JP4882978B2 - Refrigeration cycle equipment - Google Patents

Description

  The present invention relates to a refrigeration cycle apparatus configured to be able to determine a refrigerant shortage state in which a refrigerant circulating in a cycle is insufficient.

  In a vapor compression refrigeration cycle apparatus, in general, when the refrigerant circulating in the cycle is insufficient, not only the refrigeration capacity exhibited by the evaporator is significantly reduced, but also the refrigerant circulates in the cycle together with the refrigerant. Refrigerating machine oil circulation decreases, causing poor lubrication of the compressor. Therefore, conventionally, a refrigeration cycle apparatus configured to be able to determine a refrigerant shortage state in which the refrigerant circulating in the cycle is insufficient is known.

  For example, in the refrigeration cycle apparatus of Patent Document 1, a differential pressure between the discharge pressure and the suction pressure of the compressor is detected using a capacity control valve that changes the discharge capacity of the variable capacity compressor, and further, from this differential pressure, Based on the estimated refrigerant evaporation temperature of the evaporator and the actually detected evaporator outlet air temperature, the state of the refrigerant is determined with reference to a control map stored in advance in the control device.

When it is determined that the state of the refrigerant circulating in the cycle is a refrigerant shortage state, the discharge capacity of the compressor is reduced or the compressor is stopped to prevent poor lubrication of the compressor. is doing.
JP 2007-17110 A

  Incidentally, as described in paragraph 0043 of Patent Document 1, the capacity control valve of Patent Document 1 is configured to include an electromagnetic mechanism that generates an electromagnetic force corresponding to a control current value output from the control device. The valve body is displaced by a balance between the electromagnetic force generated by the electromagnetic mechanism and the force generated by the differential pressure.

  And in the refrigerating cycle device of patent documents 1, while changing discharge current of a variable capacity type compressor by changing a control current value, differential pressure itself is controlled. That is, by determining the control current value, the target differential pressure between the discharge pressure and the suction pressure is determined, so that the differential pressure can be detected based on this control current value.

  However, Patent Document 1 discloses specifications of cycle components such as a compressor and a capacity control valve, such as a refrigeration cycle apparatus having a different control mode of a capacity control valve, or a refrigeration cycle apparatus not provided with a capacity control valve. However, it does not describe details of specific differential pressure detection means in different refrigeration cycle apparatuses.

  Therefore, in the refrigeration cycle apparatus of Patent Document 1, if the specifications of the cycle constituent devices are different, there is a possibility that the refrigerant shortage state cannot be properly determined.

  In view of the above points, an object of the present invention is to provide a refrigeration cycle apparatus that can accurately determine a refrigerant shortage state regardless of the specifications of the cycle component equipment.

In order to achieve the above object, the present invention includes a compressor (11, 41) that compresses and discharges a refrigerant, a radiator (12) that radiates the refrigerant discharged from the compressor (11, 41), Decompression means (13) for decompressing and expanding the refrigerant radiated by the evaporator (12), evaporator (14) for evaporating the refrigerant decompressed and expanded by the decompression means (13), and compressors (11, 41) The discharge capacity changing means (11a, 41a) for changing the refrigerant discharge capacity of the refrigerant, the discharge capacity control means (20a) for controlling the operation of the discharge capacity changing means (11a, 41a), and the high pressure side of the cycle, A flow rate detection means (26) for detecting the flow rate of the refrigerant circulating in the cycle, a refrigerant temperature detection means (24) for detecting the temperature of a predetermined portion of the evaporator (14), and a refrigerant temperature detection means (24) Detected refrigerant detected by The flow rate estimation means (S71, S711, S712, S72, S721) for estimating the flow rate of the refrigerant circulating in the cycle using the degree (Te) and the refrigerant shortage state in which the refrigerant circulating in the cycle is insufficient Refrigerant shortage determining means (S73) for determining
In the refrigerant shortage determining means (S73), the estimated flow rate (G2) estimated by the flow rate estimating means (S71, S711, S712, S72, S721) is larger than the detected flow rate (G1) of the flow rate detecting means (26). Sometimes, it is determined that the refrigerant is in a shortage state, and the discharge capacity control means (20a) discharges the refrigerant so as to reduce the refrigerant discharge capacity of the compressor (11, 41) when the refrigerant shortage state is determined. The refrigeration cycle apparatus that controls the operation of the changing means (11a, 41a) is characterized.

  According to this, when the estimated flow rate (G2) estimated by the flow rate estimation means (S71, S711, S712, S72, S721) becomes larger than the detected flow rate (G1) of the flow rate detection means (26), the refrigerant Since the shortage determination means (S73) determines the refrigerant shortage state, the refrigerant shortage state can be reliably determined regardless of the specifications of the cycle component equipment.

  That is, in the refrigerant shortage state, the refrigerant temperature detection means (24) detects the temperature of the refrigerant having the degree of superheat, so the estimated flow rate (G2) estimated using this detected refrigerant temperature (Te) is actually the cycle. It becomes larger than the detected flow rate (G1) corresponding to the flow rate of the refrigerant circulating. Therefore, by comparing the estimated flow rate (G2) and the detected flow rate (G1), it is possible to reliably determine the refrigerant shortage state regardless of the specifications of the cycle component equipment.

  Furthermore, since the flow rate detection means (26) is arranged on the high pressure side of the cycle, the detection error of the detected flow rate (G1) with respect to the actual refrigerant flow rate can be reduced. That is, since the flow rate of the high-pressure side refrigerant having a higher density than that of the low-pressure side refrigerant is measured, an increase in pressure loss when the refrigerant passes through the flow rate detection means (26) can be suppressed. Therefore, when the flow rate detection means (26) is arranged on the low pressure side of the cycle, the detection error can be reduced and the refrigerant shortage state can be accurately determined.

  And since the discharge capacity control means (20a) controls the operation of the discharge capacity changing means (11a, 41a) so that the refrigerant discharge capacity of the compressor (11) is reduced when the refrigerant shortage state is determined. In addition, poor lubrication of the compressor (11) can be reliably prevented.

  Further, by disposing the flow rate detection means (26) on the high pressure side, it is necessary to enlarge the flow rate detection means (26) itself in order to reduce the pressure loss when the refrigerant passes through the flow rate detection means (26). Absent. As a result, it is possible to accurately determine the refrigerant shortage state, and to avoid an increase in the size of the entire refrigeration cycle apparatus.

  Further, in the refrigeration cycle apparatus having the above characteristics, the flow rate estimation means (S71, S711, S72) determines an intake refrigerant density (ρ) drawn into the compressor (11) using the detected refrigerant temperature (Te). Thus, the estimated flow rate (G2) may be estimated.

  For example, if the suction refrigerant density is determined from the saturation pressure at the detected refrigerant temperature (Te) in the refrigerant shortage state, a value higher than the density of the refrigerant actually sucked into the compressor (11, 41) is determined.

  Furthermore, if the refrigerant flow rate circulating through the cycle is estimated using a value higher than the actual refrigerant density, the estimated flow rate (G2) is also larger than the detected flow rate (G1). Therefore, if the estimated flow rate (G2) estimated in this way is used, the refrigerant shortage state can be determined more reliably.

  Further, in the refrigeration cycle apparatus having the above-described characteristics, an air temperature detecting means (21) for detecting the temperature of the air blown to the radiator (12) and a high pressure side pressure detecting means (25) for detecting the high pressure side refrigerant pressure. ), And the flow rate estimating means (S711, S712, S72, S721) is detected by the detected air temperature (Tam) detected by the air temperature detecting means (21) and the high pressure side pressure detecting means (25). The estimated flow rate (G2) may be estimated using at least one of the detected high pressures (Pd) detected.

  According to this, the flow rate estimation means (S711, S712, S72, S721) calculates the estimated flow rate (G2) using not only the detected refrigerant temperature (Te) but also the detected air temperature (Tam) and the detected high pressure (Pd). Since the estimation is performed, the estimation accuracy of the estimated flow rate (G2) can be improved. As a result, the refrigerant shortage determination means (S73) can avoid erroneously determining the refrigerant shortage state, and the refrigerant shortage state can be determined with higher accuracy.

