WO2017068649A1 - Heat pump system - Google Patents

Heat pump system Download PDF

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Publication number
WO2017068649A1
WO2017068649A1 PCT/JP2015/079571 JP2015079571W WO2017068649A1 WO 2017068649 A1 WO2017068649 A1 WO 2017068649A1 JP 2015079571 W JP2015079571 W JP 2015079571W WO 2017068649 A1 WO2017068649 A1 WO 2017068649A1
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WO
WIPO (PCT)
Prior art keywords
heat exchanger
refrigerant
refrigerant flowing
expansion valve
outdoor heat
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PCT/JP2015/079571
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French (fr)
Japanese (ja)
Inventor
千歳 田中
拓也 松田
航祐 田中
Original Assignee
三菱電機株式会社
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Application filed by 三菱電機株式会社 filed Critical 三菱電機株式会社
Priority to PCT/JP2015/079571 priority Critical patent/WO2017068649A1/en
Priority to JP2017546315A priority patent/JPWO2017068649A1/en
Publication of WO2017068649A1 publication Critical patent/WO2017068649A1/en

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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B1/00—Compression machines, plants or systems with non-reversible cycle
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F24—HEATING; RANGES; VENTILATING
    • F24F—AIR-CONDITIONING; AIR-HUMIDIFICATION; VENTILATION; USE OF AIR CURRENTS FOR SCREENING
    • F24F11/00—Control or safety arrangements
    • F24F11/89—Arrangement or mounting of control or safety devices

Definitions

  • the present invention relates to a heat pump system, and more particularly to a heat pump system including a bypass flow path configured such that at least a part of a refrigerant bypasses an outdoor heat exchanger.
  • condensation pressure of the refrigerant in the outdoor heat exchanger is excessively low during the cooling operation of the heat pump system, frost adheres to the indoor heat exchanger (evaporator) or excessive dehumidification occurs. Malfunctions can occur. Therefore, a technique for suppressing a decrease in the condensation pressure has been proposed.
  • Patent Document 1 discloses an air conditioner including a three-way valve that adjusts a ratio between a flow rate of refrigerant passing through a condenser and a flow rate of refrigerant bypassing the aggregator.
  • This air conditioner control means is used when a predetermined condition is satisfied (specifically, when the expansion valve is fully opened, when the compression ratio of the compressor is less than the allowable minimum compression ratio, or when the reheat amount of the reheater is a predetermined value).
  • the three-way valve is controlled in the direction of increasing the condensation pressure.
  • the refrigerant bypasses the condenser by changing the opening of the three-way valve as compared to when the predetermined condition is not satisfied.
  • the flow rate is set large. Thereby, the fall of a condensation pressure can be suppressed.
  • the flow rate of the refrigerant flowing through the condenser decreases as the flow rate of the refrigerant bypassing the condenser increases. For this reason, the amount of heat released from the refrigerant to the outside air in the condenser is reduced, so that the amount of decrease in the refrigerant temperature in the condenser can be reduced. If it does so, there exists a possibility that required supercooling degree cannot be ensured about the refrigerant
  • the present invention has been made to solve the above-described problems, and its purpose is to ensure a degree of supercooling of the refrigerant flowing into the expansion valve while suppressing a decrease in condensation pressure during cooling operation of the heat pump system. Is to provide new technology.
  • the same problem can occur during the heating operation of the heat pump system. That is, for example, when the heating operation is performed when the outside air temperature is high, the evaporation pressure of the refrigerant in the outdoor heat exchanger (evaporator) becomes higher than when the outside air temperature is low. That is, since the pressure of the refrigerant flowing out from the outdoor heat exchanger and flowing into the compressor increases, the compression ratio in the compressor decreases. In this case, by increasing the flow rate of the refrigerant that bypasses the evaporator by changing the opening of the three-way valve, an increase in the evaporation pressure can be suppressed.
  • the flow rate of the refrigerant flowing through the evaporator is reduced by the increase in the flow rate of the refrigerant bypassing the evaporator. For this reason, the amount of heat absorbed from the outside air to the refrigerant in the evaporator becomes small, so that the amount of increase in the refrigerant temperature in the evaporator can be small. If it does so, there exists a possibility that required superheat degree cannot be ensured about the refrigerant
  • Another object of the present invention is to provide a technology capable of ensuring the degree of superheat of the refrigerant flowing into the compressor while suppressing an increase in evaporation pressure during heating operation of the heat pump system.
  • a heat pump system includes a main circuit, a bypass flow path, a flow rate regulator, a branch flow path, a second expansion valve, and a sub heat exchanger.
  • the main circuit includes a compressor, an outdoor heat exchanger, a first expansion valve, and an indoor heat exchanger, and is configured to be able to circulate the refrigerant.
  • the bypass flow path is configured such that the refrigerant flowing through the main circuit bypasses the outdoor heat exchanger.
  • the flow rate adjuster is configured to be able to adjust the ratio of the flow rate of the refrigerant flowing through the outdoor heat exchanger to the flow rate of the refrigerant discharged from the compressor.
  • the branch flow path is connected between the outdoor heat exchanger and the first expansion valve, and is configured to be able to branch a part of the refrigerant flowing through the main circuit.
  • the second expansion valve is configured to be able to adjust the flow rate of the refrigerant flowing through the branch flow path.
  • the auxiliary heat exchanger exchanges heat between the refrigerant flowing through the branch flow path and the refrigerant flowing from the outdoor heat exchanger to the first expansion valve.
  • a heat pump system includes a main circuit, a bypass flow channel, a flow rate regulator, a branch flow channel, a second expansion valve, and a sub heat exchanger.
  • the main circuit includes a compressor, an outdoor heat exchanger, a first expansion valve, and an indoor heat exchanger, and is configured to be able to circulate the refrigerant.
  • the bypass flow path is configured such that the refrigerant flowing through the main circuit bypasses the outdoor heat exchanger.
  • the flow rate adjuster is configured to be able to adjust the ratio of the flow rate of the refrigerant flowing through the outdoor heat exchanger to the flow rate of the refrigerant discharged from the compressor.
  • the branch flow path is connected between the compressor and the indoor heat exchanger, and is configured to be able to branch a part of the refrigerant flowing through the main circuit.
  • the second expansion valve is configured to be able to adjust the flow rate of the refrigerant flowing through the branch flow path.
  • the auxiliary heat exchanger exchanges heat between the refrigerant flowing through the branch flow path and the refrigerant flowing from the outdoor heat exchanger to the compressor.
  • the ratio can be reduced by the flow rate regulator as compared with the case where the condensing pressure is within an appropriate range.
  • the flow rate of the refrigerant flowing through the branch flow path via the second expansion valve can be increased.
  • the amount of heat exchanged between the refrigerant flowing through the main circuit and the refrigerant decompressed by the second expansion valve is increased.
  • the temperature of the refrigerant flowing through the main circuit decreases, so that the degree of supercooling of the refrigerant flowing into the expansion valve can be ensured more reliably. Therefore, during the cooling operation of the heat pump system, it is possible to ensure the degree of supercooling while suppressing a decrease in the condensation pressure.
  • the degree of superheat of the refrigerant flowing into the compressor can be ensured while suppressing the increase in the evaporation pressure even during the heating operation of the heat pump system.
  • FIG. 1 is a block diagram schematically showing a configuration of a heat pump system according to Embodiment 1.
  • FIG. It is a figure which shows the structure of a three-way valve typically.
  • 3 is a flowchart for illustrating control of a three-way valve and a sub-expansion valve that is executed in the first embodiment.
  • FIG. 6 is a Ph diagram corresponding to the control of the three-way valve and the sub-expansion valve executed in the first embodiment.
  • It is a block diagram which shows schematically the structure of the heat pump system which concerns on the modification of Embodiment 1.
  • FIG. It is a block diagram which shows schematically the structure of the heat pump system which concerns on Embodiment 2.
  • FIG. 6 is a flowchart for illustrating control of a three-way valve and a sub-expansion valve that is executed in a second embodiment.
  • FIG. 1 is a block diagram schematically showing the configuration of the heat pump system according to the first embodiment.
  • heat pump system 100 includes a main circuit 110 and a control device 500.
  • Main circuit 110 includes compressor 10, outdoor heat exchanger 20, main expansion valve 30, indoor heat exchanger 40, and pipe 50.
  • the refrigerant flow during the cooling operation of the heat pump system 100 is indicated by an arrow REF.
  • the compressor 10 is a variable capacity compressor driven by, for example, an inverter (not shown).
  • the gas refrigerant that has been compressed by the compressor 10 to a high temperature and high pressure flows into the outdoor heat exchanger 20.
  • the outdoor heat exchanger 20 is a heat exchanger configured to include, for example, heat transfer tubes and heat radiating fins (not shown). During the cooling operation of the heat pump system 100, the outdoor heat exchanger 20 functions as a condenser. In the outdoor heat exchanger 20, the gas refrigerant condenses into a liquid refrigerant by dissipating heat to the outside air.
  • the main expansion valve (first expansion valve) 30 is a throttle valve whose opening degree can be controlled by, for example, a stepping motor (not shown).
  • the main expansion valve 30 is used for adjusting the flow rate of refrigerant (amount of refrigerant flowing per unit time).
  • the main expansion valve 30 decompresses the liquid refrigerant by expanding the high-pressure liquid refrigerant condensed by the outdoor heat exchanger 20. As a result, the refrigerant becomes a gas-liquid two-phase refrigerant and flows into the indoor heat exchanger 40.
  • the indoor heat exchanger 40 is a heat exchanger configured to include heat transfer tubes and radiating fins (not shown), like the outdoor heat exchanger 20.
  • the indoor heat exchanger 40 functions as an evaporator.
  • the air is cooled by the refrigerant.
  • the refrigerant is heated to change from a gas-liquid two-phase refrigerant to a low-pressure gas refrigerant. Thereafter, the gas refrigerant returns to the compressor 10 and is compressed again by the compressor 10 and discharged.
  • the main circuit 110 is formed by connecting the compressor 10, the outdoor heat exchanger 20, the main expansion valve 30, and the indoor heat exchanger 40 through the pipe 50 in this order.
  • the main circuit 110 may further include an accumulator (low pressure receiver), a receiver (high pressure receiver), or an oil separator (oil separator).
  • the heat pump system 100 further includes a bypass passage 120, a three-way valve 60, a branch passage 130, an internal heat exchanger 80, and a sub-expansion valve 70.
  • the heat pump system 100 further includes a pressure sensor 91 and temperature sensors 92 and 93.
  • the bypass channel 120 is connected between the upstream side and the downstream side of the outdoor heat exchanger 20.
  • the bypass passage 120 is configured such that the refrigerant flowing through the main circuit 110 bypasses the outdoor heat exchanger 20.
  • the three-way valve 60 is connected to a connection portion between the main circuit 110 and the bypass flow path 120.
  • the opening degree of the three-way valve 60 is changed by a control signal from the control device 500.
  • the configuration of the three-way valve 60 will be described in detail with reference to FIG.
  • the branch flow path 130 is connected between the outdoor heat exchanger 20 and the main expansion valve 30.
  • the branch flow path 130 is configured to be able to branch a part of the refrigerant flowing through the main circuit 110.
  • the secondary expansion valve (second expansion valve) 70 is provided in the branch flow path 130.
  • the sub-expansion valve 70 is a throttle valve whose opening degree can be controlled by, for example, a stepping motor (not shown). That is, the sub expansion valve 70 is configured to be able to adjust the flow rate of the refrigerant flowing through the branch flow path 130.
  • the internal heat exchanger (sub heat exchanger) 80 exchanges heat between the refrigerant flowing through the branch flow path 130 and the refrigerant flowing from the outdoor heat exchanger 20 to the main expansion valve 30.
  • the refrigerant that has undergone heat exchange in the internal heat exchanger 80 joins the refrigerant that flows from the indoor heat exchanger 40 to the compressor 10.
  • the pressure sensor 91 is provided on the discharge side of the compressor 10.
  • the pressure sensor 91 detects the pressure (discharge pressure) of the refrigerant discharged from the compressor 10 and outputs a signal indicating the detection result to the control device 500.
  • the discharge pressure of the compressor 10 is substantially equal to the condensation pressure Pc in the outdoor heat exchanger 20. Therefore, the condensation pressure Pc can be acquired by detecting the discharge pressure with the pressure sensor 91.
  • the temperature sensor 92 is provided on the discharge side of the compressor 10.
  • the temperature sensor 92 includes, for example, a thermistor, and detects the temperature (discharge temperature) of the refrigerant discharged from the compressor 10.
  • the temperature sensor 93 is provided on the outlet side of the internal heat exchanger 80.
  • the temperature sensor 93 includes, for example, a thermistor, and detects the condensation temperature of the refrigerant in the internal heat exchanger 80. Each sensor outputs a signal indicating the detection result to the control device 500.
  • control device 500 includes a CPU (Central Processing Unit), a memory such as a RAM (Random Access Memory) and a ROM (Read Only Memory), and an input / output interface. Based on the detection signals from the above-described sensors, the control device 500 controls each device by causing the CPU to read a program stored in advance in a ROM or the like into the RAM and execute it.
  • CPU Central Processing Unit
  • RAM Random Access Memory
  • ROM Read Only Memory
  • control device 500 executes discharge temperature control for controlling the discharge temperature of the refrigerant to a target value by adjusting the opening of the main expansion valve 30. Moreover, the control apparatus 500 controls the three-way valve 60 and the sub expansion valve 70 based on the detection signal from each sensor. Details of control of the three-way valve 60 and the sub-expansion valve 70 by the control device 500 will be described later.
  • FIG. 2 is a diagram schematically showing the configuration of the three-way valve 60.
  • a three-way valve 60 includes an input port IN into which the refrigerant discharged from the compressor 10 flows, and an output port OUT1 through which the refrigerant flows out to the outdoor heat exchanger 20. And an output port OUT2 for allowing the refrigerant to flow out to the bypass channel 120, and a valve body 61.
  • the angle ⁇ of the valve body 61 can be changed within a range of 0 ° to 90 °, for example.
  • the ratio R of the flow rate of the refrigerant flowing out from the output port OUT1 with respect to the flow rate of the refrigerant flowing into the input port IN can be adjusted.
  • the ratio R is the ratio of the flow rate of the refrigerant flowing through the outdoor heat exchanger 20 to the flow rate of the refrigerant discharged from the compressor 10.
  • the ratio R is 1.
  • the ratio R becomes zero. In this way, the ratio R can be set by adjusting the angle ⁇ of the valve body 61 between the state shown in FIG. 2B and the state shown in FIG.
  • the three-way valve 60 corresponds to a “flow regulator” according to the present invention.
  • the configuration of the “flow regulator” according to the present invention is not limited to the three-way valve 60.
  • one two-way valve may be provided in each of the flow path between the compressor 10 and the outdoor heat exchanger 20 and the bypass flow path 120.
  • the ratio R can be set by adjusting the opening degree of each of the two two-way valves.
  • valve element 61 is set at a desired angle ⁇ without depending on the refrigerant temperature. Can be controlled. Therefore, the degree of freedom when setting the ratio R can be improved.
  • the condensation pressure saturated pressure when the refrigerant condenses in the outdoor heat exchanger 20
  • various problems may occur.
  • frost may adhere to the indoor heat exchanger 40 or excessive dehumidification may occur. Therefore, it is desirable to maintain the condensation pressure Pc within an appropriate range.
