path 114, and the flow path formed by the refrigerant pipe 115 is referred to as an outward recirculation path 115, respectively.

  The return flow path 115 is not provided with a valve for regulating the flow direction of the refrigerant or adjusting the flow rate of the refrigerant. Therefore, when the pressure in the gas-liquid separator 107 is higher than the discharge refrigerant pressure of the low-pressure stage compressor 113, the refrigerant flows from the gas-liquid separator 107 to the intermediate pressure flow path 114 through the forward recirculation path 115 (first Distribution state: injection). In other words, the refrigerant separated by the gas-liquid separator 107 is guided to the inlet of the high-pressure stage compressor 101 without passing through the evaporator 111 and the low-pressure stage compressor 113. On the other hand, when the pressure in the gas-liquid separator 107 is lower than the discharge refrigerant pressure of the low-pressure compressor 113, the refrigerant flows from the intermediate pressure channel 114 to the gas-liquid separator 107 through the forward recirculation channel 115. In other words, a part of the refrigerant preliminarily compressed by the low-pressure stage compressor 113 is returned to the gas-liquid separator 107 (second circulation state: reflux). In this way, the forward recirculation path 115 is configured to allow refrigerant to flow in both directions, the direction from the gas-liquid separator 107 toward the intermediate pressure flow path 114 and the direction from the intermediate pressure flow path 114 toward the gas-liquid separator 107. ing.

  The refrigeration cycle apparatus 100 further includes a first temperature sensor 121, a second temperature sensor 122, a third temperature sensor 123, a fourth temperature sensor 124, and a controller 118. The first temperature sensor 121 detects the intake refrigerant temperature of the expander 105. The second temperature sensor 122 detects the refrigerant temperature in the evaporator 111. The third temperature sensor 123 detects the suction refrigerant temperature of the low-pressure compressor 113. The fourth temperature sensor 124 detects the refrigerant temperature in the gas-liquid separator 107. A specific example of each temperature sensor is a thermocouple or a thermistor. Each temperature sensor is connected to the controller 118. A specific example of the controller 118 is a DSP (digital signal processor). The controller 118 controls the opening degree of the expansion valve 109 based on the signal acquired from each temperature sensor.

In the following description, the suction volume of the low-pressure compressor 113 is V _lc , the suction volume of the expander 105 is V _ex , the volume ratio of the low-pressure compressor 113 and the expander 105 is (V _lc / V _ex ), and the low pressure _lci the suction refrigerant density of the stage compressor 113 [rho, _exi the suction refrigerant density of the expander 105 [rho, thirst of the discharge refrigerant of the expander 105 is expressed as Q _exo. The pressure in the gas-liquid separator 107 is referred to as an intermediate pressure.

In the present embodiment, the volume ratio (V _lc / V _ex) is _{(1-Q exo) × (} ρ exi / ρ lci) or more and the density ratio (ρ _exi / ρ _lci) as to become less, the low-pressure compressor 113 and the expander 105 are designed.

First, if the following volume ratio (V _lc / V _ex) is the density ratio (ρ _exi / ρ _lci), the refrigerant mass flow rate of the low-pressure compressor 113 is less than the refrigerant mass flow rate of the expander 105, the forward The refrigerant flows from the gas-liquid separator 107 to the intermediate pressure channel 114 through the reflux path 115 (injection). In this case, since it is not necessary to compress the refrigerant that has passed through the forward recirculation path 115, the load on the low-pressure compressor 113 can be reduced. Further, since the flow rate of the refrigerant passing through the evaporator 111 is reduced, the pressure loss when passing through the evaporator 111 is reduced.

Conversely, when the volume ratio (V _lc / V _ex) exceeds the density ratio (ρ _exi / ρ _lci), the refrigerant mass flow rate of the low-pressure compressor 113 is larger than the refrigerant mass flow rate of the expander 105, the forward The refrigerant flows from the intermediate pressure channel 114 to the gas-liquid separator 107 through the reflux channel 115 (reflux). In this case, the refrigerant pre-compressed by the low-pressure compressor 113 is re-expanded by the expansion valve 109. Since the recovery power of the expander 105 is consumed for the circulation of the refrigerant, the COP improvement effect is diminished.

From these facts, it is desirable to design so as to satisfy the relationship represented by the following formula so that reflux does not occur. In other words, the following relationship is satisfied when driving to cause injection.
_{(V lc / V ex) ≦} (ρ exi / ρ lci)

The refrigerant flowing through the long volume ratio (V _lc / V _ex) is _{(1-Q exo) × (} ρ exi / ρ lci) above, from the gas-liquid separator 107 via shuttle passage 115 to the intermediate pressure passage 114 Is equal to or less than the thirst degree Q _{exo of the} refrigerant discharged from the expander 105. That is, only the gas refrigerant is injected into the high pressure compressor 101. The ratio R _i of the refrigerant flowing from the gas-liquid separator 107 to the intermediate pressure flow path 114 through the return flow path 115 is the refrigerant mass flow rate of the expander 105 (V _ex × ρ _exi ), and the refrigerant mass flow rate of the low pressure compressor 113. as (V _lc × ρ _lci), it can be represented by _{(V ex × ρ exi -V lc} × ρ lci) / (V ex × ρ exi). Volume ratio when the ratio R _i is smaller than the thirst of _{Q exo (V lc / V ex} ) is larger than _{(1-Q exo) × (} ρ exi / ρ lci). Therefore, the refrigerant flowing if the volume ratio (V _lc / V _ex) is _{(1-Q exo) × (} ρ exi / ρ lci) above, from the gas-liquid separator 107 via shuttle passage 115 to the intermediate pressure passage 114 The ratio R _i does not exceed the thirsty degree Q _{exo of the} refrigerant discharged from the expander 105, and the liquid refrigerant injection can be avoided. In other words, when driving to avoid potential liquid injection is satisfied relation _{(1-Q exo) × (} ρ exi / ρ lci) ≦ (V lc / V ex).

Thus, according to this embodiment, when operating as gas injection to _{occur, (1-Q exo) ×} (ρ exi / ρ lci) ≦ (V lc / V ex) ≦ (ρ exi / Ρ _lci ) is satisfied.

  As described below, when the refrigeration cycle apparatus 100 having the same design is used for two or more different uses, or when one refrigeration cycle apparatus 100 is used for two or more uses, the reflux is intentionally performed. Some scenes allow. When liquid refrigerant injection occurs, the actual COP may fall below the COP when power recovery is not performed, so liquid refrigerant injection should be avoided. On the other hand, even if reflux occurs, theoretically, the COP of the refrigeration cycle apparatus 100 does not fall below the COP when power recovery is not performed.

  Specific applications of the refrigeration cycle apparatus 100 include, for example, a heat pump water heater and an air conditioner. Some heat pump water heaters have a hot water supply function that can supply hot water to a faucet and / or a floor heating function that heats a room by circulating hot water through a pipe routed around the floor of a house. The air conditioner is configured to adjust the temperature in the room by exchanging heat between the indoor air and the refrigerant, and typically has a cooling function and a heating function.

The inventors calculated the volume ratio that can sufficiently reduce the annual power consumption when the refrigeration cycle apparatus 100 is applied to a heat pump water heater or an air conditioner. Specifically, under the floor heating conditions of the heat pump water heater, it is assumed that the outside air temperature is 7 ° C., the return temperature of the warm water for floor heating is 25 ° C., the intake refrigerant temperature of the low-pressure compressor 113 is 7 ° C., and the refrigerant is carbon dioxide. did. Thirst of Q _exo and density ratio at the intermediate pressure determined for any volume ratio (ρ _exi / ρ _lci) desirable volume ratio derived from the relationship (V _lc / V _ex) is from 4.7 to 7 .1.

In the cooling condition of the air conditioner, it was assumed that the outside air temperature was 35 ° C, the intake air temperature of the indoor unit (evaporator 111) was 27 ° C, the suction refrigerant temperature of the low-pressure compressor 113 was 27 ° C, and the refrigerant was carbon dioxide. Thirst of Q _exo and density ratio at the intermediate pressure determined for any volume ratio (ρ _exi / ρ _lci) desirable volume ratio derived from the relationship (V _lc / V _ex) is from 2.4 to 3 .6.

In the heating conditions of the air conditioner, it was assumed that the outside air temperature was 7 ° C., the intake air temperature of the indoor unit (radiator 103) was 20 ° C., the suction refrigerant temperature of the low-pressure compressor 113 was 7 ° C., and the refrigerant was carbon dioxide. Thirst of Q _exo and density ratio at the intermediate pressure determined for any volume ratio (ρ _exi / ρ _lci) desirable volume ratio derived from the relationship (V _lc / V _ex) is from 2.1 to 2 .9.

Here, when the volume ratio (V _lc / V _ex) is below _{(1-Q exo) × (} ρ exi / ρ lci), since the injection of the liquid refrigerant occurs, the enthalpy of refrigerant drawn into the high-pressure compressor 101 Decrease significantly. As a result, the discharge refrigerant temperature of the high-pressure compressor 101 decreases, and the heating capacity required for the floor heating function of the heat pump water heater and the heating function of the air conditioner becomes insufficient. Further, the liquid refrigerant that should normally pass through the evaporator 111 passes through the forward recirculation path 115, so that the cooling capacity required for the cooling function of the air conditioner also decreases. Therefore, when using the refrigeration cycle apparatus 100 to a plurality of applications, the volume ratio (V _lc / V _ex) for each condition _{(1-Q exo) × (} ρ exi / ρ lci) above and permit reflux It is better to set the value to prevent as much as possible. The value is 4.7 in the above example.

  When the refrigeration cycle apparatus 100 is designed exclusively for each of a heat pump water heater, a cooling air conditioner, and a heating air conditioner, a volume ratio suitable for each application can be set. However, when the refrigeration cycle apparatus 100 is used in the multi-function heat pump system as shown in FIG. In the example described above, by selecting a desirable volume ratio of 4.7 under floor heating conditions, it is possible to reduce the amount of reflux as much as possible while reliably avoiding liquid refrigerant injection. Of course, since the optimal volume ratio changes according to external conditions such as season and weather, even if the present invention is applied to a single use refrigeration cycle apparatus, the benefits can be fully enjoyed.

  The multi-function heat pump system shown in FIG. 2 includes a heat pump water heater 12 with a floor heating function and an air conditioner 14, and a refrigeration cycle apparatus 100 common to the water heater 12 and the air conditioner 14 is used. However, a dedicated radiator 103 (FIG. 1) is provided for each of the water heater 12 and the air conditioner 14.

  Next, the operation of the refrigeration cycle apparatus 100 will be described.

  First, when starting, the expansion valve 109 is fully closed. Next, power supply to the motor of the high-pressure compressor 101 is started, and the high-pressure compressor 101 is driven. The high-pressure compressor 101 sucks and compresses the refrigerant in the intermediate pressure channel 114. The compressed refrigerant is sent to the expander 105 through the refrigerant pipe 102, the radiator 103 and the refrigerant pipe 104. The refrigerant pipe 104 on the inlet side of the expander 105 is filled with the refrigerant discharged from the high-pressure compressor 101. For this reason, the pressure in the refrigerant pipe 104 increases. Further, since the high-pressure compressor 101 sucks the refrigerant from the gas-liquid separator 107 through the forward recirculation path 115, the liquid refrigerant in the gas-liquid separator 107 evaporates. Therefore, the temperature and pressure in the refrigerant pipe 106 on the outlet side of the expander 105 are reduced. That is, when the high pressure compressor 101 is operated with the expansion valve 109 closed, the pressure difference between the inlet and the outlet of the expander 105 increases. The expander 105 is driven by the pressure difference thus generated.

  On the other hand, the high-pressure stage compressor 101, the forward recirculation path 115, and the gas-liquid separator 107 are connected to each other via the refrigerant pipe 114 (intermediate pressure flow path), so that the refrigerant pipe 114 on the outlet side of the low-pressure stage compressor 113 is connected. The pressure inside falls. In addition, the refrigerant pipe 112 on the inlet side of the low-pressure stage compressor 113 is filled with a refrigerant having an evaporation pressure corresponding to the ambient temperature (heat source temperature) at the place where the evaporator 111 is installed (for example, outdoors). Therefore, at the time of startup, the pressure in the refrigerant pipe 112 temporarily exceeds the pressure in the intermediate pressure flow path 114. Then, the low-pressure compressor 113 behaves as an expander and is driven by a pressure difference between the refrigerant pipe 112 and the intermediate pressure flow path 114.

  As described with reference to FIG. 13, a refrigeration cycle apparatus having a configuration in which a low-pressure compressor and an expander are connected by a shaft is conventionally known. However, when a rotary expander is used as the expander 803 of the conventional refrigeration cycle apparatus shown in FIG. 13, the piston stops eccentrically on the vane side, that is, the piston stops with the suction port and the discharge port communicating with each other. There are things to do. In this case, there is a problem in that a sufficient initial pressure difference required for driving the low-pressure stage compressor 801b and the expander 803 cannot be obtained, and a start-up error occurs. On the other hand, according to the starting method in the present embodiment, the low-pressure stage compressor 113 temporarily behaves as an expander. Therefore, the piston of the expander 105 and the piston of the low-pressure stage compressor 113 are simultaneously eccentric to the vane side. Can be designed not to. That is, by making the eccentric directions of the pistons different from each other, it is possible to reliably generate a pressure difference necessary for driving between the suction side and the discharge side of at least one of the expander 105 and the low-pressure compressor 113. . Thereby, a refrigerating cycle device can be started reliably.

