JP2005219576A - Vehicular air-conditioner - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a vehicular air-conditioner using a refrigeration cycle of a carbon dioxide refrigerant capable of adequately controlling a compressor without using a high-pressure sensor. <P>SOLUTION: A control device 109 operates high pressure and compressor driving torque by using an intake refrigerant temperature detected by a compressor intake temperature sensor 113, intake pressure detected by a compressor intake pressure sensor 114, and the input duty value of an ECV 102a, the duty value given to the ECV 102a is adjusted according to the required compressor torque value, and the compressor driving torque is controlled thereby. <P>COPYRIGHT: (C)2005,JPO&NCIPI

Description

  The present invention relates to a vehicle air conditioner using a supercritical refrigerant refrigeration cycle.

  In recent air conditioners for vehicles, the environmental impact is affected even if the refrigerant leaks to the outside using a refrigeration cycle using carbon dioxide, which is a supercritical fluid maintained above the gas-liquid critical temperature and pressure. Measures to reduce this are taken.

As a vehicle air conditioner using such a carbon dioxide refrigeration cycle, it is equipped with a compressor, a gas cooler (heat radiator), an internal heat exchanger, an expansion valve, an evaporator (evaporator), and an accumulator. There is known one that adjusts the degree to control the circulating refrigerant amount and the high-pressure pressure so as to obtain the cooling power with the optimum evaporating pressure (for example, see Patent Document 1).
Japanese Patent No. 2931668

  However, in the above conventional example, the following technical problems have not been solved.

(1) Since the refrigerant flow rate and pressure are mainly controlled by the throttle valve, the compressor can control the refrigerant flow rate and pressure in the same way when a variable capacity compressor is used. Emphasis and efficiency emphasis) was difficult.

(2) Since a high pressure sensor is required to detect high pressure, and a structure capable of withstanding high pressure and high temperature is required, the cost increases for ensuring reliability. In particular, when a high pressure sensor is erroneously recognized or broken, it is difficult to detect or control high pressure. Further, since the high pressure sensor is attached to the high pressure section, it is necessary to ensure sealing performance such as gas leakage. Furthermore, since the cycle (compressor capacity adjustment) is controlled after the high pressure is detected, a time loss occurs until the control device is activated after making a control judgment after detecting the pressure.

(3) Since the refrigerating capacity can only be determined by the temperature at which the evaporator blows out, not only does the heat capacity of the evaporator and the heat capacity of the temperature sensor be sensed, it will result in excessive performance, leading to excessive compressor torque consumption, This causes a decrease in fuel consumption, a decrease in acceleration performance, hunting of the blowing temperature, and the like. In addition, the required power and required torque could not be calculated.

  An object of the present invention is to provide a vehicle air conditioner that can appropriately control a compressor driving torque without using a high-pressure sensor.

  The invention according to claim 1 is a compressor driven by an engine and configured to be able to control the discharge refrigerant capacity by an external electric signal, a differential pressure control type control valve for adjusting the capacity of the compressor, and compressed by the compressor A radiator that dissipates the radiated refrigerant, a decompression unit that decompresses the refrigerant radiated by the radiator, an evaporator that evaporates the refrigerant decompressed by the decompression unit, and an engine speed that detects the engine speed Means, a suction refrigerant temperature detection means for detecting a suction refrigerant temperature of the compressor, a suction pressure detection means for detecting a suction pressure of the compressor, and a discharge refrigerant capacity of the compressor based on a detection value from each detection means And a control device for controlling the air conditioner, wherein the control device is detected by the suction refrigerant temperature detecting means. The current compressor driving torque is calculated using the suction refrigerant temperature, the suction pressure detected by the suction pressure detection means, and the high-low pressure difference obtained from the input duty value of the differential pressure control type control valve, The current compressor driving torque is controlled by adjusting a duty value given to the differential pressure control type control valve in accordance with the current compressor driving torque and a torque request value.