  • the angle ⁇ of the valve body 61 of the three-way valve 60 is adjusted as compared with the case where the condensation pressure Pc is greater than or equal to the reference value P1.
  • the ratio R can be reduced.
  • the flow rate of the refrigerant flowing through the outdoor heat exchanger 20 is decreased by the increase in the flow rate of the refrigerant bypassing the outdoor heat exchanger 20.
  • the amount of heat released from the refrigerant to the outside air in the outdoor heat exchanger 20 is reduced, and the amount of decrease in the refrigerant temperature in the outdoor heat exchanger 20 is reduced. If it does so, there exists a possibility that required supercooling degree (DELTA) Tc cannot be ensured about the refrigerant
  • DELTA required supercooling degree
  • the fact that the degree of supercooling ⁇ Tc cannot be secured means that the refrigerant flowing into the main expansion valve 30 is in a gas-liquid two-phase state.
  • the average density of the gas-liquid two-phase refrigerant is lower than the average density of the liquid refrigerant. Therefore, in order to realize a desired required cooling capacity for the gas-liquid two-phase refrigerant, it is required to increase the flow rate of the refrigerant flowing through the main expansion valve 30 as compared with the case of the liquid refrigerant. That is, it is necessary to increase the opening of the main expansion valve 30. However, generally, there is an upper limit for the opening of the expansion valve.
  • the opening degree of the main expansion valve 30 reaches the upper limit value, the flow rate of the refrigerant flowing through the main expansion valve 30 cannot be increased further. Therefore, in the case of a gas-liquid two-phase refrigerant, there is a possibility that a desired refrigerant capacity cannot be realized.
  • the opening degree of the sub expansion valve 70 is further controlled. More specifically, the refrigerant flowing through the internal heat exchanger 80 when the degree of supercooling ⁇ Tc of the refrigerant flowing into the main expansion valve 30 is lower than the reference value T2 compared to when the degree of supercooling ⁇ Tc is higher than the reference value T2.
  • the degree of opening of the secondary expansion valve 70 is controlled so that the flow rate of As a result, the amount of heat exchanged between the refrigerant flowing through the main circuit 110 and the refrigerant cooled by the pressure reduction at the sub expansion valve 70 increases.
  • the degree of supercooling of the refrigerant flowing into the main expansion valve 30 is more reliably ensured. can do.
  • FIG. 3 is a flowchart for explaining the control of the three-way valve 60 and the sub-expansion valve 70 executed in the first embodiment.
  • Each step (hereinafter abbreviated as S) of the flowchart shown in FIG. 3 and FIG. 7 described later is called from the main routine and executed every time a predetermined time elapses or a predetermined condition is satisfied.
  • control device 500 calculates condensing pressure Pc in outdoor heat exchanger 20.
  • the condensation pressure Pc can be calculated based on the discharge pressure detected by the pressure sensor 91.
  • the control device 500 determines whether or not the condensation pressure Pc is less than the reference value P1 (first reference value).
  • the reference value P1 is a value that does not cause the above-described problems such as frost adhesion in the indoor heat exchanger 40, and a value that does not cause an abnormality of the compressor 10 due to an excessively high compression ratio of the compressor 10. It is preferable that
  • the control device 500 adjusts the angle ⁇ of the valve body 61 of the three-way valve 60 in a direction in which the flow rate of the refrigerant toward the bypass flow path 120 increases.
  • the angle ⁇ is increased.
  • the ratio R of the flow rate of the refrigerant flowing through the outdoor heat exchanger 20 with respect to the flow rate of the refrigerant discharged from the compressor 10 decreases (S30).
  • the ratio of the refrigerant that reaches the downstream side of the outdoor heat exchanger 20 without passing through the outdoor heat exchanger 20 via the bypass flow path 120 increases, so that the condensation pressure Pc increases. Therefore, the fall of the condensation pressure Pc can be suppressed.
  • control device 500 calculates the degree of supercooling ⁇ Tc of the refrigerant flowing into main expansion valve 30.
  • the degree of supercooling ⁇ Tc can be calculated based on the refrigerant temperature detected by the temperature sensor 93 (the refrigerant temperature at the outlet side of the internal heat exchanger 80).
  • control device 500 determines whether or not the degree of supercooling ⁇ Tc is less than the reference value T2 (second reference value). A method for setting the reference value T2 will be described later.
  • control device 500 advances the process to S70.
  • the control device 500 maintains the opening degree of the sub-expansion valve 70 (or sets the opening degree of the sub-expansion valve 70), assuming that a sufficient degree of subcooling ⁇ Tc is secured for the refrigerant flowing into the main expansion valve 30. Adjust it small).
  • control device 500 advances the process to S60.
  • control device 500 sets the opening degree of sub expansion valve 70 to be larger than that in the case where degree of supercooling ⁇ Tc is equal to or greater than reference value T2.
  • the flow rate of the refrigerant flowing through the branch flow path 130 increases, so that in the internal heat exchanger 80, the refrigerant is exchanged between the refrigerant flowing through the main circuit 110 and the refrigerant cooled by the decompression at the sub expansion valve 70. The amount of heat to be increased.
  • the temperature of the refrigerant flowing through the main circuit 110 is lower than when the degree of supercooling ⁇ Tc is greater than or equal to the reference value T2. Therefore, the degree of supercooling ⁇ Tc of the refrigerant flowing into the main expansion valve 30 can be ensured more reliably.
  • the control device 500 When the condensing pressure Pc is equal to or higher than the reference value P1 in S20 (YES in S20), the control device 500 maintains the angle ⁇ of the valve body 61 of the three-way valve 60, or refrigerant that goes to the bypass flow path 120.
  • the angle ⁇ is adjusted in the direction in which the flow rate decreases (S80). In the example shown in FIG. 2, the angle ⁇ is maintained or decreased. Thereby, the ratio R is maintained or increased.
  • the process returns to the main routine.
  • the “state value” is not limited to this.
  • the “state value” according to the present invention may be the temperature of the refrigerant flowing through the outdoor heat exchanger 20, the discharge temperature of the refrigerant from the compressor 10, or the compression ratio of the refrigerant in the compressor 10.
  • FIG. 4 is a Ph diagram corresponding to the control of the three-way valve 60 and the sub-expansion valve 70 executed in the first embodiment.
  • the horizontal axis represents specific enthalpy h [unit: kJ / kg], and the vertical axis represents pressure P [unit: MPa].
  • point A indicates the state of a low-pressure gas refrigerant (superheated steam).
  • the process from the point A to the point B is an adiabatic compression process by the compressor 10.
  • Point B indicates a state in which the refrigerant is compressed by the compressor 10.
  • Point D shows a state where the refrigerant bypassing the outdoor heat exchanger 20 via the three-way valve 60 and the refrigerant condensed by the outdoor heat exchanger 20 are mixed.
  • Point E indicates a supercooled state obtained by condensing the refrigerant in the outdoor heat exchanger 20 and further cooling a part of the refrigerant by the internal heat exchanger 80.
  • the process from the point E to the point F shows the expansion process of the refrigerant by the main expansion valve 30.
  • the process from the point E to the point G shows the expansion process of the refrigerant by the sub expansion valve 70.
  • a process from the point E to the point G and a process from the point F to the point G indicate an evaporation process in the indoor heat exchanger 40.
  • the necessary supercooling degree ⁇ Tc is secured for the refrigerant flowing into the main expansion valve 30 while suppressing the decrease in the condensation pressure Pc. can do.
  • the compression ratio of the compressor 10 can also be lower than an appropriate value.
  • the compression ratio is less than the appropriate value, it is conceivable to increase the compression ratio by increasing the drive frequency of the compressor 10 as compared with the case where the compression ratio is equal to or more than the appropriate value.
  • the driving frequency of the compressor 10 increases, the refrigerant evaporation pressure Pe decreases.
  • the compressor is stopped. Thereafter, when the evaporation pressure recovers until it exceeds the reference value, the compressor is driven again. That is, the compressor may be repeatedly driven and stopped (such control is also referred to as “low pressure cut control”).
  • the compressor 10 is prevented from reaching a state where the low pressure cut control of the compressor 10 is executed. In other words, the operation of the compressor 10 can be stabilized.
  • the reference value T2 (see S50 in FIG. 3) set for the degree of supercooling ⁇ Tc of the refrigerant flowing into the main expansion valve 30 will be described.
  • the reference value T2 is set so that the supercooled state of the refrigerant flowing into the main expansion valve 30 is maintained even when the indoor heat exchanger 40 is provided at a position higher than the outdoor heat exchanger 20. It is preferable. The reason for this will be described below.
  • the indoor heat exchanger 40 is provided at a position lower than the outdoor heat exchanger 20, the refrigerant flow from the outdoor heat exchanger 20 toward the indoor heat exchanger 40 is a downward flow. Therefore, the pressure generated by the gravity applied to the liquid refrigerant is applied to the main expansion valve 30. Therefore, even if the pressure of the liquid refrigerant decreases due to expansion at the main expansion valve 30, boiling of the refrigerant is suppressed.
  • the refrigerant flow from the outdoor heat exchanger 20 toward the indoor heat exchanger 40 is an upward flow. Therefore, unlike the case where the refrigerant flow is a downward flow, pressure due to gravity is not applied to the main expansion valve 30. Therefore, when the pressure of the liquid refrigerant is reduced at the main expansion valve 30, the refrigerant is boiled under reduced pressure, and the refrigerant may be in a gas-liquid two-phase state. Then, the above-mentioned abnormal noise or vibration problem may occur. Therefore, it is preferable to set the reference value T2 so that the supercooled state of the refrigerant is maintained even after the pressure is reduced in the main expansion valve 30.
  • FIG. 5 is a block diagram schematically showing a configuration of a heat pump system according to a modification of the first embodiment.
  • the configurations of outdoor heat exchanger 20A, bypass flow path 120A, and three-way valve 60A correspond to the configurations in heat pump system 100 according to Embodiment 1 (see FIG. 1). Different.
  • connection part C1 is provided in the predetermined position along the flow path provided in the outdoor heat exchanger 20A.
  • the outdoor heat exchanger 20A is divided into two parts: a heat exchanger upstream of the connection part C1 and a heat exchanger downstream of the connection part C1.
  • the input port IN of the three-way valve 60A is connected to the connection portion C1 by a bypass flow path 120A.
  • a part of the refrigerant flowing through the outdoor heat exchanger 20 passes through only the upstream heat exchanger and flows out of the outdoor heat exchanger 20 through the connection portion C1.
  • the remaining refrigerant flows through both the upstream heat exchanger and the downstream heat exchanger.
  • the output port OUT1 of the three-way valve 60A is connected to the connection part C2 on the downstream side of the connection part C1 with respect to the flow path inside the outdoor heat exchanger 60 by the bypass flow path 120A.
  • the output port OUT2 of the three-way valve 60A is connected to the main circuit 110 on the downstream side of the outdoor heat exchanger 20A by a bypass passage 120A.
  • the configuration of heat pump system 100A other than outdoor heat exchanger 20A, bypass flow path 120A, and three-way valve 60A is the same as the corresponding configuration of heat pump system 100 according to Embodiment 1, and therefore detailed description will be repeated. Absent.
  • the opening degree of the three-way valve is adjusted by the angle of the valve body.
  • the structure which can change every 1 degree in the movable range whose angle of a valve body is 0 degree or more and 90 degrees or less is assumed.
  • the flow rate ratio that can be adjusted by the three-way valve (the ratio between the flow rate of the refrigerant flowing out from the first output port and the flow rate of the refrigerant flowing out from the second output port) can be adjusted only in 90 ways. . Therefore, when the flow rate of the refrigerant flowing into the input port of the three-way valve is relatively large, it is difficult to finely adjust the flow rate between the first output port and the second output port.
  • the outdoor heat exchanger 20 is not divided
  • FIG. Therefore, in Embodiment 1, the cost of the outdoor heat exchanger 20 can be reduced.
  • a throttle capillary tube that causes a pressure loss equivalent to that of the outdoor heat exchanger 20A may be provided in the bypass flow path 120A on the downstream side of the three-way valve 60A. Due to the pressure loss caused by the throttle, it is possible to prevent the refrigerant flow rate from rapidly increasing when the opening of the three-way valve 60A is increased from the closed state.
  • two two-way valves may be provided instead of the three-way valve 60A.
  • one is provided between the output port OUT1 of the three-way valve 60A in FIG. 5 and the connection C2.
  • the other is provided between the output port OUT2 of the three-way valve 60A and the downstream side of the main circuit 110 with respect to the outdoor heat exchanger 20A.
  • the two-way valve is generally suitable for fine adjustment of the flow rate because the adjustment range of the opening is smaller than that of the three-way valve.
  • FIG. 6 is a block diagram schematically showing the configuration of the heat pump system according to the second embodiment.
  • outdoor heat exchanger 20 functions as an evaporator
  • indoor heat exchanger 40 functions as a condenser.
  • the configuration of the branch flow path 130B, the sub expansion valve 70B, and the internal heat exchanger 80B is different from the corresponding configuration in the heat pump system 100 according to Embodiment 1 (see FIG. 1).
  • the heat pump system 100B is different from the heat pump system 100 shown in FIG. 1 in that a temperature sensor 94 is provided instead of the temperature sensor 92.
  • the branch flow path 130B is connected between the compressor 10 and the indoor heat exchanger 40.
  • the branch flow path 130 ⁇ / b> B is configured to be able to branch a part of the refrigerant flowing through the main circuit 110.
  • the sub expansion valve 70B is provided in the branch flow path 130B.
  • the internal heat exchanger 80B exchanges heat between the refrigerant flowing through the branch flow path 130B and the refrigerant flowing from the outdoor heat exchanger 20 to the compressor 10.
  • the refrigerant that has undergone heat exchange in the internal heat exchanger 80B joins the refrigerant that flows from the outdoor heat exchanger 20 to the compressor 10.
  • the temperature sensor 94 is provided in the indoor heat exchanger 40.
  • the temperature sensor 94 detects the evaporation temperature of the refrigerant in the indoor heat exchanger 40 and outputs a signal indicating the detection result to the control device 500.
  • structures other than the branch flow path 130B, the sub-expansion valve 70B, and the internal heat exchanger 80B of the heat pump system 100B are the same as the corresponding structures of the heat pump system 100 according to the first embodiment, a detailed description will be given. Do not repeat.
  • the evaporation pressure Pe of the refrigerant in the outdoor heat exchanger 20 can be higher than when the outside air temperature is low. That is, since the pressure of the refrigerant flowing out of the outdoor heat exchanger 20 and flowing into the compressor 10 becomes high, the pressure of the refrigerant sucked into the compressor 10 and the pressure of the refrigerant discharged from the compressor 10 The difference can be small. That is, the compression ratio in the compressor 10 may be lowered.
  • Embodiment 2 all or part of the refrigerant flowing to the outdoor heat exchanger 20 is bypassed through the bypass flow path 120.
  • the refrigerant sucked into the compressor 10 is a gas refrigerant (gas phase single-phase refrigerant).
  • the refrigerant that has passed through the main expansion valve 30 during the heating operation of the heat pump system 100B is a gas-liquid two-phase refrigerant. For this reason, the refrigerant from the main expansion valve 30 to the three-way valve 60 and further sucked into the compressor 10 may be in a gas-liquid two-phase state.
  • the opening of the sub expansion valve 70 is set larger than in the case where the refrigerant temperature is lower than the reference value T3.
