JP2022056869A - Refrigeration cycle device - Google Patents

Abstract

To maximize heating efficiency to the extent possible.SOLUTION: A refrigeration cycle device comprises: heat radiation sections 12, 40, 42, 27 which cause a refrigerant discharged from a compressor 11 to radiate heat to the air to be blown into an air-conditioning object space; and a heating section 43 which heats the air or a human body in the air-conditioning object space using a heat source different from the refrigerant. A first change rate, which is a rate of a change in heat radiation capacity at the heat radiation sections to consumption energy of the compressor per unit time is characterized by being: larger than a second change rate, which is a rate of the change in heating capacity at the heating section to the consumption energy of the heating section per unit time, when the consumption energy of the compressor per unit time is lower than an upper limit value Pe_limit; and lower than the second change rate when the consumption energy of the compressor per unit time is larger than the upper limit value. A control section 60 controls the compressor so that the consumption energy per unit time thereof is equal to or less than an optimal value.SELECTED DRAWING: Figure 3

Description

The present invention relates to a refrigeration cycle device capable of heating by a heat pump cycle.

Conventionally, Patent Document 1 describes a vehicle air conditioner for heating by a heat pump cycle. In this conventional technique, the compressor is controlled so that the target temperature of the heat exchanger for heating during the heating operation by the heat pump cycle becomes the target outlet temperature, and when the target outlet temperature is higher than the limit value, the heat exchanger for heating is targeted. The temperature is set to be lower than the target blowout temperature, and the compressor is controlled.

As a result, power saving operation is performed when the target blowing temperature is too high due to an erroneous operation by the user.

Japanese Unexamined Patent Publication No. 2011-012939

In the above-mentioned prior art, the cycle efficiency changes when the low pressure of the cycle changes, but since there is no mention of a control method according to the low pressure of the cycle, there is a limit to the improvement of the cycle efficiency.

In view of the above points, it is an object of the present invention to maximize the heating efficiency as much as possible.

In order to achieve the above object, the refrigeration cycle apparatus according to claim 1 is used.
A compressor (11) that sucks in the refrigerant, compresses it, and discharges it.
A heat dissipation unit (12, 40, 42, 27) that dissipates the refrigerant discharged from the compressor to the air blown into the air-conditioned space, and
The decompression unit (14a, 14c) that decompresses the refrigerant dissipated by the heat dissipation unit, and
The evaporating unit (16), which evaporates the refrigerant by absorbing heat from the decompressed refrigerant in the depressurizing unit,
A heating unit (43) that heats the human body in the air or air-conditioned space using a heat source different from the refrigerant, and
It is equipped with a control unit (60) that controls the compressor.
The first change rate, which is the rate of change in the heat dissipation capacity in the heat dissipation section with respect to the energy consumption per unit time of the compressor, and the rate of change in the heating capacity in the heating section with respect to the energy consumption per unit time of the heating section. When comparing the second change rate with the energy consumption per unit time, the first change rate exceeds the second change rate when the energy consumption per unit time is below the upper limit (Pe_limit), and the unit is It has the characteristic that when the energy consumption per hour exceeds the upper limit, it falls below the second rate of change.
The control unit controls the compressor so that the energy consumption per unit time of the compressor is equal to or less than the optimum value.

According to this, since the heat dissipation efficiency of the heat dissipation part can be made higher than the heating efficiency of the heating part, the heating efficiency can be maximized as much as possible.

The reference numerals in parentheses of each means described in this column and the scope of claims indicate the correspondence with the specific means described in the embodiments described later.

It is a figure which shows the whole structure of the refrigerating cycle apparatus in 1st Embodiment, and shows the operating state of a cooling mode. It is a block diagram which shows the electric control part of the refrigeration cycle apparatus in 1st Embodiment. It is a graph explaining the upper limit power consumption of the compressor in 1st Embodiment. It is a characteristic figure for calculating the upper limit flow rate in 1st Embodiment. It is a characteristic diagram used for calculating the intake refrigerant density in 1st Embodiment. It is a figure which shows the whole structure of the refrigerating cycle apparatus in 1st Embodiment, and shows the operating state of a heating mode. It is a characteristic diagram used for calculating the waste heat recovery amount in 1st Embodiment. It is a figure which shows the whole structure of the refrigerating cycle apparatus in 1st Embodiment, and shows the operating state of the waste heat recovery heating mode. It is a Moriel diagram which shows the change of the state of the refrigerant in the heating mode of the refrigeration cycle apparatus of 1st Embodiment. It is a characteristic diagram used for calculating the upper limit temperature in 2nd Embodiment. It is a graph which shows the relationship between the discharged refrigerant temperature and the discharged refrigerant flow rate in the second embodiment. It is a characteristic diagram used for calculating the upper limit power in 3rd Embodiment. 3 is a graph showing the relationship between the power consumption of the compressor and the flow rate of the discharged refrigerant in the third embodiment. It is a figure which shows the whole structure of the refrigerating cycle apparatus in 4th Embodiment. It is a figure which shows the whole structure of the refrigerating cycle apparatus in 5th Embodiment.

(First Embodiment)
The first embodiment will be described with reference to FIGS. 1 to 9. The refrigeration cycle device 10 is applied to a vehicle air conditioner 1 mounted on an electric vehicle that obtains a driving force for traveling from an electric motor. The vehicle air conditioner 1 is an air conditioner with a battery temperature adjusting function. The vehicle air conditioner 1 air-conditions the interior of the vehicle, which is the space to be air-conditioned, and adjusts the temperature of the battery 80.

The battery 80 is a secondary battery that stores electric power supplied to an in-vehicle device such as an electric motor. The battery 80 of this embodiment is a lithium ion battery. The battery 80 is a so-called assembled battery formed by stacking a plurality of battery cells and electrically connecting these battery cells in series or in parallel. The battery 80 is a waste heat generating device that generates waste heat as it operates.

The output of this type of battery tends to decrease at low temperatures, and deterioration tends to progress at high temperatures. Therefore, the temperature of the battery needs to be maintained within an appropriate temperature range (in this embodiment, 15 ° C. or higher and 55 ° C. or lower) in which the charge / discharge capacity of the battery can be fully utilized. ..

Therefore, in the vehicle air conditioner 1, the battery 80 can be cooled by the cooling heat generated by the refrigerating cycle device 10. The objects to be cooled in the refrigeration cycle device 10 of the present embodiment are air and a battery 80.

As shown in the overall configuration diagram of FIG. 1, the vehicle air conditioner 1 includes a refrigerating cycle device 10, an indoor air conditioner unit 30, a high temperature side heat medium circuit 40, a low temperature side heat medium circuit 50, and the like.

The refrigerating cycle device 10 cools the air blown into the vehicle interior and heats the high-temperature side heat medium circulating in the high-temperature side heat medium circuit 40 in order to air-condition the vehicle interior. The refrigeration cycle device 10 cools the low temperature side heat medium circulating in the low temperature side heat medium circuit 50 in order to cool the battery 80.

The refrigerating cycle device 10 can switch the refrigerant circuits for various operation modes in order to perform air conditioning in the vehicle interior. For example, the cooling mode refrigerant circuit, the heating mode refrigerant circuit, and the like can be switched. The refrigeration cycle device 10 can switch between an operation mode in which the battery 80 is cooled and an operation mode in which the battery 80 is not cooled in each operation mode for air conditioning.

The refrigeration cycle apparatus 10 employs an HFO-based refrigerant (specifically, R1234yf) as the refrigerant, and the pressure of the discharged refrigerant discharged from the compressor 11 does not exceed the critical pressure of the refrigerant, which is a steam compression type subcritical. It constitutes a refrigeration cycle. Refrigerant 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.

Among the constituent devices of the refrigerating cycle device 10, the compressor 11 sucks the refrigerant in the refrigerating cycle device 10, compresses it, and discharges it. The compressor 11 is arranged in the drive unit room partitioned in front of the vehicle room. An electric motor for traveling the vehicle and the like are housed in the drive unit 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 the control signal output from the control device 60 shown in FIG.

As shown in FIG. 1, the inlet side of the refrigerant passage of the refrigerant heat medium heat exchanger 12 is connected to the discharge port of the compressor 11. The refrigerant heat medium heat exchanger 12 has a refrigerant passage for circulating the high-pressure refrigerant discharged from the compressor 11 and a heat medium passage for circulating the high-temperature side heat medium circulating in the high-temperature side heat medium circuit 40. .. The refrigerant heat medium heat exchanger 12 is a heat exchanger for heating that heats the high temperature side heat medium by exchanging heat between the high pressure refrigerant flowing through the refrigerant passage and the high temperature side heat medium flowing through the heat medium passage. ..

The inlet side of the first three-way joint 13a having three inflow outlets communicating with each other is connected to the outlet of the refrigerant passage of the refrigerant heat medium heat exchanger 12. As such a three-way joint, one formed by joining a plurality of pipes or one formed by providing a plurality of refrigerant passages in a metal block or a resin block can be adopted.

The refrigeration cycle device 10 includes second to eighth three-way joints 13b to 13h. The basic configuration of these second to eighth three-way joints 13b to 13h is the same as that of the first three-way joint 13a.

One inflow port side of the second three-way joint 13b is connected to one outflow port of the first three-way joint 13a via a bypass passage 22a. A dehumidifying on-off valve 15a and a first fixed throttle 26a are arranged in the bypass passage 22a.

The dehumidifying on-off valve 15a is a solenoid valve that opens and closes the bypass passage 22a. The refrigeration cycle device 10 includes a heating on-off valve 15b and a cooling on-off valve 15d. The basic configuration of the heating on-off valve 15b and the cooling on-off valve 15d is the same as that of the dehumidifying on-off valve 15a.

The dehumidifying on-off valve 15a, the heating on-off valve 15b, and the cooling on-off valve 15d can switch the refrigerant circuit in each operation mode by opening and closing the refrigerant passage. The dehumidifying on-off valve 15a, the heating on-off valve 15b, and the cooling on-off valve 15d are refrigerant circuit switching units for switching the refrigerant circuit of the cycle. The dehumidifying on-off valve 15a, the heating on-off valve 15b, and the cooling on-off valve 15d are controlled by the control voltage output from the control device 60.

The first fixed throttle 26a is a pressure reducing portion for reducing the pressure of the refrigerant, and specifically, an orifice, a capillary tube, or the like.

One inflow port side of the 7th three-way joint 13g is connected to the other outflow port of the first three-way joint 13a via a cooling on-off valve 15d. The inlet side of the heating expansion valve 14a is connected to the outlet of the 7th three-way joint 13g.