  When the low-pressure stage compressor 113 and the expander 105 start to operate, the low-pressure stage compressor 113 is driven by the recovered power of the expander 105 thereafter. The low-pressure stage compressor 113 sucks up the refrigerant from the refrigerant pipe 112, the evaporator 111 and the refrigerant pipe 110. Thereby, the liquid refrigerant begins to evaporate in the evaporator 111, and the temperature and pressure in the evaporator 111 are reduced. When the pressure in the refrigerant pipe 110 becomes lower than the pressure in the refrigerant pipe 108, the opening degree of the expansion valve 109 is gradually increased to the initial value. In the present embodiment, the expansion valve 109 is opened when the detected temperature of the second temperature sensor 122 becomes lower than the detected temperature of the fourth temperature sensor 124.

  Thereafter, the opening degree of the expansion valve 109 is controlled by the controller 118. Specifically, it is determined based on the refrigerant evaporation pressure in the evaporator 111, the suction refrigerant temperature of the expander 105, the discharge refrigerant pressure of the high-pressure compressor 101, and the suction refrigerant temperature of the low-pressure compressor 113. Under the set target cycle conditions, the pressure in the gas-liquid separator 107 is adjusted so that the theoretical recovery power of the expander 105 is equal to the theoretical compression work of the low-pressure compressor 113. This is a control for making the actual high pressure in the refrigerant circuit coincide with the optimum high pressure in the target cycle condition. The pressure (intermediate pressure) in the gas-liquid separator 107 can be adjusted by the expansion valve 109. The theoretical recovery power and the theoretical compression work are values derived by calculation, respectively, and do not mean actual recovery power and compression work.

  This will be described in more detail with reference to the flowchart of FIG.

First, in step 101, the refrigerant temperature T ₁ of the expander 105 from the first temperature sensor 121, the evaporation temperature T ₂ of the refrigerant in the evaporator 111 from the second temperature sensor 122, and the low pressure stage compressor from the third temperature sensor 123 are shown. It acquires 113 suction refrigerant temperature T ₃ of the. From the evaporation temperature T ₂ of the refrigerant in the evaporator 111, the evaporation pressure of the refrigerant in the evaporator 111 can be known.

  Next, in step 102, the optimum high pressure at which the COP of the refrigeration cycle apparatus 100 is maximized is calculated based on the temperature and pressure acquired in step 101.

Next, in Step 103 and Step 104, a target intermediate pressure at which the theoretical recovery power and the theoretical compression work are equal is calculated. First, at step 103, a certain target intermediate pressure is set. The recovery power (theoretical recovery power) when the refrigerant is expanded by the expander 105 to the set target intermediate pressure is calculated based on the calculated optimum high pressure and the intake refrigerant temperature T ₁ of the expander 105. As shown in FIG. 4, the state of the refrigerant at the inlet of the expander 105 is indicated by a point D. The point D is specified by the optimum high pressure P _H and the suction refrigerant temperature T ₁ . The target intermediate pressure P _M is the pressure at point E. In the expander 105, the refrigerant expands along the isentropic line (point D → point E). If the efficiency of the expander 105 is added to the enthalpy (h ₂ −h ₁ ) that the refrigerant has lost in the process of changing from the point D to the point E, the theoretical recovery power can be obtained.

Further, in step 103, the compression work (theoretical compression work) when the refrigerant is compressed by the low-pressure stage compressor 113 up to the set target intermediate pressure is expressed by the evaporation pressure P _L of the evaporator 111 and the suction refrigerant of the low-pressure stage compressor 113. calculated on the basis of the temperature T _3. As shown in FIG. 4, the state of the refrigerant at the inlet of the low-pressure compressor 113 is indicated by a point A. Point A is specified by the evaporation pressure P _L and the suction refrigerant temperature T ₃ . In the low-pressure compressor 113, the refrigerant is pre-compressed along the isentropic line (point A → point B). The enthalpy (h ₄ -h ₃ ) acquired by the refrigerant in the process of changing from point A to point B is divided by the efficiency of the low-pressure stage compressor 113, and further, the low-pressure stage compressor 113 with respect to the refrigerant mass flow rate of the expander 105. If the ratio of the refrigerant mass flow rate is integrated, theoretical compression work can be obtained.

The refrigerant mass flow rate of the expander 105 can be calculated from the refrigerant density at the inlet of the expander 105 and the suction volume of the expander 105. The refrigerant density at the inlet of the expander 105 is calculated from, for example, the optimum high pressure and the intake refrigerant temperature T ₁ . Similarly, the refrigerant mass flow rate of the low-pressure compressor 113 can be calculated from the refrigerant density at the inlet of the low-pressure compressor 113 and the suction volume of the low-pressure compressor 113. The refrigerant density at the inlet of the low-pressure compressor 113 is calculated from, for example, the evaporation temperature T ₂ and the suction refrigerant temperature T ₃ . The efficiency of the expander 105 and the low-pressure compressor 113 is a design value.

Next, in step 104, it is determined whether the theoretical recovery power and the theoretical compression work match. If they match, the process proceeds to step 105. If they do not match, the process returns to step 103, another target intermediate pressure is set, and the processes of step 103 and step 104 are repeated until they match. As described above, the controller 118 calculates an arbitrary optimum high pressure P _H and target intermediate pressure P _M based on the detection result of each temperature sensor.

Next, in step 105, the pressure in the gas-liquid separator 107 (actual intermediate pressure) is calculated. Specifically, first, the evaporation temperature T ₄ of the refrigerant in the gas-liquid separator 107 is acquired from the fourth temperature sensor 124. It can be calculated the pressure of the refrigerant from the evaporator temperature T ₄ of the refrigerant. That is, the controller 108 as a means for controlling the opening degree of the expansion valve 109 calculates the actual pressure in the gas-liquid separator 107 based on the detection result of the fourth temperature sensor 124.

Next, in step 106, the actual intermediate pressure is compared with the target intermediate pressure P _M. When the actual intermediate pressure is higher than the target intermediate pressure P _M , the process proceeds to step 107, and when it is lower, the process proceeds to step 107 ′. In step 107, the set opening degree of the expansion valve 109 is increased. In step 107 ′, the set opening degree of the expansion valve 109 is decreased.

  Next, in step 108, the set opening degree is output to the expansion valve 109, and the opening degree of the expansion valve 109 is changed. When the opening degree of the expansion valve 109 changes, the pressure in the gas-liquid separator 107 also changes. By periodically performing the process of this flowchart, the pressure in the gas-liquid separator 107 is adjusted and optimized so that the recovery power of the expander 105 and the compression work of the low-pressure compressor 113 are theoretically equal. High pressure is maintained.

As described above, the controller 118 sets the target intermediate pressure P _{M at} which the theoretical recovery power of the expander 105 at an arbitrary optimum high pressure in the refrigeration cycle is equal to the theoretical compression work of the low-pressure compressor 113 at any optimum high pressure. Means for calculating, and means for controlling the opening of the expansion valve 109 so that the actual pressure in the gas-liquid separator 107 approaches the calculated target intermediate pressure P _M are provided. Specifically, the controller 118 controls the opening degree of the expansion valve 109 based on the detection result of the fourth temperature sensor 124 so that the pressure in the gas-liquid separator 107 matches the target intermediate pressure P _M.

Further, when the volume ratio (V _lc / V _ex ) is excessive with respect to the cycle condition of the refrigeration cycle apparatus 100, the high-pressure stage compressor 101 and the expander with respect to the refrigerant mass flow rate of the low-pressure stage compressor 113. The refrigerant mass flow rate of 105 is insufficient. In other words, the suction amount of the high-pressure stage compressor 101 is too small with respect to the discharge amount from the low-pressure stage compressor 113. Therefore, the pressure in the intermediate pressure flow path 114 rises, and the refrigerant that is not sucked into the high pressure compressor 101 returns from the intermediate pressure flow path 114 to the gas-liquid separator 107 through the forward recirculation path 115.

On the other hand, when the volume ratio (V _lc / V _ex ) is too small with respect to the cycle condition of the refrigeration cycle apparatus 100, the high-pressure stage compressor 101 and the expander with respect to the refrigerant mass flow rate of the low-pressure stage compressor 113. The refrigerant mass flow rate of 105 becomes excessive. In other words, the suction amount of the high-pressure compressor 101 is excessive with respect to the discharge amount from the low-pressure compressor 113. Therefore, the pressure in the intermediate pressure flow path 114 decreases, and a shortage of refrigerant is injected from the gas-liquid separator 107 to the intermediate pressure flow path 114 through the forward recirculation path 115.

With reference to the Mollier diagrams shown in FIGS. 5A to 5C, the operation of the refrigeration cycle apparatus 100 in each application will be described. It is assumed that the volume ratio (V _lc / V _ex ) is set to 4.7.

As shown in FIG. 5A, the floor heating conditions described above, the volume ratio 4.7 corresponds to the value represented by _{(1-Q exo) × (} ρ exi / ρ lci). Therefore, all of the gas refrigerant in the gas-liquid separator 107 is injected into the intermediate pressure flow path 114 through the forward recirculation path 115, and the liquid refrigerant is sent to the evaporator 111 through the expansion valve 109.

As shown in Figure 5B, in the cooling condition described above, the volume ratio 4.7, exceeds the value expressed by (ρ _exi / ρ _lci). That is, the refrigerant mass flow rate of the low-pressure compressor 113 is excessive with respect to the refrigerant mass flow rate of the expander 105. Therefore, a part of the refrigerant compressed by the low-pressure compressor 113 is returned to the gas-liquid separator 107 through the forward recirculation path 115 and re-expanded by the expansion valve 109.

As shown in FIG. 5C, even in the heating conditions described above, similarly to the cooling conditions, the volume ratio 4.7, exceeds the value expressed by (ρ _exi / ρ _lci). Therefore, a part of the refrigerant compressed by the low-pressure compressor 113 is returned to the gas-liquid separator 107 through the forward recirculation path 115 and re-expanded by the expansion valve 109.

As described above, according to the present embodiment, the refrigerant can flow bidirectionally in the forward and backward flow path 115, so that the flow rate balance of the refrigeration cycle is established. Accordingly, the low-pressure compressor 113 and the expander 105 can be designed so that the annual power consumption is minimized without being restricted by the constant density ratio. Further, the intermediate pressure can be easily adjusted by controlling the opening degree of the expansion valve 109 so that the actual high pressure in the refrigerant circuit becomes the optimum high pressure regardless of the volume ratio (V _lc / V _ex ). The refrigeration cycle apparatus 100 can be operated.

In addition, by setting the volume ratio _{(V lc / V ex) (} 1-Q exo) × (ρ exi / ρ lci) represented the value above, it is possible to avoid the injection of the liquid refrigerant.

6A and 6B show changes in the volume ratio under floor heating cycle conditions (outside temperature 7 ° C., return temperature of warm water for floor heating 25 ° C., suction refrigerant temperature of low-pressure compressor 7 ° C., CO ₂ refrigerant). It is a graph which shows the change of the various cycle characteristics (calculated value) with respect to. The vertical axis in FIG. 6A shows the intermediate pressure, the COP, and the refrigerant flow rate in the forward recirculation path as cycle characteristics. The vertical axis of FIG. 6B indicates the refrigerant discharge temperature of the high-pressure compressor 101 as cycle characteristics.

In the floor heating cycle condition, as shown in the graph of FIG. 6A, the refrigerant flow rate in the forward recirculation path 115 is switched between the volume ratio (V _lc / V _ex ) = 7.1 as a boundary. When the refrigerant flow rate is positive, the refrigerant flows from the gas-liquid separator 107 to the intermediate pressure channel 114 (injection). When the refrigerant flow rate is negative, the refrigerant flows from the intermediate pressure channel 114 to the gas-liquid separator 107. (reflux). When gas injection occurs, both COP and intermediate pressure increase. This is because the compression load of the low-pressure stage compressor 113 is reduced by reducing the refrigerant mass flow rate of the low-pressure stage compressor 113, and as a result, the compression load of the high-pressure stage compressor 101 is reduced.

Further, under the above floor heating cycle condition, as shown in FIG. 6B, the discharge refrigerant temperature of the high-pressure compressor 101 rapidly decreases with the volume ratio (V _lc / V _ex ) = 4.7 as a boundary. That is, the injection of the liquid refrigerant (including the gas refrigerant) and the injection of the gas refrigerant are switched around the volume ratio (V _lc / V _ex ) = 4.7. As described above, when (1-Q _exo) less than × (ρ _exi / ρ _lci) represented value is the volume ratio (V _lc / V _ex) does not occur injection of the liquid refrigerant . Conversely, (1-Q _exo) when × value expressed by (ρ _exi / ρ _lci) is greater than the volume ratio (V _lc / V _ex) is injection of the liquid refrigerant occurs.