  • the flow rate of the refrigerant flowing through the heat exchanger 80 is increased.
  • the amount of heat exchanged between the high-temperature and high-pressure gas refrigerant discharged from the compressor 10 and the gas-liquid two-phase refrigerant from the main expansion valve 30 to the three-way valve 60 increases.
  • the refrigerant sucked into the compressor 10 can be completely gasified, so that an abnormality of the compressor 10 can be prevented.
  • FIG. 7 is a flowchart for explaining the control of the three-way valve 60 and the sub-expansion valve 70 executed in the second embodiment.
  • control device 500 calculates evaporation pressure Pe in outdoor heat exchanger 20.
  • the evaporation pressure Pe can be calculated, for example, by converting the evaporation temperature of the indoor heat exchanger 40 detected using the temperature sensor 94 into a saturation pressure.
  • the control device 500 determines whether or not the evaporation pressure Pe is equal to or higher than the reference value P3 (third reference value).
  • the control device 500 adjusts the angle ⁇ of the valve body 61 of the three-way valve 60 in the direction in which the flow rate of the refrigerant toward the bypass flow path 120 increases. Thereby, the ratio R of the flow rate of the refrigerant flowing through the outdoor heat exchanger 20 with respect to the flow rate of the refrigerant discharged from the compressor 10 decreases (S130).
  • control device 500 calculates the heating degree ⁇ Te of the refrigerant flowing into compressor 10.
  • the degree of heating ⁇ Te can be calculated based on, for example, the refrigerant temperature detected by the temperature sensor 93 (the refrigerant temperature at the outlet side of the internal heat exchanger 80).
  • control device 500 determines whether or not the degree of superheat ⁇ Te is less than the reference value T4 (fourth reference value).
  • control device 500 maintains the opening degree of sub-expansion valve 70B on the assumption that sufficient superheat degree ⁇ Te is secured for the refrigerant flowing into compressor 10. (Or adjust the opening of the sub-expansion valve 70B to be small) (S170).
  • control device 500 sets the opening degree of sub-expansion valve 70B to be larger than when superheat degree ⁇ Te is greater than or equal to reference value T4. (S160).
  • the flow rate of the refrigerant flowing through the branch flow path 130B is increased, so that the internal heat exchanger 80B and the high-temperature and high-pressure gas refrigerant discharged from the compressor 10 and the gas from the main expansion valve 30 to the three-way valve 60 are discharged.
  • the amount of heat exchanged with the liquid two-phase refrigerant increases.
  • the temperature of the refrigerant flowing through the main circuit 110 is higher than when the degree of superheat ⁇ Te is equal to or greater than the reference value T4. Therefore, the degree of superheat ⁇ Te of the refrigerant flowing into the compressor 10 can be ensured more reliably.
  • Embodiment 2 may combine Embodiment 2 and a modification. That is, also in Embodiment 2, it is possible to adopt a configuration in which only a part of the refrigerant can bypass the outdoor heat exchanger.
  • the heat pump system including only one indoor unit has been described.
  • the present invention is also applicable to a heat pump system including a plurality of indoor units (that is, a plurality of main expansion valves and a plurality of indoor heat exchangers).
  • An example of such a system is a multi air conditioning system for buildings.
  • the present invention can also be applied to a case where mixed operation of cooling and heating is performed in the cooling main operation of the multi-air conditioning system.
  • the difference between the heat absorption amount in the cooling indoor unit and the heat dissipation amount in the heating indoor unit becomes smaller, the heat dissipation amount in the outdoor heat exchanger becomes smaller. Therefore, in some cases, it is desirable to reduce the heat radiation amount in the outdoor heat exchanger as compared with the cooling operation of the system including only one indoor unit.
  • the present invention can be applied to such a case.
  • 100, 100A, 100B heat pump system 10 compressor, 20, 20A outdoor heat exchanger, 30 main expansion valve, 40 indoor heat exchanger, 50 piping, 60, 60A three-way valve, 61 valve body, 70 sub expansion valve, 80 , 80B internal heat exchanger, 91 pressure sensor, 92-94 temperature sensor, 110 main circuit, 120, 120A bypass flow path, 130, 130B branch flow path, 500 control device.

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Abstract

In the present invention, a main circuit (110) includes a compressor (10), an outdoor heat exchanger (20), a main expansion valve (30), and an indoor heat exchanger (40), and is configured so as to allow a refrigerant to circulate therethrough. A bypass flow path (120) is configured so as to cause the refrigerant flowing through the main circuit (110) to bypass the outdoor heat exchanger (20). A three-way valve (60) is configured so as to be capable of adjusting the ratio of the flow rate of the refrigerant flowing through the outdoor heat exchanger (20) to the flow rate of the refrigerant discharged from the compressor (10). A branch flow path (130) is connected between the outdoor heat exchanger (20) and the main expansion valve (30) and is configured so as to be capable of partially branching the refrigerant flowing through the main circuit (110). A sub-expansion valve (70) is configured so as to be capable of adjusting the flow rate of the refrigerant flowing through the branch flow path (130). An internal heat exchanger (80) performs heat exchange between the refrigerant flowing through the branch flow path (130) and the refrigerant flowing from the outdoor heat exchanger (20) to the main expansion valve (30).

Description

ヒートポンプシステムHeat pump system
　本発明はヒートポンプシステムに関し、より特定的には、冷媒の少なくとも一部が室外熱交換器を迂回するように構成されたバイパス流路を備えたヒートポンプシステムに関する。 The present invention relates to a heat pump system, and more particularly to a heat pump system including a bypass flow path configured such that at least a part of a refrigerant bypasses an outdoor heat exchanger.
　ヒートポンプシステムの冷房運転時に室外熱交換器（凝縮器）における冷媒の凝縮圧力が過度に低い場合、室内熱交換器（蒸発器）に霜が付着したり、過剰な除湿が行なわれたりするなどの不具合が生じ得る。そのため、凝縮圧力の低下を抑制するための技術が提案されている。 If the condensation pressure of the refrigerant in the outdoor heat exchanger (condenser) is excessively low during the cooling operation of the heat pump system, frost adheres to the indoor heat exchanger (evaporator) or excessive dehumidification occurs. Malfunctions can occur. Therefore, a technique for suppressing a decrease in the condensation pressure has been proposed.
　たとえば特開２０００－５５４４４号公報（特許文献１）は、凝縮器を通る冷媒の流量と、凝集器をバイパスする冷媒の流量との比率を調整する三方弁を備えた空気調和器を開示する。この空気調和器の制御手段は、所定条件の成立時に（具体的には膨張弁が全開のとき、圧縮機の圧縮比が許容最低圧縮比以下のとき、または再熱器の再熱量が所定値以下のときに）、凝縮圧力を高める方向に三方弁を制御する。 For example, Japanese Patent Application Laid-Open No. 2000-55444 (Patent Document 1) discloses an air conditioner including a three-way valve that adjusts a ratio between a flow rate of refrigerant passing through a condenser and a flow rate of refrigerant bypassing the aggregator. This air conditioner control means is used when a predetermined condition is satisfied (specifically, when the expansion valve is fully opened, when the compression ratio of the compressor is less than the allowable minimum compression ratio, or when the reheat amount of the reheater is a predetermined value). In the following case, the three-way valve is controlled in the direction of increasing the condensation pressure.
特開２０００－５５４４４号公報JP 2000-55444 A
　たとえば特許文献１に開示された空気調和器が冷房運転さえる場合、上記所定条件の成立時には、所定条件の不成立時と比べて、三方弁の開度を変化させることにより、凝縮器を迂回する冷媒の流量が大きく設定される。これにより、凝縮圧力の低下を抑制することができる。 For example, when the air conditioner disclosed in Patent Document 1 is in cooling operation, when the predetermined condition is satisfied, the refrigerant bypasses the condenser by changing the opening of the three-way valve as compared to when the predetermined condition is not satisfied. The flow rate is set large. Thereby, the fall of a condensation pressure can be suppressed.
　その一方で、凝縮器を迂回する冷媒の流量が増加した分だけ、凝縮器を流れる冷媒の流量が減少する。そのため、凝縮器における冷媒から外気への放熱量が小さくなるので、凝縮器における冷媒温度の低下量が小さくなり得る。そうすると、凝縮器から流出して膨張弁に流入する冷媒について、必要な過冷却度が確保できない可能性がある。 On the other hand, the flow rate of the refrigerant flowing through the condenser decreases as the flow rate of the refrigerant bypassing the condenser increases. For this reason, the amount of heat released from the refrigerant to the outside air in the condenser is reduced, so that the amount of decrease in the refrigerant temperature in the condenser can be reduced. If it does so, there exists a possibility that required supercooling degree cannot be ensured about the refrigerant | coolant which flows out from a condenser and flows in into an expansion valve.
　本発明は上記課題を解決するためになされたものであり、その目的は、ヒートポンプシステムの冷房運転時において、凝縮圧力の低下を抑制しつつ、膨張弁に流入する冷媒の過冷却度を確保可能な技術を提供することである。 The present invention has been made to solve the above-described problems, and its purpose is to ensure a degree of supercooling of the refrigerant flowing into the expansion valve while suppressing a decrease in condensation pressure during cooling operation of the heat pump system. Is to provide new technology.
　また、ヒートポンプシステムの暖房運転時にも同様の課題が起こり得る。すなわち、たとえば外気温が高い場合に暖房運転を行なうと、外気温が低い場合と比べて、室外熱交換器（蒸発器）における冷媒の蒸発圧力が高くなる。つまり、室外熱交換器から流出して圧縮機へと流入する冷媒の圧力が高くなるので、圧縮機における圧縮比が低くなる。この場合には、三方弁の開度変化により、蒸発器を迂回する冷媒の流量を大きく設定ことによって、蒸発圧力の上昇を抑制することができる。 Also, the same problem can occur during the heating operation of the heat pump system. That is, for example, when the heating operation is performed when the outside air temperature is high, the evaporation pressure of the refrigerant in the outdoor heat exchanger (evaporator) becomes higher than when the outside air temperature is low. That is, since the pressure of the refrigerant flowing out from the outdoor heat exchanger and flowing into the compressor increases, the compression ratio in the compressor decreases. In this case, by increasing the flow rate of the refrigerant that bypasses the evaporator by changing the opening of the three-way valve, an increase in the evaporation pressure can be suppressed.
　その一方で、蒸発器を迂回する冷媒の流量が増加した分だけ、蒸発器を流れる冷媒の流量が減少する。そのため、蒸発器における外気から冷媒への吸熱量が小さくなるので、蒸発器における冷媒温度の上昇量が小さくなり得る。そうすると、蒸発器から流出して圧縮機に流入する冷媒について、必要な過熱度が確保できない可能性がある。 On the other hand, the flow rate of the refrigerant flowing through the evaporator is reduced by the increase in the flow rate of the refrigerant bypassing the evaporator. For this reason, the amount of heat absorbed from the outside air to the refrigerant in the evaporator becomes small, so that the amount of increase in the refrigerant temperature in the evaporator can be small. If it does so, there exists a possibility that required superheat degree cannot be ensured about the refrigerant | coolant which flows out from an evaporator and flows in into a compressor.
　本発明の他の目的は、ヒートポンプシステムの暖房運転時において、蒸発圧力の上昇を抑制しつつ、圧縮機に流入する冷媒の過熱度を確保可能な技術を提供することである。 Another object of the present invention is to provide a technology capable of ensuring the degree of superheat of the refrigerant flowing into the compressor while suppressing an increase in evaporation pressure during heating operation of the heat pump system.
　本発明のある局面に従うヒートポンプシステムは、主回路と、バイパス流路と、流量調整器と、分岐流路と、第２の膨張弁と、副熱交換器とを備える。主回路は、圧縮機、室外熱交換器、第１の膨張弁および室内熱交換器を含み、冷媒を循環可能に構成される。バイパス流路は、主回路を流れる冷媒が室外熱交換器を迂回するように構成される。流量調整器は、圧縮機から吐出される冷媒の流量に対する室外熱交換器を流れる冷媒の流量の比率を調整可能に構成される。分岐流路は、室外熱交換器と第１の膨張弁との間に接続され、主回路を流れる冷媒の一部を分岐可能に構成される。第２の膨張弁は、分岐流路を流れる冷媒の流量を調整可能に構成される。副熱交換器は、分岐流路を流れる冷媒と、室外熱交換器から第１の膨張弁に流れる冷媒との間で熱交換する。 A heat pump system according to an aspect of the present invention includes a main circuit, a bypass flow path, a flow rate regulator, a branch flow path, a second expansion valve, and a sub heat exchanger. The main circuit includes a compressor, an outdoor heat exchanger, a first expansion valve, and an indoor heat exchanger, and is configured to be able to circulate the refrigerant. The bypass flow path is configured such that the refrigerant flowing through the main circuit bypasses the outdoor heat exchanger. The flow rate adjuster is configured to be able to adjust the ratio of the flow rate of the refrigerant flowing through the outdoor heat exchanger to the flow rate of the refrigerant discharged from the compressor. The branch flow path is connected between the outdoor heat exchanger and the first expansion valve, and is configured to be able to branch a part of the refrigerant flowing through the main circuit. The second expansion valve is configured to be able to adjust the flow rate of the refrigerant flowing through the branch flow path. The auxiliary heat exchanger exchanges heat between the refrigerant flowing through the branch flow path and the refrigerant flowing from the outdoor heat exchanger to the first expansion valve.
　本発明の他の局面に従うヒートポンプシステムは、主回路と、バイパス流路と、流量調整器と、分岐流路と、第２の膨張弁と、副熱交換器とを備える。主回路は、圧縮機、室外熱交換器、第１の膨張弁および室内熱交換器を含み、冷媒を循環可能に構成される。バイパス流路は、主回路を流れる冷媒が室外熱交換器を迂回するように構成される。流量調整器は、圧縮機から吐出される冷媒の流量に対する室外熱交換器を流れる冷媒の流量の比率を調整可能に構成される。分岐流路は、圧縮機と室内熱交換器との間に接続され、主回路を流れる冷媒の一部を分岐可能に構成される。第２の膨張弁は、分岐流路を流れる冷媒の流量を調整可能に構成される。副熱交換器は、分岐流路を流れる冷媒と、室外熱交換器から圧縮機へと流れる冷媒との間で熱交換する。 A heat pump system according to another aspect of the present invention includes a main circuit, a bypass flow channel, a flow rate regulator, a branch flow channel, a second expansion valve, and a sub heat exchanger. The main circuit includes a compressor, an outdoor heat exchanger, a first expansion valve, and an indoor heat exchanger, and is configured to be able to circulate the refrigerant. The bypass flow path is configured such that the refrigerant flowing through the main circuit bypasses the outdoor heat exchanger. The flow rate adjuster is configured to be able to adjust the ratio of the flow rate of the refrigerant flowing through the outdoor heat exchanger to the flow rate of the refrigerant discharged from the compressor. The branch flow path is connected between the compressor and the indoor heat exchanger, and is configured to be able to branch a part of the refrigerant flowing through the main circuit. The second expansion valve is configured to be able to adjust the flow rate of the refrigerant flowing through the branch flow path. The auxiliary heat exchanger exchanges heat between the refrigerant flowing through the branch flow path and the refrigerant flowing from the outdoor heat exchanger to the compressor.