The heating expansion valve 14a reduces the pressure of the high-pressure refrigerant flowing out from the refrigerant passage of the refrigerant heat medium heat exchanger 12 at least in the operation mode of heating the vehicle interior, and the flow rate (mass flow rate) of the refrigerant flowing out to the downstream side. It is a decompression unit for heating that adjusts. The heating expansion valve 14a is an electric type having a valve body configured to change the throttle opening degree and an electric actuator (in other words, an electric mechanism) for changing the opening degree of the valve body. Variable throttle mechanism (in other words, an electric expansion valve).

The refrigeration cycle device 10 includes a cooling expansion valve 14b and a cooling expansion valve 14c. The basic configuration of the cooling expansion valve 14b and the cooling expansion valve 14c is the same as that of the heating expansion valve 14a.

The expansion valve 14a for heating, the expansion valve 14b for cooling, and the expansion valve 14c for cooling have a fully open function that functions as a mere refrigerant passage without exerting a flow rate adjusting action and a refrigerant depressurizing action by fully opening the valve opening. It also has a fully closed function that closes the refrigerant passage by fully closing the valve opening.

With the fully open function and the fully closed function, the heating expansion valve 14a, the cooling expansion valve 14b, and the cooling expansion valve 14c can switch the refrigerant circuit in each operation mode. The heating expansion valve 14a, the cooling expansion valve 14b, and the cooling expansion valve 14c function as a refrigerant circuit switching unit. The heating expansion valve 14a, the cooling expansion valve 14b, and the cooling expansion valve 14c are controlled by a control signal (control pulse) output from the control device 60.

The refrigerant inlet side of the outdoor heat exchanger 16 is connected to the outlet of the heating expansion valve 14a. The outdoor heat exchanger 16 is a heat exchanger that exchanges heat between the refrigerant flowing out from the heating expansion valve 14a and the outside air blown by a cooling fan (not shown). The outdoor heat exchanger 16 is arranged on the front side in the drive unit room. Therefore, when the vehicle is running, the running wind can be applied to the outdoor heat exchanger 16.

The inlet side of the third three-way joint 13c is connected to the refrigerant outlet of the outdoor heat exchanger 16. One inflow port side of the fourth three-way joint 13d is connected to one outflow port of the third three-way joint 13c via a heating passage 22b. A heating on-off valve 15b for opening and closing the refrigerant passage is arranged in the heating passage 22b.

In the heating passage 22b, a first check valve 17a is arranged between the heating on-off valve 15b and the fourth three-way joint 13d. The first check valve 17a allows the refrigerant to flow from the heating on-off valve 15b side to the fourth three-way joint 13d side, and prohibits the refrigerant from flowing from the fourth three-way joint 13d side to the heating on-off valve 15b side. do.

The other inlet side of the second three-way joint 13b is connected to the other outlet of the third three-way joint 13c. A second check valve 17b is arranged in a refrigerant passage connecting the other outlet side of the third three-way joint 13c and the other inlet side of the second three-way joint 13b. The second check valve 17b allows the refrigerant to flow from the third three-way joint 13c side to the second three-way joint 13b side, and prohibits the refrigerant from flowing from the second three-way joint 13b side to the third three-way joint 13c side. do.

In the refrigerant passage connecting the other outlet side of the third three-way joint 13c and the other inlet side of the second three-way joint 13b, the second check valve 17b and the other inlet of the second three-way joint 13b A second fixed diaphragm 26b is arranged between them. The second fixed throttle 26b is a pressure reducing portion for reducing the pressure of the refrigerant, and specifically, an orifice, a capillary tube, or the like.

The inlet side of the receiver 25 is connected to the outlet of the second three-way joint 13b. The receiver 25 is a liquid storage unit having a gas-liquid separation function. The receiver 25 separates the air and liquid of the refrigerant flowing out from the refrigerant heat medium heat exchanger 12 or the outdoor heat exchanger 16 (that is, the heat exchanger functioning as a condenser that condenses the refrigerant in the refrigeration cycle device 10). The receiver 25 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 eighth three-way joint 13h is connected to the refrigerant outlet of the receiver 25. The other inflow port side of the 7th three-way joint 13g is connected to one outflow port of the sixth three-way joint 13f via the outlet side passage 22d. A third check valve 17c is arranged in the outlet side passage 22d. The third check valve 17c allows the refrigerant to flow from the 8th three-way joint 13h side to the 7th three-way joint 13g side, and prohibits the refrigerant from flowing from the 7th three-way joint 13g side to the 8th three-way joint 13h side. do.

The inlet side of the fifth three-way joint 13e is connected to the other outlet of the eighth three-way joint 13h. The inlet side of the cooling expansion valve 14b is connected to one of the outlets of the fifth three-way joint 13e. The inlet side of the cooling expansion valve 14c is connected to the other outlet of the fifth three-way joint 13e.

The cooling expansion valve 14b is an air-conditioning decompression unit that reduces the pressure of the refrigerant flowing out from the outdoor heat exchanger 16 and adjusts the flow rate of the refrigerant flowing out to the downstream side at least in the operation mode for cooling the vehicle interior. The cooling expansion valve 14b is the first pressure reducing unit.

The refrigerant inlet side of the indoor evaporator 18 is connected to the outlet of the cooling expansion valve 14b. The indoor evaporator 18 is arranged in the air conditioning case 31 of the indoor air conditioning unit 30. The indoor evaporator 18 exchanges heat between the low-pressure refrigerant decompressed by the cooling expansion valve 14b and the air blown from the blower 32 to evaporate the low-pressure refrigerant, and causes the low-pressure refrigerant to exert a heat absorbing action to absorb air. It is an air conditioning evaporative unit that cools. The indoor evaporator 18 is the first evaporation unit. The inlet side of the evaporation pressure adjusting valve 20 is connected to the refrigerant outlet of the indoor evaporator 18.

The evaporation pressure adjusting valve 20 maintains the refrigerant evaporation pressure in the indoor evaporator 18 at a predetermined reference pressure or higher in order to suppress frost formation in the indoor evaporator 18. The evaporation pressure adjusting valve 20 is a mechanical variable throttle mechanism that increases the valve opening degree as the pressure of the refrigerant on the outlet side of the indoor evaporator 18 increases. As a result, the evaporation pressure adjusting valve 20 maintains the refrigerant evaporation temperature in the indoor evaporator 18 at a frost formation suppression temperature (1 ° C. in the present embodiment) that can suppress frost formation in the indoor evaporator 18. .. One inflow port side of the sixth three-way joint 13f is connected to the outlet of the evaporation pressure adjusting valve 20.

The cooling expansion valve 14c is a battery decompression unit that depressurizes the refrigerant flowing out from the outdoor heat exchanger 16 and adjusts the flow rate of the refrigerant flowing out to the downstream side at least in the operation mode for cooling the battery 80.

The inlet side of the refrigerant passage of the chiller 19 is connected to the outlet of the cooling expansion valve 14c. The chiller 19 has a refrigerant passage for circulating a low-pressure refrigerant decompressed by a cooling expansion valve 14c, and a heat medium passage for circulating a low-temperature side heat medium circulating in the low-temperature side heat medium circuit 50. The chiller 19 is a second evaporation unit that exchanges heat between the low-pressure refrigerant flowing through the refrigerant passage and the low-temperature side heat medium flowing through the heat medium passage to evaporate the low-pressure refrigerant and exert a heat absorbing action.

The other inlet side of the sixth three-way joint 13f is connected to the outlet of the refrigerant passage of the chiller 19. The other inlet side of the fourth three-way joint 13d is connected to the outlet of the sixth three-way joint 13f. The suction port side of the compressor 11 is connected to the outlet of the fourth three-way joint 13d.

The indoor evaporator 18 and the chiller 19 are connected in parallel to each other with respect to the refrigerant flow. The bypass passage 22a guides the refrigerant flowing out of the refrigerant passage of the refrigerant heat medium heat exchanger 12 to the upstream side of the receiver 25. The heating passage 22b guides the refrigerant flowing out of the outdoor heat exchanger 16 to the suction port side of the compressor 11.

The high temperature side heat medium circuit 40 is a heat medium circulation circuit that circulates the high temperature side heat medium. As the high temperature side heat medium, a solution containing ethylene glycol, dimethylpolysiloxane, a nanofluid, or the like, an antifreeze solution, or the like can be adopted. In the high temperature side heat medium circuit 40, a heat medium passage of the refrigerant heat medium heat exchanger 12, a high temperature side heat medium pump 41, a heater core 42, and the like are arranged.

The high temperature side heat medium pump 41 is a water pump that pumps the high temperature side heat medium to the inlet side of the heat medium passage of the refrigerant heat medium heat exchanger 12. The high temperature side heat medium pump 41 is an electric pump whose rotation speed (that is, pumping capacity) is controlled by a control voltage output from the control device 60.

The heat medium inlet side of the heater core 42 is connected to the outlet of the heat medium passage of the refrigerant heat medium heat exchanger 12. The heater core 42 is a heat exchanger that heats the air by exchanging heat between the high temperature side heat medium heated by the refrigerant heat medium heat exchanger 12 and the air that has passed through the indoor evaporator 18. The heater core 42 is arranged in the air conditioning case 31 of the indoor air conditioning unit 30. The suction port side of the high temperature side heat medium pump 41 is connected to the heat medium outlet of the heater core 42.

Therefore, in the high temperature side heat medium circuit 40, the high temperature side heat medium pump 41 adjusts the flow rate of the high temperature side heat medium flowing into the heater core 42 to dissipate heat of the high temperature side heat medium in the heater core 42 to the air (that is, that is). , The amount of heat of air in the heater core 42) can be adjusted.

Each component of the refrigerant heat medium heat exchanger 12 and the high temperature side heat medium circuit 40 is a heat radiating unit that dissipates heat to the air using the refrigerant discharged from the compressor 11 as a heat source.

The low temperature side heat medium circuit 50 is a heat medium circulation circuit that circulates the low temperature side heat medium. As the low temperature side heat medium, the same fluid as the high temperature side heat medium can be adopted. In the low temperature side heat medium circuit 50, a heat medium passage of the chiller 19, a low temperature side heat medium pump 51, a cooling heat exchange section 52, and the like are arranged.

The low temperature side heat medium pump 51 is a water pump that pumps the low temperature side heat medium to the inlet side of the heat medium passage of the chiller 19. The basic configuration of the low temperature side heat medium pump 51 is the same as that of the high temperature side heat medium pump 41.

