WO2022038951A1 - Refrigeration cycle device - Google Patents

Abstract

Provided is a refrigeration cycle device comprising a heating unit (12, 121, 50), a first decompression unit (13), an outdoor heat-exchange unit (14), a second decompression unit (16), and an indoor evaporation unit (17). The heating unit (12, 121, 50) heats air to be blown. The first decompression unit (13) decompresses a refrigerant flowing out from the heating unit (12, 121, 50). The outdoor heat-exchange unit (14) performs heat exchange between the refrigerant flowing out from the first decompression unit (13) and outside air. The second decompression unit (16) decompresses the refrigerant flowing out from the outdoor heat-exchange unit (14). The indoor evaporation unit (17) cools the air to be blown before said air is heated at the heating unit (12, 121, 50). A first decompression control unit (40b) controls the operation of the first decompression unit (13) such that the supercooling degree (SC1) of the refrigerant flowing into the first decompression unit (13) is a target supercooling degree (SCO) or less.

Description

Refrigeration cycle device Cross-reference of related applications

This application is based on Japanese Patent Application No. 2020-139066 filed on August 20, 2020, and the contents of the description are incorporated herein by reference.

The present disclosure is suitable for use in an air conditioner that dehumidifies and heats an air-conditioned space with respect to a refrigeration cycle device.

Conventionally, Patent Document 1 discloses a refrigeration cycle device applied to a vehicle air conditioner for dehumidifying and heating a vehicle interior, which is an air conditioning target space.

The refrigeration cycle device of Patent Document 1 includes a plurality of heat exchange units such as an indoor condenser, an outdoor heat exchanger, and an indoor evaporator. The indoor condenser is a heating unit that heats the blown air blown into the vehicle interior by using the discharged refrigerant discharged from the compressor as a heat source. Further, the refrigeration cycle apparatus of Patent Document 1 includes a first decompression unit for reducing the pressure of the refrigerant flowing out of the indoor condenser and a second decompression unit for reducing the pressure of the refrigerant flowing out of the outdoor heat exchanger.

In the refrigeration cycle device of Patent Document 1, in the dehumidifying / heating mode for dehumidifying / heating the vehicle interior, three heat exchange units are connected in series in the order of the indoor condenser, the outdoor heat exchanger, and the indoor evaporator from the upstream side of the refrigerant flow. Switch to the refrigerant circuit to be connected. In the vehicle air conditioner of Patent Document 1, dehumidifying and heating of the vehicle interior is realized by reheating the blown air cooled and dehumidified by the indoor evaporator with the indoor condenser and blowing it into the vehicle interior. is doing.

Further, in the refrigeration cycle apparatus of Patent Document 1, the heating capacity of the blown air in the indoor condenser is adjusted by changing the throttle opening of the first decompression section and the throttle opening of the second decompression section in the dehumidifying / heating mode. is doing.

Japanese Patent No. 5585549

By the way, in the refrigerating cycle device of Patent Document 1, in the dehumidifying and heating mode, the throttle opening of the first decompression unit is changed in order to adjust the heating capacity of the blown air in the indoor condenser. For example, when improving the heating capacity of the blown air in the indoor condenser, the throttle opening of the first decompression unit is reduced. Therefore, as the heating capacity of the blown air in the indoor condenser is improved, the degree of supercooling of the refrigerant on the outlet side of the indoor condenser tends to increase.

However, if the degree of supercooling of the refrigerant on the outlet side of the indoor condenser becomes unnecessarily large, a temperature distribution will occur in the blown air heated by the indoor condenser, realizing comfortable dehumidification and heating in the vehicle interior. I can't do it. Further, if the degree of supercooling of the refrigerant on the outlet side of the indoor condenser becomes unnecessarily large, the coefficient of performance of the cycle (that is, COP) may decrease.

In view of the above points, it is an object of the present disclosure to provide a refrigerating cycle apparatus capable of appropriately adjusting the degree of supercooling of the refrigerant on the outlet side of the heating unit.

In order to achieve the above object, the refrigerating cycle apparatus of the first aspect of the present disclosure includes a compressor, a heating unit, a first decompression unit, an outdoor heat exchange unit, a second decompression unit, and an indoor evaporation unit. A target supercooling degree determination unit, a supercooling degree estimation unit, and a first decompression control unit are provided.

The compressor compresses and discharges the refrigerant. The heating unit heats the blown air blown to the air-conditioned space using the refrigerant discharged from the compressor as a heat source. The first decompression unit decompresses the refrigerant flowing out of the heating unit. The outdoor heat exchange unit exchanges heat between the refrigerant flowing out from the first decompression unit and the outside air. The second decompression unit decompresses the refrigerant flowing out from the outdoor heat exchange unit. The indoor evaporating unit evaporates the refrigerant decompressed by the second decompression unit to cool the blown air before being heated by the heating unit. The target supercooling degree determining unit determines the target supercooling degree of the refrigerant flowing into the first decompression unit. The supercooling degree estimation unit estimates the supercooling degree of the refrigerant flowing into the first decompression unit. The first decompression control unit controls the operation of the first decompression unit.

Further, the first decompression control unit controls the operation of the first decompression unit so that the supercooling degree estimated by the supercooling degree estimation unit is equal to or less than the target supercooling degree.

According to this, the first decompression control unit controls the operation of the first decompression unit so that the supercooling degree becomes equal to or less than the target supercooling degree. Therefore, it is possible to prevent the degree of supercooling of the refrigerant flowing out of the heating section and flowing into the first decompression section from becoming unnecessarily high.

That is, according to the refrigerating cycle device of the first aspect, the degree of supercooling of the refrigerant on the outlet side of the heating unit can be appropriately adjusted.

Further, the refrigerating cycle apparatus according to the second aspect of the present disclosure includes a compressor, a heating unit, a first decompression unit, an outdoor heat exchange unit, a second decompression unit, an indoor evaporation unit, and a target supercooling degree. It includes a determination unit, a lower limit area calculation unit, and a first decompression control unit.

The compressor compresses and discharges the refrigerant. The heating unit heats the blown air blown to the air-conditioned space using the refrigerant discharged from the compressor as a heat source. The first decompression unit decompresses the refrigerant flowing out of the heating unit. The outdoor heat exchange unit exchanges heat between the refrigerant flowing out from the first decompression unit and the outside air. The second decompression unit decompresses the refrigerant flowing out from the outdoor heat exchange unit. The indoor evaporating unit evaporates the refrigerant decompressed by the second decompression unit to cool the blown air before being heated by the heating unit. The target supercooling degree determining unit determines the target supercooling degree of the refrigerant flowing into the first decompression unit. The lower limit area calculation unit calculates the lower limit throttle passage area of the first decompression unit where the supercooling degree of the refrigerant flowing into the first decompression unit is the target supercooling degree. The first decompression control unit controls the operation of the first decompression unit.

Further, the first decompression control unit controls the operation of the first decompression unit so that the throttle passage area of the first decompression unit is equal to or larger than the lower limit throttle passage area.

According to this, the first decompression control unit controls the operation of the first decompression unit so that the throttle passage area of the first decompression unit is equal to or larger than the lower limit throttle passage area, so that the supercooling degree is equal to or less than the target supercooling degree. Can be. Therefore, it is possible to prevent the degree of supercooling of the refrigerant flowing out of the heating section and flowing into the first decompression section from becoming unnecessarily high.

That is, according to the refrigerating cycle device of the second aspect, the degree of supercooling of the refrigerant on the outlet side of the heating unit can be appropriately adjusted.

It is an overall block diagram of the vehicle air conditioner of 1st Embodiment. It is a block diagram which shows the electric control part of the air-conditioning apparatus for a vehicle of 1st Embodiment. It is a flowchart which shows a part of the control flow of the air conditioner for a vehicle of 1st Embodiment. It is a graph which shows the relationship between the supercooling degree of a refrigerant, and the density. It is a Moriel diagram which shows the change of the state of the refrigerant in the dehumidifying heating mode of the refrigerating cycle apparatus of 1st Embodiment. It is a flowchart which shows a part of the control flow of the air conditioner for a vehicle of 2nd Embodiment. It is an overall block diagram of the vehicle air conditioner of 3rd Embodiment. It is an overall block diagram of the vehicle air conditioner of 4th Embodiment. It is a graph which shows the temperature change of a refrigerant and a heat medium in a water refrigerant heat exchanger.

Hereinafter, a plurality of embodiments for carrying out the present disclosure will be described with reference to the drawings. In each embodiment, the same reference numerals may be given to the parts corresponding to the matters described in the preceding embodiments, and duplicate explanations may be omitted. When only a part of the configuration is described in each embodiment, other embodiments described above can be applied to the other parts of the configuration. Not only the combination of the parts that clearly indicate that the combination is possible in each embodiment, but also the partial combination of the embodiments even if the combination is not specified if there is no problem in the combination. Is also possible.