Further, the change in the refrigerant temperature discharged from the high-pressure compressor 101 stops at the boundary of the volume ratio (V _lc / V _ex ) = 7.1. That is, the injection and recirculation of the gas refrigerant are switched around the volume ratio (V _lc / V _ex ) = 7.1. As described above, when the density ratio (ρ _exi / ρ _lci) is greater than the volume ratio (V _lc / V _ex) is refluxed does not occur. Conversely, if the density ratio (ρ _exi / ρ _lci) is smaller than the volume ratio (V _lc / V _ex) is refluxed occurs.

In view of the above, assuming that operation is performed under floor heating cycle conditions, the volume ratio (V _lc / V _ex ) is in the range of 4.7 to 7.1 in order to avoid liquid refrigerant injection and reflux. Should be set.

As described with reference to FIG. 2, in the multi-function heat pump system, it should be considered that a single refrigeration cycle apparatus 100 covers hot water supply, floor heating, and air conditioning from the viewpoint of cost and the like. In such a case, if the volume ratio is set with only floor heating in mind, there is a possibility that efficient operation cannot be performed in other applications. That is, as shown in FIG. 7A and FIG. 7B, under the cooling cycle conditions, the liquid refrigerant injection and reflux can be avoided by setting the volume ratio (V _lc / V _ex ) in the range of 2.4 to 3.6. . Similarly, as shown in FIG. 8A and FIG. 8B, in the heating cycle condition, by setting the volume ratio (V _lc / V _ex ) in the range of 2.1 to 2.9, liquid refrigerant injection and reflux are avoided. it can. Thus, the range of suitable volume ratios differs depending on the cycle conditions. Although reflux is acceptable, liquid refrigerant injection should be avoided, so in this example, a volume ratio (V _lc / V _ex ) = 4.7 that can avoid liquid injection under floor heating cycle conditions is reasonable.

When setting the volume ratio (V _lc / V _ex) to 4.7, the floor heating condition, since the volume ratio (V _lc / V _ex) is equal to or less than the density ratio (ρ _exi / ρ _lci), the low-pressure stage compressor All the refrigerant compressed by the machine 113 is sucked into the high-pressure compressor 101, and the refrigeration cycle apparatus 100 can be operated efficiently. Not floor heating conditions but also in the cooling and the application of the heating, it is smaller than _{(1-Q exo) × (} ρ exi / ρ lci) represented value is the volume ratio (V _lc / V _ex), Liquid refrigerant injection can be avoided.

If the volume ratio (V _lc / V _ex ) = 4.7, 23.6% of the refrigerant compressed by the low-pressure compressor 113 passes to the gas-liquid separator 107 via the forward recirculation path 115 under the cooling condition. And the expansion valve 109 is re-expanded. When the calculation is performed under the same cooling condition, 49.6% of the refrigerant bypasses the expander in the conventional example (FIG. 14) in which the restriction of the constant density ratio is avoided in the bypass circuit. Similarly, under the heating condition, 36.6% of the refrigerant compressed by the low-pressure compressor 113 returns to the gas-liquid separator 107 via the forward recirculation path 115 and is re-expanded by the expansion valve 109. When the calculation is performed under the same heating condition, 58.2% of the refrigerant bypasses the expander in the conventional example (FIG. 14) in which the restriction of the constant density ratio is avoided in the bypass circuit. Thus, according to the refrigeration cycle apparatus 100 of the present embodiment, a more efficient operation can be performed than a conventional refrigeration cycle apparatus having a bypass circuit.

  It should be noted that the present invention does not completely prohibit liquid refrigerant injection.

(Second Embodiment)
FIG. 9 is a configuration diagram of a refrigeration cycle apparatus according to the second embodiment of the present invention. The refrigeration cycle apparatus 500 of the present embodiment has a configuration that is substantially the same as that of the refrigeration cycle apparatus 100 (see FIG. 1) according to the first embodiment. The difference between the present embodiment and the first embodiment is that the temperature sensor 520 is provided and the control performed by the controller 118. In the following, the same functional parts are denoted by the same reference numerals, and the description thereof is omitted.

  As shown in FIG. 9, the refrigeration cycle apparatus 500 includes a temperature sensor 520 for detecting the discharge refrigerant temperature of the high-pressure compressor 101. Similar to the first embodiment, a temperature sensor 122 for detecting the evaporation temperature of the refrigerant in the evaporator 111 is also provided. The controller 118 controls the opening degree of the expansion valve 109 based on the detection results of the temperature sensor 520 and the temperature sensor 122.

  The operation of the refrigeration cycle apparatus 500 will be described with reference to the flowchart of FIG. First, in step 501, the outside air temperature is estimated based on the evaporation temperature of the refrigerant in the evaporator 111. Next, in step 502, the target discharge refrigerant temperature of the high-pressure compressor 101 is calculated. The target discharge refrigerant temperature is determined according to, for example, the outside air temperature and the floor heating set temperature (or the heating set temperature). Next, in step 503, the actual discharge refrigerant temperature of the high-pressure compressor 101 is acquired from the temperature sensor 520. Next, in step 504, the actual discharge refrigerant temperature is compared with the target discharge refrigerant temperature. When the actual discharge refrigerant temperature is higher than the target discharge refrigerant temperature, the process proceeds to step 505. If the actual discharge refrigerant temperature is lower than the target discharge refrigerant temperature, the process proceeds to step 505 '.

  In step 505 and step 506, the set opening degree of the expansion valve 109 is increased to lower the intermediate pressure. When the intermediate pressure decreases, the compression work of the low-pressure stage compressor 113 decreases. Thereby, an imbalance occurs between the compression work of the low-pressure stage compressor 113 and the recovered power of the expander 105. In order to eliminate this imbalance, the rotation speed of the shaft 116 increases and the high pressure of the refrigeration cycle decreases. As a result, the recovery power of the expander 105 is reduced and the unbalance is eliminated. Moreover, the discharge refrigerant | coolant temperature of the high pressure compressor 101 falls because the high pressure of a refrigerating cycle falls.

  In steps 505 'and 506, the set opening degree of the expansion valve 109 is decreased to increase the intermediate pressure. When the intermediate pressure increases, the compression work of the low-pressure compressor 113 increases. Thereby, an imbalance occurs between the compression work of the low-pressure stage compressor 113 and the recovered power of the expander 105. In order to eliminate this imbalance, the rotational speed of the shaft 116 is decreased and the high pressure of the refrigeration cycle is increased. As a result, the recovery power of the expander 105 is increased and the unbalance is eliminated. Moreover, the discharge refrigerant | coolant temperature of the high pressure compressor 101 rises because the high pressure of a refrigerating cycle rises.

  According to this embodiment, the opening degree of the expansion valve 109 is controlled based on the detection results of the temperature sensors 520 and 123. By controlling the opening degree of the expansion valve 109, the intermediate pressure can be adjusted. The rotation speed of the shaft 116 changes according to the intermediate pressure. When the rotation speed of the shaft 116 changes, the high pressure of the refrigeration cycle also changes. That is, the expansion valve 109 can adjust the high pressure of the refrigeration cycle. Therefore, the refrigeration cycle apparatus 500 of the present embodiment is suitable for applications that require heating capability such as floor heating, heating, and hot water supply.

(Third embodiment)
FIG. 11 is a configuration diagram of a refrigeration cycle apparatus according to the third embodiment of the present invention. The refrigeration cycle apparatus 700 has substantially the same configuration as the refrigeration cycle apparatus described in the first embodiment and the second embodiment. The difference between the present embodiment and the first embodiment is that the high-pressure stage compressor 701, the low-pressure stage compressor 713, and the expander 705 are housed in a common sealed container 717.

  As shown in FIG. 11, in the refrigeration cycle apparatus 700, the high-pressure stage compressor 701, the low-pressure stage compressor 713, and the expander 705 are arranged in this order from above in a single sealed container 717. The low-pressure stage compressor 713 and the expander 705 are connected by a shaft 716 so that power can be transmitted. Oil is stored at the bottom of the sealed container 717. The space above the oil level is filled with the refrigerant discharged from the high-pressure compressor 701. The periphery of the low-pressure stage compressor 713 and the expander 705 is filled with oil.

  When the high-pressure compressor 701 is driven, the space above the oil level is filled with high-pressure discharged refrigerant. Around the high-pressure compressor 701, high-temperature oil that lubricates the high-pressure compressor 701 is held. On the other hand, the low-pressure stage compressor 713 and the expander 705 operate at a lower temperature than the high-pressure stage compressor 701. Therefore, oil having a lower temperature than the oil existing around the high-pressure stage compressor 713 is held around the low-pressure stage compressor 713 and the expander 705.

  That is, high temperature oil is held around the high pressure stage compressor 701, and low temperature oil is held around the low pressure stage compressor 713 and the expander 705. The oil forms a temperature stratification along the vertical direction. By forming the temperature stratification, it becomes difficult to mix the upper layer oil and the lower layer oil. Thereby, the heat transfer from the high pressure compressor 701 to the expander 705 via the oil can be suppressed. When heat transfer occurs, the discharge refrigerant temperature of the high-pressure compressor 701 decreases and the discharge refrigerant temperature of the expander 705 increases, which is not preferable from the viewpoint of the efficiency of the refrigeration cycle apparatus. According to this embodiment, since heat transfer can be effectively suppressed, the efficiency of the refrigeration cycle apparatus 700 can be increased.

(Fourth embodiment)
FIG. 12A is a configuration diagram of a refrigeration cycle apparatus according to the fourth embodiment of the present invention. In the refrigeration cycle apparatus 600 of the present embodiment, a multistage rotary expander 605 capable of changing the suction volume is used. Furthermore, the refrigeration cycle apparatus 600 includes an expander injection flow path 630 and an expander injection valve 631. As shown in FIG. 12B, the expander injection flow path 630 connects the suction flow path (pipe 104) of the expander 605 and the expander injection port 632 opened to the expansion chamber 611 of the expander 605. The expander injection valve 631 is provided on the expander injection flow path 630. The suction volume of the expander 605 can be changed by controlling the expander injection valve 631. Other configurations are the same as those described in the first and second embodiments.

  As shown in FIG. 12B, the expander 605 is a two-stage rotary expander having a first-stage cylinder 605a and a second-stage cylinder 605b. The expander injection port 632 is provided in the first stage cylinder 605a and opens toward the expansion chamber 611 of the first stage cylinder 605a. The first-stage cylinder 605a and the second-stage cylinder 605b are separated by an intermediate plate 605c, but the expansion chamber 611 of the first-stage cylinder 605a is in two stages through a communication hole 605d formed in the intermediate plate 605c. It communicates with the expansion chamber 612 of the eye cylinder 605b. Thereby, the expansion chambers 611 and 612 form a single expansion chamber. The expander injection flow path 630 branches from the pipe 104 and is connected to the first-stage cylinder 605a so that the refrigerant can be injected from the expander injection port 632 into the expansion chamber 611. The expander injection port 632 is provided in the vicinity of the communication hole 605d in the circumferential direction of the shaft 116.

When the expander injection valve 631 is closed, the refrigerant does not flow into the expansion chamber 611 from the expander injection port 632. Therefore, the cylinder volume of the first-stage cylinder 605a behaves as the suction volume V _ex. On the other hand, when the expander injection valve 631 is opened, the refrigerant flows into the expansion chamber 611 from the expander injection port 632. Therefore, the cylinder volume of the second-stage cylinder 605b behaves as the suction volume V _ex ′. For example, cylinder volume of the first-stage cylinder 605a may double the cylinder volume of the second-stage cylinder 605b, can change the suction volume V _ex optionally twice the suction volume V _ex '.

According to the first embodiment, the desirable volume ratio (V _lc / V _ex ) under the cooling condition was 2.4 to 3.6. In the present embodiment, when the expander injection valve 631 is opened and the suction volume V _ex ′ is used, the desirable volume ratio (V _lc / V _ex ′) under the same cooling condition is 2.4 to 3.6. If expressed using the suction volume V _ex , the desirable volume ratio (V _lc / V _ex ) under the same cooling condition is 4.8 to 7.2. Also, according to the first embodiment, the desired volume ratio in the heating condition (V _lc / V _ex) was 2.1 to 2.9. In the present embodiment, when the expander injection valve 631 is opened and the suction volume V _ex ′ is used, the desirable volume ratio (V _lc / V _ex ′) under the same heating condition is 2.1 to 2.9. If expressed using the suction volume V _ex , the desirable volume ratio (V _lc / V _ex ) under the same heating condition is 4.2 to 5.8. Also, according to the first embodiment, the desired volume ratio in the floor heating condition (V _lc / V _ex) was 4.7 to 7.1. Also in the present embodiment, when the expander injection valve 631 is closed and the suction volume V _ex is used, the desirable volume ratio (V _lc / V _ex ) under the same floor heating condition is 4.7 to 7.1.