The inlet side of the cooling heat exchange unit 52 is connected to the outlet of the heat medium passage of the chiller 19. The cooling heat exchange unit 52 has a plurality of metal heat medium flow paths arranged so as to be in contact with the plurality of battery cells of the battery 80. The cooling heat exchange unit 52 is a heat exchange unit that cools the battery 80 by exchanging heat between the battery cell and the low temperature side heat medium flowing through the heat medium flow path.

The cooling heat exchange unit 52 is formed by arranging a heat medium flow path between the battery cells arranged in a laminated manner. The cooling heat exchange unit 52 may be integrally formed with the battery 80. For example, the battery 80 may be integrally formed by providing a heat medium flow path in a dedicated case for accommodating the stacked battery cells. The suction port side of the low temperature side heat medium pump 51 is connected to the outlet of the cooling heat exchange unit 52.

In the low temperature side heat medium circuit 50, the low temperature side heat medium pump 51 adjusts the flow rate of the low temperature side heat medium flowing into the cooling heat exchange unit 52, so that the low temperature side heat medium in the cooling heat exchange unit 52 becomes a battery. The amount of heat absorbed from 80 can be adjusted. Each component of the chiller 19 and the low temperature side heat medium circuit 50 is a battery cooling unit that cools the battery 80 by evaporating the refrigerant flowing out from the cooling expansion valve 14c.

The indoor air-conditioning unit 30 is for blowing out air whose temperature has been adjusted by the refrigeration cycle device 10 into the vehicle interior. The indoor air conditioning unit 30 is arranged inside the instrument panel (instrument panel) at the front of the vehicle interior.

The indoor air conditioning unit 30 accommodates a blower 32, an indoor evaporator 18, a heater core 42, and the like in an air passage formed in an air conditioning case 31 forming an outer shell thereof. The air conditioning case 31 forms an air passage for air to be blown into the vehicle interior. The air-conditioning case 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 air flow of the air conditioning case 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 air conditioning case 31.

The inside / outside air switching device 33 continuously adjusts the opening areas of the inside air introduction port for introducing the inside air into the air conditioning case 31 and the outside air introduction port for introducing the outside air by the inside / outside air switching door, and the introduction air volume of the inside air and the outside air. Change the introduction ratio with the introduction air volume. The inside / outside air switching door is driven by an electric actuator for the inside / outside air switching door. The electric actuator for the inside / outside air switching door is controlled by a control signal output from the control device 60.

A blower 32 is arranged on the downstream side of the air flow of the inside / outside air switching device 33. The blower 32 blows the air sucked through the inside / outside air switching device 33 toward the vehicle interior. The 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 blower 32 is controlled by the control voltage output from the control device 60.

On the downstream side of the air flow of the blower 32, the indoor evaporator 18, the heater core 42, and the electric heater 43 are arranged in this order with respect to the air flow. The indoor evaporator 18 is arranged on the upstream side of the air flow with respect to the heater core 42. The heater core 42 is arranged on the upstream side of the air flow with respect to the electric heater 43.

The electric heater 43 is an auxiliary heating unit that auxiliaryly heats the air. The electric heater 43 is, for example, a PTC heater having a PTC element (that is, a positive characteristic thermistor). The electric heater 43 can arbitrarily adjust the amount of heat for heating the air by the control voltage output from the control device 60. The electric heater 43 is a heating unit that heats air by using a heat source different from that of the refrigerant.

The air-conditioning case 31 is provided with a cold air bypass passage 35 that allows air after passing through the indoor evaporator 18 to bypass the heater core 42 and the electric heater 43. An air mix door 34 is arranged on the downstream side of the air flow of the indoor evaporator 18 in the air conditioning case 31 and on the upstream side of the air flow of the heater core 42.

The air mix door 34 adjusts the air volume ratio of the air after passing through the indoor evaporator 18 to the air volume of the air passing through the heater core 42 and the electric heater 43 side and the air volume of the air passing through the cold air bypass passage 35. It is an adjustment part. The air mix door 34 is driven by an electric actuator for the air mix door. This electric actuator is controlled by a control signal output from the control device 60.

A mixing space is arranged on the downstream side of the air flow of the heater core 42, the electric heater 43, and the cold air bypass passage 35 in the air conditioning case 31. The mixing space is a space in which the air heated by the heater core 42 and the electric heater 43 and the unheated air passing through the cold air bypass passage 35 are mixed.

In the downstream portion of the air flow of the air conditioning case 31, an opening hole for blowing out the air mixed in the mixing space (that is, the air conditioning air) into the vehicle interior, which is the air conditioning target space, is arranged.

As the opening hole, 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-conditioned air toward the inner side surface of the front window glass of the vehicle.

These face opening holes, foot opening holes, and defroster opening holes are the face outlets, foot outlets, and defroster outlets (none of which are shown) provided in the vehicle interior via ducts forming air passages, respectively. )It is connected to the.

The temperature of the conditioned air mixed in the mixing space is adjusted by adjusting the air volume ratio between the air volume passing through the heater core 42 and the electric heater 43 and the air volume passing through the cold air bypass passage 35 by the air mix door 34. .. As a result, the temperature of the air (air-conditioned air) blown from each outlet into the vehicle interior is adjusted.

A face door, a foot door, and a defroster door (none of which are shown) are arranged on the upstream side of the air flow of the face opening hole, the foot opening hole, and the defroster opening hole, respectively. The face door adjusts the opening area of the face opening hole. The foot door adjusts the opening area of the foot opening hole. The defroster door adjusts the opening area of the defroster opening hole.

These face doors, foot doors, and defroster doors constitute an outlet mode switching device for switching the outlet mode. These doors are connected to an electric actuator for driving the outlet mode door via a link mechanism or the like, and are rotated in conjunction with each other. The operation of this electric actuator is also controlled by a control signal output from the control device 60.

Specific examples of the outlet mode that can be switched by the outlet mode switching device include a face mode, a bi-level mode, and a foot mode.

The face mode is an outlet mode in which the face outlet is fully opened and air is blown from the face outlet toward the upper body of the passenger in the passenger compartment. The bi-level mode is an outlet mode in which both the face outlet and the foot outlet are opened to blow air toward the upper body and feet of the passengers in the passenger compartment. The foot mode is an outlet mode in which the foot outlet is fully opened and the defroster outlet is opened by a small opening, and air is mainly blown out from the foot outlet.

The occupant can also switch to the defroster mode by manually operating the blowout mode changeover switch provided on the operation panel 70 shown in FIG. 2. The defroster mode is an outlet mode in which the defroster outlet is fully opened and air is blown from the defroster outlet to the inner surface of the front window glass.

Next, the outline of the electric control unit of this embodiment will be described. The control device 60 includes a well-known microcomputer including a CPU, ROM, RAM, and the like, and peripheral circuits thereof. Then, various operations and processes are performed based on the control program stored in the ROM, and various controlled devices 11, 14a to 14c, 15a, 15b, 15d, 32, 41, 51, etc. connected to the output side thereof, etc. Control the operation of.

On the input side of the control device 60, as shown in the block diagram of FIG. 2, the inside temperature sensor 61, the outside temperature sensor 62, the solar radiation sensor 63, the first to fifth refrigerant temperature sensors 64a to 64e, and the evaporator temperature sensor 64f , 1st to 3rd refrigerant pressure sensors 65a to 65c, high temperature side heat medium temperature sensor 66a, low temperature side heat medium temperature sensor 67a, air conditioning air temperature sensor 68, battery temperature sensor 69 and the like are connected. Then, the detection signals of these sensor groups are input to the control device 60.

The internal air temperature sensor 61 is an internal air temperature detection unit that detects the internal air temperature Tr (that is, the vehicle interior temperature). The outside air temperature sensor 62 is an outside air temperature detection unit that detects the outside air temperature Tam (that is, the outside air temperature of the vehicle interior). The solar radiation sensor 63 is a solar radiation amount detection unit that detects the solar radiation amount As irradiated to the vehicle interior.

The first refrigerant temperature sensor 64a is a discharge refrigerant temperature detection unit that detects the temperature T1 of the refrigerant discharged from the compressor 11. The second refrigerant temperature sensor 64b is a second refrigerant temperature detecting unit that detects the temperature T2 of the refrigerant flowing out from the refrigerant passage of the refrigerant heat medium heat exchanger 12. The third refrigerant temperature sensor 64c is a third refrigerant temperature detecting unit that detects the temperature T3 of the refrigerant flowing out from the outdoor heat exchanger 16.

The fourth refrigerant temperature sensor 64d is a fourth refrigerant temperature detecting unit that detects the temperature T4 of the refrigerant flowing out of the indoor evaporator 18. The fifth refrigerant temperature sensor 64e is a fifth refrigerant temperature detecting unit that detects the temperature T5 of the refrigerant flowing out from the refrigerant passage of the chiller 19.

The evaporator temperature sensor 64f is an evaporator temperature detection unit that detects the evaporator temperature Tefin, which is the refrigerant evaporation temperature in the indoor evaporator 18. The evaporator temperature sensor 64f of the present embodiment detects the heat exchange fin temperature of the indoor evaporator 18.

The first refrigerant pressure sensor 65a is a first refrigerant pressure detecting unit that detects the pressure P1 of the refrigerant flowing out from the refrigerant passage of the refrigerant heat medium heat exchanger 12. The second refrigerant pressure sensor 65b is a second refrigerant pressure detecting unit that detects the pressure P2 of the refrigerant flowing out from the refrigerant passage of the chiller 19. The third refrigerant pressure sensor 65c is a third refrigerant pressure detecting unit that detects the pressure P3 of the refrigerant in the heating passage 22b.

The high temperature side heat medium temperature sensor 66a is a high temperature side heat medium temperature detection unit that detects the high temperature side heat medium temperature TWH, which is the temperature of the high temperature side heat medium flowing out from the heat medium passage of the refrigerant heat medium heat exchanger 12. The low temperature side heat medium temperature sensor 67a is a low temperature side heat medium temperature detection unit that detects the low temperature side heat medium temperature TWL, which is the temperature of the low temperature side heat medium in the heat medium passage of the chiller 19.

The conditioned air temperature sensor 68 is an conditioned air temperature detecting unit that detects the air temperature TAV blown from the mixed space to the vehicle interior.

The battery temperature sensor 69 is a battery temperature detection unit that detects the battery temperature TB (that is, the temperature of the battery 80). The battery temperature sensor 69 of the present embodiment has a plurality of temperature sensors and detects the temperature of a plurality of points of the battery 80. Therefore, the control device 60 can also detect the temperature difference of each part of the battery 80. As the battery temperature TB, the average value of the detected values of a plurality of temperature sensors is adopted.

