JP2006145087A - Supercritical refrigeration cycle - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To accurately implement protection control of abnormal high pressure and high-pressure control for maximizing COP only by using one refrigerant pressure sensor in a supercritical refrigeration cycle. <P>SOLUTION: A refrigerant flow rate is calculated on the basis of a control electric current value of a flow rate control-type variable displacement compressor (S110), pressure loss of a radiator is calculated on the basis of a calculated value of the refrigerant flow rate or the like, an outlet-side refrigerant pressure of the radiator is calculated on the basis of the pressure loss of the radiator and the compressor discharge pressure detected by a pressure sensor (S160), an opening of an expansion valve is controlled to allow the calculated value of refrigerant pressure to be agreed with a target high pressure (S170), and the compressor is stopped when the compressor discharge pressure becomes higher than a high-pressure upper limit value (S30). <P>COPYRIGHT: (C)2006,JPO&NCIPI

Description

  The present invention relates to a supercritical refrigeration cycle using a refrigerant having a high pressure equal to or higher than a critical pressure (supercritical state) such as a CO2 refrigerant, and is suitable for use in a refrigeration cycle for vehicle air conditioning.

  In the supercritical refrigeration cycle, there is a high pressure at which the COP (coefficient of performance) of the cycle is maximum with respect to the refrigerant temperature at the outlet of the radiator that cools the high-pressure side refrigerant. Therefore, various methods have been proposed for controlling the high pressure by adjusting the opening of the expansion valve (decompression unit) so that the COP is maximized.

  For example, in Patent Document 1, the refrigerant temperature and the refrigerant pressure at the radiator outlet are detected by a temperature sensor and a pressure sensor, and the expansion valve is set so that the COP becomes maximum based on the refrigerant temperature and the refrigerant pressure on the radiator outlet side. It has been proposed to control the high pressure by adjusting the opening of the (pressure reducing means).

Further, in Patent Document 2, a pressure sensor for detecting the compressor discharge pressure is provided on the compressor discharge side, and a temperature sensor for detecting the refrigerant temperature is provided on the radiator outlet side, and the compressor discharge pressure and the radiator outlet are provided. It has been proposed to control the high pressure by adjusting the opening of the expansion valve (decompression unit) so that the COP becomes maximum based on the refrigerant temperature.
JP 2003-74996 A Japanese Patent No. 3479747

  By the way, in the refrigeration cycle, the high pressure of the cycle is prevented from rising abnormally to protect the cycle equipment. In order to reliably prevent this abnormal high pressure, it is necessary to detect the compressor discharge pressure at which the pressure becomes highest.

  However, in Patent Document 1, only the refrigerant pressure at the outlet side of the radiator is detected, so a pressure sensor for detecting the compressor discharge pressure is added in order to accurately control the abnormally high pressure. It is necessary to increase the number of pressure sensors.

  On the other hand, Patent Document 2 has a pressure sensor that detects the compressor discharge pressure, so that the protection control of abnormally high pressure can be performed accurately, but on the other hand, the compressor discharge side pressure and the radiator outlet side pressure Since the pressure difference, that is, the pressure loss of the radiator varies depending on the cycle operation state, the high pressure control for maximizing the COP becomes inaccurate.

  For example, under the cycle operation conditions where the cooling heat load is very large, such as at the start of cooling in the summer of a vehicle air conditioner, the compression function force is maximized and the cycle circulation refrigerant flow is maximized. The pressure difference is also maximized.

  However, if high pressure control is performed to maximize the COP using the compressor discharge side pressure instead of the radiator outlet side refrigerant pressure, the expansion valve will increase due to an increase in the pressure loss of the radiator at high loads such as during cooling start. The opening degree is controlled more than necessary. As a result, the lowering of the low-pressure pressure is delayed, and it takes time to lower the evaporator blowing air temperature, so that the start-up of the cooling effect is delayed.

  Therefore, in Patent Document 2, it is necessary to add a pressure sensor for detecting the refrigerant pressure at the outlet side of the radiator in order to accurately perform the high pressure control for maximizing the COP. Incurs an increase in costs.

  In view of the above, the present invention provides a supercritical refrigeration cycle that can accurately perform both high-pressure protection control and COP maximization control by using only one refrigerant pressure sensor. With the goal.

In order to achieve the above object, in the invention according to claim 1, in the supercritical refrigeration cycle in which the refrigerant pressure on the high pressure side is equal to or higher than the critical pressure of the refrigerant,
Pressure detection means (12) for detecting the discharge pressure of the compressor (1);
Temperature detecting means (13) for detecting the outlet side refrigerant temperature of the radiator (2) for cooling the refrigerant discharged from the compressor (1);
Pressure calculation means (S160) for calculating the outlet side refrigerant pressure of the radiator (2) from the pressure detection value of the pressure detection means (12) and the cycle operation state;
Opening control means for controlling the opening of the decompression means (3) so that the pressure calculation value of the pressure calculation means (S160) matches the target value determined based on the temperature detection value of the temperature detection means (13). (S170),
When the pressure detection value of the pressure detection means (12) reaches a preset abnormal high pressure set value, the compressor control means (S30) controls the stop of the compressor (1) or the deterioration of the capacity of the compressor (1). It is characterized by comprising.

  According to this, since the discharge pressure of the compressor (1) is detected by the pressure detection means (12) and protection control for abnormally high pressure is performed based on the detected pressure value, the discharge pressure of the compressor itself without using the calculated value. Based on the above, it is possible to accurately perform protection control of abnormally high pressure.

  Moreover, the pressure calculating means (S160) always grasps the fluctuation of the cycle operation state, and calculates the refrigerant pressure on the outlet side of the radiator (2) from this cycle operation state and the pressure detection value of the pressure detection means (12). Thus, even during a transient time when the cooling heat load greatly fluctuates, such as at the start of cooling, the opening degree of the decompression means (3) is controlled based on the calculated value of the pressure calculation means (S160), and thus the high pressure for maximizing the COP. Pressure control can be performed accurately.

  Thereby, since only one pressure detection means (12) for detecting the discharge pressure of the compressor (1) is used, the cost can be reduced by reducing the number of pressure sensors.

According to a second aspect of the present invention, in the supercritical refrigeration cycle according to the first aspect, the compressor has a capacity control means (1b) for controlling a change in capacity, and the capacity control means (1b) provides a refrigerant. A variable capacity compressor (1) of a flow rate control type that variably controls the capacity so that the discharge flow rate becomes a target flow rate,
The target flow rate is determined by the control current value (Ic) of the capacity control means (1b),
As the information value related to the cycle operation state, at least the control current value (Ic) of the capacity control means (1b) is used,
The pressure loss of the radiator (2) is calculated based on the control current value (Ic), and the outlet side of the radiator (2) is calculated based on the pressure detection value of the pressure detection means (12) and the pressure loss of the radiator (2). The refrigerant pressure is calculated.