(First Embodiment)
A first embodiment of the refrigeration cycle apparatus 10 according to the present disclosure will be described with reference to FIGS. 1 to 5. The refrigeration cycle device 10 is applied to the vehicle air conditioner 1 shown in the overall configuration diagram of FIG. The refrigeration cycle device 10 adjusts the temperature of the blown air blown into the vehicle interior, which is the space to be air-conditioned, in the vehicle air-conditioning device 1. The vehicle air conditioner 1 includes a refrigeration cycle device 10, an indoor air conditioner unit 30, a control device 40, and the like.

The refrigeration cycle apparatus 10 uses an HFO-based refrigerant (specifically, R1234yf) as the refrigerant. The refrigeration cycle device 10 constitutes a steam compression type subcritical refrigeration cycle in which the pressure of the high-pressure refrigerant discharged from the compressor 11 does not exceed the critical pressure of the refrigerant. Refrigerating machine oil (specifically, PAG oil) for lubricating the compressor 11 is mixed in the refrigerant. Some of the refrigerating machine oil circulates in the cycle together with the refrigerant.

The compressor 11 sucks in the refrigerant in the refrigerating cycle device 10, compresses it, and discharges it. The compressor 11 is arranged in the vehicle bonnet, which is an auxiliary machine room on the front side of the vehicle room. The compressor 11 is an electric compressor that rotationally drives a fixed-capacity compression mechanism having a fixed discharge capacity by an electric motor. The number of revolutions (that is, the refrigerant discharge capacity) of the compressor 11 is controlled by a control signal output from the control device 40 described later.

The refrigerant inlet side of the indoor condenser 12 is connected to the discharge port of the compressor 11. The indoor condenser 12 is arranged in the casing 31 of the indoor air conditioning unit 30, which will be described later. The indoor condenser 12 is a heat exchange unit that exchanges heat between the high-pressure refrigerant discharged from the compressor 11 and the blown air, and dissipates the heat of the high-pressure refrigerant to the blown air. Further, the indoor condenser 12 is a heating unit that heats the blown air using the high-temperature and high-pressure discharged refrigerant discharged from the compressor 11 as a heat source.

The inlet side of the heating expansion valve 13 is connected to the refrigerant outlet of the indoor condenser 12. The heating expansion valve 13 is a first decompression unit that depressurizes the refrigerant flowing out of the indoor condenser 12.

The heating expansion valve 13 is an electrically variable type having a valve body portion that changes the opening degree of the throttle passage (that is, the valve opening degree) and an electric actuator (specifically, a stepping motor) that displaces the valve body portion. It is an aperture mechanism. The operation of the heating expansion valve 13 is controlled by a control signal (specifically, a control pulse) output from the control device 40.

The heating expansion valve 13 has a fully open function in which the valve body portion fully opens the valve opening so as to function as a mere refrigerant passage without exerting a flow rate adjusting action and a refrigerant depressurizing action.

The refrigerant inlet side of the outdoor heat exchanger 14 is connected to the outlet of the heating expansion valve 13. The outdoor heat exchanger 14 is an outdoor heat exchange unit that exchanges heat between the refrigerant flowing out from the heating expansion valve 13 and the outside air blown from an outside air fan (not shown). The outdoor heat exchanger 14 is arranged on the front side in the vehicle bonnet. Therefore, when the vehicle is running, the running wind that has flowed into the vehicle bonnet through the grill can be applied to the outdoor heat exchanger 14.

The inlet side of the receiver 15 is connected to the refrigerant outlet of the outdoor heat exchanger 14. The receiver 15 is a liquid storage unit on the high pressure side having a gas-liquid separation function. The receiver 15 separates the gas and liquid of the refrigerant flowing out from the heat exchange unit that functions as a condenser that condenses the refrigerant in the refrigeration cycle device 10. Further, the receiver 15 causes a part of the separated liquid phase refrigerant to flow out to the downstream side, and stores the remaining liquid phase refrigerant as the surplus refrigerant in the cycle.

The inlet side of the cooling expansion valve 16 is connected to the outlet of the receiver 15. The cooling expansion valve 16 is a second pressure reducing unit that reduces the pressure of the refrigerant flowing out of the receiver 15. The basic configuration of the cooling expansion valve 16 is the same as that of the heating expansion valve 13.

The refrigerant inlet side of the indoor evaporator 17 is connected to the outlet of the cooling expansion valve 16. The indoor evaporator 17 is arranged in the casing 31 of the indoor air conditioning unit 30. The indoor evaporator 17 is a heat exchange unit that evaporates the low-pressure refrigerant decompressed by the cooling expansion valve 16 by exchanging heat with the blown air blown from the indoor blower 32. Further, the indoor evaporator 17 is an indoor evaporation unit that cools the blown air by evaporating the low-pressure refrigerant to exert an endothermic action. The suction port side of the compressor 11 is connected to the refrigerant outlet of the indoor evaporator 17.

Next, the indoor air conditioning unit 30 will be described. The indoor air-conditioning unit 30 is a unit in which various components are integrated in the vehicle air-conditioning device 1 in order to blow out appropriately temperature-controlled blown air to an appropriate place in the vehicle interior. The indoor air conditioning unit 30 is arranged inside the instrument panel (that is, the instrument panel) at the front of the vehicle interior.

The indoor air conditioning unit 30 has a casing 31 that forms an air passage for blown air. An indoor blower 32, an indoor evaporator 17, an indoor condenser 12, and the like are arranged in an air passage formed in the casing 31. The casing 31 is made of a resin (for example, polypropylene) having a certain degree of elasticity and excellent strength.

An inside / outside air switching device 33 is arranged on the most upstream side of the blast air flow of the casing 31. The inside / outside air switching device 33 switches and introduces the inside air (that is, the vehicle interior air) and the outside air (that is, the vehicle interior outside air) into the casing 31. The operation of the electric actuator for driving the inside / outside air switching device 33 is controlled by the control signal output from the control device 40.

An indoor blower 32 is arranged on the downstream side of the blower air flow of the inside / outside air switching device 33. The indoor blower 32 is a blower unit that blows the air sucked through the inside / outside air switching device 33 toward the vehicle interior. The indoor blower 32 is an electric blower that drives a centrifugal multi-blade fan with an electric motor. The rotation speed (that is, the blowing capacity) of the indoor blower 32 is controlled by the control voltage output from the control device 40.

On the downstream side of the blown air flow of the indoor blower 32, the indoor evaporator 17 and the indoor condenser 12 are arranged in order from the upstream side with respect to the blown air flow. That is, the indoor evaporator 17 is arranged on the upstream side of the blown air flow with respect to the indoor condenser 12. A cold air bypass passage 35 is formed in the casing 31 to allow the blown air that has passed through the indoor evaporator 17 to bypass the indoor condenser 12 and flow to the downstream side.

The air mix door 34 is arranged on the downstream side of the blown air flow of the indoor evaporator 17 and on the upstream side of the blown air flow of the indoor condenser 12. The air mix door 34 adjusts the air volume ratio between the air volume of the blown air passing through the indoor condenser 12 and the air volume of the blown air passing through the cold air bypass passage 35. The operation of the electric actuator for driving the air mix door 34 is controlled by the control signal output from the control device 40.

On the downstream side of the blown air flow of the indoor condenser 12, the blown air heated by the indoor condenser 12 and the blown air that has passed through the cold air bypass passage 35 and are not heated by the indoor condenser 12 are mixed. Space 36 is provided. Further, an opening hole (not shown) for blowing out the blown air (air-conditioned air) mixed in the mixing space 36 into the vehicle interior is arranged at the most downstream portion of the blown air flow of the casing 31.

Therefore, the temperature of the conditioned air mixed in the mixing space 36 is adjusted by adjusting the air volume ratio between the air volume passing through the indoor condenser 12 and the air volume passing through the cold air bypass passage 35 by the air mix door 34. Can be done. Then, the temperature of the blown air blown from each opening hole into the vehicle interior can be adjusted.

As the opening holes, a face opening hole, a foot opening hole, and a defroster opening hole (none of which are shown) are provided. The face opening hole is an opening hole for blowing air-conditioned air toward the upper body of the occupant in the vehicle interior. The foot opening hole is an opening hole for blowing air-conditioned air toward the feet of the occupant. The defroster opening hole is an opening hole for blowing air conditioning air toward the inner side surface of the front window glass of the vehicle.

An outlet mode switching door (not shown) is arranged on the upstream side of these opening holes. The blowing mode switching door switches the opening hole for blowing out the conditioned air by opening and closing each opening hole. The operation of the electric actuator for driving the blowout mode switching door is controlled by the control signal output from the control device 40.

Next, the outline of the electric control unit of the vehicle air conditioner 1 will be described with reference to FIG. The control device 40 includes a well-known microcomputer including a CPU, ROM, RAM, and the like, and peripheral circuits thereof. The control device 40 performs various calculations and processes based on the control program stored in the ROM, and controls the operation of various controlled target devices 11, 13, 16, 32, 33, 34, etc. connected to the output side. ..