  For example, in the refrigeration cycle apparatus having the above-described characteristics, an internal heat exchanger (16) that further exchanges heat between the refrigerant flowing out of the radiator (12) and the refrigerant sucked into the compressor (11, 41) is provided. The flow rate estimation means (S711, S72) may estimate the estimated flow rate (G2) using the detected refrigerant temperature (Te) and the detected air temperature (Tam).

  According to this, even if it is a refrigeration cycle apparatus provided with an internal heat exchanger (16), the amount of heat exchange between the high-pressure refrigerant and the low-pressure refrigerant in the internal heat exchanger (16), as will be described in an embodiment described later. It is possible to estimate the estimated flow rate (G2) with high accuracy in which the influence of the above is corrected.

  Furthermore, in the refrigeration cycle apparatus having the above-described features, the compressor is a variable displacement compressor (41) configured to be capable of changing the discharge capacity, and the discharge capacity changing means is a capacity control valve ( 41a), and the flow rate estimating means (S712, S721) estimates the estimated flow rate (G2) using the detected refrigerant temperature (Te) and the detected air temperature (Tam) or the detected high pressure (Pd). It may be.

  According to this, even in the refrigeration cycle apparatus including the variable capacity compressor (41) whose discharge capacity is changed by the capacity control valve (41a), as described in an embodiment described later, the low pressure side of the cycle A highly accurate estimated flow rate (G2) can be estimated based on the differential pressure between the refrigerant pressure and the high-pressure side refrigerant pressure.

  Further, the flow rate estimating means (S712, S713) estimates the estimated flow rate (G2) using the control current (Ic) output from the discharge capacity control means (20a) to the capacity control valve (41a). May be. This makes it possible to estimate the estimated flow rate (G2) with higher accuracy. In this case, the capacity control valve (41a) may be configured to increase or decrease the flow rate of the refrigerant circulating in the cycle according to the increase or decrease of the control current (Ic).

  Further, in the refrigeration cycle apparatus having the above-described features, the compressor is a fixed capacity compressor (11) having a fixed discharge capacity, and the discharge capacity changing means rotationally drives the fixed capacity compressor (11). The electric motor (11a) may be used.

  Moreover, in the refrigeration cycle apparatus having the above-described characteristics, the discharge capacity control means (20a) allows the compressor (11, 41) when the detected flow rate (G1) is equal to or less than a predetermined reference flow rate (KG1). The operation may be stopped.

  As described above, generally, the refrigerant circulating in the cycle is mixed with refrigerating machine oil that lubricates the compressor (11, 41), and this refrigerating machine oil circulates in the cycle together with the refrigerant. Accordingly, the reference flow rate (KG1) is set to a value larger than the refrigerant flow rate necessary for returning the refrigeration oil to the compressor (11, 41), that is, the minimum flow rate that does not cause insufficient lubrication of the compressor (11, 41). By doing so, lack of lubrication of the compressors (11, 41) can be avoided without stopping the operation of the refrigeration cycle apparatus.

  Further, in the refrigeration cycle apparatus having the above-described characteristics, the refrigerant shortage determination unit (S73) has a change amount of the refrigerant discharge capacity of the compressor (11, 41) per unit time equal to or less than a predetermined reference refrigerant discharge capacity change amount. The refrigerant shortage state determination may be performed only when

  According to this, the refrigerant shortage determination means (S73) determines the refrigerant shortage state only when the refrigerant discharge capacity change amount of the compressor (11, 41) is small, that is, when the fluctuation of the refrigerant flow rate circulating in the cycle is small. Therefore, erroneous determination due to the influence of the change in the refrigerant flow rate can be suppressed.

  Further, in the refrigeration cycle apparatus having the above-described characteristics, the refrigerant shortage determination unit (S73) determines that at least one flow rate change amount (ΔG1) among the detected flow rate (G1) and the estimated flow rate (G2) per unit time is in advance. The refrigerant shortage state may be determined only when the reference flow rate change amount (ΔKG1) is equal to or less than the predetermined reference flow rate change amount (ΔKG1).

  According to this, when the refrigerant shortage determining means (S73) has a small flow rate variation (ΔG1) of at least one of the detected flow rate (G1) and the estimated flow rate (G2), that is, a change in the flow rate of the refrigerant circulating in the cycle. Since the determination of the refrigerant shortage state is performed only when the amount of the refrigerant is small, erroneous determination due to the influence of the fluctuation of the refrigerant flow rate can be suppressed.

  In the refrigeration cycle apparatus having the above-described features, at least a part of the flow rate detection means (26) may be configured integrally with the compressors (11, 41). Thereby, size reduction of a flow volume detection means (26) and a compressor (11, 41) can be achieved.

  In the refrigeration cycle apparatus having the above-described characteristics, the compressors (11, 41) may be configured to increase the pressure of the refrigerant until the critical pressure is reached. The refrigerant may be carbon dioxide.

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

(First embodiment)
A first embodiment of the present invention will be described with reference to FIGS. In this embodiment, the refrigeration cycle apparatus 10 of the present invention is applied to a vehicle air conditioner. FIG. 1 is an overall configuration diagram of a refrigeration cycle apparatus 10 of the present embodiment.

  The refrigeration cycle apparatus 10 employs carbon dioxide as a refrigerant, and constitutes a supercritical refrigeration cycle in which the discharge refrigerant pressure of the compressor 11 is equal to or higher than the critical pressure of the refrigerant (supercritical state). Furthermore, this refrigerant is mixed with refrigerating machine oil for lubricating the compressor 11, and this refrigerating machine oil circulates in the cycle together with the refrigerant.

  In the refrigeration cycle apparatus 10, the compressor 11 sucks, compresses and discharges the refrigerant, and is driven to rotate by being transmitted with a driving force from the electric motor 11a. The compressor 11 is a fixed capacity type compressor with a fixed discharge capacity. Specifically, various compression mechanisms such as a scroll type compressor and a vane type compressor can be adopted. The compressor 11 and the electric motor 11b are accommodated in the same housing, and are integrally configured as a so-called electric compressor.

  The operation (rotation speed) of the electric motor 11a is controlled by a control signal output from the air conditioning control device 20 described later, and either an AC motor or a DC motor may be adopted. And the refrigerant | coolant discharge capability of the compressor 11 is changed by this rotation speed control. Therefore, the electric motor 11a of the present embodiment constitutes a discharge capacity changing unit.

  A radiator 12 is connected to the refrigerant discharge side of the compressor 11. The radiator 12 is a heat-dissipating heat exchanger that performs heat exchange between the high-temperature and high-pressure refrigerant discharged from the compressor 11 and the outside air (air outside the passenger compartment) blown by the cooling fan 12a to radiate the high-pressure refrigerant. The cooling fan 12a is an electric blower in which the rotation speed (the amount of blown air) is controlled by a control voltage output from an air conditioning control device 20 described later.

  As described above, since the refrigeration cycle apparatus 10 of the present embodiment constitutes a supercritical refrigeration cycle, the refrigerant passing through the radiator 12 dissipates heat in a supercritical state without condensing.

  A pressure control valve 13 is connected to the outlet side of the radiator 12. The pressure control valve 13 decompresses and expands the high-pressure refrigerant flowing out of the radiator 12 and adjusts the valve opening (throttle opening) by a mechanical mechanism so that the high-pressure side refrigerant pressure becomes the target high pressure. It is configured.

  Specifically, the pressure control valve 13 has a temperature sensing portion 13a provided on the outlet side of the radiator 12, and a pressure corresponding to the temperature of the high-pressure refrigerant on the outlet side of the radiator 12 inside the temperature sensing portion 13a. The valve opening degree of the pressure control valve 13 is adjusted by the balance between the internal pressure of the temperature sensing part 13a and the refrigerant pressure on the radiator 12 outlet side.

  Thereby, the high pressure side refrigerant pressure can be adjusted to a target high pressure determined by the high pressure side refrigerant temperature on the outlet side of the radiator 12. A pressure control valve 13 having such a high-pressure control function is known in Japanese Patent Application Laid-Open No. 2000-81157.

  An evaporator 14 is connected to the outlet side of the pressure control valve 13. The evaporator 14 performs heat exchange between the low-pressure refrigerant decompressed by the pressure control valve 13 and the blown air blown into the vehicle interior from the blower fan 14a, and evaporates the low-pressure refrigerant to exert a heat absorption effect. It is an exchanger. The blower fan 14 a is an electric blower in which the rotation speed (the amount of blown air) is controlled by a control voltage output from the air conditioning control device 20.