  By the way, in the refrigeration cycle using the flow rate control type variable capacity compressor (1), the control current value (Ic) of the capacity control means (1b) is highly correlated with the flow rate of the cycle circulation refrigerant. In view of this, in claim 2, the pressure loss of the radiator (2) is calculated based on the control current value (Ic), and the pressure detection value of the pressure detecting means (12) and the radiator (2) Based on the pressure loss, the outlet side refrigerant pressure of the radiator (2) is calculated.

  As a result, the radiator outlet side refrigerant pressure can be accurately calculated using the control current value (Ic) in the flow rate control type variable capacity compressor (1).

  According to a third aspect of the present invention, in the supercritical refrigeration cycle according to the second aspect, the pressure loss of the radiator (2) is determined based on the control current value (Ic), the temperature detection value of the temperature detection means (13), and the pressure detection. Calculation is based on the pressure detection value of the means (12).

  According to this, since the density of the refrigerant on the high pressure side can be calculated based on the temperature detection value of the temperature detection means (13) and the pressure detection value of the pressure detection means (12), the pressure loss of the radiator (2) can be accurately detected. It can be calculated. Therefore, the calculation accuracy of the radiator outlet side refrigerant pressure can be further improved.

According to a fourth aspect of the present invention, in the supercritical refrigeration cycle according to the first aspect, internal heat is exchanged between the outlet side refrigerant of the radiator (2) and the suction side refrigerant of the compressor (1). Comprising an exchanger (20),
As information values related to the cycle operation state, the high-pressure inlet refrigerant temperature (Tgc) of the internal heat exchanger (20), the low-pressure inlet refrigerant temperature (Tac) of the internal heat exchanger (20), and the internal heat exchanger ( 20) of the high-pressure side outlet refrigerant temperature (Tex) or the low-pressure side outlet refrigerant temperature (Tsx),
Calculate the internal heat exchanger temperature ratio (A) based on these temperatures, calculate the pressure loss of the radiator (2) based on the internal heat exchanger temperature ratio (A),
The outlet side refrigerant pressure of the radiator (2) is calculated based on the pressure detection value of the pressure detection means (12) and the pressure loss of the radiator (2).

  By the way, in the refrigeration cycle including the internal heat exchanger (20), the flow rate of the cycle circulation refrigerant can be calculated from the internal heat exchanger temperature ratio (A) as described later. Accordingly, in claim 4, the pressure loss of the radiator (2) is calculated based on the internal heat exchanger temperature ratio (A), and the outlet side refrigerant pressure of the radiator (2) is further calculated.

  Therefore, the pressure loss of the radiator (2) can be determined based on the internal heat exchanger temperature ratio (A) only by detecting the refrigerant temperature at the inlet and outlet of the internal heat exchanger (20) without adding an expensive pressure sensor. And the outlet side refrigerant pressure of the radiator (2) can be calculated.

According to a fifth aspect of the present invention, in the supercritical refrigeration cycle according to the first aspect, the compressor is a fixed capacity type compressor (1) that always operates at a constant capacity, and the fixed capacity type compressor (1) ), The compression function force is controlled by intermittent control of operation,
When operating the fixed capacity compressor (1), at least the compressor rotational speed is used as the information value related to the cycle operation state, and the pressure loss of the radiator (2) is calculated based on the compressor rotational speed,
The outlet side refrigerant pressure of the radiator (2) is calculated based on the pressure detection value of the pressure detection means (12) and the pressure loss of the radiator (2).

  By the way, in the refrigeration cycle in which the operation of the fixed capacity compressor (1) is intermittently controlled, the flow rate of the cycle circulation refrigerant can be calculated based on the compressor rotational speed when the fixed capacity compressor (1) is operated. Therefore, in claim 5, the pressure loss of the radiator (2) can be calculated based on the compressor rotation speed, and the radiator outlet side refrigerant pressure can be calculated.

  In addition, the code | symbol in the bracket | parenthesis of each said means shows the correspondence with the specific means as described in embodiment mentioned later.

(First embodiment)
FIG. 1 is a configuration diagram of a refrigeration cycle for vehicle air-conditioning showing a first embodiment of the present invention, and this refrigeration cycle uses CO 2 whose high pressure is not less than a critical pressure (supercritical state) as a refrigerant. Therefore, this refrigeration cycle constitutes a supercritical refrigeration cycle.

The compressor 1 obtains driving force from a vehicle travel engine (not shown) and sucks and compresses refrigerant. The compressor 1 obtains a driving force through an electromagnetic clutch 1a serving as a clutch means for intermittently driving the driving force.

The compressor 1 of the present embodiment is a variable capacity compressor that can change its capacity by an external control signal, and includes an electromagnetic capacity control valve 1b.

  A radiator 2 is provided on the discharge side of the compressor 1. The radiator 2 cools the refrigerant by exchanging heat between the refrigerant discharged from the compressor 1 in a high-temperature and high-pressure supercritical state and the outside air (outdoor air). Outside air is blown to the radiator 2 by an electric cooling fan 2a.

  On the outlet side of the radiator 2, an electric expansion valve 3 serving as a decompression unit is provided. The electric expansion valve 3 serves as a pressure control valve whose opening degree is electrically controlled so that the high pressure of the cycle becomes the target high pressure.

  An evaporator 4 is provided on the outlet side of the electric expansion valve 3. The evaporator 4 is arranged in a case 5 that forms an air passage of an indoor air conditioning unit of the vehicle air conditioner, and constitutes a cooling means for cooling the air in the case 5.

  An electric blower 6 is disposed on the upstream side of the air flow of the evaporator 4 so that the inside air or outside air introduced through an inside / outside air switching box (not shown) is blown into the case 5. In addition, a heater core (not shown) or the like that forms heating means for heating air is disposed in the case 5 on the downstream side of the air flow of the evaporator 4, and the conditioned air whose temperature is adjusted by the degree of heating of the heater core is provided in the case 5. 5 is blown out into the passenger compartment from an outlet (not shown) at the downstream end of the air flow.

  An accumulator 7 is provided on the outlet side of the evaporator 4. The accumulator 7 is a gas-liquid separation unit that separates the liquid refrigerant and gas refrigerant of the outlet refrigerant of the evaporator 4 and stores excess refrigerant in the cycle, and directs the separated gas refrigerant toward the suction side of the compressor 1. To derive.