　上記構成によれば、凝縮圧力が過度に低い場合には、凝縮圧力が適切な範囲内の場合と比べて、流量調整器により上記比率を低下させることができる。これにより、バイパス流路を介して室外熱交換器を迂回する冷媒の流量が増加するので、凝縮圧力の低下を抑制することができる。さらに、必要な過冷却度が確保できていない場合には、第２の膨張弁を介して分岐流路を流れる冷媒の流量を増加させることができる。これにより、副熱交換器において、主回路を流れる冷媒と、第２の膨張弁により減圧された冷媒との間で交換される熱量が大きくなる。その結果、主回路を流れる冷媒の温度が低下するので、膨張弁に流入する冷媒の過冷却度をより確実に確保することができる。よって、ヒートポンプシステムの冷房運転時において、凝縮圧力の低下を抑制しつつ、過冷却度を確保することができる。 According to the above configuration, when the condensing pressure is excessively low, the ratio can be reduced by the flow rate regulator as compared with the case where the condensing pressure is within an appropriate range. Thereby, since the flow volume of the refrigerant | coolant which bypasses an outdoor heat exchanger via a bypass flow path increases, the fall of a condensation pressure can be suppressed. Furthermore, when the necessary degree of supercooling cannot be ensured, the flow rate of the refrigerant flowing through the branch flow path via the second expansion valve can be increased. Thereby, in the auxiliary heat exchanger, the amount of heat exchanged between the refrigerant flowing through the main circuit and the refrigerant decompressed by the second expansion valve is increased. As a result, the temperature of the refrigerant flowing through the main circuit decreases, so that the degree of supercooling of the refrigerant flowing into the expansion valve can be ensured more reliably. Therefore, during the cooling operation of the heat pump system, it is possible to ensure the degree of supercooling while suppressing a decrease in the condensation pressure.
　また、詳細は後述するが、ヒートポンプシステムの暖房運転時においても上記と同様に、蒸発圧力の上昇を抑制しつつ、圧縮機に流入する冷媒の過熱度を確保することができる。 As will be described in detail later, the degree of superheat of the refrigerant flowing into the compressor can be ensured while suppressing the increase in the evaporation pressure even during the heating operation of the heat pump system.
実施の形態１に係るヒートポンプシステムの構成を概略的に示すブロック図である。1 is a block diagram schematically showing a configuration of a heat pump system according to Embodiment 1. FIG. 三方弁の構成を模式的に示す図である。It is a figure which shows the structure of a three-way valve typically. 実施の形態１にて実行される三方弁および副膨張弁の制御を説明するためのフローチャートである。3 is a flowchart for illustrating control of a three-way valve and a sub-expansion valve that is executed in the first embodiment. 実施の形態１にて実行される三方弁および副膨張弁の制御に対応するＰｈ線図である。FIG. 6 is a Ph diagram corresponding to the control of the three-way valve and the sub-expansion valve executed in the first embodiment. 実施の形態１の変形例に係るヒートポンプシステムの構成を概略的に示すブロック図である。It is a block diagram which shows schematically the structure of the heat pump system which concerns on the modification of Embodiment 1. FIG. 実施の形態２に係るヒートポンプシステムの構成を概略的に示すブロック図である。It is a block diagram which shows schematically the structure of the heat pump system which concerns on Embodiment 2. FIG. 実施の形態２にて実行される三方弁および副膨張弁の制御を説明するためのフローチャートである。6 is a flowchart for illustrating control of a three-way valve and a sub-expansion valve that is executed in a second embodiment.
　以下、本発明の実施の形態について、図面を参照しながら詳細に説明する。なお、図中同一または相当部分には同一符号を付してその説明は繰り返さない。 Hereinafter, embodiments of the present invention will be described in detail with reference to the drawings. In the drawings, the same or corresponding parts are denoted by the same reference numerals and description thereof will not be repeated.
　　［実施の形態１］
　図１は、実施の形態１に係るヒートポンプシステムの構成を概略的に示すブロック図である。図１を参照して、ヒートポンプシステム１００は、主回路１１０と、制御装置５００とを備える。主回路１１０は、圧縮機１０と、室外熱交換器２０と、主膨張弁３０と、室内熱交換器４０と、配管５０とを含む。ヒートポンプシステム１００の冷房運転時における冷媒の流れを矢印ＲＥＦで示す。 [Embodiment 1]
FIG. 1 is a block diagram schematically showing the configuration of the heat pump system according to the first embodiment. Referring to FIG. 1, heat pump system 100 includes a main circuit 110 and a control device 500. Main circuit 110 includes compressor 10, outdoor heat exchanger 20, main expansion valve 30, indoor heat exchanger 40, and pipe 50. The refrigerant flow during the cooling operation of the heat pump system 100 is indicated by an arrow REF.
　圧縮機１０は、たとえばインバータ（図示せず）によって駆動される容量可変型の圧縮機である。圧縮機１０により圧縮されて高温高圧となったガス冷媒は、室外熱交換器２０に流入する。 The compressor 10 is a variable capacity compressor driven by, for example, an inverter (not shown). The gas refrigerant that has been compressed by the compressor 10 to a high temperature and high pressure flows into the outdoor heat exchanger 20.
　室外熱交換器２０は、たとえば伝熱管および放熱フィン（図示せず）を含んで構成された熱交換器である。ヒートポンプシステム１００の冷房運転時には、室外熱交換器２０は凝縮器として機能する。室外熱交換器２０において、ガス冷媒は、外気に放熱することにより凝縮して液冷媒となる。 The outdoor heat exchanger 20 is a heat exchanger configured to include, for example, heat transfer tubes and heat radiating fins (not shown). During the cooling operation of the heat pump system 100, the outdoor heat exchanger 20 functions as a condenser. In the outdoor heat exchanger 20, the gas refrigerant condenses into a liquid refrigerant by dissipating heat to the outside air.
　主膨張弁（第１の膨張弁）３０は、たとえばステッピングモータ（図示せず）により開度が制御可能な絞り弁である。主膨張弁３０は冷媒の流量（単位時間当たりに冷媒が流れる量）を調整するために用いられる。また、主膨張弁３０は、室外熱交換器２０により凝縮された高圧の液冷媒を膨張させることにより、液冷媒を減圧する。これにより、冷媒は気液二相冷媒となり、室内熱交換器４０に流入する。 The main expansion valve (first expansion valve) 30 is a throttle valve whose opening degree can be controlled by, for example, a stepping motor (not shown). The main expansion valve 30 is used for adjusting the flow rate of refrigerant (amount of refrigerant flowing per unit time). The main expansion valve 30 decompresses the liquid refrigerant by expanding the high-pressure liquid refrigerant condensed by the outdoor heat exchanger 20. As a result, the refrigerant becomes a gas-liquid two-phase refrigerant and flows into the indoor heat exchanger 40.
　室内熱交換器４０は、室外熱交換器２０と同様に、伝熱管および放熱フィン（図示せず）を含んで構成された熱交換器である。ヒートポンプシステム１００の冷房運転時には、室内熱交換器４０は蒸発器として機能する。室内熱交換器４０において、空気は冷媒によって冷却される。一方、冷媒は温められて気液二相冷媒から低圧のガス冷媒となる。その後、ガス冷媒は圧縮機１０に戻り、圧縮機１０により再び圧縮されて吐出される。 The indoor heat exchanger 40 is a heat exchanger configured to include heat transfer tubes and radiating fins (not shown), like the outdoor heat exchanger 20. During the cooling operation of the heat pump system 100, the indoor heat exchanger 40 functions as an evaporator. In the indoor heat exchanger 40, the air is cooled by the refrigerant. On the other hand, the refrigerant is heated to change from a gas-liquid two-phase refrigerant to a low-pressure gas refrigerant. Thereafter, the gas refrigerant returns to the compressor 10 and is compressed again by the compressor 10 and discharged.
　このように、主回路１１０は、圧縮機１０と室外熱交換器２０と主膨張弁３０と室内熱交換器４０とが、この順に配管５０により接続されることによって形成されている。なお、主回路１１０は、いずれも図示しないが、アキュムレータ（低圧受液器）、レシーバ（高圧受液器）、またはオイルセパレータ（油分離器）をさらに含んでもよい。 Thus, the main circuit 110 is formed by connecting the compressor 10, the outdoor heat exchanger 20, the main expansion valve 30, and the indoor heat exchanger 40 through the pipe 50 in this order. Although not shown, the main circuit 110 may further include an accumulator (low pressure receiver), a receiver (high pressure receiver), or an oil separator (oil separator).
　ヒートポンプシステム１００は、バイパス流路１２０と、三方弁６０と、分岐流路１３０と、内部熱交換器８０と、副膨張弁７０とをさらに備える。また、ヒートポンプシステム１００は、圧力センサ９１と、温度センサ９２，９３とをさらに備える。 The heat pump system 100 further includes a bypass passage 120, a three-way valve 60, a branch passage 130, an internal heat exchanger 80, and a sub-expansion valve 70. The heat pump system 100 further includes a pressure sensor 91 and temperature sensors 92 and 93.
　バイパス流路１２０は、室外熱交換器２０の上流側と下流側との間に接続される。バイパス流路１２０は、主回路１１０を流れる冷媒が室外熱交換器２０を迂回するように構成されている。 The bypass channel 120 is connected between the upstream side and the downstream side of the outdoor heat exchanger 20. The bypass passage 120 is configured such that the refrigerant flowing through the main circuit 110 bypasses the outdoor heat exchanger 20.
　三方弁６０は、主回路１１０とバイパス流路１２０との接続部に接続される。三方弁６０の開度は、制御装置５００からの制御信号によって変更される。三方弁６０の構成については図２にて詳細に説明する。 The three-way valve 60 is connected to a connection portion between the main circuit 110 and the bypass flow path 120. The opening degree of the three-way valve 60 is changed by a control signal from the control device 500. The configuration of the three-way valve 60 will be described in detail with reference to FIG.
　分岐流路１３０は、室外熱交換器２０と主膨張弁３０との間に接続される。分岐流路１３０は、主回路１１０を流れる冷媒の一部を分岐可能に構成される。 The branch flow path 130 is connected between the outdoor heat exchanger 20 and the main expansion valve 30. The branch flow path 130 is configured to be able to branch a part of the refrigerant flowing through the main circuit 110.
　副膨張弁（第２の膨張弁）７０は、分岐流路１３０に設けられる。副膨張弁７０は、たとえばステッピングモータ（図示せず）により開度が制御可能な絞り弁である。すなわち、副膨張弁７０は、分岐流路１３０を流れる冷媒の流量を調整可能に構成される。 The secondary expansion valve (second expansion valve) 70 is provided in the branch flow path 130. The sub-expansion valve 70 is a throttle valve whose opening degree can be controlled by, for example, a stepping motor (not shown). That is, the sub expansion valve 70 is configured to be able to adjust the flow rate of the refrigerant flowing through the branch flow path 130.
　内部熱交換器（副熱交換器）８０は、分岐流路１３０を流れる冷媒と、室外熱交換器２０から主膨張弁３０に流れる冷媒との間で熱交換する。内部熱交換器８０にて熱交換が行なわれた冷媒は、室内熱交換器４０から圧縮機１０へと流れる冷媒に合流する。 The internal heat exchanger (sub heat exchanger) 80 exchanges heat between the refrigerant flowing through the branch flow path 130 and the refrigerant flowing from the outdoor heat exchanger 20 to the main expansion valve 30. The refrigerant that has undergone heat exchange in the internal heat exchanger 80 joins the refrigerant that flows from the indoor heat exchanger 40 to the compressor 10.
　圧力センサ９１は、圧縮機１０の吐出側に設けられる。圧力センサ９１は、圧縮機１０から吐出される冷媒の圧力（吐出圧力）を検出して、その検出結果を示す信号を制御装置５００に出力する。圧縮機１０の吐出圧力は、室外熱交換器２０における凝縮圧力Ｐｃとほぼ等しい。そのため、圧力センサ９１により吐出圧力を検出することによって、凝縮圧力Ｐｃを取得することができる。 The pressure sensor 91 is provided on the discharge side of the compressor 10. The pressure sensor 91 detects the pressure (discharge pressure) of the refrigerant discharged from the compressor 10 and outputs a signal indicating the detection result to the control device 500. The discharge pressure of the compressor 10 is substantially equal to the condensation pressure Pc in the outdoor heat exchanger 20. Therefore, the condensation pressure Pc can be acquired by detecting the discharge pressure with the pressure sensor 91.
　温度センサ９２は、圧縮機１０の吐出側に設けられる。温度センサ９２は、たとえばサーミスタを含んで構成され、圧縮機１０から吐出される冷媒の温度（吐出温度）を検出する。温度センサ９３は、内部熱交換器８０の出口側に設けられる、温度センサ９３は、たとえばサーミスタを含んで構成され、内部熱交換器８０における冷媒の凝縮温度を検出する。各センサは、その検出結果を示す信号を制御装置５００に出力する。 The temperature sensor 92 is provided on the discharge side of the compressor 10. The temperature sensor 92 includes, for example, a thermistor, and detects the temperature (discharge temperature) of the refrigerant discharged from the compressor 10. The temperature sensor 93 is provided on the outlet side of the internal heat exchanger 80. The temperature sensor 93 includes, for example, a thermistor, and detects the condensation temperature of the refrigerant in the internal heat exchanger 80. Each sensor outputs a signal indicating the detection result to the control device 500.
　制御装置５００は、いずれも図示しないが、ＣＰＵ（Central　Processing　Unit）と、ＲＡＭ（Random　Access　Memory）およびＲＯＭ（Read　Only　Memory）などのメモリと、入出力インターフェイスとを含んで構成される。制御装置５００は、上述の各センサからの検出信号に基づいて、予めＲＯＭなどに格納されたプログラムをＣＰＵがＲＡＭに読み出して実行することによって各機器を制御する。 Although not shown, the control device 500 includes a CPU (Central Processing Unit), a memory such as a RAM (Random Access Memory) and a ROM (Read Only Memory), and an input / output interface. Based on the detection signals from the above-described sensors, the control device 500 controls each device by causing the CPU to read a program stored in advance in a ROM or the like into the RAM and execute it.
　より具体的に、制御装置５００は、主膨張弁３０の開度を調整することによって、冷媒の吐出温度を目標値に制御する吐出温度制御を実行する。また、制御装置５００は、各センサからの検出信号に基づいて、三方弁６０および副膨張弁７０を制御する。制御装置５００による三方弁６０および副膨張弁７０の制御の詳細については後述する。 More specifically, the control device 500 executes discharge temperature control for controlling the discharge temperature of the refrigerant to a target value by adjusting the opening of the main expansion valve 30. Moreover, the control apparatus 500 controls the three-way valve 60 and the sub expansion valve 70 based on the detection signal from each sensor. Details of control of the three-way valve 60 and the sub-expansion valve 70 by the control device 500 will be described later.
　図２は、三方弁６０の構成を模式的に示す図である。図１および図２（Ａ）を参照して、三方弁６０は、圧縮機１０から吐出された冷媒が流入する入力ポートＩＮと、冷媒を室外熱交換器２０へと流出させるための出力ポートＯＵＴ１と、冷媒をバイパス流路１２０へと流出させるための出力ポートＯＵＴ２と、弁体６１とを有する。 FIG. 2 is a diagram schematically showing the configuration of the three-way valve 60. Referring to FIGS. 1 and 2A, a three-way valve 60 includes an input port IN into which the refrigerant discharged from the compressor 10 flows, and an output port OUT1 through which the refrigerant flows out to the outdoor heat exchanger 20. And an output port OUT2 for allowing the refrigerant to flow out to the bypass channel 120, and a valve body 61.