  As shown in the schematic external perspective view of FIG. 2, in the present embodiment, as the evaporator 14, the two first and second heat exchange units 14 b and 14 c are arranged in series with respect to the blowing air flow direction 100. We adopt what we did. The basic configurations of the first and second heat exchange units 14b and 14c are the same, and each has a well-known fin-and-tube heat exchanger structure.

  Further, the tank part 14d for the first heat exchange part 14b arranged on the windward side and the tank part 14e for the second heat exchange part 14c arranged on the leeward side are communicated with each other, and the tank parts 14d, 14e By appropriately partitioning the internal space with a separator, the refrigerant flowing out from the pressure control valve 13 is circulated in the order of the second heat exchanging portion 14c → the first heat exchanging portion 14b, as indicated by a thick solid arrow 200 in FIG.

  Thereby, the uneven distribution of the liquid-phase refrigerant in the first and second heat exchanging portions 14b and 14c that overlap in the front and rear direction of the blown air flow direction 100 is offset to effectively cool the blown air and the evaporator 14. The temperature of the air blown out from the air is made uniform.

  In addition, the evaporator 14 is disposed in a case (not shown) that forms an air passage for the air blown into the vehicle interior in the indoor air conditioning unit of the vehicle air conditioner. An inside / outside air switching box (not shown) for switching between indoor air) and outside air (vehicle compartment outside air) is provided.

  Inside the inside / outside air switching box, an inside / outside air switching door that constitutes an air volume ratio changing means for changing an air volume ratio between the air volume of the inside air and the air volume of the outside air is arranged. The inside / outside air switching door is driven by a servomotor that is controlled by a control signal output from the air conditioning controller 20.

  Therefore, the evaporator 14 can cool the inside air, the outside air, or the mixed air of the inside air and the outside air introduced through the inside / outside air switching box. Further, the blown air cooled by the evaporator 14 is reheated by a heating means such as a heater core using engine cooling water as a heat source, and the temperature is adjusted, and the blown air is blown out into the vehicle interior that is the air-conditioning target space.

  As shown in FIG. 1, an accumulator 15 is connected to the outlet side of the evaporator 14. The accumulator 15 is a gas-liquid separator that separates the refrigerant flowing out of the evaporator 14 into a liquid-phase refrigerant and a gas-phase refrigerant and stores excess refrigerant in the cycle. Further, the accumulator 15 is provided with a gas phase refrigerant outlet through which the gas phase refrigerant flows out, and the gas phase refrigerant outlet is connected to the refrigerant suction side of the compressor 11.

  Next, an outline of the electric control unit of the present embodiment will be described. The air conditioning control device 20 includes a known microcomputer including a CPU, a ROM, a RAM, and the like and peripheral circuits thereof. Then, various calculations and processes are performed based on the control program stored in the ROM, and the operations of the above-described various electric actuators 11a, 12a, 14a and the like are controlled.

  The air-conditioning control device 20 is configured integrally with control means for controlling various electric actuators. In the present embodiment, in particular, the hardware for controlling the operation of the electric motor 11a in the air-conditioning control device 20 is used. The configuration of the wear and software is the discharge capacity control means 20a.

  An air conditioning sensor group 21 to 26 and an operation panel 30 disposed in the passenger compartment are connected to the input side of the air conditioning control device 20. The air conditioning sensor group 21 to 26 is provided with a detection signal and an operation panel 30. The operation signals of the various operation switches 31 to 33 are input.

  Specifically, the air conditioning sensor group includes an outside air temperature sensor 21 that detects the outside air temperature Tam, an inside air temperature sensor 22 that detects the inside air temperature Tr, a solar radiation sensor 23 that detects the amount of solar radiation Tsun that enters the vehicle interior, and an evaporator. An evaporator temperature sensor 24 for detecting the fin temperature Te of 14, a high pressure sensor 25 for detecting the refrigerant pressure Pd discharged from the compressor 11, and a flow rate detecting means for detecting the refrigerant flow rate G 1 discharged from the compressor 11. A certain flow sensor 26 or the like is provided.

  As described above, since the radiator 12 exchanges heat between the refrigerant and the outside air, the outside air temperature sensor 21 of the present embodiment constitutes an air temperature detecting means for detecting the temperature of the air blown to the radiator 12. The outside air temperature detected by the outside air temperature sensor 21 becomes the detected air temperature Tam.

  The evaporator temperature sensor 24 is a refrigerant temperature detection unit that detects the temperature of a predetermined portion of the evaporator 14, and the fin temperature detected by the evaporator temperature sensor 24 becomes the detected refrigerant temperature Te. Further, as shown in FIG. 2, the evaporator temperature sensor 24 is disposed on the refrigerant outlet side of the second heat exchange unit 14 c disposed on the leeward side of the evaporator 14.

  The high pressure sensor 25 constitutes high pressure side pressure detecting means for detecting the high pressure side refrigerant pressure, and the discharge refrigerant pressure detected by the high pressure sensor 25 becomes the detected high pressure Pd.

  The flow rate sensor 26 constitutes a flow rate detection means for detecting the flow rate of the refrigerant circulating in the cycle, and the refrigerant flow rate detected by the flow rate sensor 26 becomes the detected flow rate G1. The flow sensor 26 is provided in a housing that forms a refrigerant discharge side passage of the compressor 11. That is, the flow sensor 26 is configured integrally with the compressor 11 so that the compressor 11 and the flow sensor 26 are downsized.

  Further, in the present embodiment, as the flow sensor 26, a throttle part that allows the refrigerant circulating in the cycle to pass through, a differential pressure detection part that detects a pressure loss (differential pressure) in the throttle part, and a downstream refrigerant of the throttle part And a temperature / pressure detector that detects the temperature and pressure of the refrigerant, and detects the refrigerant flow rate based on the detected value (differential pressure) of the differential pressure detector and the refrigerant density estimated from the detected value of the temperature / pressure detector. A differential pressure type flow sensor is used.

  Specifically, the operation switch of the operation panel 30 includes an air conditioner switch 31 that outputs an operation command signal for the vehicle air conditioner, an auto switch 32 that outputs an automatic control request signal for requesting automatic control of the air conditioning state, and a space to be cooled. A temperature setting switch 33 or the like serving as target temperature setting means for setting a target temperature Tset in the vehicle interior is provided.

  Further, an electric actuator such as an electric motor 11a of the compressor 11, an electric motor of the cooling fan 12a and the blower fan 14a, and a servo motor of the inside / outside air switching door is connected to the output side of the air conditioning control device 20, and these devices are connected. Is controlled by an output signal of the air conditioning controller 20.

  Next, the operation of this embodiment having the above configuration will be described with reference to FIGS. 3 and 4 are flowcharts showing the control processing executed by the air conditioning control device 20. This control process starts when the auto switch 32 is turned on (ON) in a state where a vehicle start switch (ignition switch) (not shown) is turned on.

  First, as shown in FIG. 3, in step S1, flags and timers are initialized, and in the next step S2, the detection signals detected by the sensor groups 21 to 26 and the operation signals of the operation panel 30 are displayed. Read.

Next, in step S3, a target blowing temperature TAO of the vehicle cabin blowing air is calculated. The target blowing temperature TAO is calculated by the following formula F1 based on the air conditioning thermal load fluctuation and the set temperature Tset set by the temperature setting switch 33.
TAO = Kset * Tset-Kr * Tr-Kam * Tam-Ksun * Tsun + C
... (F1)
Kset, Kr, Kam, and Ksun are control gains, and C is a constant for correction.

  Next, in step S4, control states of various air conditioning control devices excluding the compressor 11 are determined. That is, among various electric actuators connected to the output side of the air conditioning control device 20, control signals to be output to the electric motor of the blower fan 14a, the servo motor of the inside / outside air switching door, etc., excluding the electric motor 11a, are determined. The

  For example, with respect to the control signal (control voltage) output to the electric motor of the blower fan 14a, the control signal stored in the air-conditioning control device 20 in advance is referred to based on the target blowout temperature TAO, and appropriate according to the TAO. It decides so that it may become the amount of blast.