  Next, an outline of the electric control unit in the present embodiment will be described. The air-conditioning control device 10 is composed of a microcomputer and its peripheral circuits, etc., and performs predetermined arithmetic processing according to a preset program to control the operation of the air-conditioning equipment.

  Specifically, air-conditioning equipment such as an electromagnetic clutch 1 a of the compressor 1, a capacity control valve 1 b, a cooling fan 2 a of the radiator 2, an electric expansion valve 3, and an electric blower 6 are provided on the output side of the air-conditioning control device 10. Connected and controls the operation of these air conditioners.

  On the input side of the air-conditioning control device 10, the discharge refrigerant temperature sensor 11 of the compressor 1, the discharge refrigerant pressure sensor 12 of the compressor 1, the refrigerant temperature sensor 13 on the outlet side of the radiator 2, and the blown air temperature sensor of the evaporator 4 14 etc. are connected. As is well known, the air conditioning controller 10 also receives detection signals from a sensor group 15 including an engine rotation sensor, an outside air temperature sensor, an inside air temperature sensor, a solar radiation sensor, an engine water temperature sensor, and the like.

  Various air conditioning operation signals are input to the air conditioning control device 10 from an air conditioning operation panel 16 disposed in the vicinity of an instrument panel in the vehicle interior.

  Specifically, a temperature setting switch for setting a set temperature in the passenger compartment, an auto switch for issuing an air conditioning automatic control command, an air conditioner switch for issuing an operation command signal for the compressor 1, an air volume switching switch for the electric blower 6, an indoor air conditioning unit The air-conditioning operation panel 16 is provided with operation members such as a blow-out mode changeover switch of the unit and an inside / outside air introduction mode changeover switch of the inside / outside air switching box.

  Next, the variable capacity compressor 1 will be specifically described. The variable capacity compressor 1 of the present embodiment is known as a swash plate compressor, and by changing the control current value applied to the electromagnetic capacity control valve 1b, the control pressure Pc of the swash plate chamber is changed, Thereby, the change of the inclination angle of the swash plate → the change of the piston stroke → the change of the capacity is performed. Here, the capacity is the geometric volume of the working space where the refrigerant is sucked and compressed.

  Further, in the swash plate type variable displacement compressor 1, the discharge capacity can be continuously changed from the maximum capacity of 100% to the minimum capacity of about 0% by adjusting the control pressure Pc.

  Then, the target discharge refrigerant flow rate is set according to the control current value of the electromagnetic capacity control valve 1b, and the control pressure Pc of the swash plate chamber is changed so that the actual discharge refrigerant flow rate becomes the target discharge refrigerant flow rate. This is a variable capacity compressor of the so-called flow rate control type that changes the pressure. Such flow rate control type variable displacement compressors are known from Japanese Patent Application Laid-Open Nos. 2001-107854, 2001-173556, and the like.

  The outline of the flow control type variable capacity compressor 1 will be described with reference to FIG. 2. FIG. 2 shows the discharge side flow path portion of the swash plate type variable capacity compressor 1 and the control pressure Pc of the swash plate chamber 103. It is the schematic which shows the electromagnetic capacity | capacitance control valve 1b part to control, and the discharge chamber 100 of the compressor 1 is a part which collects the refrigerant | coolant discharged from the some piston working chamber (cylinder) which is not shown in figure.

  A throttle part 102 is provided in the outlet-side flow path 101 of the discharge chamber 100, and when a refrigerant discharged from the compressor 1 passes through the throttle part 102, a predetermined differential pressure ΔP is generated between the front and rear of the throttle part 102. I have to do it. Here, the differential pressure ΔP = PdH−PdL. PdH is the refrigerant pressure upstream of the throttle 102 and PdL is the refrigerant pressure downstream of the throttle 102.

  Since the differential pressure ΔP is proportional to the discharge refrigerant flow rate of the compressor 1, the discharge refrigerant flow rate of the compressor 1 can be controlled by controlling the differential pressure ΔP.

  On the other hand, the displacement control valve 1b includes a differential pressure responsive mechanism 111 that generates a force F1 corresponding to the differential pressure ΔP, and an electromagnetic mechanism 112 that generates an electromagnetic force F2 that opposes the force F1 of the differential pressure responsive mechanism 111. Basically, the position of the valve body 113 (the position in the left-right direction in FIG. 2) is changed by balancing the force F1 and the electromagnetic force F2 corresponding to the differential pressure ΔP.

  However, in the illustrated example of FIG. 2, the position of the valve body 113 is changed depending on the refrigerant pressure PdH (compressor discharge pressure) in the upstream portion of the throttle portion 102 by the configuration described later.

  Note that since the differential pressure ΔP is actually a minute value, the discharged refrigerant pressure sensor 12 of FIG. 1 may be provided either upstream or downstream of the throttle unit 102, but the throttle unit 102 is the compressor 1. Therefore, the discharge refrigerant pressure sensor 12 is usually provided in the downstream portion of the throttle portion 102 for the convenience of mounting the sensor.

  The differential pressure responsive mechanism 111 accommodates a bellows 111b that can be elastically expanded and contracted in the moving direction of the valve body 113 (the left-right direction in FIG. 2) in the case 111a, and a refrigerant upstream of the throttle portion 102 inside the bellows 111b. Introduce pressure PdH. On the other hand, the refrigerant pressure PdL in the downstream portion of the throttle portion 102 is introduced into the case 111a.

  The right end portion of the bellows 111b in FIG. 2 constitutes a fixed end fixed to the case 111a, and the left end portion of the bellows 111b in FIG. 2 constitutes a movable end 111c that is displaced in the left-right direction in FIG. . A spring 111d is provided inside the bellows 111b to press the bellows 111b in the extending direction (left side in FIG. 2).

  A push rod 111e is integrally connected to the movable end 111c of the bellows 111b. The push rod 111e is slidable with respect to the fitting hole 111f of the case 111a and is airtightly fitted by a seal mechanism (not shown) and protrudes outside the case 111a.

  On the other hand, the electromagnetic mechanism 112 has an electromagnetic coil 112a, and a plunger 112b is disposed on the inner peripheral portion of the electromagnetic coil 112a so as to be displaceable in the axial direction (left-right direction in FIG. 2). A movable iron core 112c is integrally formed at the end of the plunger 112b, and a fixed iron core 112d is disposed opposite to the movable iron core 112c. An electromagnetic force (attraction force) F2 corresponding to a control current Ic supplied to the electromagnetic coil 112a is generated between the movable iron core 112c and the fixed iron core 112d.