As shown in FIG. 2, an operation panel 70 arranged near the instrument panel in the front part of the vehicle interior is connected to the input side of the control device 60, and operation signals from various operation switches provided on the operation panel 70 are connected. Is entered.

As various operation switches provided on the operation panel 70, specifically, an auto switch that sets or cancels the automatic control operation of the vehicle air conditioner, and a front seat side that requires the indoor evaporator 18 to cool the air. There are an air conditioner switch, an air volume setting switch for manually setting the air volume of the blower 32, a temperature setting switch for setting the target temperature Tset in the vehicle interior, a blowout mode changeover switch for manually setting the blowout mode, and the like.

The control device 60 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 60. The configuration (hardware and software) that controls the operation of each of the control target devices in the control device 60 is a control unit that controls the operation of each control target device.

For example, in the control device 60, the configuration for controlling the refrigerant discharge capacity of the compressor 11 (specifically, the rotation speed of the compressor 11) is the compressor control unit 60a. Further, the configuration for controlling the operation of the heating expansion valve 14a, the cooling expansion valve 14b, and the cooling expansion valve 14c is the expansion valve control unit 60b. The configuration that controls the operation of the dehumidifying on-off valve 15a, the heating on-off valve 15b, and the cooling on-off valve 15d is the refrigerant circuit switching control unit 60c.

The configuration for controlling the pumping capacity of the high temperature side heat medium of the high temperature side heat medium pump 41 is the high temperature side heat medium pump control unit 60d. The configuration for controlling the pumping capacity of the low temperature side heat medium of the low temperature side heat medium pump 51 is the low temperature side heat medium pump control unit 60e.

Next, the operation of the present embodiment in the above configuration will be described. The vehicle air conditioner 1 of the present embodiment air-conditions the interior of the vehicle and adjusts the temperature of the battery 80. In the refrigerating cycle device 10, the refrigerant circuit can be switched to operate in a plurality of operation modes. As the plurality of operation modes, for example, there are the following three types of operation modes.

(1) Cooling mode: The cooling mode is an operation mode in which the interior of the vehicle is cooled by cooling the air and blowing it into the interior of the vehicle without cooling the battery 80.

(2) Heating mode: The heating mode is an operation mode in which the interior of the vehicle is heated by heating the air and blowing it into the vehicle interior without cooling the battery 80.

(3) Waste heat recovery heating mode: The waste heat recovery heating mode is an operation mode in which the inside of the vehicle is heated by heating the air and blowing it into the vehicle interior while recovering the waste heat of the battery 80.

Switching between the cooling mode and the heating mode is performed by operating various operation switches provided on the operation panel 70 by an occupant. Switching between cooling mode and heating mode is also performed by executing a control program. Switching between the heating mode and the waste heat recovery heating mode is performed by executing the control program.

For example, if it is determined that the target outlet temperature TAO is below a predetermined cooling reference temperature, the cooling mode is selected.

The target blowout temperature TAO is the target temperature of the air blown into the vehicle interior space. Specifically, the target blowout temperature TAO is calculated by the following mathematical formula F1.
TAO = Kset x Tset-Kr x Tr-Kam x Tam-Ks x As + C ... (F1)
In this formula, Tset is the vehicle interior set temperature set by the temperature setting switch. Tr is the vehicle interior temperature detected by the internal air temperature sensor 61. Tam is the outside temperature of the vehicle interior detected by the outside air temperature sensor 62. As is the amount of solar radiation detected by the solar radiation sensor 63. Kset, Kr, Kam, and Ks are control gains, and C is a correction constant.

For example, if it is determined that the target outlet temperature TAO is above a predetermined heating reference temperature and the battery 80 does not need to be cooled, the heating mode is selected. For example, if it is determined that the target outlet temperature TAO is above a predetermined heating reference temperature and the battery 80 needs to be cooled, the waste heat recovery heating mode is selected.

The control device 60 controls the operation of the high-temperature side heat medium pump 41 so as to exert the reference pumping capacity for each predetermined operation mode regardless of the operation mode of the refrigeration cycle device 10 described above.

Therefore, in the high temperature side heat medium circuit 40, when the high temperature side heat medium is heated in the heat medium passage of the refrigerant heat medium heat exchanger 12, the heated high temperature side heat medium is pressure-fed to the heater core 42. The high temperature side heat medium flowing into the heater core 42 exchanges heat with air. This heats the air. The high temperature side heat medium flowing out of the heater core 42 is sucked into the high temperature side heat medium pump 41 and pressure-fed to the refrigerant heat medium heat exchanger 12.

The control device 60 controls the operation of the low temperature side heat medium pump 51 so as to exert the reference pumping capacity for each predetermined operation mode regardless of the operation mode of the refrigeration cycle device 10 described above.

Therefore, in the low temperature side heat medium circuit 50, when the low temperature side heat medium is cooled in the heat medium passage of the chiller 19, the cooled low temperature side heat medium is pressure-fed to the cooling heat exchange unit 52. The low temperature side heat medium flowing into the cooling heat exchange unit 52 absorbs heat from the battery 80. This cools the battery 80. The low temperature side heat medium flowing out of the cooling heat exchange unit 52 is sucked into the low temperature side heat medium pump 51 and pumped to the chiller 19.

The operation of the vehicle air conditioner 1 in each operation mode will be described below. In each operation mode, the control device 60 executes the control flow of each operation mode.

(1) Cooling mode In the control flow of the cooling mode, the target evaporator temperature TEO is determined in the first step. The target evaporator temperature TEO is determined based on the target blowout temperature TAO with reference to the control map stored in the control device 60. In the control map of the present embodiment, it is determined that the target evaporator temperature TEO increases as the target blowout temperature TAO increases.

In the next step, the amount of increase / decrease ΔIVO in the rotation speed of the compressor 11 is determined. The increase / decrease amount ΔIVO is set so that the evaporator temperature Tefin approaches the target evaporator temperature TEO by the feedback control method based on the deviation between the target evaporator temperature TEO and the evaporator temperature Tefin detected by the evaporator temperature sensor 64f. It is determined.

In the next step, the target superheat degree SHEO of the outlet side refrigerant of the indoor evaporator 18 is determined. As the target superheat degree SHEO, a predetermined constant (5 ° C. in this embodiment) can be adopted.

In the next step, the amount of increase / decrease ΔEVC in the throttle opening of the cooling expansion valve 14b is determined. The increase / decrease amount ΔEVC is based on the deviation between the target superheat degree SHEO and the superheat degree SH of the outlet side refrigerant of the indoor evaporator 18, and the superheat degree SH of the outlet side refrigerant of the indoor evaporator 18 is the target superheat degree by the feedback control method. Determined to approach SHEO.

The degree of superheat SHE of the outlet-side refrigerant of the indoor evaporator 18 is calculated based on the temperature T4 and the evaporator temperature Tefin detected by the fourth refrigerant temperature sensor 64d.

In the next step, the opening SW of the air mix door 34 is calculated using the following formula F2.
SW = {TAO- (Tefin + C2)} / {TWH- (Tefin + C2)} ... (F2)
In this formula, TWH is the high temperature side heat medium temperature detected by the high temperature side heat medium temperature sensor 66a. C2 is a constant for control.

In the next step, in order to switch the refrigerating cycle device 10 to the cooling mode refrigerant circuit, the dehumidifying on-off valve 15a and the heating on-off valve 15b are closed, the cooling on-off valve 15d is opened, and the heating expansion valve 14a is fully opened. Then, the cooling expansion valve 14b is set to the throttled state. Further, a control signal or a control voltage is output to each controlled device so that the control state determined in the above step can be obtained, and the process returns to the first step.

Therefore, in the refrigerating cycle device 10 in the cooling mode, as shown by the thick solid line in FIG. 1, the refrigerant discharged from the compressor 11 is the refrigerant heat medium heat exchanger 12, the heating expansion valve 14a, and the outdoor heat exchanger 16. , The second fixed throttle 26b, the receiver 25, the expansion valve 14b for cooling, the indoor evaporator 18, and the suction port of the compressor 11 circulate in this order.

That is, in the refrigerating cycle device 10 in the cooling mode, the refrigerant heat medium heat exchanger 12 and the outdoor heat exchanger 16 function as a heat radiating unit for dissipating the refrigerant discharged from the compressor 11, and the cooling expansion valve 14b dissipates the refrigerant. A steam compression type refrigeration cycle is configured which functions as a decompression unit for depressurizing and the indoor evaporator 18 functions as an evaporator.

According to this, the air can be cooled by the indoor evaporator 18, and the high temperature side heat medium can be heated by the refrigerant heat medium heat exchanger 12.

Therefore, in the vehicle air conditioner 1 in the cooling mode, a part of the air cooled by the indoor evaporator 18 is reheated by the heater core 42 by adjusting the opening degree of the air mix door 34, and approaches the target blowout temperature TAO. By blowing out the air whose temperature has been adjusted so as to be blown into the vehicle interior, the vehicle interior can be cooled.

When it becomes necessary to cool the battery 80 in the cooling mode, the cooling expansion valve 14c is set to the throttled state. As a result, the refrigerant flowing out of the receiver 25 is decompressed by the cooling expansion valve 14c and flows into the chiller 19, so that the chiller 19 functions as an evaporator and the low-pressure side heat medium can be cooled by the chiller 19. Then, the low temperature side heat medium cooled by the chiller 19 flows into the cooling heat exchange unit 52, so that the battery 80 can be cooled.

(2) Heating mode In the first step of the control flow of the heating mode, the increase / decrease amount ΔIVO of the rotation speed of the compressor 11 is determined. In the heating mode, the increase / decrease amount ΔIVO is determined so that the air temperature TAV approaches the target air temperature TAO by the feedback control method based on the deviation between the target air temperature TAO and the air temperature TAV detected by the air conditioning air temperature sensor 68. Will be done.

However, when the rotation speed of the compressor 11 increased or decreased by the increase / decrease amount ΔIVO exceeds the upper limit rotation speed Nc_limit, the rotation speed of the compressor 11 is determined to be the upper limit rotation speed Nc_limit. That is, the rotation speed of the compressor 11 is controlled to be equal to or less than the upper limit rotation speed Nc_limit.

The upper limit rotation speed Nc_limit is the rotation speed of the compressor 11 such that the flow rate of the refrigerant discharged from the compressor 11 (hereinafter referred to as the discharge refrigerant flow rate) becomes the upper limit flow rate Gr_limit.