　弁体６１の角度θは、たとえば０°以上９０°以下の範囲で変更可能である。角度θを変更することにより、入力ポートＩＮに流入する冷媒の流量に対する出力ポートＯＵＴ１から流出する冷媒の流量の比率Ｒを調整することができる。言い換えると、比率Ｒは、圧縮機１０から吐出された冷媒の流量に対する室外熱交換器２０を流れる冷媒の流量の比率である。 The angle θ of the valve body 61 can be changed within a range of 0 ° to 90 °, for example. By changing the angle θ, the ratio R of the flow rate of the refrigerant flowing out from the output port OUT1 with respect to the flow rate of the refrigerant flowing into the input port IN can be adjusted. In other words, the ratio R is the ratio of the flow rate of the refrigerant flowing through the outdoor heat exchanger 20 to the flow rate of the refrigerant discharged from the compressor 10.
　たとえば、図２（Ｂ）に示すように弁体６１の角度θを０°に調整すると、バイパス流路１２０へと向かう冷媒の流れが遮断される。つまり、すべての冷媒が室外熱交換器２０を流れることになるので、比率Ｒは１になる。一方、図２（Ｃ）に示すように弁体６１の角度θを９０°に調整すると、室外熱交換器２０へと向かう冷媒の流れが遮断される。つまり、すべての冷媒がバイパス流路１２０を流れることになるので、比率Ｒは０になる。このように、図２（Ｂ）に示す状態と図２（Ｃ）に示す状態との間で弁体６１の角度θを調整することにより、比率Ｒを設定することができる。 For example, when the angle θ of the valve body 61 is adjusted to 0 ° as shown in FIG. 2B, the flow of the refrigerant toward the bypass flow path 120 is blocked. That is, since all the refrigerant flows through the outdoor heat exchanger 20, the ratio R is 1. On the other hand, when the angle θ of the valve body 61 is adjusted to 90 ° as shown in FIG. 2C, the flow of the refrigerant toward the outdoor heat exchanger 20 is blocked. That is, since all the refrigerant flows through the bypass channel 120, the ratio R becomes zero. In this way, the ratio R can be set by adjusting the angle θ of the valve body 61 between the state shown in FIG. 2B and the state shown in FIG.
　なお、三方弁６０は、本発明に係る「流量調整器」に相当する。ただし、本発明に係る「流量調整器」の構成は三方弁６０に限られるものではない。たとえば、三方弁６０に代えて、圧縮機１０および室外熱交換器２０の間の流路と、バイパス流路１２０とに１つずつ二方弁を設けてもよい。２つの二方弁の各々の開度を調整することにより、比率Ｒを設定することができる。 The three-way valve 60 corresponds to a “flow regulator” according to the present invention. However, the configuration of the “flow regulator” according to the present invention is not limited to the three-way valve 60. For example, instead of the three-way valve 60, one two-way valve may be provided in each of the flow path between the compressor 10 and the outdoor heat exchanger 20 and the bypass flow path 120. The ratio R can be set by adjusting the opening degree of each of the two two-way valves.
　また、三方弁６０として、冷媒温度に応じて開度が自動調整される感温式の弁を採用することも考えられる。しかし、本実施の形態のように電子制御弁を採用することにより、たとえば冷媒温度の急変時（ヒートポンプシステム１００の起動時など）においても冷媒温度に依存せずに所望の角度θに弁体６１を制御することが可能である。よって、比率Ｒを設定する際の自由度を向上させることができる。 It is also conceivable to employ a temperature-sensitive valve whose opening degree is automatically adjusted according to the refrigerant temperature as the three-way valve 60. However, by adopting the electronic control valve as in the present embodiment, for example, even when the refrigerant temperature suddenly changes (such as when the heat pump system 100 is started), the valve element 61 is set at a desired angle θ without depending on the refrigerant temperature. Can be controlled. Therefore, the degree of freedom when setting the ratio R can be improved.
　ここで、凝縮圧力（室外熱交換器２０において冷媒が凝縮するときの飽和圧力）Ｐｃが過度に低い場合には種々の不具合が生じ得る。たとえば、室内熱交換器４０に霜が付着したり、過剰な除湿が行なわれたりする場合がある。そのため、凝縮圧力Ｐｃを適切な範囲内に維持することが望ましい。ヒートポンプシステム１００によれば、凝縮圧力Ｐｃが基準値Ｐ１未満になった場合には、凝縮圧力Ｐｃが基準値Ｐ１以上の場合と比べて、三方弁６０の弁体６１の角度θを調整することにより、比率Ｒを低下させることができる。これにより、バイパス流路１２０を介して室外熱交換器２０を迂回する冷媒の流量が増加するので、凝縮圧力Ｐｃの低下を抑制することができる。 Here, when the condensation pressure (saturation pressure when the refrigerant condenses in the outdoor heat exchanger 20) Pc is excessively low, various problems may occur. For example, frost may adhere to the indoor heat exchanger 40 or excessive dehumidification may occur. Therefore, it is desirable to maintain the condensation pressure Pc within an appropriate range. According to the heat pump system 100, when the condensation pressure Pc is less than the reference value P1, the angle θ of the valve body 61 of the three-way valve 60 is adjusted as compared with the case where the condensation pressure Pc is greater than or equal to the reference value P1. Thus, the ratio R can be reduced. Thereby, since the flow volume of the refrigerant | coolant which detours the outdoor heat exchanger 20 via the bypass flow path 120 increases, the fall of the condensation pressure Pc can be suppressed.
　その一方で、室外熱交換器２０を迂回する冷媒の流量が増加する分だけ、室外熱交換器２０を流れる冷媒の流量が減少する。そのため、室外熱交換器２０における冷媒から外気への放熱量が小さくなるので、室外熱交換器２０における冷媒温度の低下量が小さくなる。そうすると、室外熱交換器２０から流出して主膨張弁３０に流入する冷媒について、必要な過冷却度ΔＴｃが確保できない可能性がある。 On the other hand, the flow rate of the refrigerant flowing through the outdoor heat exchanger 20 is decreased by the increase in the flow rate of the refrigerant bypassing the outdoor heat exchanger 20. As a result, the amount of heat released from the refrigerant to the outside air in the outdoor heat exchanger 20 is reduced, and the amount of decrease in the refrigerant temperature in the outdoor heat exchanger 20 is reduced. If it does so, there exists a possibility that required supercooling degree (DELTA) Tc cannot be ensured about the refrigerant | coolant which flows out out of the outdoor heat exchanger 20, and flows in into the main expansion valve 30. FIG.
　過冷却度ΔＴｃが確保できないとは、主膨張弁３０に流入する冷媒が気液二相状態となることを意味する。気液二相冷媒の平均密度は、液冷媒の平均密度よりも低い。そのため、気液二相冷媒について所望の必要な冷却能力を実現するためには、液冷媒の場合と比べて、主膨張弁３０を流れる冷媒の流量を大きくすることが求められる。つまり、主膨張弁３０の開度を大きくする必要がある。しかし、一般に、膨張弁の開度には上限値が存在する。主膨張弁３０の開度が上限値に達すると、主膨張弁３０を流れる冷媒の流量をそれ以上大きくすることはできない。したがって、気液二相冷媒の場合には所望の冷媒能力を実現することができない可能性がある。 The fact that the degree of supercooling ΔTc cannot be secured means that the refrigerant flowing into the main expansion valve 30 is in a gas-liquid two-phase state. The average density of the gas-liquid two-phase refrigerant is lower than the average density of the liquid refrigerant. Therefore, in order to realize a desired required cooling capacity for the gas-liquid two-phase refrigerant, it is required to increase the flow rate of the refrigerant flowing through the main expansion valve 30 as compared with the case of the liquid refrigerant. That is, it is necessary to increase the opening of the main expansion valve 30. However, generally, there is an upper limit for the opening of the expansion valve. When the opening degree of the main expansion valve 30 reaches the upper limit value, the flow rate of the refrigerant flowing through the main expansion valve 30 cannot be increased further. Therefore, in the case of a gas-liquid two-phase refrigerant, there is a possibility that a desired refrigerant capacity cannot be realized.
　そこで、実施の形態１によれば、室外熱交換器２０を流れる冷媒の流量の比率Ｒを三方弁６０の制御により低下させた場合に、副膨張弁７０の開度をさらに制御する。より具体的には、主膨張弁３０に流入する冷媒の過冷却度ΔＴｃが基準値Ｔ２を下回るときには、過冷却度ΔＴｃが基準値Ｔ２を上回るときと比べて、内部熱交換器８０を流れる冷媒の流量が大きくなるように副膨張弁７０の開度が制御される。これにより、主回路１１０を流れる冷媒と、副膨張弁７０での減圧により冷却された冷媒との間で交換される熱量が大きくなる。よって、上述のように副膨張弁７０の開度が制御されない場合と比べて、主回路１１０を流れる冷媒温度が低くなるので、主膨張弁３０に流入する冷媒の過冷却度をより確実に確保することができる。 Therefore, according to the first embodiment, when the ratio R of the flow rate of the refrigerant flowing through the outdoor heat exchanger 20 is reduced by the control of the three-way valve 60, the opening degree of the sub expansion valve 70 is further controlled. More specifically, the refrigerant flowing through the internal heat exchanger 80 when the degree of supercooling ΔTc of the refrigerant flowing into the main expansion valve 30 is lower than the reference value T2 compared to when the degree of supercooling ΔTc is higher than the reference value T2. The degree of opening of the secondary expansion valve 70 is controlled so that the flow rate of As a result, the amount of heat exchanged between the refrigerant flowing through the main circuit 110 and the refrigerant cooled by the pressure reduction at the sub expansion valve 70 increases. Therefore, since the refrigerant temperature flowing through the main circuit 110 is lower than the case where the opening degree of the sub-expansion valve 70 is not controlled as described above, the degree of supercooling of the refrigerant flowing into the main expansion valve 30 is more reliably ensured. can do.
　図３は、実施の形態１にて実行される三方弁６０および副膨張弁７０の制御を説明するためのフローチャートである。図３および後述する図７に示されるフローチャートの各ステップ（以下、Ｓで略す）は、所定時間が経過する毎または所定条件が成立する毎にメインルーチンから呼び出されて実行される。 FIG. 3 is a flowchart for explaining the control of the three-way valve 60 and the sub-expansion valve 70 executed in the first embodiment. Each step (hereinafter abbreviated as S) of the flowchart shown in FIG. 3 and FIG. 7 described later is called from the main routine and executed every time a predetermined time elapses or a predetermined condition is satisfied.
　図１～図３を参照して、Ｓ１０において、制御装置５００は、室外熱交換器２０における凝縮圧力Ｐｃを算出する。凝縮圧力Ｐｃは、圧力センサ９１により検出される吐出圧力に基づき算出することができる。 Referring to FIGS. 1 to 3, in S10, control device 500 calculates condensing pressure Pc in outdoor heat exchanger 20. The condensation pressure Pc can be calculated based on the discharge pressure detected by the pressure sensor 91.
　Ｓ２０において、制御装置５００は、凝縮圧力Ｐｃが基準値Ｐ１（第１の基準値）未満であるか否かを判定する。基準値Ｐ１は、室内熱交換器４０に霜の付着等の上述の不具合が生じない値であり、かつ、圧縮機１０の圧縮比が過度に高くなることによる圧縮機１０の異常が起こらない値であることが好ましい。 In S20, the control device 500 determines whether or not the condensation pressure Pc is less than the reference value P1 (first reference value). The reference value P1 is a value that does not cause the above-described problems such as frost adhesion in the indoor heat exchanger 40, and a value that does not cause an abnormality of the compressor 10 due to an excessively high compression ratio of the compressor 10. It is preferable that
　凝縮圧力Ｐｃが基準値Ｐ１未満の場合（Ｓ２０においてＹＥＳ）、制御装置５００は、バイパス流路１２０へと向かう冷媒の流量が大きくなる方向に三方弁６０の弁体６１の角度θを調整する。図２に示した例においては角度θを増加させる。これにより、圧縮機１０から吐出される冷媒の流量に対する室外熱交換器２０を流れる冷媒の流量の比率Ｒが低下する（Ｓ３０）。その結果、バイパス流路１２０を介して室外熱交換器２０を通過せずに室外熱交換器２０の下流側へと至る冷媒の割合が増加するため、凝縮圧力Ｐｃが上昇する。よって、凝縮圧力Ｐｃの低下を抑制することができる。 When the condensation pressure Pc is less than the reference value P1 (YES in S20), the control device 500 adjusts the angle θ of the valve body 61 of the three-way valve 60 in a direction in which the flow rate of the refrigerant toward the bypass flow path 120 increases. In the example shown in FIG. 2, the angle θ is increased. Thereby, the ratio R of the flow rate of the refrigerant flowing through the outdoor heat exchanger 20 with respect to the flow rate of the refrigerant discharged from the compressor 10 decreases (S30). As a result, the ratio of the refrigerant that reaches the downstream side of the outdoor heat exchanger 20 without passing through the outdoor heat exchanger 20 via the bypass flow path 120 increases, so that the condensation pressure Pc increases. Therefore, the fall of the condensation pressure Pc can be suppressed.
　その一方で、室外熱交換器２０における冷媒から外気への放熱量が小さくなるので、主膨張弁３０に流入する冷媒の過冷却度ΔＴｃが過度に小さくなり得る。そのため、Ｓ４０において、制御装置５００は、主膨張弁３０に流入する冷媒の過冷却度ΔＴｃを算出する。過冷却度ΔＴｃは、温度センサ９３により検出された冷媒温度（内部熱交換器８０の出口側における冷媒温度）に基づき算出することができる。 On the other hand, since the amount of heat released from the refrigerant to the outside air in the outdoor heat exchanger 20 is reduced, the degree of supercooling ΔTc of the refrigerant flowing into the main expansion valve 30 can be excessively reduced. Therefore, in S40, control device 500 calculates the degree of supercooling ΔTc of the refrigerant flowing into main expansion valve 30. The degree of supercooling ΔTc can be calculated based on the refrigerant temperature detected by the temperature sensor 93 (the refrigerant temperature at the outlet side of the internal heat exchanger 80).
　Ｓ５０において、制御装置５００は、過冷却度ΔＴｃが基準値Ｔ２（第２の基準値）未満であるか否かを判定する。基準値Ｔ２の設定手法については後述する。 In S50, the control device 500 determines whether or not the degree of supercooling ΔTc is less than the reference value T2 (second reference value). A method for setting the reference value T2 will be described later.
　過冷却度ΔＴｃが基準値Ｔ２以上の場合（Ｓ５０においてＮＯ）、制御装置５００は処理をＳ７０に進める。Ｓ７０において、制御装置５００は、主膨張弁３０に流入する冷媒について十分な過冷却度ΔＴｃが確保されているとして、副膨張弁７０の開度を維持する（あるいは副膨張弁７０の開度を小さく調整する）。 When supercooling degree ΔTc is equal to or greater than reference value T2 (NO in S50), control device 500 advances the process to S70. In S70, the control device 500 maintains the opening degree of the sub-expansion valve 70 (or sets the opening degree of the sub-expansion valve 70), assuming that a sufficient degree of subcooling ΔTc is secured for the refrigerant flowing into the main expansion valve 30. Adjust it small).