  A spring 112e that generates a spring force in the opposite direction to the electromagnetic force F2 is disposed between the movable iron core 112c and the fixed iron core 112d. Of the plunger 112b, the valve body 113 described above is integrally formed at an end portion (right end portion in FIG. 2) opposite to the movable iron core 112c. Further, the valve body 113 is integrally connected to the push rod 111e via a connecting shaft portion 113a having a sufficiently smaller diameter than the valve body 113. Accordingly, the plunger 112b, the valve body 113, and the push rod 111e constitute an integral body and are integrally displaced in the axial direction of the plunger 112b (the left-right direction in FIG. 2).

  The valve body 113 is disposed in the control pressure passage 114 and increases or decreases the passage area of the control pressure passage 114. Since one end portion of the control pressure passage 114 communicates with the discharge chamber 100 of the compressor 1 via the communication passage 115, the refrigerant pressure PdH upstream of the throttle portion 102 is introduced into one end portion of the control pressure passage 114. . On the other hand, the other end of the control pressure passage 114 communicates with the swash plate chamber 103 of the compressor 1 through the communication passage 116.

  The swash plate chamber 103 communicates with the suction chamber 106 of the compressor 1 through a communication passage 105 having a throttle 104. When the valve body 113 is displaced in the right direction in FIG. 2, the passage area of the control pressure passage 114 is decreased, and when the valve body 113 is displaced in the left direction in FIG. 2, the passage area of the control pressure passage 114 is increased. Therefore, the electromagnetic force F2 is a force in the valve closing direction that displaces the valve body 113 in the right direction in FIG. 2, and conversely, the force F1 corresponding to the differential pressure ΔP displaces the valve body 113 in the left direction in FIG. This is the force in the valve opening direction.

  When the passage area of the control pressure passage 114 decreases, the amount of refrigerant discharged from the discharge chamber 100 of the compressor 1 through the communication passage 115 → the control pressure passage 114 → the communication passage 116 into the swash plate chamber 103 decreases, and the swash plate chamber 103. When the control pressure Pc decreases, that is, the passage area of the control pressure passage 114 increases, the amount of refrigerant discharged into the swash plate chamber 103 increases and the control pressure Pc of the swash plate chamber 103 increases.

  In the swash plate type variable displacement compressor 1, as is well known, the control pressure Pc decreases, the swash plate tilt angle increases, the piston stroke increases, the discharge capacity increases, and conversely the control pressure Pc increases. The discharge capacity changing mechanism is configured so that the inclination angle of the swash plate decreases, the piston stroke decreases, and the discharge capacity decreases.

  By the way, since the electromagnetic force F2 is a force that opposes the force F1 corresponding to the differential pressure ΔP, the target differential pressure is determined by increasing or decreasing the electromagnetic force F2, and the actual differential pressure ΔP is the electromagnetic force. The control pressure Pc of the swash plate chamber 103 is controlled so that the target differential pressure determined by F2 is reached, and the discharge capacity changes. Further, since the differential pressure ΔP and the discharge refrigerant flow rate are in a proportional relationship as described above, determining the target differential pressure determines the target discharge refrigerant flow rate.

  Since the electromagnetic force F2 is determined according to the control current Ic supplied to the electromagnetic coil 112a, the target differential pressure and the target discharge refrigerant flow rate are increased as the control current Ic increases.

  FIG. 3 is a characteristic diagram showing the relationship between the control current Ic and the circulating refrigerant flow rate in the cycle when the swash plate type variable displacement compressor 1 having such a flow rate control characteristic is used. PdH1 to PdH4 have a relationship of PdH1 <PdH2 <PdH3 <PdH4, and the refrigerant flow rate changes depending on the compressor discharge pressure PdH in addition to the control current Ic.

  Specifically, the compressor discharge pressure PdH is introduced into the control pressure passage 114 by the communication passage 115, and the pressure receiving area S1 of the valve body 113 in the control pressure passage 114 is changed to the push rod 111e of the differential pressure response mechanism 111. This is because the compressor discharge pressure PdH has an influence on the position control of the valve body 113 by making it larger than the pressure receiving area S2.

  The control current value Ic of the electromagnetic capacity control valve 1b is such that the actual blown air temperature (detected temperature of the temperature sensor 14) Te of the evaporator 4 becomes the evaporator target temperature TEO for air conditioning control. 10 is calculated. The evaporator target temperature TEO is calculated based on the target temperature TAO of the air blown into the passenger compartment, the outside air temperature Tam, and the like as is well known.

  Specifically, the control current Ic is usually changed by duty control because of the configuration of the current control circuit, but the value of the control current Ic is directly and continuously (analog-like) regardless of duty control. ) May be changed.

  Next, the operation of this embodiment in the above configuration will be described. First, the basic operation of the refrigeration cycle will be described. When the auto switch or the air conditioner switch of the air conditioning operation panel 16 is turned on, the electromagnetic clutch 1a is energized by the air conditioning control device 10 to be connected. Thereby, the driving force of the vehicle engine is transmitted to the compressor 1 via the electromagnetic clutch 1a, and the compressor 1 is driven.

The high-temperature and high-pressure refrigerant compressed by the compressor 1 flows into the radiator 2 in a supercritical state where the pressure is higher than the critical pressure. Here, the high-temperature and high-pressure supercritical refrigerant exchanges heat with the outside air blown by the cooling fan 2a to dissipate heat into the outside air, thereby reducing enthalpy.

  And the refrigerant | coolant of the exit of the heat radiator 2 is pressure-reduced by the throttle path of the expansion valve 3, and becomes a low-temperature low-pressure gas-liquid two phase state. Here, the opening degree of the expansion valve 3 is controlled so as to maximize the cycle COP, as will be described later.

  The low-temperature and low-pressure gas-liquid two-phase refrigerant after passing through the expansion valve 3 flows into the evaporator 4, where it absorbs heat from the blown air of the electric blower 6 and evaporates. Thereby, the air blown from the electric blower 6 can be cooled by the evaporator 4, and the cool air can be blown out into the passenger compartment.

The low-pressure refrigerant that has passed through the evaporator 4 then flows into the accumulator 7, where the low-pressure refrigerant liquid refrigerant and the gas refrigerant are separated by a density difference, and the gas refrigerant is discharged from the outlet of the accumulator 7 to the suction side of the compressor 1. Is drawn into the compressor 1, sucked into the compressor 1, and compressed again.