That is, in the cooling condition and the heating condition, the expander injection valve 631 is opened to increase the suction volume, while in the floor heating condition, the expander injection valve 631 is closed. By performing such control, the desired volume ratio (V _lc / V _ex ) under each condition can be brought closer to each other.

Specifically, the low-pressure compressor 113 and the expander 605 can be designed so that the volume ratio (V _lc / V _ex ) is 4.8. The design of volume ratio (V _lc / V _ex ) = 4.8 is below the upper limit of the volume ratio (V _lc / V _ex ) in each condition of cooling, heating and floor heating. Therefore, reflux does not occur in each of the cooling, heating, and floor heating conditions, and the compression work in the low-pressure compressor 113 can be used effectively. Furthermore, the design of the volume ratio (V _lc / V _ex ) = 4.8 exceeds the lower limit of the volume ratio (V _lc / V _ex ) for each condition of cooling, heating and floor heating. Therefore, liquid refrigerant injection can also be prevented.

  Accordingly, the gas refrigerant flows through the forward recirculation passage 115 from the gas-liquid separator 107 toward the intermediate pressure passage 114. The refrigeration cycle apparatus 600 can be always operated in a gas injection state. If the operation can be performed in the gas injection state, it is not necessary to compress the refrigerant that has flowed through the forward recirculation path 115 by the low-pressure stage compressor 113, so that the load on the low-pressure stage compressor 113 can be reduced. In addition, since the entire amount of the refrigerant cooled by the radiator 103 passes through the expander 605, high power recovery efficiency can be realized. Thus, according to the refrigeration cycle apparatus 600 of the present embodiment, a more efficient operation can be performed than a conventional refrigeration cycle apparatus having a bypass circuit.

  As the expander 605, other types of expanders such as a scroll expander and a reciprocal expander can be used. The expander injection valve 631 may be a valve whose opening degree can be changed in multiple stages, or may be a simple on-off valve.

  INDUSTRIAL APPLICATION This invention is useful about the refrigerating-cycle apparatus utilized for an air conditioner, a refrigerator-freezer, a heat pump water heater, a heat pump heater, a vending machine, a car air conditioner, etc. In particular, a high effect can be obtained when one refrigeration cycle apparatus is shared for two or more uses. Of course, the present invention can also be suitably applied to a refrigeration cycle apparatus for a single application.

  The present invention relates to a refrigeration cycle apparatus.

  As a refrigeration cycle apparatus used for an air conditioner or a hot water heater, as shown in FIG. 13, one having a first compressor 801a, a radiator 802, an expander 803, a heat absorber 804, and a second compressor 801b is known. (Patent Document 1). The second compressor 801b and the expander 803 are connected by a rotating shaft 806, and the driving force of the second compressor 801b is covered by the power accompanying the expansion of the refrigerant in the expander 803. Therefore, it is possible to reduce the power of the first compressor 801a that is consumed to increase the refrigerant pressure to a predetermined pressure.

  According to the refrigeration cycle apparatus shown in FIG. 13, the rotational speed of the expander 803 and the rotational speed of the second compressor 801b coincide. Further, since the refrigerant expanded by the expander 803 is compressed by the second compressor 801b through the heat absorber 804, the mass flow rate of the refrigerant passing through the expander 803 and the mass flow rate of the refrigerant passing through the second compressor 801b. Matches. Furthermore, since the suction volume of the expander 803 and the suction volume of the second compressor 801b are determined at each design time point, restrictions are imposed on the suction refrigerant density of the expander 803 and the suction refrigerant density of the second compressor 801b.

That is, the product of the suction refrigerant density [rho _exi the suction volume V _exi and the expander 803 of the expander 803, the product of the suction refrigerant density [rho _c2i of the suction volume V _C2i of the second compressor 801b and the second compressor 801b Is always equal to Between the suction refrigerant density of the expander 803 and the suction refrigerant density of the second compressor _{801b, (ρ exi / ρ c2i} ) = relationship always holds the (V _C2i / V _exi). This relationship is called a constraint with a constant density ratio. To operate the refrigeration cycle apparatus at optimum efficiency, it is essential to be able to freely adjust the density [rho _exi and the density [rho _c2i depending on external conditions such as seasonal and weather. However, it can not be freely adjusted certain the density [rho _exi and the density [rho _c2i is constraint of constant density ratio, realization of efficient operation becomes difficult.

  In order to solve these problems, a refrigeration cycle apparatus shown in FIG. 14 has been proposed (Patent Document 2). The refrigeration cycle apparatus shown in FIG. 14 includes a two-stage compressor 903, a radiator 904, an expander 905, a gas-liquid separator 906, an evaporator 908, a gas injection circuit 910, and a bypass circuit 911. The two-stage compressor 903 includes a low-pressure stage compressor 901 and a high-pressure stage compressor 902. The low pressure compressor 901 and the expander 905 are connected by a rotating shaft. The bypass circuit 911 is provided with a flow rate control valve 913. By appropriately controlling the opening degree of the flow control valve 913 and allowing a part of the refrigerant to flow through the bypass circuit 911, it is possible to avoid the restriction of a constant density ratio.

JP 2003-307358 A JP 2006-71257 A

However, when a part of the refrigerant flows through the bypass circuit 911, the amount of refrigerant contributing to power recovery by the expander 905 decreases, and there is a problem that power recovery efficiency deteriorates. This problem becomes remarkable when, for example, the refrigeration cycle apparatus having the same design is applied to each of the heat pump hot water floor heater and the air conditioner. Respect CO ₂ refrigeration cycle device of one design, the density ratio can realize optimum efficiency where (ρ _exi / ρ _c2i) The present inventors have calculated, at rated conditions of floor heating 7.13, 3 in rated conditions of cooling It was 2.98 in the rated condition of heating. If the low-pressure compressor 901 and the expander 905 are designed in accordance with floor heating, 49.6% refrigerant flows through the bypass circuit 911 during cooling, and 58.2% refrigerant flows through the bypass circuit 911 during heating. In other words, the power that can be recovered is reduced to about half that of floor heating.

  An object of the present invention is to provide a refrigeration cycle apparatus capable of efficient power recovery while avoiding the restriction of a constant density ratio.

That is, the present invention
A positive displacement low-pressure stage compressor for pre-compressing the refrigerant;
A high-pressure stage compressor for further compressing the refrigerant pre-compressed by the low-pressure stage compressor;
An intermediate pressure flow path connecting the low pressure stage compressor and the high pressure stage compressor in series so that the refrigerant pre-compressed by the low pressure stage compressor is sent to the high pressure stage compressor;
A radiator for cooling the refrigerant compressed by the high-pressure stage compressor;
It is coaxially connected to the low-pressure stage compressor so that power transmission is performed, and is configured so that the entire amount of the refrigerant cooled by the radiator passes, and for recovering power by expanding the refrigerant A positive displacement expander;
A gas-liquid separator for separating the refrigerant expanded by the expander into a gas refrigerant and a liquid refrigerant;
An evaporator for evaporating the liquid refrigerant separated by the gas-liquid separator;
An expansion valve with variable opening provided on a flow path between the liquid refrigerant outlet of the gas-liquid separator and the inlet of the evaporator;
The refrigerant stored in the gas-liquid separator is preliminarily compressed by the low-pressure stage compressor in a first flow state where the refrigerant is led to the inlet of the high-pressure stage compressor without passing through the evaporator and the low-pressure stage compressor. A recirculation path connecting the intermediate pressure flow path and the gas-liquid separator so as to be able to switch between a second flow state in which a part of the refrigerant circulates back to the gas-liquid separator;
A controller for adjusting the flow rate of the refrigerant in the forward / return path in each of the first circulation state and the second circulation state by controlling the opening of the expansion valve;
A refrigeration cycle apparatus is provided.

  According to the present invention, when the suction volume of the low-pressure compressor is insufficient compared to the suction volume of the expander, the gas refrigerant is sent from the gas-liquid separator to the intermediate pressure channel through the forward recirculation path. And sucked into the high-pressure stage compressor. Thereby, the flow volume balance of a refrigerating cycle is materialized. On the other hand, when the suction volume of the expander is insufficient compared to the suction volume of the low-pressure stage compressor, part of the gas refrigerant pre-compressed by the low-pressure stage compressor passes through the intermediate pressure flow path and the forward recirculation path. It is sent to the gas-liquid separator. Thereby, the flow volume balance of a refrigerating cycle is materialized. Regardless of the value of the suction volume of the low-pressure compressor and the suction volume of the expander (design value), the flow rate balance of the refrigeration cycle is established by the action of the forward recirculation path.

  On the other hand, the pressure in the gas-liquid separator can be freely adjusted by the expansion valve. The refrigerant pressure on the radiator side can be arbitrarily adjusted by adjusting the pressure in the gas-liquid separator. For example, when the expansion valve is fully opened under an arbitrary operating condition, the pressure in the gas-liquid separator approaches the evaporating pressure of the refrigerant in the evaporator as much as possible. Then, since the gas-liquid separator and the intermediate pressure flow path are connected by the forward recirculation path, the pressure difference between the inlet and the outlet of the low-pressure stage compressor approaches zero as much as possible. That is, the amount of compression work of the low-pressure stage compressor is reduced. On the other hand, since the pressure difference between the inlet and the outlet of the expander increases, the power recovery amount of the expander increases. Based on this relationship (power recovery amount)> (compression work amount), the rotation speeds of the expander and the low-pressure compressor are increased. As a result, since the discharge refrigerant flow rate of the expander becomes excessive with respect to the discharge refrigerant flow rate of the high-pressure compressor, the refrigerant pressure on the radiator side decreases. As a result, the power recovery amount of the expander is reduced, and is balanced with the compression work amount of the low-pressure compressor, so that the refrigeration cycle is stabilized. That is, when the expansion valve is opened, the refrigerant pressure on the radiator side can be reduced.

  Conversely, as the expansion valve is gradually throttled, the pressure in the gas-liquid separator gradually increases. Since the gas-liquid separator and the intermediate pressure flow path are connected by the forward recirculation path, the pressure difference between the inlet and the outlet of the low-pressure compressor is increased. That is, the compression work of the low-pressure stage compressor increases. On the other hand, since the pressure difference between the inlet and outlet of the expander decreases, the power recovery amount of the expander decreases. Based on this relationship (power recovery amount) <(compression work amount), the rotation speeds of the expander and the low-pressure compressor are reduced. As a result, since the discharge refrigerant flow rate of the expander is insufficient with respect to the discharge refrigerant flow rate of the high-pressure compressor, the refrigerant pressure on the radiator side increases. As a result, the power recovery amount of the expander is increased to balance the amount of compression work of the low-pressure compressor, and the refrigeration cycle is stabilized. That is, if the expansion valve is throttled, the refrigerant pressure on the radiator side can be increased.

  Thus, if the opening degree of the expansion valve is appropriately controlled and thereby the rotational speeds of the expander and the low-pressure compressor are controlled, the refrigerant pressure on the radiator side can always be optimally adjusted. In addition, since the entire amount of refrigerant passes through the expander, efficient power recovery can be performed. Even if the refrigerant recirculates in the forward recirculation path (second circulation state) and a part of the recovered power is consumed for the circulation of the refrigerant, according to the present invention, the conventional example of flowing the refrigerant through the bypass circuit (FIG. 14)) in terms of energy balance. Therefore, it is possible to operate a refrigeration cycle apparatus including an expander having a volume ratio suitable for an application and a low-pressure compressor at a pressure / temperature condition desirable from the viewpoint of energy efficiency.

  The theory described above can be established regardless of the value of the volume ratio between the low-pressure stage compressor and the expander. Therefore, according to the present invention, the low-pressure stage compressor and the expander can be designed to have an arbitrary volume ratio that can theoretically minimize the annual power consumption. That is, according to the present invention, the degree of freedom in designing the refrigeration cycle apparatus is also increased.

The block diagram of the refrigerating-cycle apparatus concerning 1st Embodiment of this invention. Configuration diagram of multi-function heat pump system Control flowchart of the first embodiment Mollier diagram showing intermediate pressure control Mollier diagram showing refrigeration cycle under floor heating cycle conditions Mollier diagram showing refrigeration cycle under cooling cycle conditions Mollier diagram showing refrigeration cycle under heating cycle conditions Graph showing changes in various cycle characteristics relative to changes in volume ratio under floor heating cycle conditions The graph which shows the change of the discharge refrigerant temperature of the high-pressure stage compressor to the change of the volume ratio under the floor heating cycle condition Graph showing changes in various cycle characteristics relative to changes in volume ratio under cooling cycle conditions The graph which shows the change of the discharge refrigerant temperature of the high-pressure stage compressor to the change of the volume ratio under the cooling cycle condition Graph showing changes in various cycle characteristics with respect to changes in volume ratio under heating cycle conditions The graph which shows the change of the discharge refrigerant temperature of the high-pressure stage compressor to the change of the volume ratio under the heating cycle condition The block diagram of the refrigerating-cycle apparatus concerning 2nd Embodiment of this invention. Control flowchart of the second embodiment The block diagram of the refrigerating-cycle apparatus concerning 3rd Embodiment of this invention. The block diagram of the refrigerating-cycle apparatus concerning 4th Embodiment of this invention. Partial enlarged view of FIG. 12A showing the detailed structure of the two-stage rotary expander Configuration diagram of conventional refrigeration cycle equipment Configuration diagram of another conventional refrigeration cycle apparatus

  Hereinafter, embodiments of the present invention will be described with reference to the accompanying drawings.