  More specifically, the control voltage is set to the maximum value in the extremely low temperature range (maximum cooling range) and the extremely high temperature range (maximum heating range) of TAO, and the air flow rate is set to the maximum air volume. As the TAO rises from the extremely low temperature range toward the intermediate temperature range, or as the TAO decreases from the extremely high temperature range toward the intermediate temperature range, the control voltage is decreased to reduce the air flow rate. Further, when the TAO enters a predetermined intermediate temperature range, the control voltage is set to the minimum value, and the air volume is set to the minimum air volume.

  For the control signal output to the servo motor of the inside / outside air switching door, referring to the control map stored in advance in the air conditioning control device 20 based on the target blowing temperature TAO, an appropriate inside / outside air mixing ratio according to TAO To be determined.

  Specifically, the inside air mode is set when the inside air temperature Tr is higher than a predetermined temperature with respect to the set temperature Tset (at the time of cooling high load), and as the TAO rises from the low temperature side to the high temperature side, the all inside air mode → the inside / outside air Switch from mixed mode to all outside air mode.

  Next, the target refrigerant | coolant evaporation temperature TEO in the evaporator 14 is determined in step S5. Specifically, based on the target blowing temperature TAO, the target refrigerant evaporation temperature TEO is determined with reference to a control map stored in the air conditioning control device 20 in advance. In the present embodiment, it is determined that TEO increases as TAO increases.

  Next, in step S6, the refrigerant discharge capacity of the compressor 11 is determined. Specifically, a deviation En (Te−TEO) between the fin temperature Te of the evaporator 14 detected by the evaporator temperature sensor 24 and the target refrigerant evaporation temperature TEO is calculated, and Te is calculated based on the deviation En. The control signal output to the electric motor 11a is changed by a feedback control method based on proportional-integral control (PI control) so as to approach TEO.

  In this case, specifically, when an AC motor is adopted as the electric motor 11a, a control signal for controlling the frequency of the drive current supplied from the inverter to the electric motor 11a is changed, and a DC motor is used as the electric motor 11a. If it is adopted, the control voltage supplied to the electric motor 11a is changed.

  Therefore, even if any type of electric motor is employed, the rotational speed Nc of the electric motor 11a can be detected by the control signal. Of course, if necessary, a rotation speed sensor may be provided to detect the rotation speed of the electric motor 11a.

  Next, in step S7, determination of a refrigerant shortage state in which the refrigerant circulating in the cycle is insufficient and control when a refrigerant shortage state is determined are performed. Details of step S7 will be described with reference to the flowchart of FIG. First, in step S71, the refrigerant density ρ sucked into the compressor 11 is determined using the fin temperature Te of the evaporator 14 detected by the evaporator temperature sensor 24.

  Here, if the fin temperature Te is the refrigerant evaporation temperature TL in the evaporator 14 and the saturation pressure at the refrigerant evaporation temperature TL is the refrigerant evaporation pressure PL, the compressor 11 is based on the refrigerant evaporation temperature TL and the refrigerant evaporation pressure PL. It is possible to calculate the suction refrigerant density ρ of the gas-phase refrigerant sucked in. In step S71, the suction refrigerant density ρ is determined based on the refrigerant evaporation temperature TL with reference to the control map obtained by the above calculation method.

Next, in step S72, the flow rate of the refrigerant circulating in the cycle is estimated by the following mathematical formula F2. Hereinafter, the flow rate obtained by Formula F2 is referred to as an estimated flow rate G2.
G2 = ηv × Nc × Vc × ρ (F2)
Ηv is the volumetric efficiency of the compressor 11, Nc is the rotation speed of the compressor 11 (electric motor 11 a) determined from the control signal of the electric motor 11 a, and Vc is the discharge per rotation of the compressor 11. Capacity. Accordingly, steps S71 and S72 constitute a flow rate estimating means for estimating the flow rate of the refrigerant circulating in the cycle.

  Next, in step S73, it is determined whether or not the refrigerant circulating in the cycle is insufficient. Specifically, the detected flow rate G1 and the estimated flow rate G2 are compared, and if G1 <G2, it is determined that the refrigerant is in a shortage state, and the process proceeds to step S74.

  That is, in the refrigerant shortage state, the refrigerant in the evaporator 14 enters a gas phase state (superheated state) having a superheat degree. Therefore, the suction refrigerant density ρ determined based on the fin temperature Te and the saturation pressure at the fin temperature Te is higher than the density of the refrigerant separated by the accumulator 15 and actually sucked into the compressor 11.

  Further, if the refrigerant flow rate circulating in the cycle is estimated by the above formula F2 using the suction refrigerant density ρ having a value higher than the actual refrigerant density, the estimated flow rate G2 becomes larger than the detected flow rate G1. As a result, in step S73, it can be determined whether or not the refrigerant is in a shortage state. Therefore, this step S73 constitutes a refrigerant shortage determining means for determining the refrigerant shortage state.

  In step S74, the control signal output to the electric motor 11a is changed so that the rotation speed of the electric motor 11a becomes 0, that is, the operation of the compressor 11 is stopped, and the process proceeds to step S8. On the other hand, if G1 <G2 is not satisfied, it is determined that the refrigerant is not insufficient, and the process proceeds to step S8.

  In step S8, a control signal is output from the air conditioning control device 20 to the electric actuator so that the control state determined in steps S4 and S7 is obtained. In the next step S9, the process waits for the control period τ, and when it is determined that the control period τ has elapsed, the process returns to step S2.

  In the present embodiment, as described above, when the estimated flow rate G2 estimated in steps S71 and S72 constituting the flow rate estimation unit becomes larger than the detected flow rate G1 of the flow rate sensor 26, the refrigerant shortage determination unit is configured. Since S73 determines the refrigerant shortage state, for example, even in a refrigeration cycle apparatus that employs an electric compressor, the refrigerant shortage state can be reliably determined.

  That is, regardless of the specifications of the cycle component equipment, in the refrigerant shortage state, the evaporator temperature sensor 24 detects the temperature of the superheated refrigerant as the fin temperature Te, and therefore the estimated flow rate G2 estimated using the fin temperature Te. Is larger than the detected flow rate G1 corresponding to the refrigerant flow rate actually circulating in the cycle. Therefore, the refrigerant shortage state can be reliably determined by comparing the estimated flow rate G2 and the detected flow rate G1.

  When the refrigerant shortage state is determined, the discharge capacity control means 20a sets the rotation speed of the electric motor 11a to 0 and stops the operation of the compressor 11. Therefore, it is possible to reliably prevent poor lubrication of the compressor 11. it can.

  Here, in this embodiment, the evaporator temperature sensor 24 is arrange | positioned at the refrigerant | coolant exit part side of the 2nd heat exchange part 14d. This means that the evaporator temperature sensor 24 is disposed in the vicinity of the central portion of the refrigerant flow path of the evaporator 14 as a whole. Therefore, it is possible to determine a refrigerant shortage state in which the refrigerant is overheated in the vicinity of the central portion of the refrigerant flow path of the evaporator 14 as a whole.

  On the other hand, for example, if the evaporator temperature sensor 24 is arranged in the vicinity of the inlet portion of the evaporator 14, an extreme refrigerant shortage state in which the refrigerant is overheated in the vicinity of the inlet portion of the evaporator 14 can be determined. If it arrange | positions in the exit part vicinity of the evaporator 14, the refrigerant | coolant insufficiency state of the grade which a refrigerant | coolant will be in a superheated state in the vicinity of the exit part of the evaporator 14 can be determined. That is, if the arrangement of the evaporator temperature sensor 24 is changed, the degree of the refrigerant shortage state that can be determined can be changed.

  Furthermore, in this embodiment, since the flow sensor 26 is disposed on the high pressure side of the cycle, the detection error of the detected flow G1 with respect to the actual refrigerant flow can be reduced. That is, since the flow rate of the high-pressure refrigerant having a higher density than the low-pressure refrigerant is measured, an increase in pressure loss when the refrigerant passes through the flow sensor 26 can be suppressed. Therefore, detection error and detection variation can be reduced as compared with the case where the flow sensor 26 is arranged on the low pressure side of the cycle.