As shown in FIG. 3, the upper limit flow rate Gr_limit is the discharge refrigerant flow rate of the compressor 11 when the power consumption of the compressor 11 is the upper limit power Pe_limit. FIG. 3 shows a heat pump heating capacity (in other words, a heat dissipation capacity in the refrigerant heat medium heat exchanger 12) with respect to the power consumption of the compressor 11 and a heating capacity (in other words) with respect to the power consumption of the electric heater 43 under predetermined conditions. If so, it is a graph showing an example of comparison with the heating capacity of the electric heater 43).

The rate of change in the heat pump heating capacity with respect to the power consumption of the compressor 11 is defined as the first rate of change, and the rate of change in the heating capacity with respect to the power consumption of the electric heater 43 is defined as the second rate of change. That is, the first rate of change is the slope of the heat pump heating capacity in the graph of FIG. 3, and the second rate of change is the slope of the electric heater heating capacity in the graph of FIG.

As can be seen from FIG. 3, the first rate of change has a characteristic that it becomes smaller as the power consumption increases. The second rate of change has a characteristic of being constant regardless of power consumption.

The upper limit power Pe_limit shown in FIG. 3 is the power consumption when the first change rate and the second change rate match. That is, the upper limit power Pe_limit is the power consumption when the slope of the heat pump heating capacity and the slope of the electric heater heating capacity match in the graph of FIG. In other words, the magnitude relationship between the first change rate and the second change rate is reversed with the upper limit power Pe_limit as the boundary.

The upper limit flow rate Gr_limit has a correlation with the low pressure pressure Ps of the refrigeration cycle, and is substantially constant with respect to the low pressure pressure Ps of the refrigeration cycle. Therefore, the upper limit flow rate Gr_limit is calculated using the map shown in FIG. 4 based on the low pressure pressure Ps of the refrigeration cycle.

In the heating mode, the low pressure Ps of the refrigeration cycle is calculated using a control map experimentally acquired in advance based on the outside air temperature Tam detected by the outside air temperature sensor 62. This control map is a control map in which the low pressure pressure Ps is substantially proportional to the outside air temperature Tam, and differs depending on the characteristics of the outdoor heat exchanger 16.

That is, in the heating mode, the refrigerant that has been heat exchanged with the outside air in the outdoor heat exchanger 16 and evaporated is sucked into the compressor 11, so that the low pressure pressure Ps in the refrigeration cycle is the saturation temperature of the refrigerant on the outlet side of the outdoor heat exchanger 16. Is proportional to. Since the saturation temperature of the refrigerant on the outlet side of the outdoor heat exchanger 16 is substantially the same as the outside air temperature Tam, the low pressure Ps in the refrigeration cycle is substantially proportional to the outside air temperature Tam. Therefore, the low pressure Ps of the refrigeration cycle can be calculated based on the outside air temperature Tam.

The upper limit rotation speed Nc_limit is determined using the following formula F3 based on the upper limit flow rate Gr_limit.
Nc_limit = Gr_limit / (ρ × ηv × Vc × 60)… (F3)
ρ is the density of the refrigerant sucked into the compressor 11 (hereinafter referred to as the suction refrigerant density), and is calculated using the map shown in FIG. 5 based on the low pressure pressure Ps of the refrigeration cycle. ηv is the volumetric efficiency of the compressor 11, and Vc is the capacity of the compressor 11.

As can be seen from FIG. 5, the intake refrigerant density ρ becomes smaller as the low pressure pressure Ps in the refrigeration cycle becomes lower. Therefore, the upper limit rotation speed Nc_limit becomes larger as the low pressure pressure Ps in the refrigeration cycle is lower when compared at the same upper limit flow rate Gr_limit.

When the rotation speed of the compressor 11 is determined to be the upper limit rotation speed Nc_limit, the control device 60 operates the electric heater 43. The output of the electric heater 43 is determined so that the air temperature TAV approaches the target air temperature TAO by the feedback control method based on the deviation between the target air temperature TAO and the air temperature TAV detected by the air conditioning air temperature sensor 68. ..

In the next step, the target superheat degree SHDO of the refrigerant flowing out from the outdoor heat exchanger 16 is determined. As the target superheat degree SHDO of the refrigerant flowing out from the outdoor heat exchanger 16, a predetermined constant (5 ° C. in this embodiment) can be adopted.

In the next step, the amount of increase / decrease ΔEVH in the throttle opening of the heating expansion valve 14a is determined. The increase / decrease amount ΔEVH is the target superheat degree SHD of the refrigerant flowing out from the outdoor heat exchanger 16 by the feedback control method based on the deviation between the target superheat degree SHDO and the superheat degree SHD of the refrigerant flowing out from the outdoor heat exchanger 16. The degree of superheat is determined to be close to SHDO.

The degree of superheat SHD of the refrigerant flowing out of the outdoor heat exchanger 16 is calculated based on the temperature T3 detected by the third refrigerant temperature sensor 64c and the pressure P1 detected by the first refrigerant pressure sensor 65a.

In the next step, the opening SW of the air mix door 34 is calculated in the same manner as in the cooling mode. Here, in the heating mode, the target outlet temperature TAO is higher than in the cooling mode, so that the opening SW of the air mix door 34 approaches 100%. Therefore, in the heating mode, the opening degree of the air mix door 34 is determined so that substantially the entire flow rate of the air after passing through the indoor evaporator 18 passes through the heater core 42.

In the next step, in order to switch the refrigerating cycle device 10 to the refrigerant circuit in the heating mode, the dehumidifying on-off valve 15a and the heating on-off valve 15b are opened, the cooling on-off valve 15d is closed, and the heating expansion valve 14a is throttled. The cooling expansion valve 14b and the cooling expansion valve 14c are fully closed. Further, a control signal or a control voltage is output to each controlled device so that the control state determined in the above step can be obtained, and the process returns to the first step.

As a result, in the refrigerating cycle device 10 in the heating mode, as shown by the thick solid line in FIG. 6, the refrigerant discharged from the compressor 11 is the refrigerant heat medium heat exchanger 12, the first fixed throttle 26a, the receiver 25, and the heating. The expansion valve 14a, the outdoor heat exchanger 16, and the suction port of the compressor 11 circulate in this order.

That is, in the refrigerating cycle device 10 in the heating mode, the refrigerant heat medium heat exchanger 12 functions as a heat dissipation unit that dissipates heat from the refrigerant discharged from the compressor 11, and the heating expansion valve 14a functions as a pressure reducing unit, and outdoor heat is generated. A refrigeration cycle is configured in which the exchanger 16 functions as an evaporative unit.

According to this, heat can be absorbed from the outside air by the outdoor heat exchanger 16 and the high temperature side heat medium can be heated by the refrigerant heat medium heat exchanger 12. Therefore, in the vehicle air conditioner 1 in the heating mode, the interior of the vehicle can be heated by blowing out the air heated by the heater core 42 into the interior of the vehicle.

(3) Waste heat recovery heating mode In the first step of the control flow of the waste heat recovery heating mode, as in the heating mode, the target high temperature side heat medium temperature TWHO and the amount of increase / decrease in the rotation speed of the compressor 11 are the same as in the heating mode. ΔIVO and the amount of increase / decrease in the throttle opening of the heating expansion valve 14a ΔEVH are determined.

Similar to the heating mode, when the rotation speed of the compressor 11 increased / decreased by the increase / decrease amount ΔIVO exceeds the upper limit rotation speed Nc_limit, the rotation speed of the compressor 11 is determined to be the upper limit rotation speed Nc_limit.

The upper limit rotation speed Nc_limit is determined based on the upper limit flow rate Gr_limit as in the heating mode. However, in the waste heat recovery heating mode, the upper limit flow rate Gr_limit is increased according to the waste heat absorption amount Qw (hereinafter referred to as waste heat recovery amount Qw) in the chiller 19 with respect to the heating mode.

That is, in the waste heat recovery heating mode, the upper limit flow rate Gr_limit is determined using the following mathematical formula F4.
Gr_limit = Gr_limit0 × (1 + α × Qw) ... (F4)
Gr_limit0 is the upper limit flow rate in the heating mode (that is, the upper limit flow rate calculated using the map shown in FIG. 4).

α is an experimentally determined constant. The waste heat recovery amount Qw in the chiller 19 is calculated using the map shown in FIG. 7 based on the difference between the temperature TWL of the low temperature side heat medium in the chiller 19 and the saturation temperature Ts of the outlet side refrigerant of the chiller 19. ..

The temperature TWL of the low temperature side heat medium flowing into the chiller 19 is the temperature of the low temperature side heat medium detected by the low temperature side heat medium temperature sensor 67a. The saturation temperature Ts of the outlet-side refrigerant of the chiller 19 is calculated based on the refrigerant pressure P2 detected by the second refrigerant pressure sensor 65b.

As described above, in the waste heat recovery heating mode, the upper limit flow rate Gr_limit is increased in proportion to the waste heat recovery amount Qw with respect to the heating mode. As can be seen from the equation F4, the difference between the upper limit flow rate Gr_limit in the waste heat recovery heating mode and the upper limit flow rate Gr_limit0 in the heating mode increases as the waste heat recovery amount Qw increases.

In the waste heat recovery heating mode, the low pressure pressure Ps of the refrigeration cycle is calculated using the following formula F5 based on the outside air temperature Tam detected by the outside air temperature sensor 62 and the waste heat recovery amount Qw in the chiller 19.
Ps = Ps0 + ΔPs ... (F5)
Ps0 is the base low pressure, which is the low pressure calculated by the same calculation method as in the heating mode. That is, the base low-pressure pressure Ps0 is a low-pressure pressure calculated based on the outside air temperature Tam using a control map that is substantially proportional to the outside air temperature Tam.

ΔPs is a correction amount of the base low pressure pressure Ps0 based on the waste heat recovery amount Qw in the chiller 19. The correction amount ΔPs is calculated using a control map experimentally acquired in advance based on the waste heat recovery amount Qw in the chiller 19. This control map is a control map in which the correction amount ΔPs is substantially proportional to the waste heat recovery amount Qw, and differs depending on the characteristics of the outdoor heat exchanger 16.

The reason why the correction amount ΔPs is calculated in this way will be described. In the waste heat recovery heating mode, the amount of heat absorbed by the outdoor heat exchanger 16 is reduced by the amount of the recovered waste heat as compared with the heating mode. This is because the low pressure of the refrigeration cycle rises in proportion to the waste heat recovery amount Qw, so that the difference between the outside air temperature Tam and the refrigerant temperature in the outdoor heat exchanger 16 becomes small. Therefore, the low pressure pressure Ps of the refrigeration cycle in the waste heat recovery heating mode can be calculated by correcting the base low pressure pressure Ps0 corresponding to the low pressure in the heating mode with the correction amount ΔPs corresponding to the waste heat recovery amount Qw. ..