(First embodiment)
As shown in FIG. 1, the refrigeration cycle apparatus 100 includes a high-pressure stage compressor 101, a radiator 103, an expander 105, a gas-liquid separator 107, an expansion valve 109, an evaporator 111, and a low-pressure stage compressor 113. .

  The low-pressure stage compressor 113 pre-compresses the gas refrigerant evaporated in the evaporator 111. The high-pressure compressor 101 further compresses the refrigerant (working fluid) preliminarily compressed by the low-pressure compressor 113. The expander 105 performs power recovery by expanding the refrigerant cooled by the radiator 103. The expander 105 is configured so that the entire amount of the refrigerant cooled by the radiator 103 passes through. That is, no bypass circuit is provided for bypassing the expander 105 and flowing the refrigerant. Since the total amount of refrigerant contributes to power recovery, the effect of improving COP (coefficient of performance) based on power recovery is high. Note that, in a normal operation that exhibits refrigeration or heating capacity, the entire amount of refrigerant passes through the expander 105, but in a special operation such as defrosting, the refrigerant may not pass through the expander 105.

  The low-pressure stage compressor 113 and the expander 105 are both constituted by positive displacement fluid machines. The low-pressure stage compressor 113 and the expander 105 are connected by a shaft 116 so that the power recovered from the refrigerant in the expander 105 is transmitted to the low-pressure stage compressor 113, and are housed in a common hermetic container 117. Yes. The cylinder volume of the low-pressure stage compressor 113 and the cylinder volume of the expander 105 are constant. Specifically, in the present embodiment, the suction volume of the low-pressure compressor 113 and the suction volume of the expander 105 are constant. The suction volume of the low-pressure stage compressor 113 is larger than the suction volume of the expander 105. Conventionally, it is known that a constant density ratio can be avoided by using a variable volume fluid machine. However, the structure of the variable volume fluid machine is complicated and causes an increase in cost. Therefore, as in the present embodiment, it is preferable to use a fluid machine having a constant suction capacity for the compressor and the expander.

  The “cylinder volume” means the volume of the working chamber (expansion chamber or compression chamber) when the suction stroke is completed, and is often referred to as “confined volume”. “Suction volume” means the volume of refrigerant sucked into the compressor or expander in one cycle (suction → compression or expansion → discharge) of the compressor or expander. In the present embodiment, the second embodiment, and the third embodiment, the “cylinder volume” is equal to the “suction volume”. However, as will be described later in the fourth embodiment, the high-pressure refrigerant may be injected into the expansion chamber while the refrigerant is expanding. In this case, the volume (intake volume) of the refrigerant sucked into the expander in one cycle of the expander exceeds the cylinder volume.

  In this embodiment, a rotary compressor is used as the high-pressure stage compressor 101, but the type of the high-pressure stage compressor 101 is not limited at all, and other positive displacement compressors such as a scroll compressor or a turbo compressor. A centrifugal compressor such as may be used. The type of the low-pressure compressor 113 and the type of the expander 105 are not limited to the rotary type as long as they can be connected to each other by the shaft 116 so that power can be transmitted. For the low-pressure compressor 113 and the expander 105, other positive displacement fluid machines such as a scroll fluid machine may be used.

  The radiator 103 is a device for cooling the refrigerant compressed by the high-pressure compressor 101, and typically includes a water-refrigerant heat exchanger or an air-refrigerant heat exchanger. The gas-liquid separator 107 is a device for separating the refrigerant expanded by the expander 105 into a gas refrigerant and a liquid refrigerant. The gas-liquid separator 107 has a liquid refrigerant outlet at the bottom, a refrigerant inlet / outlet at the top, and a refrigerant inlet at the side. The evaporator 111 is a device for evaporating the liquid refrigerant separated by the gas-liquid separator 107, and typically includes an air-refrigerant heat exchanger. The expansion valve 109 is a valve having a variable opening, for example, an electric expansion valve, and is provided on a flow path between the liquid refrigerant outlet of the gas-liquid separator 107 and the inlet of the evaporator 111.

  The above devices are connected to each other by a refrigerant pipe so that a refrigerant circuit is formed. Specifically, the outlet of the high-pressure compressor 101 and the inlet of the radiator 103 are connected by the refrigerant pipe 102. The outlet of the radiator 103 and the inlet of the expander 105 are connected by a refrigerant pipe 104. The outlet of the expander 105 and the inlet of the gas-liquid separator 107 are connected by a refrigerant pipe 106. The liquid refrigerant outlet of the gas-liquid separator 107 and the inlet of the expansion valve 109 are connected by a refrigerant pipe 108. The outlet of the expansion valve 109 and the inlet of the evaporator 111 are connected by a refrigerant pipe 110. The outlet of the evaporator 111 and the inlet of the low-pressure compressor 113 are connected by a refrigerant pipe 112. The outlet of the low-pressure stage compressor 113 and the inlet of the high-pressure stage compressor 101 are connected by a refrigerant pipe 114.

  Further, the refrigerant inlet / outlet at the upper part of the gas-liquid separator 107 and the refrigerant pipe 114 are connected by the refrigerant pipe 115. Hereinafter, in this specification, the flow path formed by the refrigerant pipe 114 is referred to as an intermediate pressure flow path 114, and the flow path formed by the refrigerant pipe 115 is referred to as an outward recirculation path 115, respectively.

  The return flow path 115 is not provided with a valve for regulating the flow direction of the refrigerant or adjusting the flow rate of the refrigerant. Therefore, when the pressure in the gas-liquid separator 107 is higher than the discharge refrigerant pressure of the low-pressure stage compressor 113, the refrigerant flows from the gas-liquid separator 107 to the intermediate pressure flow path 114 through the forward recirculation path 115 (first Distribution state: injection). In other words, the refrigerant separated by the gas-liquid separator 107 is guided to the inlet of the high-pressure stage compressor 101 without passing through the evaporator 111 and the low-pressure stage compressor 113. On the other hand, when the pressure in the gas-liquid separator 107 is lower than the discharge refrigerant pressure of the low-pressure compressor 113, the refrigerant flows from the intermediate pressure channel 114 to the gas-liquid separator 107 through the forward recirculation channel 115. In other words, a part of the refrigerant preliminarily compressed by the low-pressure stage compressor 113 is returned to the gas-liquid separator 107 (second circulation state: reflux). In this way, the forward recirculation path 115 is configured to allow refrigerant to flow in both directions, the direction from the gas-liquid separator 107 toward the intermediate pressure flow path 114 and the direction from the intermediate pressure flow path 114 toward the gas-liquid separator 107. ing.

  The refrigeration cycle apparatus 100 further includes a first temperature sensor 121, a second temperature sensor 122, a third temperature sensor 123, a fourth temperature sensor 124, and a controller 118. The first temperature sensor 121 detects the intake refrigerant temperature of the expander 105. The second temperature sensor 122 detects the refrigerant temperature in the evaporator 111. The third temperature sensor 123 detects the suction refrigerant temperature of the low-pressure compressor 113. The fourth temperature sensor 124 detects the refrigerant temperature in the gas-liquid separator 107. A specific example of each temperature sensor is a thermocouple or a thermistor. Each temperature sensor is connected to the controller 118. A specific example of the controller 118 is a DSP (digital signal processor). The controller 118 controls the opening degree of the expansion valve 109 based on the signal acquired from each temperature sensor.

In the following description, the suction volume of the low-pressure compressor 113 is V _lc , the suction volume of the expander 105 is V _ex , the volume ratio of the low-pressure compressor 113 and the expander 105 is (V _lc / V _ex ), and the low pressure _lci the suction refrigerant density of the stage compressor 113 [rho, _exi the suction refrigerant density of the expander 105 [rho, thirst of the discharge refrigerant of the expander 105 is expressed as Q _exo. The pressure in the gas-liquid separator 107 is referred to as an intermediate pressure.

In the present embodiment, the volume ratio (V _lc / V _ex) is _{(1-Q exo) × (} ρ exi / ρ lci) or more and the density ratio (ρ _exi / ρ _lci) as to become less, the low-pressure compressor 113 and the expander 105 are designed.

First, if the following volume ratio (V _lc / V _ex) is the density ratio (ρ _exi / ρ _lci), the refrigerant mass flow rate of the low-pressure compressor 113 is less than the refrigerant mass flow rate of the expander 105, the forward The refrigerant flows from the gas-liquid separator 107 to the intermediate pressure channel 114 through the reflux path 115 (injection). In this case, since it is not necessary to compress the refrigerant that has passed through the forward recirculation path 115, the load on the low-pressure compressor 113 can be reduced. Further, since the flow rate of the refrigerant passing through the evaporator 111 is reduced, the pressure loss when passing through the evaporator 111 is reduced.

Conversely, when the volume ratio (V _lc / V _ex) exceeds the density ratio (ρ _exi / ρ _lci), the refrigerant mass flow rate of the low-pressure compressor 113 is larger than the refrigerant mass flow rate of the expander 105, the forward The refrigerant flows from the intermediate pressure channel 114 to the gas-liquid separator 107 through the reflux channel 115 (reflux). In this case, the refrigerant pre-compressed by the low-pressure compressor 113 is re-expanded by the expansion valve 109. Since the recovery power of the expander 105 is consumed for the circulation of the refrigerant, the COP improvement effect is diminished.

From these facts, it is desirable to design so as to satisfy the relationship represented by the following formula so that reflux does not occur. In other words, the following relationship is satisfied when driving to cause injection.
_{(V lc / V ex) ≦} (ρ exi / ρ lci)

The refrigerant flowing through the long volume ratio (V _lc / V _ex) is _{(1-Q exo) × (} ρ exi / ρ lci) above, from the gas-liquid separator 107 via shuttle passage 115 to the intermediate pressure passage 114 Is equal to or less than the thirst degree Q _{exo of the} refrigerant discharged from the expander 105. That is, only the gas refrigerant is injected into the high pressure compressor 101. The ratio R _i of the refrigerant flowing from the gas-liquid separator 107 to the intermediate pressure flow path 114 through the return flow path 115 is the refrigerant mass flow rate of the expander 105 (V _ex × ρ _exi ), and the refrigerant mass flow rate of the low pressure compressor 113. as (V _lc × ρ _lci), it can be represented by _{(V ex × ρ exi -V lc} × ρ lci) / (V ex × ρ exi). Volume ratio when the ratio R _i is smaller than the thirst of _{Q exo (V lc / V ex} ) is larger than _{(1-Q exo) × (} ρ exi / ρ lci). Therefore, the refrigerant flowing if the volume ratio (V _lc / V _ex) is _{(1-Q exo) × (} ρ exi / ρ lci) above, from the gas-liquid separator 107 via shuttle passage 115 to the intermediate pressure passage 114 The ratio R _i does not exceed the thirsty degree Q _{exo of the} refrigerant discharged from the expander 105, and the liquid refrigerant injection can be avoided. In other words, when driving to avoid potential liquid injection is satisfied relation _{(1-Q exo) × (} ρ exi / ρ lci) ≦ (V lc / V ex).

Thus, according to this embodiment, when operating as gas injection _{occurs, (1-Q exo) ×} (ρ exi / ρ lci) ≦ (V lc / V ex) ≦ (ρ exi / Ρ _lci ) is satisfied.

  As described below, when the refrigeration cycle apparatus 100 having the same design is used for two or more different uses, or when one refrigeration cycle apparatus 100 is used for two or more uses, the reflux is intentionally performed. Some scenes allow. When liquid refrigerant injection occurs, the actual COP may fall below the COP when power recovery is not performed, so liquid refrigerant injection should be avoided. On the other hand, even if reflux occurs, theoretically, the COP of the refrigeration cycle apparatus 100 does not fall below the COP when power recovery is not performed.

  Specific applications of the refrigeration cycle apparatus 100 include, for example, a heat pump water heater and an air conditioner. Some heat pump water heaters have a hot water supply function that can supply hot water to a faucet and / or a floor heating function that heats a room by circulating hot water through a pipe routed around the floor of a house. The air conditioner is configured to adjust the temperature in the room by exchanging heat between the indoor air and the refrigerant, and typically has a cooling function and a heating function.

The inventors calculated the volume ratio that can sufficiently reduce the annual power consumption when the refrigeration cycle apparatus 100 is applied to a heat pump water heater or an air conditioner. Specifically, under the floor heating conditions of the heat pump water heater, it is assumed that the outside air temperature is 7 ° C., the return temperature of the warm water for floor heating is 25 ° C., the intake refrigerant temperature of the low-pressure compressor 113 is 7 ° C., and the refrigerant is carbon dioxide. did. Thirst of Q _exo and density ratio at the intermediate pressure determined for any volume ratio (ρ _exi / ρ _lci) desirable volume ratio derived from the relationship (V _lc / V _ex) is from 4.7 to 7 .1.