　これに対し、過冷却度ΔＴｃが基準値Ｔ２未満の場合（Ｓ５０においてＹＥＳ）、制御装置５００は処理をＳ６０に進める。Ｓ６０において、制御装置５００は、過冷却度ΔＴｃが基準値Ｔ２以上の場合と比べて、副膨張弁７０の開度を大きく設定する。これにより、分岐流路１３０の流れる冷媒の流量が増加するので、内部熱交換器８０において、主回路１１０を流れる冷媒と、副膨張弁７０での減圧により冷却された冷媒との間で交換される熱量が大きくなる。よって、過冷却度ΔＴｃが基準値Ｔ２以上の場合と比べて、主回路１１０を流れる冷媒温度が低くなる。したがって、主膨張弁３０に流入する冷媒の過冷却度ΔＴｃをより確実に確保することができる。 On the other hand, when subcooling degree ΔTc is less than reference value T2 (YES in S50), control device 500 advances the process to S60. In S60, control device 500 sets the opening degree of sub expansion valve 70 to be larger than that in the case where degree of supercooling ΔTc is equal to or greater than reference value T2. As a result, the flow rate of the refrigerant flowing through the branch flow path 130 increases, so that in the internal heat exchanger 80, the refrigerant is exchanged between the refrigerant flowing through the main circuit 110 and the refrigerant cooled by the decompression at the sub expansion valve 70. The amount of heat to be increased. Therefore, the temperature of the refrigerant flowing through the main circuit 110 is lower than when the degree of supercooling ΔTc is greater than or equal to the reference value T2. Therefore, the degree of supercooling ΔTc of the refrigerant flowing into the main expansion valve 30 can be ensured more reliably.
　なお、Ｓ２０にて凝縮圧力Ｐｃが基準値Ｐ１以上の場合（Ｓ２０においてＹＥＳ）、制御装置５００は、三方弁６０の弁体６１の角度θを維持するか、あるいはバイパス流路１２０へと向かう冷媒の流量が小さくなる方向に角度θを調整する（Ｓ８０）。図２に示した例においては角度θを維持するか減少させる。これにより、比率Ｒは、維持されるか増加する。Ｓ６０，Ｓ７０，Ｓ８０のいずれかの処理が終了すると、処理はメインルーチンへと戻される。 When the condensing pressure Pc is equal to or higher than the reference value P1 in S20 (YES in S20), the control device 500 maintains the angle θ of the valve body 61 of the three-way valve 60, or refrigerant that goes to the bypass flow path 120. The angle θ is adjusted in the direction in which the flow rate decreases (S80). In the example shown in FIG. 2, the angle θ is maintained or decreased. Thereby, the ratio R is maintained or increased. When any one of S60, S70, and S80 ends, the process returns to the main routine.
　ここでは本発明に係る「冷媒の凝縮度合を示す状態値」として凝縮圧力Ｐｃが用いられる例について説明したが、「状態値」はこれに限定されるものではない。本発明に係る「状態値」は、室外熱交換器２０を流れる冷媒温度、圧縮機１０からの冷媒の吐出温度、または圧縮機１０における冷媒の圧縮比であってもよい。 Here, although the example in which the condensation pressure Pc is used as the “state value indicating the degree of condensation of the refrigerant” according to the present invention has been described, the “state value” is not limited to this. The “state value” according to the present invention may be the temperature of the refrigerant flowing through the outdoor heat exchanger 20, the discharge temperature of the refrigerant from the compressor 10, or the compression ratio of the refrigerant in the compressor 10.
　図４は、実施の形態１にて実行される三方弁６０および副膨張弁７０の制御に対応するＰｈ線図である。図４において、横軸は比エンタルピーｈ［単位：ｋＪ／ｋｇ］を表し、縦軸は圧力Ｐ［単位：ＭＰａ］を表す。 FIG. 4 is a Ph diagram corresponding to the control of the three-way valve 60 and the sub-expansion valve 70 executed in the first embodiment. In FIG. 4, the horizontal axis represents specific enthalpy h [unit: kJ / kg], and the vertical axis represents pressure P [unit: MPa].
　図１および図４を参照して、Ａ点は、低圧のガス冷媒（過熱蒸気）の状態を示す。Ａ点からＢ点への過程は、圧縮機１０による断熱圧縮過程である。Ｂ点は、冷媒が圧縮機１０により圧縮された状態を示す。Ｄ点は、三方弁６０を介して室外熱交換器２０を迂回した冷媒と、室外熱交換器２０により凝縮された冷媒とが混合した状態を示す。Ｅ点は、冷媒が室外熱交換器２０にて凝縮され、さらに冷媒の一部が内部熱交換器８０により冷却されることによって得られた過冷却状態を示す。Ｅ点からＦ点への過程は、主膨張弁３０による冷媒の膨張過程を示す。Ｅ点からＧ点への過程は、副膨張弁７０による冷媒の膨張過程を示す。Ｅ点からＧ点への過程、およびＦ点からＧ点への過程は、室内熱交換器４０における蒸発過程を示す。 1 and 4, point A indicates the state of a low-pressure gas refrigerant (superheated steam). The process from the point A to the point B is an adiabatic compression process by the compressor 10. Point B indicates a state in which the refrigerant is compressed by the compressor 10. Point D shows a state where the refrigerant bypassing the outdoor heat exchanger 20 via the three-way valve 60 and the refrigerant condensed by the outdoor heat exchanger 20 are mixed. Point E indicates a supercooled state obtained by condensing the refrigerant in the outdoor heat exchanger 20 and further cooling a part of the refrigerant by the internal heat exchanger 80. The process from the point E to the point F shows the expansion process of the refrigerant by the main expansion valve 30. The process from the point E to the point G shows the expansion process of the refrigerant by the sub expansion valve 70. A process from the point E to the point G and a process from the point F to the point G indicate an evaporation process in the indoor heat exchanger 40.
　以上のように、実施の形態１によれば、ヒートポンプシステム１００の冷房運転時において、凝縮圧力Ｐｃの低下を抑制しつつ、主膨張弁３０に流入する冷媒について、必要な過冷却度ΔＴｃを確保することができる。 As described above, according to the first embodiment, during the cooling operation of the heat pump system 100, the necessary supercooling degree ΔTc is secured for the refrigerant flowing into the main expansion valve 30 while suppressing the decrease in the condensation pressure Pc. can do.
　また、多くの場合、膨張弁に含まれる狭い流路を気液二相冷媒が通過すると、異音または振動が発生し得る。これは、気液二相状態の冷媒では気相部分と液相部分とが均一に混合されていないので、通過する冷媒の密度が時間的に変化し得るためであると考えられる。実施の形態１によれば、必要な過冷却度ΔＴｃが確保されることにより、液冷媒が主膨張弁３０に流入することになるので、異音および振動の発生を防止することができる。 In many cases, when the gas-liquid two-phase refrigerant passes through a narrow flow path included in the expansion valve, abnormal noise or vibration may occur. This is presumably because the gas-liquid two-phase refrigerant is not uniformly mixed with the gas phase portion and the liquid phase portion, so that the density of the refrigerant passing therethrough can change over time. According to the first embodiment, since the necessary degree of supercooling ΔTc is ensured, the liquid refrigerant flows into the main expansion valve 30, so that it is possible to prevent the generation of abnormal noise and vibration.
　さらに、凝縮圧力Ｐｃが過度に低い場合には、圧縮機１０の圧縮比も適正値よりも低くなり得る。圧縮比が適正値未満の場合には、圧縮比が適正値以上の場合と比べて、圧縮機１０の駆動周波数を増加させることにより、圧縮比を高めることも考えられる。しかし、圧縮機１０の駆動周波数の増加に伴い、冷媒の蒸発圧力Ｐｅが低下する。一般に、蒸発圧力が基準値を下回ると、圧縮機は停止される。その後、蒸発圧力が基準値を上回るまで回復すると、圧縮機は再び駆動される。つまり、圧縮機の駆動および停止が繰り返される可能性がある（このような制御は「低圧カット制御」とも称される）。実施の形態１によれば、凝縮圧力Ｐｃの低下を抑制し、それにより圧縮比の過度の低下を抑制することができる。したがって、圧縮機１０の低圧カット制御が実行される状態に圧縮機１０が至ることが防止される。言い換えると、圧縮機１０の運転を安定化させることができる。 Furthermore, when the condensation pressure Pc is excessively low, the compression ratio of the compressor 10 can also be lower than an appropriate value. When the compression ratio is less than the appropriate value, it is conceivable to increase the compression ratio by increasing the drive frequency of the compressor 10 as compared with the case where the compression ratio is equal to or more than the appropriate value. However, as the driving frequency of the compressor 10 increases, the refrigerant evaporation pressure Pe decreases. Generally, when the evaporation pressure is below a reference value, the compressor is stopped. Thereafter, when the evaporation pressure recovers until it exceeds the reference value, the compressor is driven again. That is, the compressor may be repeatedly driven and stopped (such control is also referred to as “low pressure cut control”). According to the first embodiment, it is possible to suppress a decrease in the condensation pressure Pc, thereby suppressing an excessive decrease in the compression ratio. Therefore, the compressor 10 is prevented from reaching a state where the low pressure cut control of the compressor 10 is executed. In other words, the operation of the compressor 10 can be stabilized.
　ここで、主膨張弁３０に流入する冷媒の過冷却度ΔＴｃに関して設定される基準値Ｔ２（図３のＳ５０参照）について説明する。基準値Ｔ２は、室内熱交換器４０が室外熱交換器２０よりも高い位置に設けられた場合であっても、主膨張弁３０に流入する冷媒の過冷却状態が維持されるように設定することが好ましい。この理由について以下に説明する。 Here, the reference value T2 (see S50 in FIG. 3) set for the degree of supercooling ΔTc of the refrigerant flowing into the main expansion valve 30 will be described. The reference value T2 is set so that the supercooled state of the refrigerant flowing into the main expansion valve 30 is maintained even when the indoor heat exchanger 40 is provided at a position higher than the outdoor heat exchanger 20. It is preferable. The reason for this will be described below.
　仮に室内熱交換器４０が室外熱交換器２０よりも低い位置に設けられた場合、室外熱交換器２０から室内熱交換器４０へと向かう冷媒の流れは下降流となる。そのため、液冷媒にかかる重力により生じた圧力が主膨張弁３０に印加される。よって、主膨張弁３０での膨張により液冷媒の圧力が減少しても冷媒の沸騰は抑制される。 If the indoor heat exchanger 40 is provided at a position lower than the outdoor heat exchanger 20, the refrigerant flow from the outdoor heat exchanger 20 toward the indoor heat exchanger 40 is a downward flow. Therefore, the pressure generated by the gravity applied to the liquid refrigerant is applied to the main expansion valve 30. Therefore, even if the pressure of the liquid refrigerant decreases due to expansion at the main expansion valve 30, boiling of the refrigerant is suppressed.
　これに対し、室内熱交換器４０が室外熱交換器２０よりも高い位置に設けられた場合、室外熱交換器２０から室内熱交換器４０へと向かう冷媒の流れは上昇流となる。そのため、冷媒の流れが下降流の場合と異なり、重力による圧力が主膨張弁３０に印加されることはない。よって、主膨張弁３０にて液冷媒の圧力が減少した際に冷媒の減圧沸騰が起こり、冷媒が気液二相状態となる可能性がある。そうすると、上述の異音または振動の問題が起こり得る。したがって、基準値Ｔ２は、主膨張弁３０において、圧力低下後であっても冷媒の過冷却状態が維持されるように設定することが好ましい。 On the other hand, when the indoor heat exchanger 40 is provided at a position higher than the outdoor heat exchanger 20, the refrigerant flow from the outdoor heat exchanger 20 toward the indoor heat exchanger 40 is an upward flow. Therefore, unlike the case where the refrigerant flow is a downward flow, pressure due to gravity is not applied to the main expansion valve 30. Therefore, when the pressure of the liquid refrigerant is reduced at the main expansion valve 30, the refrigerant is boiled under reduced pressure, and the refrigerant may be in a gas-liquid two-phase state. Then, the above-mentioned abnormal noise or vibration problem may occur. Therefore, it is preferable to set the reference value T2 so that the supercooled state of the refrigerant is maintained even after the pressure is reduced in the main expansion valve 30.
　　［変形例］
　実施の形態１では、圧縮機から吐出された冷媒の全部がバイパス流路を介して室外熱交換器を迂回可能な構成について説明した。変形例においては、冷媒の一部のみが室外熱交換器を迂回可能な構成例を説明する。 [Modification]
In the first embodiment, the configuration in which all of the refrigerant discharged from the compressor can bypass the outdoor heat exchanger via the bypass flow path has been described. In the modification, a configuration example in which only a part of the refrigerant can bypass the outdoor heat exchanger will be described.
　図５は、実施の形態１の変形例に係るヒートポンプシステムの構成を概略的に示すブロック図である。図５を参照して、ヒートポンプシステム１００Ａは、室外熱交換器２０Ａ、バイパス流路１２０Ａ、および三方弁６０Ａの構成が、実施の形態１に係るヒートポンプシステム１００（図１参照）における対応する構成と異なる。 FIG. 5 is a block diagram schematically showing a configuration of a heat pump system according to a modification of the first embodiment. Referring to FIG. 5, in heat pump system 100A, the configurations of outdoor heat exchanger 20A, bypass flow path 120A, and three-way valve 60A correspond to the configurations in heat pump system 100 according to Embodiment 1 (see FIG. 1). Different.
　室外熱交換器２０Ａ内部に設けられた流路に沿う所定の位置に接続部Ｃ１が設けられている。図示しないが、室外熱交換器２０Ａは、接続部Ｃ１よりも上流側の熱交換器と、接続部Ｃ１よりも下流側の熱交換器とに２つに分割されている。三方弁６０Ａの入力ポートＩＮは、接続部Ｃ１にバイパス流路１２０Ａによって接続されている。室外熱交換器２０を流れる冷媒の一部は、上流側の熱交換器のみを通り、接続部Ｃ１を介して室外熱交換器２０外部へと流出する。一方、残りの冷媒は、上流側の熱交換器および下流側の熱交換器の両方を流れる。 The connection part C1 is provided in the predetermined position along the flow path provided in the outdoor heat exchanger 20A. Although not shown, the outdoor heat exchanger 20A is divided into two parts: a heat exchanger upstream of the connection part C1 and a heat exchanger downstream of the connection part C1. The input port IN of the three-way valve 60A is connected to the connection portion C1 by a bypass flow path 120A. A part of the refrigerant flowing through the outdoor heat exchanger 20 passes through only the upstream heat exchanger and flows out of the outdoor heat exchanger 20 through the connection portion C1. On the other hand, the remaining refrigerant flows through both the upstream heat exchanger and the downstream heat exchanger.
　三方弁６０Ａの出力ポートＯＵＴ１は、室外熱交換器６０内部の流路について接続部Ｃ１よりも下流側の接続部Ｃ２にバイパス流路１２０Ａによって接続されている。三方弁６０Ａの出力ポートＯＵＴ２は、主回路１１０について室外熱交換器２０Ａよりも下流側にバイパス流路１２０Ａによって接続されている。なお、ヒートポンプシステム１００Ａの室外熱交換器２０Ａ、バイパス流路１２０Ａ、および三方弁６０Ａ以外の構成は、実施の形態１に係るヒートポンプシステム１００の対応する構成と同等であるため、詳細な説明は繰り返さない。 The output port OUT1 of the three-way valve 60A is connected to the connection part C2 on the downstream side of the connection part C1 with respect to the flow path inside the outdoor heat exchanger 60 by the bypass flow path 120A. The output port OUT2 of the three-way valve 60A is connected to the main circuit 110 on the downstream side of the outdoor heat exchanger 20A by a bypass passage 120A. The configuration of heat pump system 100A other than outdoor heat exchanger 20A, bypass flow path 120A, and three-way valve 60A is the same as the corresponding configuration of heat pump system 100 according to Embodiment 1, and therefore detailed description will be repeated. Absent.