  In the variable capacity compressor 1, the target discharge refrigerant flow rate is changed by the control current value of the electromagnetic capacity control valve 1b so that the actual blown air temperature Te of the evaporator 4 becomes the evaporator target temperature TEO. To control the capacity.

  Specifically, when the evaporator outlet temperature Te is higher than the evaporator target temperature TEO, the control current value output to the capacity control valve 1b of the compressor 1 is increased to increase the target discharge refrigerant flow rate. The capacity of the compressor 1 can be increased. As a result, the cooling capacity of the evaporator 4 is increased by increasing the circulating refrigerant flow rate to the evaporator 5.

  On the contrary, when the evaporator outlet temperature Te is lower than the evaporator target temperature TEO, the control current value output to the capacity control valve 1b of the compressor 1 is decreased to decrease the target discharge refrigerant flow rate, thereby compressing. The capacity of the machine 1 can be reduced. As a result, the circulating refrigerant flow rate to the evaporator 4 is reduced and the cooling capacity of the evaporator 4 is reduced.

  The minimum temperature of the evaporator target temperature TEO is determined to be slightly higher than 0 ° C. (about 1 ° C.) in order to prevent the evaporator 4 from being frosted.

  Next, the abnormal high pressure protection control and the cycle COP control based on the opening degree control of the expansion valve 3 in this embodiment will be described in detail with reference to FIG. FIG. 4 is a control routine executed by the microcomputer of the air-conditioning control apparatus 10, and this control routine starts, for example, when an auto switch of the air-conditioning operation panel 16 is turned on.

  First, in step S10, the discharge pressure of the compressor 1 (detected pressure of the pressure sensor 12) is read, and in step S20, it is determined whether the compressor discharge pressure is equal to or higher than the high pressure upper limit value. Here, the high pressure upper limit value is a preset value of an abnormal high pressure, and is, for example, 14.6 MPa.

  And when an actual compressor discharge pressure is more than a high pressure upper limit, it progresses to Step S30 and stops compressor 1. Specifically, the control current Ic of the capacity control valve 1b of the compressor 1 is set to 0, the energization of the electromagnetic clutch 1a of the compressor 1 is interrupted, and the compressor 1 is stopped. As a result, the compressor discharge pressure starts to decrease.

  In step S40, the compressor discharge pressure is read again. In step S50, it is determined whether the compressor discharge pressure has decreased below the return pressure. Here, the return pressure is a pressure lower than the high pressure upper limit by a certain value, for example, 10 MPa.

  While the compressor discharge pressure is equal to or higher than the return pressure, the stopped state of the compressor 1 is maintained. Thereby, since it can avoid that a compressor discharge pressure further exceeds a high pressure upper limit, the compressor 1 can be reliably protected from the overload state by abnormally high pressure.

  In this case, because the compressor discharge pressure is not based on the estimated value but based on the detected pressure of the pressure sensor 12 that directly detects the compressor discharge pressure, the abnormal high pressure protection control is performed. Can be executed reliably with value.

  Then, when the compressor discharge pressure is lower than the return pressure, the process proceeds to step S60 and the compressor 1 is returned to the operating state. Specifically, the electromagnetic clutch 1a is energized to bring the electromagnetic clutch 1a into a connected state, and the compressor 1 is rotationally driven by the vehicle engine. The capacity control valve 1b is supplied with a control current Ic calculated based on the deviation between the actual evaporator outlet temperature Te and the evaporator target temperature TEO, whereby the capacity of the compressor 1 is controlled by the control current Ic. It is controlled to a predetermined capacity according to

  On the other hand, when the compressor discharge pressure is less than the high pressure upper limit value (when the high pressure is normal), the process proceeds from step S20 to step S70, and the cycle temperature is determined based on the outlet refrigerant temperature of the radiator 2 (detected temperature of the temperature sensor 13) Tgc. A target high pressure that maximizes the COP is calculated. Specifically, a map that defines the relationship between the outlet refrigerant temperature Tgc of the radiator 2 and the high pressure at which the COP corresponding to the temperature is maximum is set in advance, and the outlet refrigerant of the radiator 2 is set in this map. By applying the detected value of the temperature Tgc, the target high pressure that maximizes the COP can be calculated.

  In the next step S80, it is determined whether the calculated target high pressure is equal to or higher than a preset control upper limit pressure. Here, the control upper limit pressure is slightly lower than the high pressure upper limit pressure in step S20, and is, for example, 13 MPa. When the calculated target high pressure is equal to or higher than the control upper limit pressure, the process proceeds to step S90, and the control upper limit pressure is set as the target high pressure. On the other hand, when the calculated target high pressure is less than the control upper limit pressure, the calculated target high pressure is determined as it is as the target high pressure in step S100.

  Next, in step S110, based on the control current value Ic of the capacity control valve 1b of the compressor 1 and the discharge pressure of the compressor 1 (detected pressure of the pressure sensor 12), the refrigerant flow rate when the capacity is variable is calculated. Here, the control current value Ic of the capacity control valve 1b sets the target pressure loss corresponding to the pressure loss generated in the throttle portion on the discharge side of the compressor 1 as described above, and finally sets the target discharge refrigerant flow rate. Is. Therefore, the control current value Ic of the capacity control valve 1b is basically an information value representative of the cycle refrigerant flow rate.

  In the capacity control valve 1b of the present embodiment, as described above, the compressor capacity is changed under the influence of the compressor discharge pressure in addition to the control current value I, and as a result, the refrigerant flow rate is changed. Therefore, the refrigerant flow rate when the capacity is variable is calculated based on the control current value Ic and the discharge pressure of the compressor 1.

  More specifically, a control map corresponding to FIG. 3 is set in advance, and by applying the control current value Ic and the detected value of the compressor discharge pressure to this control map, the refrigerant flow rate at the time of variable capacity is set. Easy to calculate.

  Next, in step S120, the refrigerant flow rate when the capacity is 100% (at the maximum capacity) is calculated. The compressor discharge refrigerant flow rate when the capacity is 100% can be calculated by the following equation (1).

Refrigerant flow rate = compressor suction refrigerant density x compressor capacity x rotation speed x compressor volumetric efficiency (1)
Here, the compressor capacity is a fixed value, and the compressor volumetric efficiency can be determined based on the number of revolutions because the influence of the number of revolutions is large.

  The compressor suction refrigerant density is determined by the temperature and pressure of the suction refrigerant. In this embodiment, in order to avoid the addition of a new sensor to the compressor suction portion, the evaporator blown air temperature Te (temperature sensor 14) is used. The refrigerant suction refrigerant density is estimated on the basis of the detected temperature).