In the cooling condition of the air conditioner, it was assumed that the outside air temperature was 35 ° C, the intake air temperature of the indoor unit (evaporator 111) was 27 ° C, the suction refrigerant temperature of the low-pressure compressor 113 was 27 ° C, and the refrigerant was carbon dioxide. Thirst of Q _exo and density ratio at the intermediate pressure determined for any volume ratio (ρ _exi / ρ _lci) desirable volume ratio derived from the relationship (V _lc / V _ex) is from 2.4 to 3 .6.

In the heating conditions of the air conditioner, it was assumed that the outside air temperature was 7 ° C., the intake air temperature of the indoor unit (radiator 103) was 20 ° C., the suction refrigerant temperature of the low-pressure compressor 113 was 7 ° C., and the refrigerant was carbon dioxide. Thirst of Q _exo and density ratio at the intermediate pressure determined for any volume ratio (ρ _exi / ρ _lci) desirable volume ratio derived from the relationship (V _lc / V _ex) is from 2.1 to 2 .9.

Here, when the volume ratio (V _lc / V _ex) is below _{(1-Q exo) × (} ρ exi / ρ lci), since the injection of the liquid refrigerant occurs, the enthalpy of refrigerant drawn into the high-pressure compressor 101 Decrease significantly. As a result, the discharge refrigerant temperature of the high-pressure compressor 101 decreases, and the heating capacity required for the floor heating function of the heat pump water heater and the heating function of the air conditioner becomes insufficient. Further, the liquid refrigerant that should normally pass through the evaporator 111 passes through the forward recirculation path 115, so that the cooling capacity required for the cooling function of the air conditioner also decreases. Therefore, when using the refrigeration cycle apparatus 100 to a plurality of applications, the volume ratio (V _lc / V _ex) for each condition _{(1-Q exo) × (} ρ exi / ρ lci) above and permit reflux It is better to set the value to prevent as much as possible. The value is 4.7 in the above example.

  When the refrigeration cycle apparatus 100 is designed exclusively for each of a heat pump water heater, a cooling air conditioner, and a heating air conditioner, a volume ratio suitable for each application can be set. However, when the refrigeration cycle apparatus 100 is used in the multi-function heat pump system as shown in FIG. In the example described above, by selecting a desirable volume ratio of 4.7 under floor heating conditions, it is possible to reduce the amount of reflux as much as possible while reliably avoiding liquid refrigerant injection. Of course, since the optimal volume ratio changes according to external conditions such as season and weather, even if the present invention is applied to a single use refrigeration cycle apparatus, the benefits can be fully enjoyed.

  The multi-function heat pump system shown in FIG. 2 includes a heat pump water heater 12 with a floor heating function and an air conditioner 14, and a refrigeration cycle apparatus 100 common to the water heater 12 and the air conditioner 14 is used. However, a dedicated radiator 103 (FIG. 1) is provided for each of the water heater 12 and the air conditioner 14.

  Next, the operation of the refrigeration cycle apparatus 100 will be described.

  First, when starting, the expansion valve 109 is fully closed. Next, power supply to the motor of the high-pressure compressor 101 is started, and the high-pressure compressor 101 is driven. The high-pressure compressor 101 sucks and compresses the refrigerant in the intermediate pressure channel 114. The compressed refrigerant is sent to the expander 105 through the refrigerant pipe 102, the radiator 103 and the refrigerant pipe 104. The refrigerant pipe 104 on the inlet side of the expander 105 is filled with the refrigerant discharged from the high-pressure compressor 101. For this reason, the pressure in the refrigerant pipe 104 increases. Further, since the high-pressure compressor 101 sucks the refrigerant from the gas-liquid separator 107 through the forward recirculation path 115, the liquid refrigerant in the gas-liquid separator 107 evaporates. Therefore, the temperature and pressure in the refrigerant pipe 106 on the outlet side of the expander 105 are reduced. That is, when the high pressure compressor 101 is operated with the expansion valve 109 closed, the pressure difference between the inlet and the outlet of the expander 105 increases. The expander 105 is driven by the pressure difference thus generated.

  On the other hand, the high-pressure stage compressor 101, the forward recirculation path 115, and the gas-liquid separator 107 are connected to each other via the refrigerant pipe 114 (intermediate pressure flow path). The pressure inside falls. In addition, the refrigerant pipe 112 on the inlet side of the low-pressure stage compressor 113 is filled with a refrigerant having an evaporation pressure corresponding to the ambient temperature (heat source temperature) at the place where the evaporator 111 is installed (for example, outdoors). Therefore, at the time of startup, the pressure in the refrigerant pipe 112 temporarily exceeds the pressure in the intermediate pressure flow path 114. Then, the low-pressure compressor 113 behaves as an expander and is driven by a pressure difference between the refrigerant pipe 112 and the intermediate pressure flow path 114.

  As described with reference to FIG. 13, a refrigeration cycle apparatus having a configuration in which a low-pressure compressor and an expander are connected by a shaft is conventionally known. However, when a rotary expander is used as the expander 803 of the conventional refrigeration cycle apparatus shown in FIG. 13, the piston stops eccentrically on the vane side, that is, the piston stops with the suction port and the discharge port communicating with each other. There are things to do. In this case, there is a problem in that a sufficient initial pressure difference required for driving the low-pressure stage compressor 801b and the expander 803 cannot be obtained, and a start-up error occurs. On the other hand, according to the starting method in the present embodiment, the low-pressure stage compressor 113 temporarily behaves as an expander. Therefore, the piston of the expander 105 and the piston of the low-pressure stage compressor 113 are simultaneously eccentric to the vane side. Can be designed not to. That is, by making the eccentric directions of the pistons different from each other, it is possible to reliably generate a pressure difference necessary for driving between the suction side and the discharge side of at least one of the expander 105 and the low-pressure compressor 113. . Thereby, a refrigerating cycle device can be started reliably.

  When the low-pressure stage compressor 113 and the expander 105 start to operate, the low-pressure stage compressor 113 is driven by the recovered power of the expander 105 thereafter. The low-pressure stage compressor 113 sucks up the refrigerant from the refrigerant pipe 112, the evaporator 111 and the refrigerant pipe 110. Thereby, the liquid refrigerant begins to evaporate in the evaporator 111, and the temperature and pressure in the evaporator 111 are reduced. When the pressure in the refrigerant pipe 110 becomes lower than the pressure in the refrigerant pipe 108, the opening degree of the expansion valve 109 is gradually increased to the initial value. In the present embodiment, the expansion valve 109 is opened when the detected temperature of the second temperature sensor 122 becomes lower than the detected temperature of the fourth temperature sensor 124.

  Thereafter, the opening degree of the expansion valve 109 is controlled by the controller 118. Specifically, it is determined based on the refrigerant evaporation pressure in the evaporator 111, the suction refrigerant temperature of the expander 105, the discharge refrigerant pressure of the high-pressure compressor 101, and the suction refrigerant temperature of the low-pressure compressor 113. Under the set target cycle conditions, the pressure in the gas-liquid separator 107 is adjusted so that the theoretical recovery power of the expander 105 is equal to the theoretical compression work of the low-pressure compressor 113. This is a control for making the actual high pressure in the refrigerant circuit coincide with the optimum high pressure in the target cycle condition. The pressure (intermediate pressure) in the gas-liquid separator 107 can be adjusted by the expansion valve 109. The theoretical recovery power and the theoretical compression work are values derived by calculation, respectively, and do not mean actual recovery power and compression work.

  This will be described in more detail with reference to the flowchart of FIG.

First, in step 101, the refrigerant temperature T ₁ of the expander 105 from the first temperature sensor 121, the evaporation temperature T ₂ of the refrigerant in the evaporator 111 from the second temperature sensor 122, and the low pressure stage compressor from the third temperature sensor 123 are shown. It acquires 113 suction refrigerant temperature T ₃ of the. From the evaporation temperature T ₂ of the refrigerant in the evaporator 111, the evaporation pressure of the refrigerant in the evaporator 111 can be known.

  Next, in step 102, the optimum high pressure at which the COP of the refrigeration cycle apparatus 100 is maximized is calculated based on the temperature and pressure acquired in step 101.

Next, in Step 103 and Step 104, a target intermediate pressure at which the theoretical recovery power and the theoretical compression work are equal is calculated. First, at step 103, a certain target intermediate pressure is set. The recovery power (theoretical recovery power) when the refrigerant is expanded by the expander 105 to the set target intermediate pressure is calculated based on the calculated optimum high pressure and the intake refrigerant temperature T ₁ of the expander 105. As shown in FIG. 4, the state of the refrigerant at the inlet of the expander 105 is indicated by a point D. The point D is specified by the optimum high pressure P _H and the suction refrigerant temperature T ₁ . The target intermediate pressure P _M is the pressure at point E. In the expander 105, the refrigerant expands along the isentropic line (point D → point E). If the efficiency of the expander 105 is added to the enthalpy (h ₂ −h ₁ ) that the refrigerant has lost in the process of changing from the point D to the point E, the theoretical recovery power can be obtained.

Further, in step 103, the compression work (theoretical compression work) when the refrigerant is compressed by the low-pressure stage compressor 113 up to the set target intermediate pressure is expressed by the evaporation pressure P _L of the evaporator 111 and the suction refrigerant of the low-pressure stage compressor 113. calculated on the basis of the temperature T _3. As shown in FIG. 4, the state of the refrigerant at the inlet of the low-pressure compressor 113 is indicated by a point A. Point A is specified by the evaporation pressure P _L and the suction refrigerant temperature T ₃ . In the low-pressure compressor 113, the refrigerant is pre-compressed along the isentropic line (point A → point B). The enthalpy (h ₄ -h ₃ ) acquired by the refrigerant in the process of changing from point A to point B is divided by the efficiency of the low-pressure stage compressor 113, and further, the low-pressure stage compressor 113 with respect to the refrigerant mass flow rate of the expander 105. If the ratio of the refrigerant mass flow rate is integrated, theoretical compression work can be obtained.

The refrigerant mass flow rate of the expander 105 can be calculated from the refrigerant density at the inlet of the expander 105 and the suction volume of the expander 105. The refrigerant density at the inlet of the expander 105 is calculated from, for example, the optimum high pressure and the intake refrigerant temperature T ₁ . Similarly, the refrigerant mass flow rate of the low-pressure compressor 113 can be calculated from the refrigerant density at the inlet of the low-pressure compressor 113 and the suction volume of the low-pressure compressor 113. The refrigerant density at the inlet of the low-pressure compressor 113 is calculated from, for example, the evaporation temperature T ₂ and the suction refrigerant temperature T ₃ . The efficiency of the expander 105 and the low-pressure compressor 113 is a design value.

Next, in step 104, it is determined whether the theoretical recovery power and the theoretical compression work match. If they match, the process proceeds to step 105. If they do not match, the process returns to step 103, another target intermediate pressure is set, and the processes of step 103 and step 104 are repeated until they match. As described above, the controller 118 calculates an arbitrary optimum high pressure P _H and target intermediate pressure P _M based on the detection result of each temperature sensor.

Next, in step 105, the pressure in the gas-liquid separator 107 (actual intermediate pressure) is calculated. Specifically, first, the evaporation temperature T ₄ of the refrigerant in the gas-liquid separator 107 is acquired from the fourth temperature sensor 124. It can be calculated the pressure of the refrigerant from the evaporator temperature T ₄ of the refrigerant. That is, the controller 108 as a means for controlling the opening degree of the expansion valve 109 calculates the actual pressure in the gas-liquid separator 107 based on the detection result of the fourth temperature sensor 124.

Next, in step 106, the actual intermediate pressure is compared with the target intermediate pressure P _M. When the actual intermediate pressure is higher than the target intermediate pressure P _M , the process proceeds to step 107, and when it is lower, the process proceeds to step 107 ′. In step 107, the set opening degree of the expansion valve 109 is increased. In step 107 ′, the set opening degree of the expansion valve 109 is decreased.

  Next, in step 108, the set opening degree is output to the expansion valve 109, and the opening degree of the expansion valve 109 is changed. When the opening degree of the expansion valve 109 changes, the pressure in the gas-liquid separator 107 also changes. By periodically performing the process of this flowchart, the pressure in the gas-liquid separator 107 is adjusted and optimized so that the recovery power of the expander 105 and the compression work of the low-pressure compressor 113 are theoretically equal. High pressure is maintained.

As described above, the controller 118 sets the target intermediate pressure P _{M at} which the theoretical recovery power of the expander 105 at an arbitrary optimum high pressure in the refrigeration cycle is equal to the theoretical compression work of the low-pressure compressor 113 at any optimum high pressure. Means for calculating, and means for controlling the opening of the expansion valve 109 so that the actual pressure in the gas-liquid separator 107 approaches the calculated target intermediate pressure P _M are provided. Specifically, the controller 118 controls the opening degree of the expansion valve 109 based on the detection result of the fourth temperature sensor 124 so that the pressure in the gas-liquid separator 107 matches the target intermediate pressure P _M.