  Moreover, it is not necessary to increase the size of the flow sensor 26 itself in order to reduce the pressure loss when the refrigerant passes through the flow sensor 26. As a result, it is possible to accurately determine the refrigerant shortage state, and to avoid an increase in the size of the entire refrigeration cycle apparatus. Furthermore, since the flow sensor 26 is provided in the housing that forms the refrigerant discharge side passage of the compressor 11, the flow sensor 26 and the compressor 11 can be reduced in size.

  In the above-described embodiment, when it is determined in step S73 that the refrigerant is in a shortage state, the operation of the compressor 11 is stopped in step S74. Concern is leaking into the passenger compartment. Therefore, in step S74, the control signal output to the servomotor of the inside / outside air switching door may be forcibly changed to the all outside air mode.

(Second Embodiment)
In the refrigeration cycle apparatus 10 of the present embodiment, as shown in FIG. 5, an internal heat exchanger 16 is provided for the first embodiment described above. FIG. 5 is an overall configuration diagram of the refrigeration cycle apparatus 10 of the present embodiment. In FIG. 5, the same or equivalent parts as those in the first embodiment are denoted by the same reference numerals. The same applies to the following drawings.

  The internal heat exchanger 16 exchanges heat between the radiator 12 outlet-side refrigerant passing through the high-pressure side refrigerant flow path 16a and the compressor 11 suction-side refrigerant passing through the low-pressure side refrigerant flow path 16b. The refrigerant passing through 16a is cooled. Thereby, the enthalpy difference (refrigeration capacity) between the enthalpy of the inlet side refrigerant and the enthalpy of the outlet side refrigerant in the evaporator 14 can be increased.

  As a specific configuration of the internal heat exchanger 16, a configuration in which the refrigerant pipes forming the high-pressure side refrigerant flow path 16a and the low-pressure side refrigerant flow path 16b are brazed and joined to exchange heat, or the high-pressure side refrigerant flow is set. A double-pipe heat exchanger configuration in which the low-pressure side refrigerant flow path 16b is disposed outside the inner pipe forming the path 16a can be employed.

  By the way, in the cycle including the internal heat exchanger 16 as in this embodiment, the refrigerant flowing out of the evaporator 14 is heated by the refrigerant on the outlet side of the radiator 12 and then sucked into the compressor 11. Therefore, the fin temperature Te detected by the evaporator temperature sensor 24 is lower than the temperature of the refrigerant sucked by the compressor 11. For this reason, when the suction refrigerant density ρ is determined by the same calculation method as in the first embodiment, a value smaller than the actual suction refrigerant density of the compressor 11 is determined.

  Further, if the estimated flow rate G2 is estimated using a value smaller than the actual suction refrigerant density, a value smaller than the estimated flow rate G2 actual refrigerant flow rate is estimated. There is a risk that the shortage cannot be determined. Therefore, in the present embodiment, as shown in FIG. 6, the control processing in step S7 in FIG. 3 is changed with respect to the first embodiment, and the flow rate estimation means is configured by steps S711 and S72.

First, in step S711 of this embodiment, the intake refrigerant temperature Ts is calculated by the following mathematical formula F3.
Ts = f (TL, Tam, G1, φ) (F3)
Note that φ is the heat exchange efficiency of the internal heat exchanger 16, and is determined by the length of each refrigerant flow path 16a, 16b of the internal heat exchanger 16, the heat transfer coefficient, and the like.

  As described above, in the internal heat exchanger 16, heat is exchanged between the evaporator 14 outlet-side refrigerant and the radiator 12 outlet-side refrigerant. Accordingly, the refrigerant evaporation temperature TL corresponding to the refrigerant temperature at the outlet side of the evaporator 14, the outside air temperature (detected air temperature) Tam corresponding to the refrigerant temperature at the outlet side of the radiator 12, the detected flow rate G1 of the refrigerant circulating in the cycle, and the heat exchange efficiency From φ, the temperature of the refrigerant flowing out of the low-pressure side refrigerant passage 16b of the internal heat exchanger 16, that is, the intake refrigerant temperature Ts can be calculated.

  Then, if the saturation pressure at the intake refrigerant temperature Ts calculated by the above formula F3 is the refrigerant evaporation pressure PL, the compressor 11 is controlled based on the intake refrigerant temperature Ts and the refrigerant evaporation pressure PL as in the first embodiment. The suction refrigerant density ρ of the sucked gas phase refrigerant can be determined. Other configurations and controls are the same as those in the first embodiment.

  In the present embodiment, as described above, in steps S71 and S72 constituting the flow rate estimation means, the estimated flow rate G2 is estimated using not only the detected refrigerant temperature Te but also the detected air temperature Tam, so the internal heat exchanger 16 is provided. Even in the refrigeration cycle apparatus 10, the estimation accuracy of the estimated flow rate G2 can be improved.

  That is, in step S711, the influence of the heat exchange amount between the high-pressure refrigerant and the low-pressure refrigerant in the internal heat exchanger 16 is corrected, so that the estimated flow rate G2 can be estimated with high accuracy. As a result, the refrigerant shortage state can be determined reliably and accurately, and the same effect as in the first embodiment can be obtained.

(Third embodiment)
In 1st Embodiment, when step S73 which comprises a refrigerant | coolant shortage determination means determines the refrigerant | coolant shortage state, although the action | operation of the compressor 11 is stopped, in this embodiment, even if a refrigerant | coolant shortage state is determined. An example of continuing the operation of the compressor 11 when there is no insufficient lubrication of the compressor 11 will be described. Specifically, in this embodiment, as shown in FIG. 7, the control process in step S7 of FIG. 3 is changed.

  If it is determined in step S73 that the refrigerant is in a shortage state, the process proceeds to step S75, and it is determined whether or not the detected refrigerant flow rate G1 ≦ reference flow rate KG1. The reference flow rate KG1 is set to a value higher than the refrigerant flow rate necessary for returning the refrigeration oil circulating in the cycle together with the refrigerant to the compressor 11, that is, the minimum refrigerant circulation flow rate that does not cause insufficient lubrication of the compressor 11. Has been.

  If G1 ≦ KG1 in step S75, it is determined that the compressor 11 is insufficiently lubricated, and the process proceeds to step S76 so that the number of revolutions of the electric motor 11a becomes zero, that is, the compressor The control signal output to the electric motor 11a is changed so as to stop the operation of No. 11, and the process proceeds to Step S8. On the other hand, if G1 ≦ KG1 is not satisfied in step S75, it is determined that there is no insufficient lubrication of the compressor 11 and the process proceeds to step S8.

  Other controls and configurations are the same as those in the first embodiment. Therefore, according to the refrigeration cycle apparatus 10 of the present embodiment, not only can the same effect as in the first embodiment be obtained, but also when the compressor 11 is not insufficiently lubricated, the operation of the compressor 11 is continued. Thus, the air conditioning of the air conditioning target space can be continued.

  Of course, you may apply control by step S75, S76 mentioned above to the refrigerating-cycle apparatus 10 provided with the internal heat exchanger 16 demonstrated in 2nd Embodiment. Further, even when the compressor 11 is not insufficiently lubricated, if a refrigerant shortage state is determined in step S73, an alarm means for notifying the user that a refrigerant shortage state has occurred is provided. It may be provided.

(Fourth embodiment)
In the above-described first to third embodiments, an example in which a fixed capacity compressor is employed as the compressor 11 has been described. However, in this embodiment, a variable capacity compressor 41 is employed as shown in FIG. An example will be described. FIG. 8 is an overall configuration diagram of the refrigeration cycle apparatus 10 of the present embodiment.

  The compressor 41 of the present embodiment is driven to rotate by a driving force transmitted from a vehicle travel engine (not shown) via a pulley and a belt, and is discharged by a control signal output from the air conditioning control device 20. This is a well-known swash plate type variable displacement compressor configured to be capable of continuously changing the capacity. The discharge capacity is the geometric volume of the working space where the refrigerant is sucked and compressed, that is, the cylinder volume between the top dead center and the bottom dead center of the piston stroke.

  Specifically, the compressor 41 includes a swash plate chamber (not shown) that introduces suction refrigerant and discharge refrigerant, and an electromagnetic capacity control valve 41a that adjusts the ratio of suction refrigerant and discharge refrigerant introduced into the swash plate chamber. It has a swash plate (not shown) that displaces the tilt angle in accordance with the pressure in the swash plate chamber. The piston stroke (discharge capacity) is changed according to the inclination angle of the swash plate.