  That is, in the accumulator cycle in which the accumulator 7 is arranged on the outlet side of the evaporator 4 as shown in FIG. 1, the gas-liquid interface of the saturated refrigerant is formed inside the accumulator 7, so that the outlet refrigerant of the evaporator 4 is always saturated. Maintained in a state. For this reason, the temperature and pressure of the saturated refrigerant in the evaporator 4 can be accurately estimated from the evaporator blowing air temperature Te, and consequently the density of the intake refrigerant can be accurately estimated.

  As a result, the refrigerant flow rate at a capacity of 100% can be calculated based on the above equation (1) only by using the evaporator blown air temperature Te and the compressor rotational speed as variables. The compressor speed can be obtained based on the detection value of the engine rotation sensor provided in the sensor group 15 of FIG. 1, so that a dedicated sensor is unnecessary.

  Note that the refrigerant flow rate at a capacity of 100% can be calculated (estimated) from the compressor rotation speed and the compressor discharge pressure, although the accuracy is lowered. That is, since the compressor discharge pressure is determined by the radiator outlet refrigerant temperature, there is a correlation with the outside air temperature, and since the outside temperature and the cooling heat load of the evaporator 4 are correlated, the evaporator discharge pressure is based on the compressor discharge pressure. This is because the blown air temperature Te can be estimated to some extent.

  In the next step S130, the refrigerant flow rate at the time of variable capacity and the refrigerant flow rate at the capacity of 100% are compared. When the refrigerant flow rate at the time of variable capacity is larger than the refrigerant flow rate at the capacity of 100%, the process proceeds to step S140, and the refrigerant flow rate at the capacity of 100% is set as the final refrigerant flow rate.

  In contrast, when the refrigerant flow rate at the time of variable capacity is smaller than the refrigerant flow rate at the capacity of 100%, the process proceeds to step S150, and the refrigerant flow rate at the time of variable capacity is set as the final refrigerant flow rate. By performing such control processing of S110 to S150, it is possible to reliably avoid the unreasonable situation that the final refrigerant flow rate becomes larger than the refrigerant flow rate when the capacity is 100%.

  In the next step S160, the pressure loss (pressure loss) of the radiator 2 is calculated based on the final refrigerant flow rate, and the outlet pressure of the radiator 2 is calculated based on the pressure loss of the radiator 2.

  Here, since the pressure loss of the radiator 2 has a correlation with the refrigerant flow rate and the density of the high-pressure side refrigerant, the density of the high-pressure side refrigerant is expressed as the refrigerant outlet refrigerant temperature (detected value Tgc of the temperature sensor 13) and the compressor discharge pressure. The pressure loss of the radiator 2 can be calculated based on the density of the high-pressure side refrigerant and the final refrigerant flow rate.

  Next, the outlet pressure of the radiator 2 can be calculated by subtracting the calculated value of the radiator pressure loss from the compressor discharge pressure.

  As described above, the calculation of the pressure loss of the radiator 2 and the calculation of the outlet pressure of the radiator 2 are not performed in two stages, and the final refrigerant flow rate, the radiator outlet refrigerant temperature, and the compressor discharge are not processed. By preparing in advance a control map for determining the radiator outlet pressure from the three parties, the final refrigerant flow rate, the radiator outlet refrigerant temperature, and the compressor discharge pressure to this control map, heat dissipation The vessel outlet pressure may be calculated all at once.

  In addition, when the cycle operating conditions are limited to a specific range, the change in the density of the high-pressure side refrigerant is small, so it is possible to calculate the radiator outlet pressure by calculating the radiator pressure loss only from the final refrigerant flow rate. Is possible.

  In the next step S170, the opening degree control of the expansion valve 3 for maximizing the COP is performed. That is, the opening degree of the expansion valve 3 is controlled so that the calculated outlet pressure value of the radiator 2 matches the target high pressure in step S90 or step S100. That is, when the calculated value of the outlet pressure of the radiator 2 is higher than the target high pressure, the opening degree of the expansion valve 3 is increased. Conversely, when the calculated value of the outlet pressure of the radiator 2 is lower than the target high pressure, the expansion valve 3 is opened. Decrease the degree. By controlling the opening of the expansion valve 3 as described above, high pressure control is performed so that the calculated outlet pressure of the radiator 2 matches the target high pressure.

  By the way, in this embodiment, only one pressure sensor 12 for detecting the compressor discharge pressure is provided as the pressure sensor, but the pressure loss of the radiator 2 is calculated based on at least the cycle refrigerant flow rate as described above. Since the outlet pressure of the radiator 2 is calculated based on the pressure loss of the radiator 2, the outlet pressure of the radiator 2 can be calculated with high accuracy.

  That is, the outlet of the radiator 2 can be calculated by calculating the pressure loss of the radiator 2 even under the operating conditions in which the cycle refrigerant flow rate is maximized and the pressure loss of the radiator 2 is maximized as in the cooling start in summer. The pressure can be calculated accurately.

  For this reason, even if only one pressure sensor 12 for detecting the compressor discharge pressure is provided, the COP maximization control based on the opening degree control of the expansion valve 3 can be executed accurately and accurately. Therefore, it is possible to skillfully achieve both cost reduction by reducing the number of pressure sensors 12 and accuracy of COP maximization control.

  The detection signal of the discharge refrigerant temperature sensor 11 shown in FIG. 1 is used for controlling the discharge refrigerant temperature of the compressor 1. That is, when the discharge refrigerant temperature of the compressor 1 (detected value of the sensor 11) is higher than the discharge temperature target value, the capacity of the compressor 1 is forcibly set to a predetermined amount regardless of the detected value of the blown air temperature sensor 14. The discharge pressure is decreased to reduce the discharge refrigerant temperature, thereby controlling the discharge refrigerant temperature to the discharge temperature target value (limit temperature from heat resistance).

  Each step in FIG. 4 constitutes a function realization means, and the correspondence between each step in FIG. 4 and the means of the present invention will be described. Step S30 is a compressor control for stopping the compressor 1 at an abnormally high pressure. Configure the means. Step S160 constitutes pressure calculation means for calculating the outlet side refrigerant pressure of the radiator 2.

  Step S170 constitutes an opening degree control means for controlling the opening degree of the expansion valve (decompression means) 3 so that the pressure calculated value in step S160 becomes the target high pressure.

(Second Embodiment)
The second embodiment relates to a supercritical refrigeration cycle including an internal heat exchanger 20 as shown in FIG. The internal heat exchanger 20 has a high-pressure channel 20 a provided on the outlet side of the radiator 2 and a low-pressure channel 20 b provided on the outlet side of the accumulator 7. The low pressure side flow path 20 b is connected to the suction side of the compressor 1.