As shown in FIG. 2, various sensors used for air conditioning control are connected to the input side of the control device 40. Various sensors include an inside temperature sensor 41a, an outside temperature sensor 41b, a solar radiation amount sensor 41c, a high pressure temperature sensor 41d, a high pressure pressure sensor 41e, a low pressure temperature sensor 41f, a low pressure pressure sensor 41g, and an air conditioning air temperature sensor 41h.

The internal air temperature sensor 41a is an internal air temperature detection unit that detects the internal air temperature Tr, which is the temperature inside the vehicle. The outside air temperature sensor 41b is an outside air temperature detection unit that detects the outside air temperature Tam, which is the temperature outside the vehicle interior. The solar radiation amount sensor 41c is a solar radiation amount detection unit that detects the solar radiation amount As irradiated to the vehicle interior.

The high-pressure temperature sensor 41d is a high-pressure temperature detection unit that detects the discharge temperature Td of the discharged refrigerant discharged from the compressor 11. The high-pressure pressure sensor 41e is a high-pressure pressure detection unit that detects the discharge pressure Pd of the discharge refrigerant discharged from the compressor 11.

The high pressure temperature sensor 41d and the high pressure pressure sensor 41e are also used to detect an abnormal rise in the discharge temperature Td or the discharge pressure Pd. When the control device 40 detects an abnormal rise in the discharge temperature Td or the discharge pressure Pd, the control device 40 stops the compressor 11 to perform compressor protection control for protecting the compressor 11.

The low pressure temperature sensor 41f is a low pressure temperature detection unit that detects the suction temperature Ts of the suction refrigerant sucked into the compressor 11. The low pressure pressure sensor 41g is a low pressure pressure detection unit that detects the suction pressure Ps of the suction refrigerant sucked into the compressor 11.

The low-pressure temperature sensor 41f of the present embodiment specifically detects the temperature of the outer surface of the refrigerant pipe from the refrigerant outlet of the indoor evaporator 17 to the suction port of the compressor 11. As the low-pressure temperature sensor, an evaporator temperature detection unit that detects the evaporator temperature Tefin, which is the temperature of the heat exchange fin temperature of the indoor evaporator 17, may be adopted.

The air-conditioned air temperature sensor 41h is an air-conditioned air temperature detection unit that detects the air blown air temperature TAV blown out from the mixing space 36 into the vehicle interior.

Further, an operation panel 42 arranged near the instrument panel in the front part of the vehicle interior is connected to the input side of the control device 40, and operation signals from various operation switches provided on the operation panel 42 are input. Specific examples of the various operation switches provided on the operation panel 42 include an auto switch, an air conditioner switch, an air volume setting switch, and a temperature setting switch.

The auto switch is an automatic control requesting unit for requesting that the occupant set or cancel the automatic control operation of the vehicle air conditioner 1. The air conditioner switch is a cooling requesting unit for requiring the occupant to cool the blown air with the indoor evaporator 17. The air volume setting switch is an air volume setting unit in which the occupant manually sets the air volume of the indoor blower 32. The temperature setting switch is a temperature setting unit in which the occupant sets the target temperature Tset in the vehicle interior.

Further, the control device 40 of the present embodiment is integrally configured with a control unit that controls various controlled devices connected to the output side of the control device 40. Therefore, a configuration (that is, hardware and software) that controls the operation of each controlled device constitutes a control unit that controls the operation of each controlled device.

For example, among the control devices 40, the configuration for controlling the rotation speed of the compressor 11 is the compressor control unit 40a. Further, among the control devices 40, the configuration for controlling the operation of the heating expansion valve 13 is the first decompression control unit 40b. Further, among the control devices 40, the configuration for controlling the operation of the cooling expansion valve 16 is the second decompression control unit 40c.

Next, the operation of the vehicle air conditioner 1 of the present embodiment having the above configuration will be described. In the vehicle air conditioner 1, the operation mode is switched for proper air conditioning in the vehicle interior. Specifically, in the vehicle air conditioner 1, the cooling mode and the dehumidifying / heating mode can be switched. The operation mode is switched by executing the control program stored in the control device 40.

The control program is executed when the auto switch is turned on while the air conditioner switch of the operation panel 42 is turned on. In the control program, the operation mode is determined based on the target blowout temperature TAO, the detection signals of various sensors, and the operation signal of the operation panel 42. The target blowout temperature TAO is the target temperature of the blown air blown into the vehicle interior.

The target blowout temperature TAO is calculated by the following formula F1.
TAO = Kset x Tset-Kr x Tr-Kam x Tam-Ks x As + C ... (F1)
In addition, Tset is the vehicle interior set temperature set by the temperature setting switch. Tr is the vehicle interior temperature detected by the inside air sensor. Tam is the temperature outside the vehicle interior detected by the outside air sensor. As is the amount of solar radiation detected by the solar radiation sensor. Kset, Kr, Kam, and Ks are control gains, and C is a correction constant.

Furthermore, in the control program, the control state of various controlled devices is determined according to the determined operation mode. Then, the control device 40 outputs a control signal or a control voltage to various control target devices so that the control state determined by the control program can be obtained.

Then, in the control program, the detection signal and the operation signal are read, the target blowout temperature TAO is calculated, and the control state of various controlled devices is determined at each predetermined control cycle until the vehicle air conditioner 1 is requested to be stopped. , Repeat control routines such as outputting control signals to various controlled devices. Each operation mode will be described below.

(A) Cooling mode In the cooling mode, the control state of various controlled devices is determined as described below. For the compressor 11, the refrigerant discharge capacity is determined so that the suction temperature Ts detected by the low pressure temperature sensor 41f approaches the target evaporator temperature TEO. The target evaporator temperature TEO is determined based on the target blowout temperature TAO with reference to a control map stored in advance in the control device 40.

In the control map, the target evaporator temperature TEO is lowered as the target blowout temperature TAO is lowered. Further, the target evaporator temperature TEO is determined to be a value within a range that does not cause frost formation in the indoor evaporator 17.

Further, the heating expansion valve 13 is determined to be fully open. Further, for the expansion valve 16 for cooling, the throttle opening is determined so that the superheat degree SH of the outlet side refrigerant of the indoor evaporator 17 approaches the predetermined standard superheat degree KSH (3 ° C. in this embodiment). To. The degree of superheat SH is calculated using the suction temperature Ts detected by the low pressure temperature sensor 41f and the suction pressure Ps detected by the low pressure pressure sensor 41g.

Further, for the indoor blower 32, the blower capacity is determined based on the target blowout temperature TAO with reference to the control map stored in the control device 40 in advance. In the control map, the amount of air blown is maximized when the target blowout temperature TAO is in the extremely low temperature region or extremely high temperature region, and the amount of air blown gradually increases from the extremely low temperature region or extremely high temperature region to the intermediate temperature region. The ventilation capacity is determined so that it decreases.

Regarding the electric actuator for the air mix door, it is decided that the ventilation path on the indoor condenser 12 side is fully closed and the cold air bypass passage 35 is fully opened. In the cooling mode, the opening degree of the air mix door 34 may be determined so that the blown air temperature TAV detected by the air conditioning air temperature sensor 41h approaches the target blown temperature TAO.

Therefore, in the refrigerating cycle device 10 in the cooling mode, the discharged refrigerant discharged from the compressor 11 flows into the indoor condenser 12. In the cooling mode, since the air mix door 34 closes the ventilation path on the indoor condenser 12 side, the refrigerant flowing into the indoor condenser 12 flows out from the indoor condenser 12 with almost no heat exchange with the blown air. do.

The refrigerant flowing out of the indoor condenser 12 flows into the outdoor heat exchanger 14 via the fully opened heating expansion valve 13. The refrigerant flowing into the outdoor heat exchanger 14 dissipates heat to the outside air and condenses. The refrigerant flowing out of the outdoor heat exchanger 14 flows into the receiver 15 and is separated into gas and liquid.

The liquid phase refrigerant flowing out of the receiver 15 flows into the cooling expansion valve 16 and is depressurized. At this time, the throttle opening of the cooling expansion valve 16 is determined so that the superheat degree SH of the outlet-side refrigerant of the indoor evaporator 17 approaches the reference superheat degree KSH. The refrigerant decompressed by the cooling expansion valve 16 flows into the indoor evaporator 17.

The refrigerant flowing into the indoor evaporator 17 absorbs heat from the blown air and evaporates. As a result, the blown air is cooled. The refrigerant flowing out of the indoor evaporator 17 is sucked into the compressor 11 and compressed again.

In the indoor air conditioning unit 30 in the cooling mode, the blown air cooled by the indoor evaporator 17 is blown into the vehicle interior. As a result, cooling of the passenger compartment is realized.