Further, when the volume ratio (V _lc / V _ex ) is excessive with respect to the cycle condition of the refrigeration cycle apparatus 100, the high-pressure stage compressor 101 and the expander with respect to the refrigerant mass flow rate of the low-pressure stage compressor 113. The refrigerant mass flow rate of 105 is insufficient. In other words, the suction amount of the high-pressure stage compressor 101 is too small with respect to the discharge amount from the low-pressure stage compressor 113. Therefore, the pressure in the intermediate pressure flow path 114 rises, and the refrigerant that is not sucked into the high pressure compressor 101 returns from the intermediate pressure flow path 114 to the gas-liquid separator 107 through the forward recirculation path 115.

On the other hand, when the volume ratio (V _lc / V _ex ) is too small with respect to the cycle condition of the refrigeration cycle apparatus 100, the high-pressure stage compressor 101 and the expander with respect to the refrigerant mass flow rate of the low-pressure stage compressor 113. The refrigerant mass flow rate of 105 becomes excessive. In other words, the suction amount of the high-pressure compressor 101 is excessive with respect to the discharge amount from the low-pressure compressor 113. Therefore, the pressure in the intermediate pressure flow path 114 decreases, and a shortage of refrigerant is injected from the gas-liquid separator 107 to the intermediate pressure flow path 114 through the forward recirculation path 115.

With reference to the Mollier diagrams shown in FIGS. 5A to 5C, the operation of the refrigeration cycle apparatus 100 in each application will be described. It is assumed that the volume ratio (V _lc / V _ex ) is set to 4.7.

As shown in FIG. 5A, the floor heating conditions described above, the volume ratio 4.7 corresponds to the value represented by _{(1-Q exo) × (} ρ exi / ρ lci). Therefore, all of the gas refrigerant in the gas-liquid separator 107 is injected into the intermediate pressure flow path 114 through the forward recirculation path 115, and the liquid refrigerant is sent to the evaporator 111 through the expansion valve 109.

As shown in Figure 5B, in the cooling condition described above, the volume ratio 4.7, exceeds the value expressed by (ρ _exi / ρ _lci). That is, the refrigerant mass flow rate of the low-pressure compressor 113 is excessive with respect to the refrigerant mass flow rate of the expander 105. Therefore, a part of the refrigerant compressed by the low-pressure compressor 113 is returned to the gas-liquid separator 107 through the forward recirculation path 115 and re-expanded by the expansion valve 109.

As shown in FIG. 5C, even in the heating conditions described above, similarly to the cooling conditions, the volume ratio 4.7, exceeds the value expressed by (ρ _exi / ρ _lci). Therefore, a part of the refrigerant compressed by the low-pressure compressor 113 is returned to the gas-liquid separator 107 through the forward recirculation path 115 and re-expanded by the expansion valve 109.

As described above, according to the present embodiment, the refrigerant can flow bidirectionally in the forward and backward flow path 115, so that the flow rate balance of the refrigeration cycle is established. Accordingly, the low-pressure compressor 113 and the expander 105 can be designed so that the annual power consumption is minimized without being restricted by the constant density ratio. Further, the intermediate pressure can be easily adjusted by controlling the opening degree of the expansion valve 109 so that the actual high pressure in the refrigerant circuit becomes the optimum high pressure regardless of the volume ratio (V _lc / V _ex ). The refrigeration cycle apparatus 100 can be operated.

In addition, by setting the volume ratio _{(V lc / V ex) (} 1-Q exo) × (ρ exi / ρ lci) represented the value above, it is possible to avoid the injection of the liquid refrigerant.

6A and 6B show changes in the volume ratio under floor heating cycle conditions (outside temperature 7 ° C., return temperature of warm water for floor heating 25 ° C., suction refrigerant temperature of low-pressure compressor 7 ° C., CO ₂ refrigerant). It is a graph which shows the change of the various cycle characteristics (calculated value) with respect to. The vertical axis in FIG. 6A shows the intermediate pressure, the COP, and the refrigerant flow rate in the forward recirculation path as cycle characteristics. The vertical axis of FIG. 6B indicates the refrigerant discharge temperature of the high-pressure compressor 101 as cycle characteristics.

In the floor heating cycle condition, as shown in the graph of FIG. 6A, the refrigerant flow rate in the forward recirculation path 115 is switched between the volume ratio (V _lc / V _ex ) = 7.1 as a boundary. When the refrigerant flow rate is positive, the refrigerant flows from the gas-liquid separator 107 to the intermediate pressure channel 114 (injection). When the refrigerant flow rate is negative, the refrigerant flows from the intermediate pressure channel 114 to the gas-liquid separator 107. (reflux). When gas injection occurs, both COP and intermediate pressure increase. This is because the compression load of the low-pressure stage compressor 113 is reduced by reducing the refrigerant mass flow rate of the low-pressure stage compressor 113, and as a result, the compression load of the high-pressure stage compressor 101 is reduced.

Further, under the above floor heating cycle condition, as shown in FIG. 6B, the discharge refrigerant temperature of the high-pressure compressor 101 rapidly decreases with the volume ratio (V _lc / V _ex ) = 4.7 as a boundary. That is, the injection of the liquid refrigerant (including the gas refrigerant) and the injection of the gas refrigerant are switched around the volume ratio (V _lc / V _ex ) = 4.7. As described above, when (1-Q _exo) less than × (ρ _exi / ρ _lci) represented value is the volume ratio (V _lc / V _ex) does not occur injection of the liquid refrigerant . Conversely, (1-Q _exo) when × value expressed by (ρ _exi / ρ _lci) is greater than the volume ratio (V _lc / V _ex) is injection of the liquid refrigerant occurs.

Further, the change in the refrigerant temperature discharged from the high-pressure compressor 101 stops at the boundary of the volume ratio (V _lc / V _ex ) = 7.1. That is, the injection and recirculation of the gas refrigerant are switched around the volume ratio (V _lc / V _ex ) = 7.1. As described above, when the density ratio (ρ _exi / ρ _lci) is greater than the volume ratio (V _lc / V _ex) is refluxed does not occur. Conversely, if the density ratio (ρ _exi / ρ _lci) is smaller than the volume ratio (V _lc / V _ex) is refluxed occurs.

In view of the above, assuming that operation is performed under floor heating cycle conditions, the volume ratio (V _lc / V _ex ) is in the range of 4.7 to 7.1 in order to avoid liquid refrigerant injection and reflux. Should be set.

As described with reference to FIG. 2, in the multi-function heat pump system, it should be considered that a single refrigeration cycle apparatus 100 covers hot water supply, floor heating, and air conditioning from the viewpoint of cost and the like. In such a case, if the volume ratio is set with only floor heating in mind, there is a possibility that efficient operation cannot be performed in other applications. That is, as shown in FIG. 7A and FIG. 7B, under the cooling cycle conditions, the liquid refrigerant injection and reflux can be avoided by setting the volume ratio (V _lc / V _ex ) in the range of 2.4 to 3.6. . Similarly, as shown in FIG. 8A and FIG. 8B, in the heating cycle condition, by setting the volume ratio (V _lc / V _ex ) in the range of 2.1 to 2.9, liquid refrigerant injection and reflux are avoided. it can. Thus, the range of suitable volume ratios differs depending on the cycle conditions. Although reflux is acceptable, liquid refrigerant injection should be avoided, so in this example, a volume ratio (V _lc / V _ex ) = 4.7 that can avoid liquid injection under floor heating cycle conditions is reasonable.

When setting the volume ratio (V _lc / V _ex) to 4.7, the floor heating condition, since the volume ratio (V _lc / V _ex) is equal to or less than the density ratio (ρ _exi / ρ _lci), the low-pressure stage compressor All the refrigerant compressed by the machine 113 is sucked into the high-pressure compressor 101, and the refrigeration cycle apparatus 100 can be operated efficiently. Not floor heating conditions but also in the cooling and the application of the heating, it is smaller than _{(1-Q exo) × (} ρ exi / ρ lci) represented value is the volume ratio (V _lc / V _ex), Liquid refrigerant injection can be avoided.

If the volume ratio (V _lc / V _ex ) = 4.7, 23.6% of the refrigerant compressed by the low-pressure compressor 113 passes to the gas-liquid separator 107 via the forward recirculation path 115 under the cooling condition. And the expansion valve 109 is re-expanded. When the calculation is performed under the same cooling condition, 49.6% of the refrigerant bypasses the expander in the conventional example (FIG. 14) in which the restriction of the constant density ratio is avoided in the bypass circuit. Similarly, under the heating condition, 36.6% of the refrigerant compressed by the low-pressure compressor 113 returns to the gas-liquid separator 107 via the forward recirculation path 115 and is re-expanded by the expansion valve 109. When the calculation is performed under the same heating condition, 58.2% of the refrigerant bypasses the expander in the conventional example (FIG. 14) in which the restriction of the constant density ratio is avoided in the bypass circuit. Thus, according to the refrigeration cycle apparatus 100 of the present embodiment, a more efficient operation can be performed than a conventional refrigeration cycle apparatus having a bypass circuit.

  It should be noted that the present invention does not completely prohibit liquid refrigerant injection.

(Second Embodiment)
FIG. 9 is a configuration diagram of a refrigeration cycle apparatus according to the second embodiment of the present invention. The refrigeration cycle apparatus 500 of the present embodiment has a configuration that is substantially the same as that of the refrigeration cycle apparatus 100 (see FIG. 1) according to the first embodiment. The difference between the present embodiment and the first embodiment is that the temperature sensor 520 is provided and the control performed by the controller 118. In the following, the same functional parts are denoted by the same reference numerals, and the description thereof is omitted.

  As shown in FIG. 9, the refrigeration cycle apparatus 500 includes a temperature sensor 520 for detecting the discharge refrigerant temperature of the high-pressure compressor 101. Similar to the first embodiment, a temperature sensor 122 for detecting the evaporation temperature of the refrigerant in the evaporator 111 is also provided. The controller 118 controls the opening degree of the expansion valve 109 based on the detection results of the temperature sensor 520 and the temperature sensor 122.

  The operation of the refrigeration cycle apparatus 500 will be described with reference to the flowchart of FIG. First, in step 501, the outside air temperature is estimated based on the evaporation temperature of the refrigerant in the evaporator 111. Next, in step 502, the target discharge refrigerant temperature of the high-pressure compressor 101 is calculated. The target discharge refrigerant temperature is determined according to, for example, the outside air temperature and the floor heating set temperature (or the heating set temperature). Next, in step 503, the actual discharge refrigerant temperature of the high-pressure compressor 101 is acquired from the temperature sensor 520. Next, in step 504, the actual discharge refrigerant temperature is compared with the target discharge refrigerant temperature. When the actual discharge refrigerant temperature is higher than the target discharge refrigerant temperature, the process proceeds to step 505. If the actual discharge refrigerant temperature is lower than the target discharge refrigerant temperature, the process proceeds to step 505 '.

  In step 505 and step 506, the set opening degree of the expansion valve 109 is increased to lower the intermediate pressure. When the intermediate pressure decreases, the compression work of the low-pressure stage compressor 113 decreases. Thereby, an imbalance occurs between the compression work of the low-pressure stage compressor 113 and the recovered power of the expander 105. In order to eliminate this imbalance, the rotation speed of the shaft 116 increases and the high pressure of the refrigeration cycle decreases. As a result, the recovery power of the expander 105 is reduced and the unbalance is eliminated. Moreover, the discharge refrigerant | coolant temperature of the high pressure compressor 101 falls because the high pressure of a refrigerating cycle falls.

  In steps 505 'and 506, the set opening degree of the expansion valve 109 is decreased to increase the intermediate pressure. When the intermediate pressure increases, the compression work of the low-pressure compressor 113 increases. Thereby, an imbalance occurs between the compression work of the low-pressure stage compressor 113 and the recovered power of the expander 105. In order to eliminate this imbalance, the rotational speed of the shaft 116 is decreased and the high pressure of the refrigeration cycle is increased. As a result, the recovery power of the expander 105 is increased and the unbalance is eliminated. Moreover, the discharge refrigerant | coolant temperature of the high pressure compressor 101 rises because the high pressure of a refrigerating cycle rises.

  According to this embodiment, the opening degree of the expansion valve 109 is controlled based on the detection results of the temperature sensors 520 and 123. By controlling the opening degree of the expansion valve 109, the intermediate pressure can be adjusted. The rotation speed of the shaft 116 changes according to the intermediate pressure. When the rotation speed of the shaft 116 changes, the high pressure of the refrigeration cycle also changes. That is, the expansion valve 109 can adjust the high pressure of the refrigeration cycle. Therefore, the refrigeration cycle apparatus 500 of the present embodiment is suitable for applications that require heating capability such as floor heating, heating, and hot water supply.