  The electromagnetic capacity control valve 41a includes a pressure responsive mechanism that generates a force due to a differential pressure between the suction refrigerant pressure and the discharge refrigerant pressure of the compressor 41, and an electromagnetic mechanism that generates an electromagnetic force opposite to the force due to the differential pressure. It is built in and adjusts the valve opening (ratio of suction refrigerant and discharge refrigerant) by the balance between the force due to the differential pressure and the electromagnetic force to change the pressure in the swash plate chamber.

  Further, the electromagnetic force of the electromagnetic mechanism is determined by the control current Ic output from the air conditioning controller 20, and when the control current Ic is increased, the pressure in the swash plate chamber decreases and the inclination angle of the swash plate increases. Thereby, piston stroke (discharge capacity) increases. Conversely, when the control current Ic is decreased, the pressure in the swash plate chamber increases and the tilt angle of the swash plate decreases. Thereby, piston stroke (discharge capacity) decreases.

  Since the discharge flow rate of the compressor 41 increases and decreases according to the increase and decrease of the discharge capacity, the electromagnetic capacity control valve 41a constitutes the discharge capacity changing means in this embodiment. Note that the relationship between the control current Ic and the discharge flow rate of the compressor 41 in the present embodiment is that the discharge flow rate of the compressor 41 increases as the control current Ic increases as shown in the characteristic diagram of FIG. It is like that. A capacity control valve configured to have such a circulating refrigerant flow rate control function is known in Japanese Patent Laid-Open No. 2001-173556.

  Here, the relationship between the control current Ic and the discharge flow rate of the compressor 41 will be described with reference to FIG. Note that Icmax in FIG. 9 indicates a control current value for discharging the maximum discharge flow rate at a predetermined rotation speed. As shown in FIG. 9, the increase rate of the discharge flow rate accompanying the increase in the control current Ic decreases as the differential pressure between the suction refrigerant pressure and the discharge refrigerant pressure of the compressor 41 increases.

  The reason for this relationship is that the control current Ic required to discharge a predetermined flow rate when the differential pressure is large, as compared to the control current Ic required to discharge a predetermined flow rate when the differential pressure is small. This is because Ic becomes small.

  Further, the maximum discharge flow rate of the compressor 41 varies depending on the rotation speed of the compressor 41. When the rotation speed Nc2 is lower than the rotation speed Nc1, the rotation speed is higher than the maximum flow rate at the rotation speed Nc1, as shown in FIG. The maximum flow rate at several Nc2 is reduced. For this reason, Icmax also decreases as the rotational speed of the compressor decreases.

  Specifically, the output of the control current Ic is normally changed by duty control because of the configuration of the current control circuit, but the value of the control current Ic is directly or continuously (analog) regardless of duty control. May be changed. By adjusting the control current Ic in this way, the compressor 41 can continuously change the discharge capacity in a range of approximately 0% to 100%.

  In the compressor 41 of the present embodiment, the discharge capacity can be reduced to about 0%. Therefore, as described above, the compressor 41 has a clutchless configuration in which the compressor 41 is always connected to the vehicle travel engine via a pulley and a belt. can do. Of course, power may be transmitted from the vehicle running engine via the electromagnetic clutch. Further, in the present embodiment, a rotation speed sensor that detects the rotation speed Nc of the compressor 41 or the vehicle running engine (not shown) is also provided. Other configurations are the same as those of the first embodiment.

  Next, the operation of the refrigeration cycle apparatus 10 of the present embodiment will be described. In this embodiment, since the electromagnetic capacity control valve 41a constitutes the discharge capacity changing means, the control processing executed by the air conditioning control device 20 is as follows with respect to the flowchart shown in FIG. 3 of the first embodiment. Has been changed.

  First, in step S6, a deviation En (Te-TEO) between the fin temperature Te of the evaporator 14 detected by the evaporator temperature sensor 24 and the target refrigerant evaporation temperature TEO is calculated, and Te is calculated based on the deviation En. The control current Ic output to the electromagnetic capacity control valve 41a is changed by a feedback control method based on proportional-integral control (PI control) so as to be close to TEO.

  Next, step S7 of FIG. 3 in this embodiment is changed to step S70 shown in FIG. As shown in FIG. 10, first, in step S712, the suction refrigerant pressure Ps sucked into the compressor 41 is determined using the fin temperature Te of the evaporator 14 detected by the evaporator temperature sensor 24. Specifically, the fin temperature Te is the refrigerant evaporation temperature TL in the evaporator 14, and the saturation pressure at the refrigerant evaporation temperature TL is the suction refrigerant pressure Ps.

  Next, in step S721, based on the suction refrigerant pressure Ps determined in step S712, the detected high pressure (discharged refrigerant pressure) Pd detected by the high pressure sensor 25, the rotational speed Nc, and the control current Ic, the above-described figure. With reference to the control map obtained from 9, the estimated flow rate G2 is estimated. Therefore, in this embodiment, steps S712 and S721 constitute a flow rate estimating means. The control after step S73 is the same as in the first embodiment.

  That is, in the refrigerant shortage state, the refrigerant in the evaporator 14 is in a gas phase state (superheated state) having a superheat degree, so the suction refrigerant pressure Ps that is the saturation pressure at the fin temperature Te is sucked into the compressor 41. The value is higher than the pressure of the refrigerant. Therefore, the differential pressure between the discharge refrigerant pressure of the compressor 41 and the suction refrigerant pressure is also smaller than the actual differential pressure.

  Therefore, when the estimated flow rate G2 is estimated from the control current Ic with reference to the characteristic diagram of FIG. 9, the estimated flow rate G2 is larger than the detected flow rate G1. As a result, in step S73, it is possible to determine whether or not the refrigerant is insufficient by comparing the estimated flow rate G2 and the detected flow rate G1.

  In the present embodiment, as described above, in steps S712 and S721 constituting the flow rate estimation means, the estimated flow rate G2 is estimated using not only the detected refrigerant temperature Te but also the detected high pressure Pd, so that the electromagnetic capacity control valve 41a Even in the refrigeration cycle apparatus 10 including the variable capacity compressor 41 whose discharge capacity is changed, the estimation accuracy of the estimated flow rate G2 can be improved, and the same effect as in the first embodiment can be obtained. .

  In the present embodiment, the estimated flow rate G2 is estimated using the detected high pressure Pd as the refrigerant discharge pressure of the compressor 41, but the outside air temperature (detected air temperature) Tam detected by the outside air temperature sensor 21 is used. The refrigerant discharge pressure may be determined.

  As described above, the outside air temperature (detected air temperature) Tam is a temperature corresponding to the refrigerant temperature at the outlet side of the radiator 12. In the pressure control valve 13, the high-pressure side refrigerant pressure is determined by the high-pressure side refrigerant temperature at the outlet side of the radiator 12. The target high pressure is adjusted. Therefore, the high-pressure side refrigerant pressure, that is, the refrigerant discharge pressure can be estimated from the detected air temperature Tam. Therefore, the refrigerant shortage state can be accurately determined even if the detected air temperature Tam is used in step S721.

  Further, as in the present embodiment, the internal heat exchanger 16 is provided in the refrigeration cycle apparatus 10 including the variable capacity compressor 41 whose discharge capacity is changed by the electromagnetic capacity control valve 41a, as in the second embodiment. You may control. Furthermore, when the lack of lubrication of the compressor 41 is not caused by the same control as in the third embodiment, the operation of the refrigeration cycle apparatus 10 may be continued.

(Fifth embodiment)
In the above-described fourth embodiment, while the refrigeration cycle apparatus 10 is operating, it is repeatedly determined in step S73 whether or not the refrigerant is in a shortage state. In this embodiment, the compressor per unit time is determined. The refrigerant shortage state is determined only when the refrigerant discharge capacity change amount 41 is small, that is, when the refrigerant flow rate circulating through the cycle is small.

  Specifically, as shown in the flowchart of FIG. 11, after the refrigerant discharge capacity of the compressor 41 is determined in step S6, the control current Ic output to the electromagnetic capacity control valve 41a is compressed in step S61. It is determined whether or not the discharge flow rate of the machine 41 is greater than the maximum control current value Icmax.