(B) Dehumidifying / heating mode In the dehumidifying / heating mode, the control state of various controlled devices is determined as follows. The compressor 11 and the indoor blower 32 are determined in the same manner as in the cooling mode.

Further, the throttle opening of the heating expansion valve 13 is determined by executing the control flow shown in FIG. The control flow shown in FIG. 3 is executed as a subroutine of the main routine of the control program. Further, each control step shown in the flowchart of FIG. 3 or the like is a function realization unit of the control device 40.

First, in step S1 of FIG. 3, the target supercooling degree SCO of the refrigerant flowing out of the indoor condenser 12 and flowing into the heating expansion valve 13, that is, the refrigerant on the outlet side of the indoor condenser 12 is determined. Therefore, step S1 is a target supercooling degree determination unit.

Specifically, in step S1, the target supercooling degree SCO is determined with reference to the control map stored in advance in the control device 40 based on the inlet-side pressure P1 of the refrigerant flowing into the heating expansion valve 13. do. In the control map, the target supercooling degree SCO is determined so that the coefficient of performance (that is, COP) of the cycle becomes a maximum value.

For the inlet side pressure P1, a value obtained by subtracting the pressure loss generated when the refrigerant passes through the indoor condenser 12 from the discharge pressure Pd detected by the high pressure pressure sensor 41e can be used. In the present embodiment, since the ratio of the pressure loss to the discharge pressure Pd is relatively small, the discharge pressure Pd is adopted as the inlet side pressure P1.

Next, in step S2, the degree of supercooling SC1 of the refrigerant flowing out of the indoor condenser 12 and flowing into the heating expansion valve 13 is estimated. Therefore, step S2 is a supercooling degree estimation unit. In step S2, the degree of supercooling SC1 is estimated using the refrigerant discharge flow rate Gr (mass flow rate) of the compressor 11, the throttle passage area A of the heating expansion valve 13, the discharge pressure Pd, and the outside temperature Tam.

The refrigerant discharge flow rate Gr of the compressor 11 can be calculated from the suction density ρcin of the suction refrigerant sucked into the compressor 11, the rotation speed of the compressor 11, the discharge capacity of the compressor 11, and the volumetric efficiency of the compressor 11. can.

The suction density ρcin can be determined from the suction temperature Ts and the suction pressure Ps based on the physical properties of the refrigerant. The rotation speed of the compressor 11 can be determined from the control signal output from the control device 40 to the compressor 11. The discharge capacity of the compressor 11 and the volumetric efficiency of the compressor 11 can be grasped from the specifications of the compressor 11, test data, and the like.

The throttle passage area A of the heating expansion valve 13 is determined based on the specifications of the heating expansion valve 13 and the control signal (specifically, the control pulse) output from the control device 40 to the heating expansion valve 13. Can be done.

The discharge pressure Pd is used to determine the inlet side pressure P1 as in the target supercooling degree determination unit.

The outside air temperature Tam is used to determine the outlet pressure P2 of the refrigerant flowing out from the heating expansion valve 13. The outlet side pressure P2 is equivalent to the saturation pressure of the refrigerant in the outdoor heat exchanger 14 connected to the outlet side of the heating expansion valve 13, and the temperature of the refrigerant in the outdoor heat exchanger 14 is substantially equivalent to the outside temperature Tam. Become. Therefore, the saturation pressure of the refrigerant at the outside air temperature Tam can be adopted as the outlet side pressure P2.

Then, in step S2, the inlet-side density ρin of the refrigerant flowing into the heating expansion valve 13 is calculated using the following mathematical formula F2.
ρin = Gr ² / (2 × (P1-P2) × A ² )… (F2)
The inlet-side density ρin and the supercooling degree SC1 are correlated as shown in FIG. Therefore, the supercooling degree SC1 can be estimated by calculating the inlet-side density ρin by the mathematical formula F2.

Next, in step S3, it is determined whether or not the supercooling degree SC1 estimated in step S2 is larger than the target supercooling degree SCO determined in step S1. Therefore, step S3 is a supercooling degree determining unit for determining whether or not the supercooling degree SC1 is larger than the target supercooling degree SCO.

If it is determined in step S3 that the supercooling degree SC1 is not larger than the target supercooling degree SCO, that is, if the supercooling degree SC1 is equal to or less than the target supercooling degree SCO, the process proceeds to step S4. move on. On the other hand, if it is determined in step S3 that the supercooling degree SC1 is larger than the target supercooling degree SCO, the process proceeds to step S5.

In step S4, the blowout temperature control is executed and the process returns to the main routine. In the blowout temperature control, the throttle opening of the heating expansion valve 13 is controlled so that the blowout air temperature TAV approaches the target blowout temperature TAO.

Specifically, when the blowout air temperature TAV is lower than the target blowout temperature TAO, the heating expansion valve is in a range where the temperature of the refrigerant flowing into the outdoor heat exchanger 14 is higher than the outside air temperature Tam. The throttle opening of 13 is reduced. Further, when the blowing air temperature TAV is higher than the target blowing temperature TAO, the throttle opening degree of the heating expansion valve 13 is expanded.

As a result, in the blowout temperature control, the amount of heat exchange between the refrigerant and the outside air in the outdoor heat exchanger 14 is adjusted, and the amount of heat radiated from the refrigerant in the indoor condenser 12 to the blown air, that is, the blown air in the indoor condenser 12. The heating capacity of the can be adjusted.

In step S5, the throttle opening of the heating expansion valve 13 is expanded by a predetermined amount, and the process returns to the main routine. The degree of supercooling SC1 of the refrigerant flowing out of the indoor condenser 12 and flowing into the heating expansion valve 13, that is, the refrigerant on the outlet side of the indoor condenser 12 is lowered.

Further, in the dehumidifying / heating mode, the throttle opening of the cooling expansion valve 16 is controlled in the same manner as in the cooling mode. In the dehumidifying and heating mode, the heating expansion valve 13 is in a throttled state in which the refrigerant decompressing action is exerted, so that the throttle opening of the cooling expansion valve 16 is larger than that in the cooling mode.

Further, in the dehumidifying / heating mode, for the electric actuator for the air mix door, it is determined that the ventilation passage on the indoor condenser 12 side is fully opened and the cold air bypass passage 35 is fully closed. The opening degree of the air mix door 34 may be determined so that the blown air temperature TAV approaches the target blown temperature TAO, as in the cooling mode.

Therefore, in the refrigeration cycle device 10, as shown in the Moriel diagram of FIG. 5, the discharged refrigerant (point a5 in FIG. 5) discharged from the compressor 11 flows into the indoor condenser 12. In the dehumidifying / heating mode, the air mix door 34 opens the ventilation path on the indoor condenser 12 side, so that the refrigerant flowing into the indoor condenser 12 dissipates heat to the blown air passing through the indoor evaporator 17 and condenses. (Points a5 to b5 in FIG. 5). As a result, the blown air that has passed through the indoor evaporator 17 is heated.

The refrigerant flowing out of the indoor condenser 12 flows into the heating expansion valve 13 and is depressurized (points b5 to c5 in FIG. 5). At this time, the throttle opening of the heating expansion valve 13 is such that at least the supercooling degree SC1 of the refrigerant (point b5 in FIG. 5) flowing out of the indoor condenser 12 and flowing into the heating expansion valve 13 is the target supercooling degree. It is adjusted to be less than or equal to SCO.

Further, the throttle opening of the heating expansion valve 13 is blown out in a range where the supercooling degree SC1 of the refrigerant flowing out of the indoor condenser 12 and flowing into the heating expansion valve 13 is equal to or less than the target supercooling degree SCO. The air temperature TAV is adjusted to approach the target blowout temperature TAO. The refrigerant decompressed by the heating expansion valve 13 flows into the outdoor heat exchanger 14. The refrigerant flowing into the outdoor heat exchanger 14 dissipates heat to the outside air and condenses (points c5 to d5 in FIG. 5).

The refrigerant flowing out of the outdoor heat exchanger 14 flows into the receiver 15 and is separated into gas and liquid. The liquid phase refrigerant (point d5 in FIG. 5) flowing out from the receiver 15 flows into the expansion valve 16 for cooling and is depressurized (points d5 to e5 in FIG. 5). At this time, the throttle opening of the cooling expansion valve 16 is determined so that the superheat degree SH of the outlet-side refrigerant (point f5 in FIG. 5) of the indoor evaporator 17 approaches the reference superheat degree KSH. The refrigerant decompressed by the cooling expansion valve 16 flows into the indoor evaporator 17.

The refrigerant flowing into the indoor evaporator 17 absorbs heat from the blown air and evaporates (points e5 to f5 in FIG. 5). As a result, the blown air is cooled and dehumidified. The refrigerant flowing out of the indoor evaporator 17 is sucked into the compressor 11 and compressed again (points f5 to a5 in FIG. 5).