　図２にて説明したように、一般に、三方弁の開度は、弁体の角度によって調節される。たとえば弁体の角度が０°以上９０°以下の可動範囲で１°毎に変更可能な構成を想定する。この場合、三方弁によって調整可能な流量比（第１の出力ポートから流出する冷媒の流量と、第２の出力ポートから流出する冷媒の流量との比率）は、９０通りしか調整できないことになる。そのため、三方弁の入力ポートに流入する冷媒の流量が比較的大きい場合には、第１の出力ポートと第２の出力ポートとの間で流量の微調整を行なうことは難しい。 As described in FIG. 2, generally, the opening degree of the three-way valve is adjusted by the angle of the valve body. For example, the structure which can change every 1 degree in the movable range whose angle of a valve body is 0 degree or more and 90 degrees or less is assumed. In this case, the flow rate ratio that can be adjusted by the three-way valve (the ratio between the flow rate of the refrigerant flowing out from the first output port and the flow rate of the refrigerant flowing out from the second output port) can be adjusted only in 90 ways. . Therefore, when the flow rate of the refrigerant flowing into the input port of the three-way valve is relatively large, it is difficult to finely adjust the flow rate between the first output port and the second output port.
　これに対し、本変形例によれば、圧縮機１０から吐出された冷媒のうち、一部のみが室外熱交換器２０内部にて分岐して三方弁６０Ａの入力ポートＩＮ１に流入する。そのため、実施の形態１のように、圧縮機１０から吐出された冷媒の全部について、室外熱交換器２０を迂回させることはできない。一方で、圧縮機１０から吐出された冷媒の流量が等しい場合、実施の形態１と比べて、三方弁６０Ａの入力ポートＩＮに流入する冷媒の流量が小さくなるので、流量の微調整を三方弁６０Ａによって行なうことができる。その結果、凝縮圧力Ｐｃの調整幅を、より小さくすることが可能になる。 On the other hand, according to this modification, only a part of the refrigerant discharged from the compressor 10 branches inside the outdoor heat exchanger 20 and flows into the input port IN1 of the three-way valve 60A. Therefore, the outdoor heat exchanger 20 cannot be bypassed for all of the refrigerant discharged from the compressor 10 as in the first embodiment. On the other hand, when the flow rate of the refrigerant discharged from the compressor 10 is equal, the flow rate of the refrigerant flowing into the input port IN of the three-way valve 60A is smaller than that in the first embodiment, so that the flow rate is finely adjusted. This can be done with 60A. As a result, the adjustment range of the condensation pressure Pc can be further reduced.
　逆の観点から説明すると、実施の形態１に係るヒートポンプシステム１００においては、変形例１に係るヒートポンプシステム１００Ａのように、冷媒を取り出すために室外熱交換器２０が２つに分割されていない。したがって、実施の形態１では室外熱交換器２０のコストを低減することができる。 If it demonstrates from a reverse viewpoint, in the heat pump system 100 which concerns on Embodiment 1, the outdoor heat exchanger 20 is not divided | segmented into two in order to take out a refrigerant | coolant like the heat pump system 100A which concerns on the modification 1. FIG. Therefore, in Embodiment 1, the cost of the outdoor heat exchanger 20 can be reduced.
　なお、三方弁６０Ａよりも下流側のバイパス流路１２０Ａに、室外熱交換器２０Ａと同等の圧力損失が生じる絞り（毛細管）を設けてもよい。この絞りにより生じる圧力損失によって、三方弁６０Ａが閉鎖状態から開度を増加させた場合に冷媒の流量が急激に増加することを防止できる。 In addition, a throttle (capillary tube) that causes a pressure loss equivalent to that of the outdoor heat exchanger 20A may be provided in the bypass flow path 120A on the downstream side of the three-way valve 60A. Due to the pressure loss caused by the throttle, it is possible to prevent the refrigerant flow rate from rapidly increasing when the opening of the three-way valve 60A is increased from the closed state.
　また、実施の形態１にて説明したように、三方弁６０Ａに代えて２つの二方弁を設けてもよい。図示しないが、一方は、図５における三方弁６０Ａの出力ポートＯＵＴ１と接続部Ｃ２との間に設けられる。他方は、三方弁６０Ａの出力ポートＯＵＴ２と、主回路１１０について室外熱交換器２０Ａよりも下流側との間に設けられる。この場合、三方弁を設ける場合と比べて弁の数が大きくなるものの、一般に二方弁の方が三方弁と比べて開度の調整幅が小さいので、流量の微調整に適している。 Further, as described in the first embodiment, two two-way valves may be provided instead of the three-way valve 60A. Although not shown, one is provided between the output port OUT1 of the three-way valve 60A in FIG. 5 and the connection C2. The other is provided between the output port OUT2 of the three-way valve 60A and the downstream side of the main circuit 110 with respect to the outdoor heat exchanger 20A. In this case, although the number of valves is larger than that in the case where a three-way valve is provided, the two-way valve is generally suitable for fine adjustment of the flow rate because the adjustment range of the opening is smaller than that of the three-way valve.
　　［実施の形態２］
　実施の形態１では、ヒートポンプシステムの冷房運転時の構成および制御について説明した。実施の形態２においては、ヒートポンプシステムの暖房運転時の構成および制御について説明する。 [Embodiment 2]
In the first embodiment, the configuration and control during the cooling operation of the heat pump system have been described. In the second embodiment, the configuration and control during the heating operation of the heat pump system will be described.
　図６は、実施の形態２に係るヒートポンプシステムの構成を概略的に示すブロック図である。図６を参照して、ヒートポンプシステム１００Ｂの暖房運転時においては、室外熱交換器２０が蒸発器として機能し、室内熱交換器４０が凝縮器として機能する。さらに、ヒートポンプシステム１００Ｂは、分岐流路１３０Ｂ、副膨張弁７０Ｂ、および内部熱交換器８０Ｂの構成が、実施の形態１に係るヒートポンプシステム１００（図１参照）における対応する構成と異なる。また、ヒートポンプシステム１００Ｂは、温度センサ９２に代えて温度センサ９４を備える点において、図１に示すヒートポンプシステム１００と異なる。 FIG. 6 is a block diagram schematically showing the configuration of the heat pump system according to the second embodiment. Referring to FIG. 6, during the heating operation of heat pump system 100B, outdoor heat exchanger 20 functions as an evaporator, and indoor heat exchanger 40 functions as a condenser. Furthermore, in the heat pump system 100B, the configuration of the branch flow path 130B, the sub expansion valve 70B, and the internal heat exchanger 80B is different from the corresponding configuration in the heat pump system 100 according to Embodiment 1 (see FIG. 1). The heat pump system 100B is different from the heat pump system 100 shown in FIG. 1 in that a temperature sensor 94 is provided instead of the temperature sensor 92.
　分岐流路１３０Ｂは、圧縮機１０と室内熱交換器４０との間に接続される。分岐流路１３０Ｂは、主回路１１０を流れる冷媒の一部を分岐可能に構成される。副膨張弁７０Ｂは、分岐流路１３０Ｂに設けられる。 The branch flow path 130B is connected between the compressor 10 and the indoor heat exchanger 40. The branch flow path 130 </ b> B is configured to be able to branch a part of the refrigerant flowing through the main circuit 110. The sub expansion valve 70B is provided in the branch flow path 130B.
　内部熱交換器８０Ｂは、分岐流路１３０Ｂを流れる冷媒と、室外熱交換器２０から圧縮機１０に流れる冷媒との間で熱交換する。内部熱交換器８０Ｂにて熱交換が行なわれた冷媒は、室外熱交換器２０から圧縮機１０へと流れる冷媒に合流する。 The internal heat exchanger 80B exchanges heat between the refrigerant flowing through the branch flow path 130B and the refrigerant flowing from the outdoor heat exchanger 20 to the compressor 10. The refrigerant that has undergone heat exchange in the internal heat exchanger 80B joins the refrigerant that flows from the outdoor heat exchanger 20 to the compressor 10.
　温度センサ９４は室内熱交換器４０に設けられる。温度センサ９４は、室内熱交換器４０における冷媒の蒸発温度を検出し、その検出結果を示す信号を制御装置５００に出力する。なお、ヒートポンプシステム１００Ｂの分岐流路１３０Ｂ、副膨張弁７０Ｂ、および内部熱交換器８０Ｂ以外の構成は、実施の形態１に係るヒートポンプシステム１００の対応する構成と同等であるため、詳細な説明は繰り返さない。 The temperature sensor 94 is provided in the indoor heat exchanger 40. The temperature sensor 94 detects the evaporation temperature of the refrigerant in the indoor heat exchanger 40 and outputs a signal indicating the detection result to the control device 500. In addition, since structures other than the branch flow path 130B, the sub-expansion valve 70B, and the internal heat exchanger 80B of the heat pump system 100B are the same as the corresponding structures of the heat pump system 100 according to the first embodiment, a detailed description will be given. Do not repeat.
　たとえば外気温が高い場合に暖房運転を行なうと、外気温が低い場合と比べて、室外熱交換器２０における冷媒の蒸発圧力Ｐｅが高くなり得る。つまり、室外熱交換器２０から流出して圧縮機１０へと流入する冷媒の圧力が高くなるので、圧縮機１０に吸入される冷媒の圧力と、圧縮機１０から吐出される冷媒の圧力との差が小さくなり得る。すなわち、圧縮機１０における圧縮比が低くなる可能性がある。 For example, when the heating operation is performed when the outside air temperature is high, the evaporation pressure Pe of the refrigerant in the outdoor heat exchanger 20 can be higher than when the outside air temperature is low. That is, since the pressure of the refrigerant flowing out of the outdoor heat exchanger 20 and flowing into the compressor 10 becomes high, the pressure of the refrigerant sucked into the compressor 10 and the pressure of the refrigerant discharged from the compressor 10 The difference can be small. That is, the compression ratio in the compressor 10 may be lowered.
　実施の形態２においては、室外熱交換器２０へと流れる冷媒の全部または一部について、バイパス流路１２０を介して室外熱交換器２０を迂回させる。これにより、室外熱交換器２０にて熱交換される冷媒の流量が減るので、蒸発圧力Ｐｅの上昇が抑制される。よって、外気温が比較的高い場合の暖房運転時においても圧縮機１０の圧縮比の低下を抑制することができる。 In Embodiment 2, all or part of the refrigerant flowing to the outdoor heat exchanger 20 is bypassed through the bypass flow path 120. Thereby, since the flow volume of the refrigerant | coolant heat-exchanged with the outdoor heat exchanger 20 reduces, the raise of the evaporation pressure Pe is suppressed. Therefore, a reduction in the compression ratio of the compressor 10 can be suppressed even during heating operation when the outside air temperature is relatively high.
　ここで、液冷媒が圧縮されても液冷媒の体積は変化しない。そのため、圧縮機１０により過剰な液冷媒を圧縮すると、圧縮機１０の異常が起こり得る。よって、圧縮機１０に吸入される冷媒がガス冷媒（気相単相の冷媒）であることは必須である。しかし、ヒートポンプシステム１００Ｂの暖房運転時に主膨張弁３０を通過した冷媒は気液二相冷媒である。このため、主膨張弁３０から三方弁６０に至り、さらに圧縮機１０に吸入される冷媒は、気液二相状態である場合がある。 Here, even if the liquid refrigerant is compressed, the volume of the liquid refrigerant does not change. Therefore, when an excessive liquid refrigerant is compressed by the compressor 10, an abnormality of the compressor 10 may occur. Therefore, it is essential that the refrigerant sucked into the compressor 10 is a gas refrigerant (gas phase single-phase refrigerant). However, the refrigerant that has passed through the main expansion valve 30 during the heating operation of the heat pump system 100B is a gas-liquid two-phase refrigerant. For this reason, the refrigerant from the main expansion valve 30 to the three-way valve 60 and further sucked into the compressor 10 may be in a gas-liquid two-phase state.
　そこで、実施の形態２においては、たとえば冷媒温度が基準値Ｔ３以上の場合には、冷媒温度が基準値Ｔ３未満の場合と比べて、副膨張弁７０の開度を大きく設定することにより、内部熱交換器８０を流れる冷媒の流量を増加させる。これにより、圧縮機１０から吐出された高温高圧のガス冷媒と、主膨張弁３０から三方弁６０へと至る気液二相冷媒との間で交換される熱量が大きくなる。その結果、圧縮機１０に吸入される冷媒を完全にガス化させることが可能になるので、圧縮機１０の異常を防止することができる。 Therefore, in the second embodiment, for example, when the refrigerant temperature is equal to or higher than the reference value T3, the opening of the sub expansion valve 70 is set larger than in the case where the refrigerant temperature is lower than the reference value T3. The flow rate of the refrigerant flowing through the heat exchanger 80 is increased. As a result, the amount of heat exchanged between the high-temperature and high-pressure gas refrigerant discharged from the compressor 10 and the gas-liquid two-phase refrigerant from the main expansion valve 30 to the three-way valve 60 increases. As a result, the refrigerant sucked into the compressor 10 can be completely gasified, so that an abnormality of the compressor 10 can be prevented.
　図７は、実施の形態２にて実行される三方弁６０および副膨張弁７０の制御を説明するためのフローチャートである。図６および図７を参照して、Ｓ１０において、制御装置５００は、室外熱交換器２０における蒸発圧力Ｐｅを算出する。蒸発圧力Ｐｅは、たとえば温度センサ９４を用いて検出された室内熱交換器４０の蒸発温度を飽和圧力に換算することにより算出することができる。 FIG. 7 is a flowchart for explaining the control of the three-way valve 60 and the sub-expansion valve 70 executed in the second embodiment. Referring to FIGS. 6 and 7, in S <b> 10, control device 500 calculates evaporation pressure Pe in outdoor heat exchanger 20. The evaporation pressure Pe can be calculated, for example, by converting the evaporation temperature of the indoor heat exchanger 40 detected using the temperature sensor 94 into a saturation pressure.
　Ｓ１２０において、制御装置５００は、蒸発圧力Ｐｅが基準値Ｐ３（第３の基準値）以上であるか否かを判定する。蒸発圧力Ｐｅが基準値Ｐ３以上の場合（Ｓ１２０においてＹＥＳ）、制御装置５００は、バイパス流路１２０へと向かう冷媒の流量が大きくなる方向に三方弁６０の弁体６１の角度θを調整する。これにより、圧縮機１０から吐出される冷媒の流量に対する室外熱交換器２０を流れる冷媒の流量の比率Ｒが低下する（Ｓ１３０）。その結果、室外熱交換器２０を通過せずに室外熱交換器２０の下流側へと至る冷媒の割合が増加するため、蒸発圧力Ｐｅが低下する。よって、蒸発圧力Ｐｅの上昇を抑制することができる。 In S120, the control device 500 determines whether or not the evaporation pressure Pe is equal to or higher than the reference value P3 (third reference value). When the evaporation pressure Pe is equal to or higher than the reference value P3 (YES in S120), the control device 500 adjusts the angle θ of the valve body 61 of the three-way valve 60 in the direction in which the flow rate of the refrigerant toward the bypass flow path 120 increases. Thereby, the ratio R of the flow rate of the refrigerant flowing through the outdoor heat exchanger 20 with respect to the flow rate of the refrigerant discharged from the compressor 10 decreases (S130). As a result, since the ratio of the refrigerant that does not pass through the outdoor heat exchanger 20 and reaches the downstream side of the outdoor heat exchanger 20 increases, the evaporation pressure Pe decreases. Therefore, the increase in the evaporation pressure Pe can be suppressed.