  If Ic> Icmax is satisfied in step S61, the process proceeds to step S7. As shown in FIG. 9 described above, when Ic> Icmax, the discharge flow rate of the compressor 41 is maximized, so the discharge flow rate does not change. That is, in the present embodiment, the refrigerant discharge capacity change amount of the compressor 41 per unit time is equal to or less than the reference refrigerant discharge capacity change amount (change amount 0 in this embodiment), and the refrigerant flow rate circulating in the cycle does not vary.

  Then, the process proceeds to step S7, and similarly to the first embodiment, the determination as to whether or not the refrigerant is in a shortage state and the control when the refrigerant shortage state is determined are performed. On the other hand, if Ic> Icmax is not satisfied, the discharge flow rate of the compressor 41 can be changed, and the process proceeds to step S8. That is, the control process in step S7 is not performed. Other controls and configurations are the same as in the first embodiment.

  In the present embodiment, as described above, only when the change amount of the refrigerant discharge capacity of the compressor 41 per unit time is equal to or less than the reference refrigerant discharge capacity change amount, that is, there is little fluctuation in the refrigerant flow rate circulating in the cycle. Since the determination of the refrigerant shortage state is performed only at the time, erroneous determination due to the influence of the fluctuation of the refrigerant flow rate can be suppressed.

  Furthermore, when the refrigerant shortage state is determined, the refrigerant shortage state can be reliably and accurately determined as in the first embodiment. In other words, when Ic> Icmax, even when a variable capacity compressor is employed as the compressor 41, the capacity is 100%, so that the fixed capacity compressor of the first embodiment employing the fixed capacity compressor is employed. Control similar to the control step S7 can be applied. Of course, the same control as the control step S7 of the second and third embodiments may be applied.

(Sixth embodiment)
In the above fifth embodiment, only when Ic> Icmax, that is, when the flow rate of the refrigerant circulating in the cycle does not fluctuate, it is determined whether or not the refrigerant is short, and control when the refrigerant shortage is determined. In this embodiment, as shown in FIG. 12, the magnitude relationship between Ic and Icmax is compared, and the flow rate of refrigerant circulating in the cycle can vary, and when it does not vary Control according to each of these.

  Specifically, as shown in FIG. 12, when Ic> Icmax is satisfied in step S61, it is determined whether or not the refrigerant is in a shortage state by the same control as in step S7 of the fifth embodiment. Control is performed when the refrigerant shortage state is determined. On the other hand, if Ic> Icmax is not satisfied, determination is made as to whether or not the refrigerant is in a shortage state by the same control as in step S70 of the fourth embodiment, and control is performed when the refrigerant shortage state is determined.

  Other controls and configurations are the same as in the fifth embodiment. Therefore, in the present embodiment, not only can the refrigerant shortage state be reliably and accurately determined when the flow rate of the refrigerant circulating in the cycle is small, but the amount of change in the refrigerant discharge capacity of the compressor 41 per unit time can be determined. Even when the change amount of the reference refrigerant discharge capacity is exceeded, that is, when the flow rate of the refrigerant circulating in the cycle is largely varied, the refrigerant shortage state can be determined.

(Seventh embodiment)
In the above-described first embodiment, while the refrigeration cycle apparatus 10 is operating, it is repeatedly determined whether or not the refrigerant is in a shortage state in step S73. In this embodiment, the cycle is circulated per unit time. The refrigerant shortage state is determined only when the change amount of the refrigerant flow rate to be performed is smaller than the reference change amount, that is, when the change in the refrigerant flow rate circulating in the cycle is small.

  Specifically, as shown in the flowchart of FIG. 13, after the refrigerant discharge capacity of the compressor 11 is determined in step S6, the process proceeds to step S62. In step S62, the absolute value of the value obtained by subtracting the detected flow rate G1n-1 detected before the unit time stored in the memory of the air conditioning controller 20 from the detected flow rate G1n detected this time, that is, the detected flow rate per unit time. It is determined whether or not the flow rate change amount ΔG1 of G1 is equal to or less than the reference flow rate change amount ΔKG1.

  Here, the reference flow rate change amount ΔKG1 is a value determined so that the fluctuation of the refrigerant flow rate circulating in the cycle becomes a sufficiently small value. In the present embodiment, specifically, the detected flow rate detected before unit time. A value of 10% of G1n-1 is set as a reference flow rate change amount ΔKG1. Further, the unit time is 30 seconds.

  If ΔG1 ≦ ΔKG1 in step S62, it is determined that the flow rate of the refrigerant circulating in the cycle is small, and the process proceeds to step S7, and whether or not the refrigerant is in a shortage state as in the first embodiment. And the control when the refrigerant shortage state is determined.

  On the other hand, if ΔG1 ≦ ΔKG1 is not satisfied, the flow rate of the refrigerant circulating in the cycle can be greatly changed, and thus the process proceeds to step S8. That is, the control process in step S7 is not performed. Other controls and configurations are the same as in the first embodiment.

  In the present embodiment, as described above, only when the change amount of the refrigerant discharge capacity of the compressor 11 per unit time is equal to or less than the reference flow rate change amount ΔKG1, that is, only when the change in the refrigerant flow rate circulating in the cycle is small. Since the determination of the refrigerant shortage state is performed at the same time, erroneous determination due to the influence of the fluctuation of the refrigerant flow rate can be suppressed. Furthermore, when the refrigerant shortage state is determined, the refrigerant shortage state can be reliably and accurately determined as in the first embodiment.

  In step S62 of the present embodiment, it is determined that the change in the refrigerant flow rate is small based on the amount of change in the detected flow rate G1 per unit time. For example, in step S6, steps S71 and S72 are performed. Similarly, the estimated flow rate G2 may be estimated, and it may be determined that the change in the refrigerant flow rate is small based on the amount of change in the estimated flow rate G2 per unit time.

  Further, it may be determined that the refrigerant flow rate is less varied based on the change amount of the control voltage output to the electric motor 11a and the change amount of the rotational speed Nc of the compressor 11 or the electric motor 11a.

  Moreover, you may apply control by step S62 of this embodiment to the refrigerating-cycle apparatus 10 provided with the internal heat exchanger 16 demonstrated in 2nd Embodiment. Furthermore, in the present embodiment, the operation of the refrigeration cycle apparatus 10 may be continued if the lubrication of the compressor 11 is not insufficient by the same control as in the third embodiment.

(Other embodiments)
The present invention is not limited to the above-described embodiment, and can be variously modified as follows.

  (1) In the third embodiment described above, whether or not the compressor 11 is insufficiently lubricated in step S75 even when step S73 constituting the refrigerant shortage determining unit determines that the refrigerant is in a shortage state. Although it is determined, the determination in step S75 may be executed before step 73. For example, immediately after step S2 of FIG. 3, it is determined whether or not the detected flow rate G1 ≦ reference flow rate KG1, and when G1 ≦ KG1, the operation of the compressor 11 is stopped. Good.

  (2) In the fifth embodiment described above, when Ic> Icmax, it is determined that the change in the flow rate of the refrigerant circulating in the cycle is small, but the change amount ΔIc of the control current Ic per unit time is determined. May be determined to be less fluctuating in the flow rate of the refrigerant circulating in the cycle when the reference control current change amount ΔKIc is less than or equal to a predetermined value.

  In this case, ΔIc ≦ ΔKIc, and the change amount ΔNc of the rotation speed Nc of the compressor 41 (vehicle traveling engine) per unit time is equal to or less than a predetermined reference rotation speed change amount ΔNc. In addition, if it is determined that the flow rate of the refrigerant circulating through the cycle is small, the determination accuracy can be improved.

  (3) In the above-described sixth embodiment, an example in which an electric compressor constituted by the fixed capacity type compressor 11 and the electric motor 11a is described, but of course, the variable capacity type compressor 41 and the electromagnetic A type capacity control valve 41a may be employed.

  (4) In the above-described embodiment, an example in which a differential pressure type flow sensor is employed as the flow sensor 26 has been described, but the type of the flow sensor 26 is not limited thereto. For example, a mass flow sensor such as a hot wire flow sensor may be employed.