In the indoor air conditioning unit 30 in the dehumidifying / heating mode, the blown air cooled by the indoor evaporator 17 and dehumidified is reheated by the indoor condenser 12 and blown out into the vehicle interior. As a result, dehumidifying and heating of the vehicle interior is realized.

As described above, according to the vehicle air conditioner 1 of the present embodiment, it is possible to realize cooling and dehumidifying heating in the vehicle interior.

Here, in the refrigerating cycle device 10, in the dehumidifying and heating mode, in order to adjust the heating capacity of the blown air in the indoor condenser 12, the outlet temperature control for changing the throttle opening of the heating expansion valve 13 is executed. In the blowout temperature control, the throttle opening of the heating expansion valve 13 is reduced when the heating capacity of the blown air in the indoor condenser 12 is improved. Therefore, as the heating capacity of the indoor condenser 12 is improved, the supercooling degree SC1 is likely to increase.

However, if the supercooling degree SC1 is unnecessarily increased, a temperature distribution will occur in the blown air heated by the indoor condenser 12, and it may not be possible to realize comfortable dehumidifying and heating in the vehicle interior. be. Further, if the supercooling degree SC1 greatly exceeds the target supercooling degree SCO, the COP also decreases.

On the other hand, in the refrigerating cycle apparatus 10 of the present embodiment, in the dehumidifying and heating mode, the supercooling degree SC1 estimated in step S2 is heated so as to be equal to or less than the target supercooling degree SCO determined in step S1. The throttle opening of the expansion valve 13 is controlled. Therefore, it is possible to prevent the degree of supercooling SC1 of the refrigerant flowing out of the indoor condenser 12 and flowing into the heating expansion valve 13 from becoming unnecessarily large.

That is, according to the refrigerating cycle device 10 of the present embodiment, the supercooling degree SC1 of the refrigerant on the outlet side of the indoor condenser 12 which is a heating unit can be appropriately adjusted. As a result, comfortable dehumidifying and heating of the vehicle interior can be realized, and a decrease in COP can be suppressed.

Further, in the supercooling degree estimation unit of the present embodiment, the refrigerant discharge flow rate Gr of the compressor 11, the throttle passage area A of the heating expansion valve 13, the inlet side pressure P1, and the outside temperature Tam are used to use the supercooling degree SC1. Is estimated. Therefore, as described using the mathematical formula F2, the supercooling degree SC1 can be estimated accurately.

Furthermore, the parameters used to estimate the supercooling degree SC1 can be detected by the detector essential for blowout temperature control and compressor protection control. Therefore, in the supercooling degree estimation unit of the present embodiment, it is not necessary to add a new detection unit in order to estimate the supercooling degree SC1.

Further, since the supercooling degree estimation unit of the present embodiment estimates the supercooling degree SC1, even if the refrigerant actually flowing into the heating expansion valve 13 is in a gas-liquid two-phase state having a dryness. , The operation in the dehumidifying and heating mode can be continued.

(Second Embodiment)
In this embodiment, an example in which the control mode of the heating expansion valve 13 in the dehumidifying and heating mode is changed as shown in the flowchart of FIG. 6 will be described with respect to the first embodiment.

Specifically, in step S11 of FIG. 6, the target supercooling degree SCO is determined as in the first embodiment. Therefore, step S11 is a target supercooling degree determination unit.

Next, in step S12, the lower limit throttle passage area Amin of the heating expansion valve 13 is calculated. The lower limit throttle passage area Amin is the throttle passage area of the heating expansion valve 13 in which the supercooling degree SC1 of the refrigerant flowing into the heating expansion valve 13 is the target supercooling degree SCO. Therefore, step S12 is a lower limit area calculation unit.

Specifically, in step S12, the target supercooling degree SCO, the refrigerant discharge flow rate Gr of the compressor 11, the discharge pressure Pd, and the outside temperature Tam are used to calculate the lower limit throttle passage area Amin, and the minimum throttle passage area is calculated. Calculate Amin. The refrigerant discharge flow rate Gr of the compressor 11 can be obtained in the same manner as in the first embodiment. The discharge pressure Pd is used to determine the inlet side pressure P1 as in the first embodiment. The outside air temperature Tam is used to determine the outlet side pressure P2 as in the first embodiment.

Then, in step S12, the lower limit throttle passage area Amin is calculated by the following mathematical formula F3.
Amin = Gr / ρmax × ( ^ρmax / (2 × (P1-P2))) 1/2… (F3)
Note that ρmax is the inlet-side density of the refrigerant flowing into the heating expansion valve 13 at the target supercooling degree SCO. ρmax can be determined using FIG. 4 described in the first embodiment.

The mathematical formula F3 is a modified number of the mathematical formula F2 described in the first embodiment. That is, the lower limit area calculation unit calculates the lower limit throttle passage area Amin using the same formula as the supercooling degree estimation unit of the first embodiment.

Next, in step S13, it is determined whether or not the throttle passage area A of the actual heating expansion valve 13 is smaller than the lower limit throttle passage area Amin determined in step S12. Therefore, step S13 is a throttle passage area determination unit for determining whether or not the throttle passage area A is smaller than the lower limit throttle passage area Amin.

If it is determined in step S13 that the throttle passage area A is not smaller than the lower limit throttle passage area Amin, that is, if the throttle passage area A is equal to or larger than the lower limit throttle passage area Amin, the process proceeds to step S14. move on. On the other hand, if it is determined in step S13 that the throttle passage area A is smaller than the lower limit throttle passage area Amin, the process proceeds to step S15.

In step S14, the blowout temperature control is executed and the process returns to the main routine, as in step S4 of the first embodiment. Further, in step S15, as in step S5 of the first embodiment, the throttle opening degree of the heating expansion valve 13 is expanded by a predetermined amount, and the process returns to the main routine.

The configuration and operation of the other refrigeration cycle device 10 and the vehicle air conditioner 1 are the same as those in the first embodiment. Therefore, also in the vehicle air conditioner 1 of the present embodiment, cooling and dehumidifying heating of the vehicle interior can be realized as in the first embodiment.

Further, in the refrigerating cycle device 10 of the present embodiment, the throttle opening of the heating expansion valve 13 is controlled so that the throttle passage area A is equal to or larger than the lower limit throttle passage area Amin in the dehumidifying / heating mode. According to this, the supercooling degree SC1 can be set to the target supercooling degree SCO or less. Therefore, the same effect as that of the first embodiment can be obtained.

That is, even in the refrigerating cycle device 10 of the present embodiment, the supercooling degree SC1 of the refrigerant on the outlet side of the indoor condenser 12 which is a heating unit can be appropriately adjusted.

(Third Embodiment)
In the vehicle air conditioner 1 of the present embodiment, as shown in the overall configuration diagram of FIG. 7, the electric heater 37 is located downstream of the blown air flow of the indoor condenser 12 in the indoor air conditioner unit 30 with respect to the first embodiment. Is placed.

The electric heater 37 supplementarily heats the blown air when the heating capacity of the refrigerating cycle device 10 alone cannot raise the blown air temperature TAV to the target blown temperature TAO in the dehumidifying and heating mode. It is an auxiliary heating part. As the electric heater 37, a PTC heater or the like that generates heat by being supplied with electric power can be adopted. The calorific value of the electric heater 37 is controlled by the control voltage output from the control device 40.

Further, a suction temperature sensor 41i is connected to the input side of the control device 40 of the present embodiment as a sensor used for air conditioning control. The suction temperature sensor 41i is a suction temperature detection unit that detects the suction air temperature Tein of the suction air flowing into the indoor evaporator 17 via the inside / outside air switching device 33.

Further, in the supercooling degree estimation unit of the present embodiment, the refrigerant discharge flow rate Gr of the compressor 11, the discharge temperature Td, the discharge pressure Pd, the air flow rate Airf (mass flow rate) of the indoor blower 32, the suction air temperature Ten, and the electric heater The supercooling degree SC1 is estimated using the heating amount Qh of 37.

The refrigerant discharge flow rate Gr of the compressor 11 can be calculated in the same manner as in the first embodiment. Further, the air volume Airf of the indoor blower 32 can be determined from the specifications of the indoor blower 32 and the control voltage output from the control device 40 to the indoor blower 32. The heating amount Qh of the electric heater 37 can be determined based on the specifications of the electric heater 37 and the amount of electric power supplied from the control device 40 to the electric heater 37.

Then, the supercooling degree estimation unit of the present embodiment calculates the outlet-side enthalpy Hcout of the outlet-side refrigerant of the indoor condenser 12 based on the following mathematical formulas F4 to F6.
Qc = Qex + Qh ... (F4)
Qc = ρain × Airfc × Air × (TAV-Tein)… (F5)
Qex = Gr × (Hcin-Hcout) ... (F6)
Here, Qc is the total heating amount of the blast air (that is, the total heat absorption amount of the blast air). Qex is the amount of heat dissipated to the blown air of the refrigerant condensed by the indoor condenser 12.