　その一方で、室外熱交換器２０における外気から冷媒への吸熱量が小さくなるため、圧縮機１０に流入する冷媒の加熱度ΔＴｅは小さくなる。そのため、Ｓ１４０において、制御装置５００は、圧縮機１０に流入する冷媒の加熱度ΔＴｅを算出する。加熱度ΔＴｅは、たとえば温度センサ９３により検出された冷媒温度（内部熱交換器８０の出口側における冷媒温度）に基づき算出することができる。 On the other hand, since the amount of heat absorbed from the outside air to the refrigerant in the outdoor heat exchanger 20 becomes small, the degree of heating ΔTe of the refrigerant flowing into the compressor 10 becomes small. Therefore, in S140, control device 500 calculates the heating degree ΔTe of the refrigerant flowing into compressor 10. The degree of heating ΔTe can be calculated based on, for example, the refrigerant temperature detected by the temperature sensor 93 (the refrigerant temperature at the outlet side of the internal heat exchanger 80).
　Ｓ１５０において、制御装置５００は、過熱度ΔＴｅが基準値Ｔ４（第４の基準値）未満であるか否かを判定する。過熱度ΔＴｅが基準値Ｔ４以上の場合（Ｓ１５０においてＮＯ）、制御装置５００は、圧縮機１０に流入する冷媒について十分な過熱度ΔＴｅが確保されているとして、副膨張弁７０Ｂの開度を維持する（あるいは副膨張弁７０Ｂの開度を小さく調整する）（Ｓ１７０）。 In S150, the control device 500 determines whether or not the degree of superheat ΔTe is less than the reference value T4 (fourth reference value). When superheat degree ΔTe is equal to or greater than reference value T4 (NO in S150), control device 500 maintains the opening degree of sub-expansion valve 70B on the assumption that sufficient superheat degree ΔTe is secured for the refrigerant flowing into compressor 10. (Or adjust the opening of the sub-expansion valve 70B to be small) (S170).
　これに対し、過熱度ΔＴｅが基準値Ｔ４未満の場合（Ｓ１５０においてＹＥＳ）、制御装置５００は、過熱度ΔＴｅが基準値Ｔ４以上の場合と比べて、副膨張弁７０Ｂの開度を大きく設定する（Ｓ１６０）。これにより、分岐流路１３０Ｂを流れる冷媒の流量が増加するので、内部熱交換器８０Ｂにおいて、圧縮機１０から吐出された高温高圧のガス冷媒と、主膨張弁３０から三方弁６０へと至る気液二相冷媒との間で交換される熱量が大きくなる。よって、過熱度ΔＴｅが基準値Ｔ４以上の場合と比べて、主回路１１０を流れる冷媒の温度が高くなる。したがって、圧縮機１０に流入する冷媒の過熱度ΔＴｅをより確実に確保することができる。 In contrast, when superheat degree ΔTe is less than reference value T4 (YES in S150), control device 500 sets the opening degree of sub-expansion valve 70B to be larger than when superheat degree ΔTe is greater than or equal to reference value T4. (S160). As a result, the flow rate of the refrigerant flowing through the branch flow path 130B is increased, so that the internal heat exchanger 80B and the high-temperature and high-pressure gas refrigerant discharged from the compressor 10 and the gas from the main expansion valve 30 to the three-way valve 60 are discharged. The amount of heat exchanged with the liquid two-phase refrigerant increases. Therefore, the temperature of the refrigerant flowing through the main circuit 110 is higher than when the degree of superheat ΔTe is equal to or greater than the reference value T4. Therefore, the degree of superheat ΔTe of the refrigerant flowing into the compressor 10 can be ensured more reliably.
　なお、実施の形態２と変形例とを組み合わせてもよい。すなわち、実施の形態２においても、冷媒の一部のみが室外熱交換器を迂回可能な構成を採用することができる。 In addition, you may combine Embodiment 2 and a modification. That is, also in Embodiment 2, it is possible to adopt a configuration in which only a part of the refrigerant can bypass the outdoor heat exchanger.
　また、実施の形態１，２では、１台の室内機のみが含まれるヒートポンプシステムについて説明した。しかし、本発明は、複数台の室内機（すなわち複数の主膨張弁および複数の室内熱交換器）を含むヒートポンプシステムにも適用可能である。そのようなシステムの例としては、ビル用のマルチ空調システムが挙げられる。 In the first and second embodiments, the heat pump system including only one indoor unit has been described. However, the present invention is also applicable to a heat pump system including a plurality of indoor units (that is, a plurality of main expansion valves and a plurality of indoor heat exchangers). An example of such a system is a multi air conditioning system for buildings.
　本発明は、マルチ空調システムの冷房主体運転において、冷房および暖房の混在運転が行なわれる場合にも適用することができる。この場合、冷房中の室内機における吸熱量と暖房中の室内機における放熱量との差が小さくなるに従い、室外熱交換器における放熱量が小さくなる。そのため、場合によっては、１台の室内機のみを含むシステムの冷房運転時と比べて、室外熱交換器での放熱量を小さくすることが望ましい。本発明は、このような場合に適用することができる。 The present invention can also be applied to a case where mixed operation of cooling and heating is performed in the cooling main operation of the multi-air conditioning system. In this case, as the difference between the heat absorption amount in the cooling indoor unit and the heat dissipation amount in the heating indoor unit becomes smaller, the heat dissipation amount in the outdoor heat exchanger becomes smaller. Therefore, in some cases, it is desirable to reduce the heat radiation amount in the outdoor heat exchanger as compared with the cooling operation of the system including only one indoor unit. The present invention can be applied to such a case.
　今回開示された実施の形態はすべての点で例示であって制限的なものではないと考えられるべきである。本発明の範囲は、上記した説明ではなく、請求の範囲によって示され、請求の範囲と均等の意味および範囲内でのすべての変更が含まれることが意図される。 The embodiment disclosed this time should be considered as illustrative in all points and not restrictive. The scope of the present invention is defined by the terms of the claims, rather than the description above, and is intended to include any modifications within the scope and meaning equivalent to the terms of the claims.
　１００，１００Ａ，１００Ｂ　ヒートポンプシステム、１０　圧縮機、２０，２０Ａ　室外熱交換器、３０　主膨張弁、４０　室内熱交換器、５０　配管、６０，６０Ａ　三方弁、６１　弁体、７０　副膨張弁、８０，８０Ｂ　内部熱交換器、９１　圧力センサ、９２～９４　温度センサ、１１０　主回路、１２０，１２０Ａ　バイパス流路、１３０，１３０Ｂ　分岐流路、５００　制御装置。 100, 100A, 100B heat pump system, 10 compressor, 20, 20A outdoor heat exchanger, 30 main expansion valve, 40 indoor heat exchanger, 50 piping, 60, 60A three-way valve, 61 valve body, 70 sub expansion valve, 80 , 80B internal heat exchanger, 91 pressure sensor, 92-94 temperature sensor, 110 main circuit, 120, 120A bypass flow path, 130, 130B branch flow path, 500 control device.

Claims (7)

  1. 　圧縮機、室外熱交換器、第１の膨張弁および室内熱交換器を含み、冷媒を循環可能に構成された主回路と、
    　前記主回路を流れる冷媒が前記室外熱交換器の少なくとも一部を迂回するように構成されたバイパス流路と、
    　前記圧縮機から吐出される冷媒の流量に対する前記室外熱交換器を流れる冷媒の流量の比率を調整可能に構成された流量調整器と、
    　前記室外熱交換器と前記第１の膨張弁との間に接続され、前記主回路を流れる冷媒の一部を分岐可能に構成された分岐流路と、
    　前記分岐流路を流れる冷媒の流量を調整可能に構成された第２の膨張弁と、
    　前記分岐流路を流れる冷媒と、前記室外熱交換器から前記第１の膨張弁へと流れる冷媒との間で熱交換する副熱交換器とを備える、ヒートポンプシステム。     A main circuit including a compressor, an outdoor heat exchanger, a first expansion valve, and an indoor heat exchanger configured to circulate refrigerant;
    A bypass passage configured such that the refrigerant flowing through the main circuit bypasses at least a part of the outdoor heat exchanger;
    A flow rate regulator configured to be capable of adjusting a ratio of a flow rate of the refrigerant flowing through the outdoor heat exchanger to a flow rate of the refrigerant discharged from the compressor;
    A branch flow path connected between the outdoor heat exchanger and the first expansion valve and configured to be able to branch a part of the refrigerant flowing through the main circuit;
    A second expansion valve configured to be capable of adjusting the flow rate of the refrigerant flowing through the branch flow path;
    A heat pump system comprising: a refrigerant flowing through the branch flow path; and a sub heat exchanger that exchanges heat between the refrigerant flowing from the outdoor heat exchanger to the first expansion valve.
  2. 　前記流量調整器および前記第２の膨張弁を制御する制御装置をさらに備え、
    　前記制御装置は、前記室外熱交換器における冷媒の凝縮度合を示す状態値が第１の基準値を下回る場合には、
    　　前記状態値が前記第１の基準値を上回る場合と比べて、前記室外熱交換器を流れる冷媒の比率が小さくなるように前記流量調整器を制御し、かつ、
    　　前記第１の膨張弁に流入する冷媒の過冷却度が第２の基準値を下回るときには、前記過冷却度が前記第２の基準値を上回るときと比べて、前記副熱交換器を流れる冷媒の流量が大きくなるように前記第２の膨張弁を制御する、請求項１に記載のヒートポンプシステム。     A controller for controlling the flow regulator and the second expansion valve;
    When the state value indicating the degree of condensation of the refrigerant in the outdoor heat exchanger is lower than the first reference value, the control device,
    Controlling the flow regulator so that the ratio of the refrigerant flowing through the outdoor heat exchanger is smaller than when the state value exceeds the first reference value; and
    When the degree of supercooling of the refrigerant flowing into the first expansion valve is lower than the second reference value, the refrigerant flowing through the auxiliary heat exchanger is smaller than when the degree of supercooling is higher than the second reference value. The heat pump system according to claim 1, wherein the second expansion valve is controlled so that the flow rate of the gas increases.
  3. 　前記第２の基準値は、前記室内熱交換器が前記室外熱交換器よりも高い位置に設けられた場合であっても、前記第１の膨張弁に流入する冷媒が過冷却状態を維持するように設定される、請求項２に記載のヒートポンプシステム。 The second reference value is that the refrigerant flowing into the first expansion valve maintains a supercooled state even when the indoor heat exchanger is provided at a position higher than the outdoor heat exchanger. The heat pump system according to claim 2 set up as follows.
  4. 　前記状態値は、前記室外熱交換器を流れる冷媒の温度、前記室外熱交換器における冷媒の凝縮圧力、前記圧縮機から吐出される冷媒の温度、および、前記圧縮機における冷媒の圧縮比のうちの少なくとも１つを含む、請求項２または３に記載のヒートポンプシステム。 The state value includes the temperature of the refrigerant flowing through the outdoor heat exchanger, the condensation pressure of the refrigerant in the outdoor heat exchanger, the temperature of the refrigerant discharged from the compressor, and the compression ratio of the refrigerant in the compressor The heat pump system according to claim 2 or 3, comprising at least one of the following.
  5. 　圧縮機、室外熱交換器、第１の膨張弁および室内熱交換器を含み、冷媒を循環可能に構成された主回路と、
    　前記主回路を流れる冷媒が前記室外熱交換器の少なくとも一部を迂回するように構成されたバイパス流路と、
    　前記圧縮機から吐出される冷媒の流量に対する前記室外熱交換器を流れる冷媒の流量の比率を調整可能に構成された流量調整器と、
    　前記圧縮機と前記室内熱交換器との間に接続され、前記主回路を流れる冷媒の一部を分岐可能に構成された分岐流路と、
    　前記分岐流路を流れる冷媒の流量を調整可能に構成された第２の膨張弁と、
    　前記分岐流路を流れる冷媒と、前記室外熱交換器から前記圧縮機へと流れる冷媒との間で熱交換する副熱交換器とを備える、ヒートポンプシステム。     A main circuit including a compressor, an outdoor heat exchanger, a first expansion valve, and an indoor heat exchanger configured to circulate refrigerant;
    A bypass passage configured such that the refrigerant flowing through the main circuit bypasses at least a part of the outdoor heat exchanger;
    A flow rate regulator configured to be capable of adjusting a ratio of a flow rate of the refrigerant flowing through the outdoor heat exchanger to a flow rate of the refrigerant discharged from the compressor;
    A branch passage connected between the compressor and the indoor heat exchanger and configured to be able to branch a part of the refrigerant flowing through the main circuit;
    A second expansion valve configured to be capable of adjusting the flow rate of the refrigerant flowing through the branch flow path;
    A heat pump system comprising: a refrigerant flowing through the branch flow path; and a sub heat exchanger that exchanges heat between the refrigerant flowing from the outdoor heat exchanger to the compressor.
  6. 　前記流量調整器および前記第２の膨張弁を制御する制御装置をさらに備え、
    　前記制御装置は、前記冷媒の蒸発圧力が第３の基準値を上回る場合には、
    　前記蒸発圧力が前記第３の基準値を下回る場合と比べて、前記室外熱交換器を流れる冷媒の比率が小さくなるように前記流量調整器を制御し、かつ、
    　前記第１の膨張弁に流入する冷媒の過熱度が第４の基準値を下回るときには、前記過熱度が前記第４の基準値を上回るときと比べて、前記副熱交換器を流れる冷媒の流量が大きくなるように前記第２の膨張弁を制御する、請求項５に記載のヒートポンプシステム。     A controller for controlling the flow regulator and the second expansion valve;
    The control device, when the evaporation pressure of the refrigerant exceeds a third reference value,
    Controlling the flow regulator so that the ratio of the refrigerant flowing through the outdoor heat exchanger is smaller than when the evaporation pressure is lower than the third reference value; and
    When the superheat degree of the refrigerant flowing into the first expansion valve is lower than the fourth reference value, the flow rate of the refrigerant flowing through the auxiliary heat exchanger is smaller than when the superheat degree is higher than the fourth reference value. The heat pump system according to claim 5, wherein the second expansion valve is controlled so as to increase.
  7. 　前記第４の基準値は、前記圧縮機に流入する冷媒がガス冷媒となるように設定される、請求項６に記載のヒートポンプシステム。 The heat pump system according to claim 6, wherein the fourth reference value is set so that the refrigerant flowing into the compressor becomes a gas refrigerant.
PCT/JP2015/079571 2015-10-20 2015-10-20 Heat pump system WO2017068649A1 (en)

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CN111383963A (en) * 2018-12-26 2020-07-07 东京毅力科创株式会社 Temperature adjustment device and method for controlling temperature adjustment device
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