In the next step, the target superheat degree SHCO of the refrigerant flowing out of the chiller 19 is determined. As the target superheat degree SHCO of the refrigerant flowing out of the chiller 19, a predetermined constant (5 ° C. in the present embodiment) can be adopted.

In the next step, the amount of increase / decrease ΔEVB in the throttle opening of the cooling expansion valve 14c is determined. The increase / decrease amount ΔEVB is such that the superheat degree SHCO of the refrigerant flowing out of the chiller 19 approaches the target superheat degree SHCO by the feedback control method based on the deviation between the target superheat degree SHCO and the superheat degree SHCO of the refrigerant flowing out from the chiller 19. Will be decided. The degree of superheat SHC of the refrigerant flowing out of the chiller 19 is calculated based on the temperature T5 detected by the fifth refrigerant temperature sensor 64e and the pressure P2 detected by the second refrigerant pressure sensor 65b.

In the next step, the opening SW of the air mix door 34 is calculated in the same manner as in the cooling mode. Here, in the waste heat recovery heating mode, the target blowing temperature TAO is higher than in the cooling mode, so that the opening SW of the air mix door 34 approaches 100%. Therefore, in the heating mode, the opening degree of the air mix door 34 is determined so that substantially the entire flow rate of the air after passing through the indoor evaporator 18 passes through the heater core 42.

In the next step, in order to switch the refrigerating cycle device 10 to the refrigerant circuit in the waste heat recovery heating mode, the dehumidifying on-off valve 15a and the heating on-off valve 15b are opened, the cooling on-off valve 15d is closed, and the heating on-off valve 14a is closed. The cooling expansion valve 14c is set to the throttled state, and the cooling expansion valve 14b is set to the fully closed state. Further, a control signal or a control voltage is output to each controlled device so that the control state determined in the above step can be obtained, and the process returns to the first step.

As a result, in the refrigeration cycle device 10 in the waste heat recovery heating mode, as shown by the thick solid line in FIG. 8, the refrigerant discharged from the compressor 11 is the refrigerant heat medium heat exchanger 12, the first fixed throttle 26a, and the receiver. 25, heating expansion valve 14a, outdoor heat exchanger 16, compressor 11, suction port, refrigerant heat medium heat exchanger 12, first fixed throttle 26a, receiver 25, cooling expansion valve 14c, chiller It circulates in the order of 19 and the suction port of the compressor 11.

That is, in the refrigerating cycle device 10 in the waste heat recovery heating mode, the refrigerant heat medium heat exchanger 12 functions as a heat radiating unit for dissipating the refrigerant discharged from the compressor 11, and the heating expansion valve 14a and the cooling expansion valve 14c. Functions as a decompression unit, and a refrigeration cycle in which the outdoor heat exchanger 16 and the chiller 19 function as an evaporation unit is configured.

According to this, the outdoor heat exchanger 16 can absorb heat from the outside air, the chiller 19 can absorb heat from the low temperature side heat medium, and the refrigerant heat medium heat exchanger 12 can heat the high temperature side heat medium. Therefore, in the vehicle air conditioner 1 in the waste heat recovery heating mode, the vehicle interior can be heated by blowing out the air heated by the heater core 42 into the vehicle interior.

As described above, in the refrigerating cycle apparatus 10 of the present embodiment, various operation modes can be switched. As a result, the vehicle air conditioner 1 can realize comfortable air conditioning in the vehicle interior while appropriately adjusting the temperature of the battery 80.

In the heating mode and the waste heat recovery heating mode, the rotation speed of the compressor 11 is controlled to be equal to or lower than the upper limit rotation speed Nc_limit, so that the discharge refrigerant flow rate of the compressor 11 does not exceed the upper limit flow rate Gr_limit. Therefore, the compressor 11 can be operated efficiently with the upper limit power Pe_limit as much as possible. Since the electric heater 43 efficiently compensates for the insufficient heating capacity when the compressor 11 is operated with the upper limit power Pe_limit, the refrigerating cycle device 10 and the electric heater 43 can be efficiently linked and heated.

Further, since the refrigeration cycle apparatus 10 of the present embodiment includes the first fixed throttle 26a and the second fixed throttle 26b, the coefficient of performance of the cycle can be improved.

This will be described with reference to FIG. FIG. 9 is a Moriel diagram showing the state of the refrigerant in the refrigeration cycle device 10 in the heating mode. In the heating mode, the refrigerant heat medium heat exchanger 12 serves as a heat exchange unit for condensing the refrigerant. Further, the outdoor heat exchanger 16 serves as a heat exchange unit for evaporating the refrigerant.

In FIG. 9, the change in the state of the refrigerant in the refrigerating cycle apparatus 10 of the present embodiment provided with the first fixed throttle 26a is shown by a thick solid line. Further, the change in the state of the refrigerant in the refrigeration cycle apparatus of the comparative example not provided with the first fixed throttle 26a is shown by a broken line.

In FIG. 9, the state of the refrigerant in the receiver 25 in the refrigeration cycle apparatus 10 of the present embodiment is shown by the point Lq1. Further, in FIG. 9, the state of the refrigerant in the receiver 25 in the refrigeration cycle device of the comparative example is shown by a point Lqex.

Since the refrigerating cycle device 10 of the present embodiment includes the first fixed throttle 26a, the pressure of the refrigerant in the receiver 25 is lower than the pressure of the high-pressure refrigerant in the heat exchange section that condenses the refrigerant. Therefore, as shown in FIG. 9, the pressure of the refrigerant at the point Lq1 of the refrigeration cycle apparatus 10 of the present embodiment is lower than the pressure of the refrigerant at the point Lqex of the refrigeration cycle apparatus of the comparative example.

Along the slope of the saturated liquid line in the Moriel diagram, the enthalpy of the refrigerant at the point Lq1 of the refrigeration cycle apparatus 10 of the present embodiment is lower than the enthalpy of the refrigerant at the point Lqex of the refrigeration cycle apparatus of the comparative example. Therefore, in the refrigerating cycle apparatus 10 of the present embodiment, the refrigerant on the outlet side of the heat exchange unit that condenses the refrigerant is the supercooled liquid phase refrigerant SC1.

Therefore, in the refrigerating cycle device 10 of the present embodiment, the enthalpy of the refrigerant flowing into the heat exchange unit that evaporates the refrigerant can be lowered as compared with the refrigerating cycle device 10 of the comparative example. As a result, the amount of heat absorbed by the refrigerant in the heat exchange unit that evaporates the refrigerant can be increased, and the coefficient of performance can be improved. This effect can also be obtained in other operating modes. That is, the first fixed throttle 26a and the second fixed throttle 26b are supercooling decompression portions that give the refrigerant a degree of supercooling.

The first fixed throttle 26a and the second fixed throttle 26b need to be selected according to the flow rate of the refrigerant. That is, as the first fixed throttle 26a and the second fixed throttle 26b, a fixed throttle having a large opening is required as the refrigerant flow rate is large, and a fixed throttle having a small opening is required as the refrigerant flow rate is small. When the range of change in the refrigerant flow rate of the refrigeration cycle device 10 is large, it becomes difficult to select an appropriate fixed throttle as the first fixed throttle 26a and the second fixed throttle 26b.

In that respect, in the present embodiment, since the discharge refrigerant flow rate of the compressor 11 is limited to the upper limit flow rate Gr_limit or less, the range of change in the refrigerant flow rate of the refrigeration cycle device 10 is relatively small. As a result, it becomes easy to select an appropriate fixed diaphragm as the first fixed diaphragm 26a and the second fixed diaphragm 26b.

In the present embodiment, the control device 60 controls the compressor 11 so that the power consumption of the compressor 11 is equal to or less than the upper limit power Pe_limit in the heating mode and the waste heat recovery heating mode. According to this, since the heat dissipation efficiency of the refrigerant heat medium heat exchanger 12 can be made higher than the heating efficiency of the electric heater 43, the heating efficiency can be maximized as much as possible.

In the present embodiment, the control device 60 controls the compressor 11 so that the flow rate of the refrigerant discharged by the compressor 11 is equal to or less than the upper limit flow rate Gr_limit in the heating mode and the waste heat recovery heating mode. The power consumption of is set to be equal to or less than the upper limit power Pe_limit. As a result, the power consumption of the compressor 11 can be surely reduced to the upper limit power Pe_limit or less.

In the present embodiment, the control device 60 increases the upper limit flow rate Gr_limit in the waste heat recovery heating mode as compared with the heating mode. According to this, the power consumption of the compressor 11 can be appropriately reduced to the upper limit power Pe_limit or less not only in the heating mode but also in the waste heat recovery heating mode.

In the present embodiment, the control device 60 increases the amount ΔGr_limit that increases the upper limit flow rate Gr_limit in the waste heat recovery heating mode as compared with the heating mode, as the waste heat recovery amount Qw in the chiller 19 increases. According to this, the power consumption of the compressor 11 can be more appropriately reduced to the upper limit power Pe_limit or less in the waste heat recovery heating mode.

In the present embodiment, the control device 60 controls the compressor 11 so that the rotation speed of the compressor 11 is equal to or less than the upper limit rotation speed Nc_limit, so that the flow rate of the refrigerant discharged by the compressor 11 is set to the upper limit flow rate Gr_limit or less. The upper limit rotation rate Nc_limit is determined based on the pressure of the refrigerant sucked by the compressor 11. According to this, the flow rate of the refrigerant discharged by the compressor 11 can be set to the upper limit flow rate Gr_limit or less by a simple method.

In the present embodiment, in the waste heat recovery heating mode, the control device 60 calculates the pressure Ps of the refrigerant sucked by the compressor 11 based on the temperature Tam of the outside air and the waste heat recovery amount Qw in the chiller 19. As a result, the pressure Ps of the refrigerant sucked by the compressor 11 (that is, the low pressure in the refrigeration cycle) can be obtained without using the pressure sensor.

In the present embodiment, the control device 60 calculates the waste heat recovery amount Qw in the chiller 19 based on the difference between the temperature TWL of the low temperature side heat medium in the chiller 19 and the saturation temperature Ts of the refrigerant. Thereby, the waste heat recovery amount Qw in the chiller 19 can be easily obtained.

In the present embodiment, the first fixed throttle 26a is arranged on the downstream side of the refrigerant flow of the refrigerant heat medium heat exchanger 12 and the upstream side of the refrigerant flow of the receiver 25 in the heating mode and the waste heat recovery heating mode, and the first fixed throttle 26a is arranged. In 26a, the pressure of the refrigerant is reduced to give the refrigerant a degree of supercooling. According to this, the cycle coefficient of performance can be improved by giving the refrigerant a degree of supercooling. Since the refrigerant is depressurized using a fixed throttle, the refrigerant can be supercooled with a simple configuration.