(Third embodiment)
FIG. 11 is a configuration diagram of a refrigeration cycle apparatus according to the third embodiment of the present invention. The refrigeration cycle apparatus 700 has substantially the same configuration as the refrigeration cycle apparatus described in the first embodiment and the second embodiment. The difference between the present embodiment and the first embodiment is that the high-pressure stage compressor 701, the low-pressure stage compressor 713, and the expander 705 are housed in a common sealed container 717.

  As shown in FIG. 11, in the refrigeration cycle apparatus 700, the high-pressure stage compressor 701, the low-pressure stage compressor 713, and the expander 705 are arranged in this order from above in a single sealed container 717. The low-pressure stage compressor 713 and the expander 705 are connected by a shaft 716 so that power can be transmitted. Oil is stored at the bottom of the sealed container 717. The space above the oil level is filled with the refrigerant discharged from the high-pressure compressor 701. The periphery of the low-pressure stage compressor 713 and the expander 705 is filled with oil.

  When the high-pressure compressor 701 is driven, the space above the oil level is filled with high-pressure discharged refrigerant. Around the high-pressure compressor 701, high-temperature oil that lubricates the high-pressure compressor 701 is held. On the other hand, the low-pressure stage compressor 713 and the expander 705 operate at a lower temperature than the high-pressure stage compressor 701. Therefore, oil having a lower temperature than the oil existing around the high-pressure stage compressor 713 is held around the low-pressure stage compressor 713 and the expander 705.

  That is, high temperature oil is held around the high pressure stage compressor 701, and low temperature oil is held around the low pressure stage compressor 713 and the expander 705. The oil forms a temperature stratification along the vertical direction. By forming the temperature stratification, it becomes difficult to mix the upper layer oil and the lower layer oil. Thereby, the heat transfer from the high pressure compressor 701 to the expander 705 via the oil can be suppressed. When heat transfer occurs, the discharge refrigerant temperature of the high-pressure compressor 701 decreases and the discharge refrigerant temperature of the expander 705 increases, which is not preferable from the viewpoint of the efficiency of the refrigeration cycle apparatus. According to this embodiment, since heat transfer can be effectively suppressed, the efficiency of the refrigeration cycle apparatus 700 can be increased.

(Fourth embodiment)
FIG. 12A is a configuration diagram of a refrigeration cycle apparatus according to the fourth embodiment of the present invention. In the refrigeration cycle apparatus 600 of the present embodiment, a multistage rotary expander 605 capable of changing the suction volume is used. Furthermore, the refrigeration cycle apparatus 600 includes an expander injection flow path 630 and an expander injection valve 631. As shown in FIG. 12B, the expander injection flow path 630 connects the suction flow path (pipe 104) of the expander 605 and the expander injection port 632 opened to the expansion chamber 611 of the expander 605. The expander injection valve 631 is provided on the expander injection flow path 630. The suction volume of the expander 605 can be changed by controlling the expander injection valve 631. Other configurations are the same as those described in the first and second embodiments.

  As shown in FIG. 12B, the expander 605 is a two-stage rotary expander having a first-stage cylinder 605a and a second-stage cylinder 605b. The expander injection port 632 is provided in the first stage cylinder 605a and opens toward the expansion chamber 611 of the first stage cylinder 605a. The first-stage cylinder 605a and the second-stage cylinder 605b are separated by an intermediate plate 605c, but the expansion chamber 611 of the first-stage cylinder 605a is in two stages through a communication hole 605d formed in the intermediate plate 605c. It communicates with the expansion chamber 612 of the eye cylinder 605b. Thereby, the expansion chambers 611 and 612 form a single expansion chamber. The expander injection flow path 630 branches from the pipe 104 and is connected to the first-stage cylinder 605a so that the refrigerant can be injected from the expander injection port 632 into the expansion chamber 611. The expander injection port 632 is provided in the vicinity of the communication hole 605d in the circumferential direction of the shaft 116.

When the expander injection valve 631 is closed, the refrigerant does not flow into the expansion chamber 611 from the expander injection port 632. Therefore, the cylinder volume of the first-stage cylinder 605a behaves as the suction volume V _ex. On the other hand, when the expander injection valve 631 is opened, the refrigerant flows into the expansion chamber 611 from the expander injection port 632. Therefore, the cylinder volume of the second-stage cylinder 605b behaves as the suction volume V _ex ′. For example, cylinder volume of the first-stage cylinder 605a may double the cylinder volume of the second-stage cylinder 605b, can change the suction volume V _ex optionally twice the suction volume V _ex '.

According to the first embodiment, the desirable volume ratio (V _lc / V _ex ) under the cooling condition was 2.4 to 3.6. In the present embodiment, when the expander injection valve 631 is opened and the suction volume V _ex ′ is used, the desirable volume ratio (V _lc / V _ex ′) under the same cooling condition is 2.4 to 3.6. If expressed using the suction volume V _ex , the desirable volume ratio (V _lc / V _ex ) under the same cooling condition is 4.8 to 7.2. Also, according to the first embodiment, the desired volume ratio in the heating condition (V _lc / V _ex) was 2.1 to 2.9. In the present embodiment, when the expander injection valve 631 is opened and the suction volume V _ex ′ is used, the desirable volume ratio (V _lc / V _ex ′) under the same heating condition is 2.1 to 2.9. If expressed using the suction volume V _ex , the desirable volume ratio (V _lc / V _ex ) under the same heating condition is 4.2 to 5.8. Also, according to the first embodiment, the desired volume ratio in the floor heating condition (V _lc / V _ex) was 4.7 to 7.1. Also in the present embodiment, when the expander injection valve 631 is closed and the suction volume V _ex is used, the desirable volume ratio (V _lc / V _ex ) under the same floor heating condition is 4.7 to 7.1.

That is, in the cooling condition and the heating condition, the expander injection valve 631 is opened to increase the suction volume, while in the floor heating condition, the expander injection valve 631 is closed. By performing such control, the desired volume ratio (V _lc / V _ex ) under each condition can be brought closer to each other.

Specifically, the low-pressure compressor 113 and the expander 605 can be designed so that the volume ratio (V _lc / V _ex ) is 4.8. The design of volume ratio (V _lc / V _ex ) = 4.8 is below the upper limit of the volume ratio (V _lc / V _ex ) in each condition of cooling, heating and floor heating. Therefore, reflux does not occur in each of the cooling, heating, and floor heating conditions, and the compression work in the low-pressure compressor 113 can be used effectively. Furthermore, the design of the volume ratio (V _lc / V _ex ) = 4.8 exceeds the lower limit of the volume ratio (V _lc / V _ex ) for each condition of cooling, heating and floor heating. Therefore, liquid refrigerant injection can also be prevented.

  Accordingly, the gas refrigerant flows through the forward recirculation passage 115 from the gas-liquid separator 107 toward the intermediate pressure passage 114. The refrigeration cycle apparatus 600 can be always operated in a gas injection state. If the operation can be performed in the gas injection state, it is not necessary to compress the refrigerant that has flowed through the forward recirculation path 115 by the low-pressure stage compressor 113, so that the load on the low-pressure stage compressor 113 can be reduced. In addition, since the entire amount of the refrigerant cooled by the radiator 103 passes through the expander 605, high power recovery efficiency can be realized. Thus, according to the refrigeration cycle apparatus 600 of the present embodiment, a more efficient operation can be performed than a conventional refrigeration cycle apparatus having a bypass circuit.

  As the expander 605, other types of expanders such as a scroll expander and a reciprocal expander can be used. The expander injection valve 631 may be a valve whose opening degree can be changed in multiple stages, or may be a simple on-off valve.

  INDUSTRIAL APPLICATION This invention is useful about the refrigerating-cycle apparatus utilized for an air conditioner, a refrigerator-freezer, a heat pump water heater, a heat pump heater, a vending machine, a car air conditioner, etc. In particular, a high effect can be obtained when one refrigeration cycle apparatus is shared for two or more uses. Of course, the present invention can also be suitably applied to a refrigeration cycle apparatus for a single application.

Claims (11)

A positive displacement low-pressure stage compressor for pre-compressing the refrigerant;
A high-pressure stage compressor for further compressing the refrigerant pre-compressed by the low-pressure stage compressor;
An intermediate pressure flow path connecting the low pressure stage compressor and the high pressure stage compressor in series so that the refrigerant pre-compressed by the low pressure stage compressor is sent to the high pressure stage compressor;
A radiator for cooling the refrigerant compressed by the high-pressure stage compressor;
It is coaxially connected to the low-pressure stage compressor so that power transmission is performed, and is configured so that the entire amount of the refrigerant cooled by the radiator passes, and for recovering power by expanding the refrigerant A positive displacement expander;
A gas-liquid separator for separating the refrigerant expanded by the expander into a gas refrigerant and a liquid refrigerant;
An evaporator for evaporating the liquid refrigerant separated by the gas-liquid separator;
An expansion valve with variable opening provided on a flow path between the liquid refrigerant outlet of the gas-liquid separator and the inlet of the evaporator;
The refrigerant stored in the gas-liquid separator is preliminarily compressed by the low-pressure stage compressor in a first flow state where the refrigerant is led to the inlet of the high-pressure stage compressor without passing through the evaporator and the low-pressure stage compressor. A recirculation path connecting the intermediate pressure flow path and the gas-liquid separator so as to be able to switch between a second flow state in which a part of the refrigerant circulates back to the gas-liquid separator;
A controller for adjusting the flow rate of the refrigerant in the forward / return path in each of the first circulation state and the second circulation state by controlling the opening of the expansion valve;
A refrigeration cycle apparatus comprising: The cylinder volume of the low-pressure stage compressor and the cylinder volume of the expander are each constant,
The refrigeration cycle apparatus according to claim 1, wherein a cylinder volume of the low-pressure stage compressor is larger than a cylinder volume of the expander. The controller is
Means for calculating a target intermediate pressure at which the theoretical recovery power of the expander at any optimum high pressure of the refrigeration cycle is equal to the theoretical compression work of the low pressure stage compressor at any optimum high pressure;
Means for controlling the opening of the expansion valve so that the actual pressure in the gas-liquid separator approaches the calculated target intermediate pressure,
The refrigeration cycle apparatus according to claim 1 or 2. A first temperature sensor for detecting an inlet refrigerant temperature of the expander;
A second temperature sensor for detecting a refrigerant evaporation temperature of the evaporator;
A third temperature sensor for detecting an inlet refrigerant temperature of the low-pressure stage compressor;
Further comprising
The refrigeration cycle apparatus according to claim 3, wherein the controller calculates the arbitrary optimum high pressure and the target intermediate pressure based on detection results of the first to third temperature sensors. A fourth temperature sensor for detecting a refrigerant temperature in the gas-liquid separator;
The refrigeration cycle apparatus according to claim 4, wherein the means for controlling the opening degree of the expansion valve calculates an actual pressure in the gas-liquid separator based on a detection result of the fourth temperature sensor. The suction volume of the expander is V _ex , the suction volume of the low-pressure stage compressor is V _lc , the thirst degree of the refrigerant discharged from the expander is Q _exo , the density of the suction refrigerant of the expander is ρ _exi , and the low-pressure stage Compressor suction refrigerant density ρ _lci
The refrigeration cycle apparatus according to any one of claims 1 to 5, wherein the relationship represented by the following expression (1) is satisfied.
_{(1-Q exo) × (} ρ exi / ρ lci) ≦ (V lc / V ex) ··· (1) The suction volume V _ex expander, the suction volume V _lc of the low-pressure compressor, the density of the suction refrigerant of the expander [rho _exi, when the density [rho _lci of suction refrigerant of the low-pressure compressor,
The refrigeration cycle apparatus according to any one of claims 1 to 5, which satisfies the relationship represented by the following expression (2).
_{(V lc / V ex) ≦} (ρ exi / ρ lci) ··· (2) The suction volume of the expander is V _ex , the suction volume of the low-pressure stage compressor is V _lc , the thirst degree of the refrigerant discharged from the expander is Q _exo , the density of the suction refrigerant of the expander is ρ _exi , and the low-pressure stage Compressor suction refrigerant density ρ _lci
The refrigeration cycle apparatus according to any one of claims 1 to 5, which satisfies the relationship represented by the following expression (3).
_{(1-Q exo) × (} ρ exi / ρ lci) ≦ (V lc / V ex) ≦ (ρ exi / ρ lci) ··· (3) An expander injection flow path connecting the suction flow path of the expander and an expander injection port opened toward the expansion chamber of the expander;
An expander injection valve provided on the expander injection flow path,
The refrigeration cycle apparatus according to any one of claims 1 to 8, wherein a suction volume of the expander can be changed by controlling the expander injection valve.   The refrigeration cycle apparatus according to any one of claims 1 to 9, wherein the low-pressure stage compressor and the expander are arranged in a common sealed container. A heat pump water heater having a hot water supply function that can supply hot water to the faucet and / or a floor heating function that heats the interior of the house by circulating hot water through a piping around the floor of the house;
An air conditioner configured to adjust indoor temperature by exchanging heat between indoor air and refrigerant;
A multifunction heat pump system in which the refrigeration cycle apparatus according to any one of claims 1 to 10 is used as a refrigeration cycle apparatus common to the water heater and the air conditioner.

Images