Also, ρair is the density of the suction air. In this embodiment, the density of air in a predetermined reference state (for example, 25 ° C., 101.3 kPa) is adopted as the density of the suction air. Further, Air is the specific heat of air in the reference state.

Further, Airfc is the amount of air blown air passing through the indoor condenser 12 on the indoor condenser side. In the dehumidifying / heating mode of the present embodiment, the air mix door 34 completely closes the cold air bypass passage 35, so that the air volume Airfc on the indoor condenser side is the same as the air volume Airf. When the air mix door 34 opens the cold air bypass passage 35, the indoor condenser side is according to the opening ratio between the opening degree of the ventilation passage on the indoor condenser 12 side and the opening degree of the cold air bypass passage 35. The air volume Airfc may be determined.

Hcin is the inlet-side enthalpy of the inlet-side refrigerant of the indoor condenser 12. The inlet-side enthalpy Hcin can be determined from the discharge temperature Td and the discharge pressure Pd based on the physical characteristics of the refrigerant.

Therefore, in the supercooling degree estimation unit of the present embodiment, the exit-side enthalpy Hcout can be calculated using the formulas F4 to F6. Then, based on the outlet-side enthalpy Hcout and the discharge pressure Pd, the degree of supercooling SC1 of the refrigerant flowing out of the indoor condenser 12 and flowing into the heating expansion valve 13 can be estimated. Of course, if the refrigerant flowing out of the indoor condenser 12 and flowing into the heating expansion valve 13 is in a gas-liquid two-phase state, the dryness can be estimated.

The configuration and operation of the other refrigeration cycle device 10 and the vehicle air conditioner 1 are the same as those in the first embodiment. Therefore, also in the vehicle air conditioner 1 of the present embodiment, cooling and dehumidifying heating of the vehicle interior can be realized as in the first embodiment. Further, the refrigerating cycle apparatus 10 of the present embodiment can also obtain the same effect as that of the first embodiment.

That is, even in the refrigerating cycle device 10 of the present embodiment, the supercooling degree SC1 of the refrigerant on the outlet side of the indoor condenser 12 which is a heating unit can be appropriately adjusted.

Further, the vehicle air conditioner 1 of the present embodiment includes an electric heater 37 for heating the blown air as an auxiliary heating unit. According to this, in order to adjust the supercooling degree SC1 to the target supercooling degree SCO or less, when the throttle opening of the heating expansion valve 13 cannot be reduced, the blown air is heated by the electric heater 37. Can be done. As a result, comfortable dehumidifying and heating can be realized by raising the blown air temperature TAV until the blown air temperature TAO is reached.

Further, in the supercooling degree estimation unit of the present embodiment, the refrigerant discharge flow rate Gr of the compressor 11, the discharge temperature Td, the discharge pressure Pd, the air flow rate Airf of the indoor blower 32, the suction air temperature Ten, and the heating amount of the electric heater 37. The degree of supercooling SC1 is estimated using Qh. Therefore, as described using the formulas F4 to F6, the supercooling degree SC1 can be estimated accurately only by adding the suction temperature sensor 41i.

Further, the supercooling degree estimation unit of the present embodiment is also effective when applied to the refrigeration cycle device 10 not provided with the electric heater 37. In that case, the heating amount Qh of the electric heater 37 may be set to 0.

(Fourth Embodiment)
In this embodiment, the refrigerating cycle apparatus 10a shown in FIG. 8 in which the configuration of the heating unit is changed with respect to the refrigerating cycle apparatus 10 described in the first embodiment will be described. The refrigeration cycle device 10a is applied to the vehicle air conditioner 1 similar to the first embodiment. The heating unit of the refrigeration cycle device 10a has a water-refrigerant heat exchanger 121, a heater core 53 arranged in the heat medium circuit 50, and the like.

The heat medium circuit 50 is a heat medium circulation circuit that circulates a heat medium. In the heat medium circuit 50, an ethylene glycol aqueous solution is used as the heat medium. In the heat medium circuit 50, a water passage of the water refrigerant heat exchanger 121, a heat medium pump 51, an electric heater 52, a heater core 53, and the like are arranged.

The water refrigerant heat exchanger 121 is a heat dissipation unit that exchanges heat between the high pressure refrigerant discharged from the compressor 11 and the heat medium to dissipate the heat of the high pressure refrigerant to the blown air. In this embodiment, a so-called countercurrent type heat exchanger is adopted as the water refrigerant heat exchanger 121. In the countercurrent type heat exchanger, the flow direction of the refrigerant flowing through the refrigerant passage and the flow direction of the heat medium flowing through the heat medium passage are opposite to each other.

The heat medium pump 51 is a heat medium pumping unit that pumps the heat medium flowing out of the heater core 53 to the water refrigerant heat exchanger 121. The heat medium pump 51 is an electric water pump that rotationally drives an impeller (that is, an impeller) with an electric motor. The rotation speed (that is, the pumping capacity) of the heat medium pump 51 is controlled by the control voltage output from the control device 40.

The electric heater 52 heats the heat medium flowing out of the water refrigerant heat exchanger 121. When the electric heater 52 cannot raise the blown air temperature TAV of the blown air to the target blown temperature TAO only by the heating capacity of the refrigerating cycle device 10 in the dehumidifying and heating mode, the blown air through the heat medium It is an auxiliary heating part that auxiliaryly heats. As the electric heater 52, a PTC heater or the like having the same configuration as the electric heater 37 for blown air can be adopted.

The heater core 53 is a heat exchange unit for heating that heats the blown air by exchanging heat between the heat medium flowing out from the water refrigerant heat exchanger 121 and the blown air. The heater core 53 is arranged in the indoor air conditioning unit 30 in the same manner as the indoor condenser 12.

Further, a heat medium temperature sensor 41j is connected to the input side of the control device 40 of the present embodiment as a sensor for air conditioning control. The heat medium temperature sensor 41j is an inlet side heat medium temperature detection unit that detects the heater core inlet side heat medium temperature Twin of the heat medium flowing into the heater core 53.

Further, in the supercooling degree estimation unit of the present embodiment, the refrigerant discharge flow rate Gr of the compressor 11, the discharge temperature Td, the discharge pressure Pd, the heat medium flow rate LQf (mass flow rate) pressure-fed from the heat medium pump 51, and the heater core inlet side. The degree of supercooling SC1 is estimated using the heat medium temperature Twin, the air flow rate Airf of the indoor blower 32, the suction temperature Ts, and the heating amount Qh2 of the electric heater 52.

The supercooling degree SC1 of the present embodiment is the supercooling degree of the refrigerant flowing out from the water refrigerant heat exchanger 121 forming the heating portion and flowing into the heating expansion valve 13.

The refrigerant discharge flow rate Gr of the compressor 11 can be calculated in the same manner as in the first embodiment. Further, the discharge temperature Td and the discharge pressure Pd are used to determine the inlet-side enthalpy Hwcin of the inlet-side refrigerant of the water-refrigerant heat exchanger 121, as in the third embodiment. The heat medium flow rate LQf of the heat medium pump 51 can be determined from the specifications of the heat medium pump 51 and the control voltage output from the control device 40 to the heat medium pump 51.

The air volume Airf of the indoor blower 32 can be determined in the same manner as in the third embodiment. The suction temperature Ts is used to determine the cooling air temperature Tae that is cooled by the indoor evaporator 17 and flows into the heater core 53. Specifically, the cooling air temperature Tae may be a value obtained by subtracting the degree of superheat (3 ° C. in this embodiment) of the refrigerant on the outlet side of the indoor evaporator 17 from the suction temperature Ts.

The heating amount Qh2 of the electric heater 52 can be determined based on the specifications of the electric heater 52 and the amount of electric power output from the control device 40 to the electric heater 52.

Then, the supercooling degree estimation unit of the present embodiment estimates the outlet-side enthalpy Hwcout of the outlet-side refrigerant of the water-refrigerant heat exchanger 121 based on the following mathematical formulas F7 to F9.
Qwr = Qwex + Qh2 ... (F7)
Qwr = f (LQf, Twin, Airfh, Tae) ... (F8)
Qwex = Gr × (Hwcin-Hwcout) ... (F9)
Here, Qwr is the amount of heat radiated from the heat medium to the blown air by the heater core 53. Qwex is the amount of heat released from the refrigerant condensed in the water refrigerant heat exchanger 121 to the heat medium.

Further, Airfh is the air volume on the heater core side of the blown air passing through the heater core 53. In the dehumidifying and heating mode of the present embodiment, the air mix door 34 completely closes the cold air bypass passage 35, so that the air volume Airfh on the heater core side is the same as the air volume Airf. When the air mix door 34 opens the cold air bypass passage 35, the air volume Airfh on the heater core side is determined according to the opening ratio between the opening degree of the ventilation passage on the heater core 53 side and the opening degree of the cold air bypass passage 35. do it.