(Second Embodiment)
In the first embodiment described above, the discharge refrigerant flow rate of the compressor 11 is limited to the upper limit flow rate Gr_limit or less by limiting the rotation speed of the compressor 11 to the upper limit rotation speed Nc_limit or less in the heating mode and the waste heat recovery heating mode. Will be done. On the other hand, in the present embodiment, the heating temperature (in other words, the heat pump heating temperature) by the refrigerant heat medium heat exchanger 12 is limited to the upper limit temperature TAO_HP or less in the heating mode and the waste heat recovery heating mode. The discharge refrigerant flow rate of the compressor 11 is limited to the upper limit flow rate Gr_limit or less.

In the present embodiment, the target blowout temperature TAO is calculated in the same manner as in the first embodiment described above. In the heating mode and the waste heat recovery heating mode, the upper limit temperature TAO_HP is calculated using the map shown in FIG. 10 based on the low pressure Ps of the refrigeration cycle. The upper limit temperature TAO_HP is a value of the heating temperature (in other words, the temperature T1 of the refrigerant discharged from the compressor 11) by the refrigerant heat medium heat exchanger 12 such that the flow rate of the refrigerant discharged from the compressor 11 becomes the upper limit flow rate Gr_limit. Is.

The temperature T1 of the refrigerant discharged from the compressor 11 and the discharged refrigerant flow rate Gr have the relationship shown in the graph of FIG. The upper limit temperature TAO_HP becomes a smaller value as the low pressure pressure Ps in the refrigeration cycle becomes lower when the discharge refrigerant flow rate Gr is constant. Therefore, the upper limit temperature TAO_HP can be calculated based on the low pressure pressure Ps of the refrigeration cycle.

When the upper limit temperature TAO_HP is smaller than the target blowout temperature TAO, the rotation speed of the compressor 11 is such that the discharge refrigerant temperature T1 becomes the upper limit temperature TAO_HP by the feedback control method based on the deviation between the upper limit temperature TAO_HP and the discharge refrigerant temperature T1. Determined to approach.

When the upper limit temperature TAO_HP is smaller than the target blowout temperature TAO, the electric heater 43 is operated.

The output of the electric heater 43 is determined so that the air temperature TAV approaches the target air temperature TAO by the feedback control method based on the deviation between the target air temperature TAO and the air temperature TAV detected by the air conditioning air temperature sensor 68. ..

Also in this embodiment, as in the first embodiment described above, in the heating mode and the waste heat recovery heating mode, the flow rate of the discharged refrigerant of the compressor 11 does not exceed the upper limit flow rate Gr_limit, so that the refrigeration cycle device 10 is used as much as possible. , It can be operated efficiently with the upper limit power Pe_limit. Then, since the electric heater 43 efficiently compensates for the insufficient heating capacity when the refrigerating cycle device 10 is operated with the upper limit power Pe_limit, the refrigerating cycle device 10 and the electric heater 43 can be efficiently linked and heated.

In the present embodiment, the control device 60 controls the compressor 11 so that the temperature of the air heated by the refrigerant heat medium heat exchanger 12 is equal to or lower than the upper limit temperature TAO_HP, so that the refrigerant discharged by the compressor 11 is discharged. The flow rate is set to the upper limit flow rate Gr_limit or less, and the upper limit temperature TAO_HP is determined based on the pressure P3 of the refrigerant sucked by the compressor 11.

According to this, the flow rate of the refrigerant discharged by the compressor 11 can be set to the upper limit flow rate Gr_limit or less by a simple method.

(Third Embodiment)
In the first embodiment described above, the discharge refrigerant flow rate of the compressor 11 is limited to the upper limit flow rate Gr_limit or less by limiting the rotation speed of the compressor 11 to the upper limit rotation speed Nc_limit or less in the heating mode and the waste heat recovery heating mode. Will be done. On the other hand, in the present embodiment, the power consumption of the compressor 11 is limited to the upper limit power Pe_limit or less in the heating mode and the waste heat recovery heating mode, so that the discharge refrigerant flow rate of the compressor 11 becomes the upper limit flow Gr_limit or less. Be restricted.

In the present embodiment, as in the first embodiment described above, the increase / decrease amount ΔIVO of the number of revolutions of the compressor 11 is based on the deviation between the target blowing temperature TAO and the air temperature TAV detected by the air conditioning air temperature sensor 68. The feedback control method determines that the air temperature TAV approaches the target outlet temperature TAO.

In the heating mode and the waste heat recovery heating mode, the upper limit power Pe_limit is calculated using the map shown in FIG. 12 based on the low pressure Ps of the refrigeration cycle. The upper limit power Pe_limit is a value of the power consumption of the compressor 11 such that the flow rate of the discharged refrigerant from the compressor 11 becomes the upper limit flow rate Gr_limit.

The power consumption Pe_comp of the compressor 11 and the discharged refrigerant flow rate Gr have the relationship shown in the graph of FIG. The graph of FIG. 13 shows an example when the low pressure pressure Ps of the refrigeration 0 cycle is a specific value, but the power consumption Pe_comp and the discharge refrigerant flow rate of the compressor 11 regardless of the value of the low pressure pressure Ps of the refrigeration cycle. The relationship with Gr generally tends to be as shown in the graph of FIG. Therefore, as shown in FIG. 12, the upper limit power Pe_limit becomes a substantially constant value with respect to the low pressure pressure Ps of the refrigeration cycle.

In the heating mode and the waste heat recovery heating mode, when the power consumption of the compressor 11 becomes the upper limit power Pe_limit or more, the compressor 11 is controlled so that the power consumption of the compressor 11 becomes the upper limit power Pe_limit. That is, the rotation speed of the compressor 11 is controlled to be equal to or less than the upper limit rotation speed Nc_limit.

For example, the power consumption of the compressor 11 can be acquired based on the voltage value and the current value in the compressor 11.

The output of the electric heater 43 is determined so that the air temperature TAV approaches the target air temperature TAO by the feedback control method based on the deviation between the target air temperature TAO and the air temperature TAV detected by the air conditioning air temperature sensor 68. ..

Also in this embodiment, as in the first embodiment described above, in the heating mode and the waste heat recovery heating mode, the flow rate of the discharged refrigerant of the compressor 11 does not exceed the upper limit flow rate Gr_limit, so that the refrigeration cycle device 10 is used as much as possible. , It can be operated efficiently with the upper limit power Pe_limit. Then, since the electric heater 43 efficiently compensates for the insufficient heating capacity when the refrigerating cycle device 10 is operated with the upper limit power Pe_limit, the refrigerating cycle device 10 and the electric heater 43 can be efficiently linked and heated.

(Fourth Embodiment)
In the above-described embodiment, the refrigerating cycle device 10 includes the refrigerant heat medium heat exchanger 12 and the indoor air conditioning unit 30 includes the heater core 42. However, in the present embodiment, as shown in FIG. 14, the refrigerant heat medium heat is provided. An indoor condenser 27 is provided in place of the exchanger 12 and the heater core 42.

The indoor condenser 27 is a heat exchanger that heats the air by exchanging heat between the high-pressure refrigerant discharged from the compressor 11 and the air that has passed through the indoor evaporator 18.

Also in this embodiment, the operation in the cooling mode, the heating mode, and the waste heat recovery heating mode can be performed as in the first embodiment described above. Therefore, also in this embodiment, the same action and effect as those in the above-described embodiment can be obtained.

(Fifth Embodiment)
In the above embodiment, the refrigerant flows in parallel to the outdoor heat exchanger 16 and the chiller 19 in the waste heat recovery heating mode, but in the present embodiment, the outdoor heat exchanger 16 and the chiller 19 flow in the waste heat recovery heating mode. The refrigerant flows in series.

Specifically, as shown in FIG. 15, the cooling expansion valve 14c and the chiller 19 are arranged in the heating passage 22b.

In the present embodiment, the cooling expansion valve 14c is fully opened in the heating mode. As a result, the refrigerant flows through the cooling expansion valve 14c and the chiller 19 in the heating mode, but heat exchange is hardly performed in the chiller 19.

Also in this embodiment, the operation in the cooling mode, the heating mode, and the waste heat recovery heating mode can be performed in the same manner as in the above-described embodiment. Therefore, also in this embodiment, the same action and effect as those in the above-described embodiment can be obtained.

(Other embodiments)
The present invention is not limited to the above-described embodiment, and can be variously modified as follows without departing from the spirit of the present invention. In addition, the means disclosed in each of the above embodiments may be appropriately combined to the extent feasible.

(A) In the above-described embodiment, the refrigeration cycle device 10 capable of switching to a plurality of operation modes has been described, but the switching of the operation mode of the refrigeration cycle device 10 is not limited to this. It suffices if at least the heating mode or the waste heat recovery heating mode can be executed.

(B) The components of the refrigeration cycle apparatus are not limited to those disclosed in the above-described embodiment. A plurality of cycle components may be integrated so that the above-mentioned effects can be exhibited. For example, a four-way joint structure in which the second three-way joint 13b and the fifth three-way joint 13e are integrated may be adopted. Further, as the cooling expansion valve 14b and the cooling expansion valve 14c, those in which the electric expansion valve having no fully closed function and the on-off valve are directly connected may be adopted.

Further, in the above-described embodiment, an example in which R1234yf is adopted as the refrigerant has been described, but the refrigerant is not limited to this. For example, R134a, R600a, R410A, R404A, R32, R407C, etc. may be adopted. Alternatively, a mixed refrigerant or the like in which a plurality of types of these refrigerants are mixed may be adopted. Further, carbon dioxide may be adopted as the refrigerant to form a supercritical refrigeration cycle in which the pressure of the refrigerant on the high pressure side is equal to or higher than the critical pressure of the refrigerant.

(C) The configuration of the heating unit is not limited to that disclosed in the above-described embodiment. For example, in a vehicle equipped with an internal combustion engine (engine) such as a hybrid vehicle, engine cooling water may be circulated in the high temperature side heat medium circuit 40.

(D) The configuration of the battery cooling unit is not limited to that disclosed in the above-described embodiment. For example, as the battery cooling unit, a thermosiphon may be adopted in which the chiller 19 of the low temperature side heat medium circuit 50 described in the first embodiment is used as a condensing unit and the cooling heat exchange unit 52 functions as an evaporating unit. According to this, the low temperature side heat medium pump 51 can be abolished.