  (5) In the above-described embodiment, an example in which the refrigeration cycle apparatus 10 of the present invention is applied to a vehicle air conditioner has been described, but the application of the present invention is not limited to this. For example, the present invention may be applied to a commercial refrigeration apparatus, a household refrigerator, and the like.

  Further, the refrigerant is not limited to carbon dioxide, and a chlorofluorocarbon refrigerant or an HC refrigerant may be employed. In this case, the accumulator 15 may be abolished and a liquid receiver that separates the gas-liquid of the refrigerant on the downstream side of the radiator 12 may be provided.

  (6) In the above-described embodiment, the radiator 12 is an outdoor heat exchanger that exchanges heat between the refrigerant and the outside air, and the evaporator 14 is an indoor heat exchanger that is applied for cooling the vehicle interior. The present invention is applied to a heat pump cycle in which the evaporator 14 is configured as an outdoor heat exchanger that absorbs heat from a heat source such as outside air, and the radiator 12 is configured as an indoor heat exchanger that heats a heated fluid such as air or water. May be.

1 is an overall configuration diagram of a refrigeration cycle apparatus applied to a vehicle air conditioner according to a first embodiment. It is an external appearance perspective view of the evaporator of 1st Embodiment. It is a flowchart which shows control of the vehicle air conditioner of 1st Embodiment. It is a flowchart which shows the principal part of control of the vehicle air conditioner of 1st Embodiment. It is a whole block diagram of the refrigerating-cycle apparatus applied to the vehicle air conditioner of 2nd Embodiment. It is a flowchart which shows the principal part of control of the vehicle air conditioner of 2nd Embodiment. It is a flowchart which shows the principal part of control of the vehicle air conditioner of 3rd Embodiment. It is a whole block diagram of the refrigerating-cycle apparatus applied to the vehicle air conditioner of 4th Embodiment. It is a characteristic view which shows the relationship between the control current of 4th Embodiment, and a compressor discharge refrigerant | coolant flow rate. It is a flowchart which shows the principal part of control of the vehicle air conditioner of 4th Embodiment. It is a flowchart which shows control of the vehicle air conditioner of 5th Embodiment. It is a flowchart which shows control of the vehicle air conditioner of 6th Embodiment. It is a flowchart which shows control of the vehicle air conditioner of 7th Embodiment.

Explanation of symbols

DESCRIPTION OF SYMBOLS 11, 41 ... Compressor, 11a ... Electric motor, 12 ... Radiator, 13 ... Pressure control valve,
14 ... Evaporator, 16 ... Internal heat exchanger, 20a ... Discharge capacity control means,
21 ... Outside air temperature sensor, 24 ... Evaporator temperature sensor, 25 ... High pressure sensor,
26 ... Flow rate sensor, 41a ... Electromagnetic capacity control valve,
S71, S711, S712, S72, S721 ... flow rate estimation means,
S73: refrigerant shortage determining means, G1: detected flow rate, G2: estimated flow rate.

Claims (15)

A compressor (11, 41) for compressing and discharging the refrigerant;
A radiator (12) for radiating the refrigerant discharged from the compressor (11, 41);
Decompression means (13) for decompressing and expanding the refrigerant radiated by the radiator (12);
An evaporator (14) for evaporating the refrigerant decompressed and expanded by the decompression means (13);
Discharge capacity changing means (11a, 41a) for changing the refrigerant discharge capacity of the compressor (11, 41);
A discharge capacity control means (20a) for controlling the operation of the discharge capacity changing means (11a, 41a);
A flow rate detection means (26) disposed on the high pressure side of the cycle for detecting the flow rate of the refrigerant circulating in the cycle;
Refrigerant temperature detection means (24) for detecting the temperature of a predetermined portion of the evaporator (14);
Flow rate estimation means (S71, S711, S712, S72, S721) for estimating the flow rate of the refrigerant circulating in the cycle using the detected refrigerant temperature (Te) detected by the refrigerant temperature detection means (24);
A refrigerant shortage determining means (S73) for determining that the refrigerant circulating in the cycle is short of refrigerant.
The refrigerant shortage determining means (S73) is configured such that the estimated flow rate (G2) estimated by the flow rate estimating means (S71, S711, S712, S72, S721) is greater than the detected flow rate (G1) of the flow rate detecting means (26). When it becomes large, it is determined that the refrigerant is in a shortage state,
The discharge capacity control means (20a) operates the discharge capacity change means (11a, 41a) so as to reduce the refrigerant discharge capacity of the compressor (11, 41) when the refrigerant shortage state is determined. A refrigeration cycle apparatus that controls the refrigeration. The flow rate estimation means (S71, S711, S72) determines the intake refrigerant density (ρ) drawn into the compressor (11) using the detected refrigerant temperature (Te), thereby determining the estimated flow rate (G2 The refrigeration cycle apparatus according to claim 1, wherein: At least one of air temperature detection means (21) for detecting the temperature of air blown to the radiator (12) and high pressure side pressure detection means (25) for detecting high pressure side refrigerant pressure is provided,
The flow rate estimating means (S711, S712, S72, S721) includes a detected air temperature (Tam) detected by the air temperature detecting means (21) and a detected high pressure detected by the high pressure side pressure detecting means (25) ( The refrigeration cycle apparatus according to claim 1, wherein the estimated flow rate (G2) is estimated using at least one of Pd). And an internal heat exchanger (16) for exchanging heat between the refrigerant flowing out of the radiator (12) and the refrigerant sucked into the compressor (11, 41),
The refrigeration according to claim 3, wherein the flow rate estimation means (S711, S72) estimates the estimated flow rate (G2) using the detected refrigerant temperature (Te) and the detected air temperature (Tam). Cycle equipment. The compressor is a variable capacity compressor (41) configured to be capable of changing the discharge capacity,
The discharge capacity changing means is a capacity control valve (41a) for changing the discharge capacity,
The flow rate estimation means (S712, S721) estimates the estimated flow rate (G2) using the detected refrigerant temperature (Te) and the detected air temperature (Tam). Refrigeration cycle equipment. The compressor is a variable capacity compressor (41) configured to be capable of changing the discharge capacity,
The discharge capacity changing means is a capacity control valve (41a) for changing the discharge capacity,
The flow rate estimation means (S712, S721) estimates the estimated flow rate (G2) using the detected refrigerant temperature (Te) and the detected high pressure (Pd). Refrigeration cycle equipment. The flow rate estimation means (S712, S713) further estimates the estimated flow rate (G2) using a control current (Ic) output from the discharge capacity control means (20a) to the capacity control valve (41a). The refrigeration cycle apparatus according to claim 5 or 6, wherein The refrigeration cycle according to claim 7, wherein the capacity control valve (41a) is configured to increase or decrease the flow rate of refrigerant circulating in the cycle in accordance with increase or decrease of the control current (Ic). apparatus. The compressor is a fixed capacity compressor (11) having a fixed discharge capacity,
The refrigeration cycle apparatus according to any one of claims 1 to 4, wherein the discharge capacity changing means is an electric motor (11a) that rotationally drives the fixed capacity compressor (11). The discharge capacity control means (20a) stops the operation of the compressor (11, 41) when the detected flow rate (G1) is equal to or lower than a predetermined reference flow rate (KG1). The refrigeration cycle apparatus according to any one of claims 1 to 9. The refrigerant shortage determining means (S73) determines that the change amount of the refrigerant discharge capacity of the compressor (11, 41) per unit time is equal to or less than a predetermined reference refrigerant discharge capacity change amount. The refrigeration cycle apparatus according to any one of claims 1 to 10, wherein a refrigerant shortage state is determined. The refrigerant shortage determining means (S73) determines that at least one of the detected flow rate (G1) and the estimated flow rate (G2) per unit time is a predetermined reference flow rate change amount (ΔKG1). The refrigeration cycle apparatus according to any one of claims 1 to 11, wherein the refrigerant shortage state is determined only when: The refrigeration cycle apparatus according to any one of claims 1 to 12, wherein at least a part of the flow rate detection means (26) is configured integrally with the compressor (11, 41). The refrigeration cycle apparatus according to any one of claims 1 to 13, wherein the compressor (11, 41) pressurizes the refrigerant until the refrigerant reaches a critical pressure or higher. The refrigeration cycle apparatus according to any one of claims 1 to 14, wherein the refrigerant is carbon dioxide.

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