  The internal heat exchanger 20 exchanges heat between the low-temperature refrigerant (compressor suction refrigerant) flowing out from the accumulator 7 and the high-temperature refrigerant at the outlet side of the radiator 2 to reduce the enthalpy of the refrigerant flowing into the evaporator 4. 4 increases the refrigerant enthalpy difference (refrigeration capacity) between the refrigerant inlet and outlet, and prevents the liquid refrigerant from being sucked into the compressor 1.

  When the internal heat exchanger 20 is installed in this way, the refrigerant enthalpy difference (refrigeration capacity) between the refrigerant inlet and outlet of the evaporator 4 can be increased, and the cycle COP can be improved.

  In the second embodiment, as the internal heat exchanger 20 is installed, a temperature sensor 21 that detects the outlet refrigerant temperature Tex of the high-pressure side passage 3a and a temperature sensor 22 that detects the inlet refrigerant temperature Tac of the low-pressure side passage 20b. Has been added.

  The amount of heat exchange Qih in the internal heat exchanger 20 is roughly proportional to the temperature difference between the high-pressure side refrigerant and the low-pressure side refrigerant. The refrigerant enthalpy difference (Δi) between the inlet and outlet of the high-pressure side passage 3a of the internal heat exchanger 20 and the refrigerant enthalpy difference (Δi) between the inlet and outlet of the low-pressure side passage 20b are the same value. The refrigerant enthalpy difference (Δi), the refrigerant flow rate G, and the heat exchange amount Qih are in the relationship of the following equation (2).

Qih = G × Δi (2)
Since the refrigerant enthalpy difference (Δi) decreases as the refrigerant flow rate G increases from the equation (2), the refrigerant temperature difference between the inlet and outlet of the high-pressure side channel 3a and the low-pressure side channel 20b decreases.

  In other words, from equation (2), the refrigerant flow rate G can be obtained by G = Qih / Δi. Qih can be obtained from the temperature difference (Tgc-Tac) between the high-pressure side refrigerant and the low-pressure side refrigerant as described above, and Δi is the refrigerant temperature difference (Tgc-Tex) between the inlet and outlet of the high-pressure side flow path 3a. ).

  Therefore, if the temperature ratio A = (Tgc−Tac) / (Tgc−Tex) is calculated, the refrigerant flow rate G can be calculated based on the temperature ratio A.

  In addition, calculation of the refrigerant | coolant flow volume G in 2nd Embodiment replaces the refrigerant | coolant flow volume calculation of step S110-step S150 of 1st Embodiment, Comprising: The other control of 1st Embodiment, ie, step S10-step. The control in S100 and the control in steps S160 to S170 are performed in the same manner in the second embodiment.

  In the second embodiment, the example in which the temperature ratio A = (Tgc−Tac) / (Tgc−Tex) is calculated and the refrigerant flow rate G is calculated based on the temperature ratio A has been described. Instead of the temperature sensor 21 for detecting the outlet refrigerant temperature Tex, a temperature sensor 23 for detecting the outlet refrigerant temperature Tsx of the low-pressure channel 20b of the internal heat exchanger 20 is provided, and the temperature ratio A = (Tgc−Tac) / (Tsx−Tac) may be calculated, and the refrigerant flow rate G may be calculated based on the temperature ratio A.

  Further, since the inlet refrigerant temperature Tac of the low-pressure side passage 20b of the internal heat exchanger 20 has a strong correlation with the evaporator blown air temperature Te, the temperature sensor 22 is eliminated and the inlet refrigerant temperature Tac of the low-pressure side passage 20b. Instead of the above, the temperature ratio A may be obtained using the evaporator blown air temperature Te detected by the temperature sensor 14.

(Third embodiment)
In the above-described first and second embodiments, the refrigeration cycle in which the capacity control of the compressor 1 is controlled by the capacity control of the compressor 1 using the variable capacity compressor 1 has been described. In the third embodiment, The present invention relates to a refrigeration cycle in which a fixed capacity compressor whose capacity is always maintained constant is used as the compressor 1 and the operation of the fixed capacity compressor 1 is intermittently controlled by an electromagnetic clutch 1a to control the capacity of the compressor 1. .

  FIG. 6 shows a supercritical refrigeration cycle of the third embodiment using this fixed capacity compressor 1, and the fixed capacity compressor 1 includes only the electromagnetic clutch 1a and does not include the capacity control valve 1b. In the refrigeration cycle configuration having the fixed capacity compressor 1, the evaporator blown air temperature Te and the target evaporator temperature TEO are compared as is well known, and when the evaporator blown air temperature Te exceeds the target evaporator temperature TEO. The compressor 1 is operated by energizing the electromagnetic clutch 1a.

  When the evaporator blown air temperature Te falls below the off-side temperature TEO ′ lower than the target evaporator temperature TEO by the operation of the compressor 1, the electromagnetic clutch 1a is cut off and the compressor 1 is stopped. Let By such intermittent operation control of the compressor 1, the evaporator blown air temperature Te is controlled to the target evaporator temperature TEO.

  By the way, in 3rd Embodiment, the refrigerant | coolant flow rate at the time of the action | operation of the fixed capacity type compressor 1 is computed by the same method as the refrigerant | coolant flow rate calculation at the time of 100% of capacity | capacitance in step S120 of 1st Embodiment. That is, the refrigerant flow rate during the operation of the compressor is calculated based on the evaporator blown air temperature Te and the compressor rotational speed.

  Then, the pressure loss (pressure loss) of the radiator 2 is calculated based on the calculated refrigerant flow rate as in step S160 of the first embodiment, and the outlet pressure of the radiator 2 is calculated based on the pressure loss of the radiator 2. To do.

  Next, similarly to step S170 of the first embodiment, the opening degree control of the expansion valve 3 is performed so that the calculated value of the outlet pressure of the radiator 2 matches the target high pressure of steps S90 and S100.

  Note that when the fixed capacity compressor 1 is stopped by intermittent control, the discharge pressure of the compressor 1 rapidly decreases, so that the opening degree of the expansion valve 3 is controlled in the valve closing direction.

  Also in the third embodiment, the control in steps S10 to S100 and the control in steps S160 to S170 of the first embodiment are performed in the same manner.