Further, in the formula F8, it is shown that Qwr is determined based on the heat medium flow rate LQf, the heater core inlet side heat medium temperature Twin, the heater core side air volume Airfh, and the cooling air temperature Tae. That is, the amount of heat exchange between the refrigerant and the heat medium in the heater core 53 is determined based on the temperature and air volume of the refrigerant flowing into the heater core 53, the temperature and air volume of the heat medium flowing into the heater core 53, and the heat exchange performance of the heater core 53. can do.

Therefore, in the present embodiment, the water refrigerant heat exchanger is referred to in advance based on the heat medium flow rate LQf, the heater core inlet side heat medium temperature Twin, the heater core side air volume Airfh, and the cooling air temperature Tae. The amount of heat released from the refrigerant condensed at 121 to the heat medium is determined. The heat exchange performance of the heater core 53 can be grasped from the specifications of the heater core 53, test data, and the like.

Therefore, in the present embodiment, the exit-side enthalpy Hwcout can be calculated using the formulas F7 to F9. Then, based on the outlet-side enthalpy Hwcout and the discharge pressure Pd, the degree of supercooling SC1 of the refrigerant flowing out of the heater core 53 and flowing into the heating expansion valve 13 can be estimated. Of course, if the refrigerant flowing out of the heater core 53 and flowing into the heating expansion valve 13 is in a gas-liquid two-phase state, the dryness can be estimated.

The configuration and operation of the other refrigeration cycle device 10a are the same as those of the refrigeration cycle device 10 of the first embodiment. Therefore, also in the vehicle air conditioner 1 of the present embodiment, cooling and dehumidifying heating of the vehicle interior can be realized as in the first embodiment. Further, the refrigerating cycle apparatus 10a of the present embodiment can also obtain the same effect as that of the first embodiment.

That is, even in the refrigerating cycle device 10 of the present embodiment, the supercooling degree SC1 of the refrigerant on the outlet side of the water refrigerant heat exchanger 121 forming the heating portion can be appropriately adjusted. As a result, it is possible to suppress a decrease in COP of the refrigeration cycle device 10.

Further, the vehicle air conditioner 1 of the present embodiment includes an electric heater 52 that heats a heat medium as an auxiliary heating unit. Therefore, in order to adjust the supercooling degree SC1 to the target supercooling degree SCO or less, the heat medium can be heated by the electric heater 52 when the throttle opening of the heating expansion valve 13 cannot be reduced. As a result, comfortable dehumidifying and heating can be realized by raising the blown air temperature TAV until the blown air temperature TAO is reached.

Further, in the supercooling degree estimation unit of the present embodiment, the refrigerant discharge flow rate Gr of the compressor 11, the discharge temperature Td, the discharge pressure Pd, the heat medium flow rate LQf of the heat medium pump 51, the heat medium temperature Twin on the heater core inlet side, and the indoor blower. The degree of supercooling SC1 is estimated using the air flow amount Airf of 32, the suction temperature Ts, and the heating amount Qh2 of the electric heater 52. Therefore, as described using the mathematical formulas F7 to F9, the supercooling degree SC1 can be estimated accurately only by adding the heat medium temperature sensor 41j.

Further, the supercooling degree estimation unit of the present embodiment is effective even when applied to the refrigeration cycle device 10a not provided with the electric heater 52. In that case, the heating amount Qh2 of the electric heater 52 may be set to 0.

(Fifth Embodiment)
In this embodiment, an example in which the estimation mode of the supercooling degree SC1 in the supercooling degree estimation unit is changed with respect to the fourth embodiment will be described.

As described in the fourth embodiment, in the refrigerating cycle apparatus 10a, a countercurrent type heat exchanger is adopted as the water refrigerant heat exchanger 121. In the countercurrent type heat exchanger, the flow direction of the refrigerant flowing through the refrigerant passage and the flow direction of the heat medium flowing through the heat medium passage are opposite to each other. Therefore, as shown in FIG. 9, the temperatures of the refrigerant and the heat medium change. In FIG. 9, the thick solid line shows the temperature change of the refrigerant, and the thick broken line shows the temperature change of the heat medium.

Therefore, in the water refrigerant heat exchanger 121, the heater core outlet side heat medium temperature Twout of the heat medium flowing out from the heater core 53 and flowing into the heat medium passage is compared with the water refrigerant outlet side refrigerant temperature Tdout of the refrigerant flowing out from the refrigerant passage. It will be a close value. Further, the heat medium temperature Twout on the outlet side of the heater core becomes a value lower than the refrigerant temperature Tdout on the outlet side of the water refrigerant.

Therefore, assuming that the heater core outlet side heat medium temperature Twout is the water refrigerant outlet side refrigerant temperature Tdout, it is possible to estimate the supercooling degree SC1 that is larger than the actual value, that is, the supercooling degree SCO1 on the worst crossing side. can.

Therefore, in the supercooling degree estimation unit of the present embodiment, the discharge pressure Pd, the heat medium flow rate LQf pressure-fed from the heat medium pump 51, the heat medium temperature Twin on the heater core inlet side, the air flow amount Airf of the indoor blower 32, and the suction temperature Ts are set. It is used to estimate the degree of supercooling SC1. The air volume Airf is used to determine the air volume Airfh on the heater core side, as in the fourth embodiment. The suction temperature Ts is used to determine the cooling air temperature Tae, as in the fourth embodiment.

Then, as in the equation F8 of the fourth embodiment, refer to the control map stored in advance based on the heat medium flow rate LQf, the heater core inlet side heat medium temperature Twin, the heater core side air volume Airfh, and the cooling air temperature Tae. , Determines the heat medium temperature Twout on the outlet side of the heater core. Further, based on the heater core outlet side heat medium temperature Twout and the discharge pressure Pd, the supercooling degree SC1 of the refrigerant flowing out of the water refrigerant heat exchanger 121 and flowing into the heating expansion valve 13 is estimated.

The configuration and operation of the other refrigeration cycle device 10a are the same as those in the fourth embodiment. Therefore, also in the vehicle air conditioner 1 of the present embodiment, cooling and dehumidifying heating of the vehicle interior can be realized as in the fourth embodiment. Further, the refrigerating cycle apparatus 10a of the present embodiment can also obtain the same effect as that of the fourth embodiment.

That is, even in the refrigerating cycle device 10 of the present embodiment, the supercooling degree SC1 of the refrigerant on the outlet side of the water refrigerant heat exchanger 121 forming the heating portion can be appropriately adjusted.

In the supercooling degree estimation unit of the present embodiment, an example in which the heater core outlet side heat medium temperature Twout is used as the water refrigerant outlet side refrigerant temperature Tdout has been described, but the present invention is not limited to this. For example, a value obtained by adding a predetermined value to the heater core outlet side heat medium temperature Twout may be used as the water refrigerant outlet side refrigerant temperature Tdout.

The present disclosure is not limited to the above-described embodiment, and can be variously modified as follows without departing from the spirit of the present disclosure.

The circuit configuration of the refrigeration cycle device according to the present disclosure is not limited to the configuration of the refrigeration cycle devices 10 and 10a disclosed in the above-described embodiment.

For example, it may be a refrigerating cycle device configured so that the refrigerant circuit can be switched, and a refrigerating cycle device in which the same refrigerant circuit as the above-described embodiment is formed in a predetermined operation mode may be used. Then, when the refrigerant circuit is switched to the same as that of the above-described embodiment, the same effect as that of the above-mentioned embodiment can be obtained by performing the same control as that of the above-mentioned embodiment.

Further, in the above-described embodiment, an example in which the receiver 15 is connected to the refrigerant outlet of the outdoor heat exchanger 14 has been described, but the present invention is not limited to this. For example, in the refrigerating cycle apparatus provided with the supercooling degree estimation unit described in the third to fifth embodiments, the receiver 15 is abolished and the refrigerant flow path from the refrigerant outlet of the indoor evaporator 17 to the suction port of the compressor 11 An accumulator may be placed in.

The accumulator separates the gas and liquid of the refrigerant flowing out from the indoor evaporator 17 and causes the separated gas phase refrigerant to flow out to the suction port side of the compressor 11, and the separated liquid phase refrigerant is used as the surplus refrigerant in the cycle. It is a liquid storage part on the low pressure side that stores. In the refrigeration cycle device including the accumulator, the outdoor heat exchanger 14 may function as an evaporator for evaporating the refrigerant when the outlet temperature is controlled in the dehumidification / heating mode.

Each component of the refrigeration cycle devices 10 and 10a is not limited to the components disclosed in the above-described embodiment.

For example, as the compressor 11, an engine-driven compressor driven by a rotational driving force transmitted from an internal combustion engine (that is, an engine) may be adopted. In an engine-driven compressor, the refrigerant discharge flow rate Gr can be calculated by considering the engine speed, the discharge capacity, the operating rate, and the like.