The thermosiphon has an evaporating part for evaporating the refrigerant and a condensing part for condensing the refrigerant, and is configured by connecting the evaporating part and the condensing part in a closed loop shape (that is, in a ring shape). Then, the temperature difference between the temperature of the refrigerant in the evaporating part and the temperature of the refrigerant in the condensing part causes a difference in specific gravity in the refrigerant in the circuit, and the action of gravity naturally circulates the refrigerant to transport heat together with the refrigerant. It is a circuit.

Further, in the above-described embodiment, the example in which the cooling target to be cooled by the battery cooling unit is the battery 80 has been described, but the cooling target is not limited to this. An inverter that converts direct current and alternating current, a charger that charges the battery 80 with electric power, and a motor generator that outputs driving force for driving by supplying electric power and generates regenerative electric power during deceleration. It may be an electric device that generates heat during operation.

(E) In the above-mentioned first embodiment, the number of revolutions of the compressor 11 is determined based on the deviation between the target outlet temperature TAO and the air temperature TAV in the heating mode and the waste heat recovery heating mode, but the target outlet temperature is determined. The rotation speed of the compressor 11 may be determined based on the deviation between the TAO and the temperature T1 of the refrigerant detected by the first refrigerant temperature sensor 64a (that is, the temperature of the refrigerant discharged from the compressor 11).

(F) In the above-mentioned second embodiment, when the upper limit temperature TAO_HP is smaller than the target outlet temperature TAO in the heating mode and the waste heat recovery heating mode, the upper limit temperature TAO_HP and the discharge refrigerant temperature T1 (detected by the first refrigerant temperature sensor 64a). The rotation speed of the compressor 11 is determined based on the deviation from the temperature T1) of the refrigerant.

On the other hand, the rotation speed of the compressor 11 is based on the deviation between the upper limit temperature TAO_HP and the high temperature side heat medium temperature TWH (that is, the temperature of the air heated by the heater core 42) detected by the high temperature side heat medium temperature sensor 66a. May be determined. The rotation speed of the compressor 11 may be determined based on the deviation between the upper limit temperature TAO_HP and the temperature of the air heated by the heater core 42.

(G) In each of the above-described embodiments, the refrigerating cycle apparatus 10 includes a first fixed throttle 26a and a second fixed throttle 26b as a supercooling degree generating unit for giving a supercooling degree to the condensed refrigerant, but the supercooling degree generation. As a unit, instead of the first fixed diaphragm 26a and the second fixed diaphragm 26b, a variable diaphragm capable of changing the diaphragm opening degree may be arranged.

In this configuration, the control device 60 controls the throttle opening of the variable throttle based on the temperature of the refrigerant on the downstream side of the refrigerant flow of the variable throttle and on the upstream side of the refrigerant flow of the receiver 25, whereby the degree of supercooling appropriate for the refrigerant is controlled. It is possible to surely improve the cycle coefficient of performance.

(H) In each of the above-described embodiments, the compressor 11 is an electric compressor, but the compressor 11 may be a compressor driven by the power of an engine.

In each of the above-described embodiments, the auxiliary heating unit for auxiliary heating the air is the electric heater 43, but the auxiliary heating unit may be a heater core using the waste heat of the engine as a heat source.

When the power source of the compressor 11 and the heat source of the auxiliary heating unit are other than electric power, the upper limit flow rate Gr_limit is the discharge refrigerant of the compressor 11 when the energy consumption per unit time of the compressor 11 is the upper limit value. The flow rate. The upper limit is the per-unit time when the magnitude relationship between the rate of change in the heating capacity with respect to the energy consumption of the compressor 11 per unit time and the rate of change in the heating capacity with respect to the energy consumption of the electric heater 43 is reversed. It is the value of energy consumption.

The electric heater as the auxiliary heating unit may be a seat heater that heats the surface of the seat on which the occupant sits. The auxiliary heating unit may be a radiant electric heater that heats the occupant. That is, the auxiliary heating unit is a heating unit that heats the human body in the air or the air-conditioned space by heat conduction or radiation using a heat source different from the refrigerant.

(I) In each of the above-described embodiments, the refrigeration cycle device 10 according to the present invention is applied to the vehicle air conditioner 1, but the application of the refrigeration cycle device 10 is not limited to this. For example, it may be applied to an air conditioner having a battery cooling function that air-conditions a room while appropriately adjusting the temperature of a stationary battery.

11 Compressor 12 Refrigerant heat medium heat exchanger (heat dissipation part)
14a Expansion valve for heating (pressure reducing part)
14c Cooling expansion valve (pressure reducing part)
16 Outdoor heat exchanger (evaporation part, first evaporation part)
18 Indoor evaporator (1st evaporator)
19 Chiller (2nd evaporation part)
40 High temperature side heat medium circuit (heat dissipation part)
42 Heater core (heat dissipation part)
43 Electric heater (heating part)
60 Control device (control unit)

Claims (15)

A compressor (11) that sucks in the refrigerant, compresses it, and discharges it.
A heat radiating unit (12, 40, 42, 27) that dissipates the refrigerant discharged from the compressor to the air blown into the air-conditioned space, and
The decompression unit (14a, 14c) that decompresses the refrigerant dissipated by the heat dissipation unit, and the decompression unit.
The evaporating unit (16), which evaporates the refrigerant by absorbing heat from the refrigerant decompressed by the depressurizing unit,
A heating unit (43) that heats the air or the human body in the air-conditioned space by using a heat source different from that of the refrigerant.
A control unit (60) for controlling the compressor is provided.
The first change rate, which is the ratio of the change in the heat dissipation capacity in the heat dissipation unit to the energy consumption per unit time of the compressor, and the change in the heating capacity in the heating unit with respect to the energy consumption per unit time of the heating unit. When the second change rate, which is the rate of, is compared with the energy consumption per unit time, the first change rate is the second change rate when the energy consumption per unit time is below the upper limit (Pe_limit). It has the characteristic that it exceeds the rate of change and falls below the second rate of change when the energy consumption per unit time exceeds the upper limit.
The control unit is a refrigeration cycle device that controls the compressor so that the energy consumption per unit time of the compressor is equal to or less than the upper limit value. The control unit controls the compressor so that the flow rate of the refrigerant discharged by the compressor is equal to or less than the upper limit flow rate (Gr_limit), whereby the energy consumption per unit time of the compressor is equal to or less than the upper limit value. The refrigeration cycle apparatus according to claim 1. The evaporation unit is a first evaporation unit that absorbs heat from the outside air to the refrigerant.
Further, a second evaporation unit (19) for evaporating the refrigerant by absorbing the waste heat generated by the waste heat generator (80) into the refrigerant decompressed by the decompression unit.
A switching unit (14c, 15b) for switching between a waste heat recovery heating mode in which the refrigerant flows in the second evaporation unit and a heating mode in which the refrigerant does not flow in the second evaporation unit is provided.
The refrigeration cycle device according to claim 2, wherein the control unit increases the upper limit flow rate in the waste heat recovery heating mode as compared with the heating mode. The control unit increases the amount (ΔGr_limit) of increasing the upper limit flow rate in the waste heat recovery heating mode as compared with the heating mode, as the amount of heat absorption (Qw) of the waste heat in the second evaporation unit increases. The refrigeration cycle apparatus according to claim 3. The control unit
By controlling the compressor so that the rotation speed of the compressor is equal to or less than the upper limit rotation speed (Nc_limit), the flow rate of the refrigerant discharged by the compressor is set to be equal to or less than the upper limit flow rate.
The refrigerating cycle apparatus according to claim 3 or 4, wherein the upper limit rotation speed is determined based on the pressure (Ps) of the refrigerant sucked by the compressor. The control unit
By controlling the compressor so that the temperature of the air heated by the heat radiating unit is equal to or lower than the upper limit temperature (TAO_HP), the flow rate of the refrigerant discharged by the compressor is set to be equal to or lower than the upper limit flow rate.
The refrigerating cycle apparatus according to claim 3 or 4, wherein the upper limit temperature is determined based on the pressure (Ps) of the refrigerant sucked by the compressor. In the waste heat recovery heating mode, the control unit calculates the pressure of the refrigerant sucked by the compressor based on the temperature of the outside air (Tam) and the heat absorption amount of the waste heat in the second evaporation unit. The refrigeration cycle apparatus according to claim 5 or 6. The second evaporation unit absorbs the waste heat to the refrigerant via a heat medium.
The control unit calculates the amount of waste heat absorbed by the second evaporation unit based on the difference between the temperature of the heat medium (TWL) and the saturation temperature (Ts) of the refrigerant in the second evaporation unit. The refrigeration cycle apparatus according to claim 4 or 7. A liquid storage unit (25) arranged on the downstream side of the refrigerant flow of the heat dissipation unit and upstream of the refrigerant flow of the decompression unit and storing the excess refrigerant.
A claim including a supercooling decompression unit (26a, 26b) arranged on the downstream side of the refrigerant flow of the heat dissipation unit and on the upstream side of the refrigerant flow of the liquid storage unit, and decompressing the refrigerant to give the refrigerant a degree of supercooling. Item 6. The refrigerating cycle apparatus according to any one of Items 2 to 8. The refrigeration cycle device according to claim 9, wherein the supercooling decompression unit is a fixed throttle having a fixed throttle opening. The supercooling decompression unit is a variable throttle capable of changing the throttle opening.
9. The control unit controls the throttle opening of the supercooling decompression unit based on the temperature of the refrigerant on the downstream side of the refrigerant flow of the supercooling decompression unit and on the upstream side of the refrigerant flow of the liquid storage unit. The refrigeration cycle device described in. The compressor sucks, compresses, and discharges the refrigerant by being supplied with electric power.
The heating unit heats the air or the human body in the air-conditioned space by being supplied with electric power.
The first change rate is the rate of change in the heat dissipation capacity of the heat dissipation unit with respect to the power consumption of the compressor.
The refrigeration cycle apparatus according to any one of claims 1 to 11, wherein the second change rate is the rate of change in the heating capacity in the heating section with respect to the power consumption of the heating section. The control unit
The refrigeration according to claim 12, wherein the energy consumption per unit time of the compressor is set to be equal to or lower than the upper limit by controlling the compressor so that the power consumption of the compressor is equal to or less than the upper limit power (Pe_limit). Cycle device. The refrigeration cycle apparatus according to claim 12, wherein the heating unit heats the air heated by the heat radiation unit. The refrigerating cycle device according to claim 12 or 13, wherein the heating unit heats a human body in the air-conditioned space by heat conduction or radiation.

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