(Other embodiments)
In the first embodiment, the variable displacement compressor 1 is provided with the electromagnetic clutch 1a, and the example in which the electromagnetic clutch 1a is disengaged and the variable displacement compressor 1 is stopped at the time of abnormally high pressure protection control has been described. When using a variable capacity compressor 1 whose capacity can be reduced from 100% capacity (maximum capacity) to a minimum capacity of near 0%, the electromagnetic clutch 1a is eliminated and the engine rotation is always compressed. A so-called clutchless type that is transmitted to the rotating shaft of the machine 1 may be employed.

  In such a refrigeration cycle using the clutchless type variable displacement compressor 1, the abnormally high pressure protection control may be performed by forcibly reducing the capacity to a minimum capacity of around 0% at the abnormally high pressure. .

  In the first embodiment, as shown in FIG. 3, the refrigerant flow rate is determined depending on the compressor discharge pressure (high-pressure side pressure) in addition to the control current Ic of the capacity control valve 1b. However, the refrigerant flow rate may be determined only by the control current Ic of the capacity control valve 1b without depending on the compressor discharge pressure (high pressure side pressure).

  In the first to third embodiments described above, the case where the refrigeration cycle according to the present invention is applied to a cycle dedicated to cooling operation has been described. However, the present invention is not limited to this, and heating operation or dehumidification operation is performed. Of course, the heat pump cycle may be applicable.

  In the first and second embodiments described above, the refrigeration cycle has been described in which the variable capacity compressor 1 is used to change the discharge refrigerant flow rate of the compressor 1 by the capacity control of the compressor 1. The present invention may be applied to a refrigeration cycle in which an electric compressor capable of continuously controlling the rotational speed as 1 is used and the discharge refrigerant flow rate is changed by controlling the rotational speed of the electric compressor. In the refrigeration cycle using this electric compressor, the refrigerant flow rate can be calculated from the compressor rotation speed.

  In addition to CO2, a refrigerant such as ethylene, ethane, or nitrogen oxide may be used as the supercritical cycle refrigerant.

It is a block diagram which shows the supercritical refrigerating cycle for vehicles by 1st Embodiment of this invention. It is typical sectional drawing which shows the specific example of the capacity | capacitance control valve of the variable displacement compressor of FIG. It is a characteristic view of the refrigerant | coolant flow control of the capacity | capacitance control valve of FIG. It is a flowchart which shows the operation control by 1st Embodiment of this invention. It is a block diagram which shows the supercritical refrigerating cycle for vehicles by 2nd Embodiment of this invention. It is a block diagram which shows the supercritical refrigerating cycle for vehicles by 3rd Embodiment of this invention.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 ... Variable capacity type compressor, 2 ... Radiator, 3 ... Expansion valve (pressure reduction means), 4 ... Evaporator,
DESCRIPTION OF SYMBOLS 10 ... Control apparatus, 12 ... Refrigerant pressure sensor (pressure detection means),
13 ... Refrigerant temperature sensor (temperature detection means), 20 ... Internal heat exchanger.

Claims (5)

A compressor (1) for sucking and compressing refrigerant;
A radiator (2) for cooling the refrigerant discharged from the compressor (1);
Decompression means (3) for decompressing the outlet side refrigerant of the radiator (2);
An evaporator (4) for evaporating the low-pressure refrigerant decompressed by the decompression means (3),
The refrigerant that has passed through the evaporator (4) is sucked into the compressor (1),
Furthermore, in the supercritical refrigeration cycle where the refrigerant pressure on the high pressure side is equal to or higher than the critical pressure of the refrigerant,
Pressure detection means (12) for detecting the discharge pressure of the compressor (1);
Temperature detecting means (13) for detecting the outlet side refrigerant temperature of the radiator (2);
Pressure calculating means (S160) for calculating the outlet side refrigerant pressure of the radiator (2) from the pressure detection value of the pressure detecting means (12) and the cycle operation state;
An opening for controlling the opening of the pressure reducing means (3) so that the pressure calculated value of the pressure calculating means (S160) matches a target value determined based on the temperature detected value of the temperature detecting means (13). Degree control means (S170),
When the pressure detection value of the pressure detection means (12) reaches a preset abnormal high pressure set value, the compressor control means performs control of stopping the compressor (1) or reducing the capacity of the compressor (1). (S30). A supercritical refrigeration cycle comprising: The compressor has a capacity control means (1b) for controlling a change in capacity, and the capacity control means (1b) variably controls the capacity so that the refrigerant discharge flow rate becomes a target flow rate. Type compressor (1),
The target flow rate is determined by the control current value (Ic) of the capacity control means (1b),
As the information value related to the cycle operation state, at least the control current value (Ic) of the capacity control means (1b) is used.
The pressure loss of the radiator (2) is calculated based on the control current value (Ic), and the radiator (2) is calculated based on the pressure detection value of the pressure detection means (12) and the pressure loss of the radiator (2). The supercritical refrigeration cycle according to claim 1, wherein the outlet side refrigerant pressure of 2) is calculated. The pressure loss of the radiator (2) is calculated based on the control current value (Ic), the temperature detection value of the temperature detection means (13), and the pressure detection value of the pressure detection means (12). The supercritical refrigeration cycle according to claim 2. An internal heat exchanger (20) for exchanging heat between the outlet side refrigerant of the radiator (2) and the suction side refrigerant of the compressor (1);
As information values related to the cycle operation state, the high-pressure inlet refrigerant temperature (Tgc) of the internal heat exchanger (20), the low-pressure inlet refrigerant temperature (Tac) of the internal heat exchanger (20), and the internal Using the high-pressure side outlet refrigerant temperature (Tex) or the low-pressure side outlet refrigerant temperature (Tsx) of the heat exchanger (20),
Calculate the internal heat exchanger temperature ratio (A) based on these temperatures, calculate the pressure loss of the radiator (2) based on the internal heat exchanger temperature ratio (A),
The superfluid pressure sensor according to claim 1, wherein an outlet side refrigerant pressure of the radiator (2) is calculated based on a pressure detection value of the pressure detection means (12) and a pressure loss of the radiator (2). Critical refrigeration cycle. The compressor is a fixed capacity compressor (1) that always operates at a constant capacity, and the compression function force is controlled by intermittent control of the operation of the fixed capacity compressor (1).
At the time of operation of the fixed capacity compressor (1), at least the compressor rotational speed is used as the information value related to the cycle operation state, and the pressure loss of the radiator (2) is calculated based on the compressor rotational speed. ,
The superfluid pressure sensor according to claim 1, wherein an outlet side refrigerant pressure of the radiator (2) is calculated based on a pressure detection value of the pressure detection means (12) and a pressure loss of the radiator (2). Critical refrigeration cycle.

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