Further, in the above-mentioned third to fifth embodiments, an example in which the minimum detection unit is added to the detection unit essential for the blowout temperature control and the compressor protection control has been described, but the addition of the detection unit has been described. Is not limited to this. For example, a flow rate sensor as a flow rate detection unit for directly detecting the refrigerant discharge flow rate Gr (mass flow rate) of the compressor 11 may be added.

Further, in the above-described embodiment, an example in which R1234yf is adopted as the refrigerant of the refrigeration cycle devices 10 and 10a has been described, but the present invention is not limited to this. For example, R134a, R600a, R410A, R404A, R32, R407C and the like may be adopted. Alternatively, a mixed refrigerant or the like in which a plurality of these refrigerants are mixed may be adopted.

Further, in the above-described embodiment, an example in which an ethylene glycol aqueous solution is used as a heat medium has been described, but the present invention is not limited to this. For example, a solution containing dimethylpolysiloxane, a nanofluid or the like, an antifreeze solution, an aqueous liquid medium containing alcohol or the like, or a liquid medium containing oil or the like may be adopted.

The control modes of the refrigeration cycle devices 10 and 10a are not limited to the control modes disclosed in the above-described embodiment.

For example, the target supercooling degree determining unit of the first to third embodiments may determine the target supercooling degree SCO capable of suppressing the temperature distribution generated in the blown air. Specifically, the target supercooling degree SCO may be determined such that the temperature difference obtained by subtracting the minimum temperature from the maximum temperature of the blown air after passing through the indoor condenser 12 or the heater core 53 is equal to or less than a predetermined reference temperature difference.

The means disclosed in each of the above embodiments may be appropriately combined to the extent practicable.

For example, as the heating unit of the refrigeration cycle apparatus 10 described in the first and second embodiments, various constituent devices arranged in the water-refrigerant heat exchanger 121 and the heat medium circuit 50 described in the fourth embodiment are adopted. May be good. In other words, the supercooling degree estimation unit described in the first embodiment or the lower limit area calculation unit described in the second embodiment may be applied to the refrigeration cycle apparatus 10a described in the fourth embodiment.

Although the present disclosure has been described in accordance with the examples, it is understood that the present disclosure is not limited to the examples and structures. The present disclosure also includes various variations and variations within a uniform range. In addition, various combinations and forms, as well as other combinations and forms that include only one element, more, or less, are within the scope and scope of the present disclosure.

Claims (9)

1. A compressor (11) that compresses and discharges the refrigerant,
   A heating unit (12, 121, 50) that heats the blown air blown to the air-conditioned space using the refrigerant discharged from the compressor as a heat source.
   A first decompression unit (13) that depressurizes the refrigerant flowing out of the heating unit, and
   An outdoor heat exchange unit (14) that exchanges heat between the refrigerant flowing out of the first decompression unit and the outside air.
   A second decompression unit (16) that depressurizes the refrigerant flowing out of the outdoor heat exchange unit, and
   An indoor evaporation unit (17) that evaporates the refrigerant decompressed by the second decompression unit and cools the blown air before being heated by the heating unit.
   A target supercooling degree determining unit (S1) for determining a target supercooling degree (SCO) of the refrigerant flowing into the first decompression unit, and a target supercooling degree determining unit (S1).
   A supercooling degree estimation unit (S2) that estimates the supercooling degree (SC1) of the refrigerant flowing into the first decompression unit, and a supercooling degree estimation unit (S2).
   A first decompression control unit (40b) for controlling the operation of the first decompression unit is provided.
   The first decompression control unit controls the operation of the first decompression unit so that the supercooling degree (SC1) estimated by the supercooling degree estimation unit is equal to or less than the target supercooling degree (SCO). Refrigeration cycle equipment.
2. The supercooling degree estimation unit includes the refrigerant discharge flow rate (Gr) of the compressor, the opening degree (A) of the first decompression unit, the discharge pressure (Pd) of the refrigerant discharged from the compressor, and the outside temperature. The refrigeration cycle apparatus according to claim 1, wherein the degree of supercooling (SC1) is estimated using (Tam).
3. Further, a blower unit (32) for blowing the blown air is provided.
   The heating unit has an indoor condenser (12) that exchanges heat between the refrigerant discharged from the compressor and the blown air.
   The supercooling degree estimation unit includes the refrigerant discharge flow rate (Gr) of the compressor, the discharge temperature (Td) of the refrigerant discharged from the compressor, and the discharge pressure (Pd) of the refrigerant discharged from the compressor. The refrigeration according to claim 1, wherein the degree of supercooling (SC1) is estimated using the air volume (Airf) of the air blower unit and the suction air temperature (Tein) of the air blown air flowing into the indoor evaporative unit. Cycle device.
4. Further, a blower unit (32) for blowing the blown air is provided.
   The heating unit includes a water refrigerant heat exchanger (121) that exchanges heat between the refrigerant discharged from the compressor and the heat medium, the heat medium heated by the water refrigerant heat exchanger, and the blown air. It has a heater core (53) for heat exchange, and a heat medium pressure feeding unit (51) for pressure-feeding the heat medium heated by the water refrigerant heat exchanger (121) to the heater core.
   The supercooling degree estimation unit includes the refrigerant discharge flow rate (Gr) of the compressor, the discharge temperature (Td) of the refrigerant discharged from the compressor, and the discharge pressure (Pd) of the refrigerant discharged from the compressor. , The heat medium flow rate (LQf) of the heat medium pressure feeding section, the heater core inlet side heat medium temperature (Twin) of the heat medium flowing into the heater core, the air volume of the blowing section (Airf), and being sucked into the compressor. The refrigeration cycle apparatus according to claim 1, wherein the degree of supercooling (SC1) is estimated using the suction temperature (Ts) of the refrigerant.
5. Further, a blower unit (32) for blowing the blown air is provided.
   The heating unit includes a water refrigerant heat exchanger (121) that exchanges heat between the refrigerant discharged from the compressor and the heat medium, the heat medium heated by the water refrigerant heat exchanger, and the blown air. It has a heater core (53) for heat exchange, and a heat medium pressure feeding unit (51) for pressure-feeding the heat medium heated by the water refrigerant heat exchanger (121) to the heater core.
   The water-refrigerant heat exchanger is a counter-flow type heat exchanger in which the flow direction of the refrigerant flowing through the refrigerant passage and the flow direction of the heat medium flowing through the heat medium passage are opposite to each other.
   The supercooling degree estimation unit includes the discharge pressure (Pd) of the refrigerant discharged from the compressor, the heat medium flow rate (LQf) of the heat medium pressure feeding unit, and the heater core inlet side heat of the heat medium flowing into the heater core. The first aspect of claim 1, wherein the degree of supercooling (SC1) is estimated using the medium temperature (Twin), the amount of air blown by the blower unit (Airf), and the suction temperature (Ts) of the refrigerant sucked into the compressor. Refrigeration cycle equipment.
6. A compressor (11) that compresses and discharges the refrigerant,
   A heating unit (12, 121, 50) that heats the blown air blown to the air-conditioned space using the refrigerant discharged from the compressor as a heat source.
   A first decompression unit (13) that depressurizes the refrigerant flowing out of the heating unit, and
   An outdoor heat exchange unit (14) that exchanges heat between the refrigerant flowing out of the first decompression unit and the outside air.
   A second decompression unit (16) that depressurizes the refrigerant flowing out of the outdoor heat exchange unit, and
   An indoor evaporation unit (17) that evaporates the refrigerant decompressed by the second decompression unit and cools the blown air before being heated by the heating unit.
   A target supercooling degree determining unit (S11) for determining a target supercooling degree (SCO) of the refrigerant flowing into the first decompression unit, and a target supercooling degree determining unit (S11).
   The lower limit area calculation unit (S12) for calculating the lower limit throttle passage area (Amin) of the first decompression unit where the supercooling degree (SC1) of the refrigerant flowing into the first decompression unit becomes the target supercooling degree (SCO). )When,
   A first decompression control unit (40b) for controlling the operation of the first decompression unit is provided.
   The first decompression control unit is a refrigeration cycle device that controls the operation of the first decompression unit so that the throttle passage area (A) of the first decompression unit is equal to or larger than the lower limit throttle passage area (Amin).
7. The lower limit area calculation unit includes the target supercooling degree (SCO), the refrigerant discharge flow rate (Gr) of the compressor, the inlet side pressure (P1) of the refrigerant flowing into the first decompression unit, and the first decompression unit. The refrigerating cycle apparatus according to claim 6, wherein the lower limit throttle passage area (Amin) is calculated by using the outlet side pressure (P2) of the refrigerant flowing out from.
8. The refrigeration cycle apparatus according to claim 7, wherein the lower limit area calculation unit uses the discharge pressure (Pd) of the refrigerant discharged from the compressor as the inlet side pressure.
9. The refrigeration cycle apparatus according to claim 7 or 8, wherein the lower limit area calculation unit uses a value determined by using the outside air temperature (Tam) as the outlet side pressure.

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