  The invention of claim 2 is a vehicle air conditioner according to claim 1, comprising a discharge temperature detecting means for detecting a discharge temperature of the refrigerant supplied from the compressor, instead of the intake refrigerant temperature detecting means. The control device includes a discharge temperature detected by the discharge temperature detecting means, a suction pressure detected by the suction pressure detecting means, and a high pressure-low pressure obtained from an input duty value of the differential pressure control type control valve. The present compressor driving torque is controlled by calculating a current compressor driving torque using a pressure difference and adjusting a duty value given to the differential pressure control type control valve according to the current compressor driving torque and a torque request value. To do.

  A third aspect of the present invention is the vehicle according to the first aspect, further comprising a discharge temperature detecting means for detecting a discharge temperature of the refrigerant supplied from the compressor in addition to the suction refrigerant temperature detecting means and the suction pressure detecting means. The control air conditioner is a suction air temperature detected by the suction refrigerant temperature detection means, a suction pressure detected by the suction pressure detection means, and a discharge detected by the discharge temperature detection means. The current compressor driving torque is calculated using the high pressure-low pressure pressure difference obtained from the temperature and the input duty value of the differential pressure control type control valve, and the differential pressure control type is calculated according to the current compressor driving torque and the torque request value. The present compressor driving torque is controlled by adjusting a duty value given to the control valve.

  According to a fourth aspect of the present invention, there is provided the control device according to any one of the first to third aspects, further comprising a blowing temperature detecting means for detecting a blowing temperature downstream of the evaporator, instead of the suction pressure detecting means. Is characterized by calculating the suction pressure of the compressor based on the blowing temperature detected by the blowing temperature detecting means.

  According to a fifth aspect of the present invention, in any one of the first to fourth aspects, the control device calculates a torque difference between a required compressor driving torque and a compressor driving torque obtained by calculation in the torque control of the compressor. In addition, the duty value given to the differential pressure control type control valve is controlled so that the torque difference falls within the range of the set allowable width X (%).

  A sixth aspect of the present invention provides the differential pressure control type control valve according to any one of the first to fifth aspects, wherein the differential pressure control type control valve has a characteristic capable of controlling a high pressure-low pressure differential pressure with respect to an input duty value. When the input duty value is high, the differential pressure increases.

  A seventh aspect of the invention is characterized in that, in any one of the first to sixth aspects, the differential pressure control type control valve has a characteristic in which a maximum control pressure is provided in a region of a specified duty value or more.

  According to an eighth aspect of the present invention, in any one of the first to seventh aspects, the control device calculates a high-low pressure differential pressure with respect to an input duty value of the differential pressure control type control valve. And

  According to a ninth aspect of the present invention, in any one of the first to eighth aspects, the control device calculates an evaporation temperature of the evaporator based on a suction pressure of the compressor.

  According to a tenth aspect of the present invention, in any one of the first to ninth aspects, the control device calculates a high pressure − based on an intake pressure of the compressor and an input duty value of the differential pressure control type control valve. The high pressure of the compressor is calculated from the low differential pressure.

  An eleventh aspect of the invention is characterized in that, in the tenth aspect, the control device controls the calculated high pressure so as not to exceed a predetermined normal maximum pressure.

  According to a twelfth aspect of the present invention, when the calculated high pressure exceeds a predetermined allowable maximum pressure, the control device decreases the input duty value using the calculation of the fifth aspect, and temporarily It is characterized by controlling to the minimum differential pressure.

  A thirteenth aspect of the present invention is the control device according to any one of the first to twelfth aspects, wherein the control device uses a correlation equation or a correlation map between a pressure difference across the pressure reducing means and a refrigerant circulation volume flow rate to calculate the compressor. The refrigerant circulation volume flow rate is calculated.

  A fourteenth aspect of the invention is characterized in that, in any one of the first to thirteenth aspects, the control device calculates an intake refrigerant density from an intake temperature and an intake pressure of the compressor.

  According to a fifteenth aspect of the present invention, in the fourteenth aspect, the control device calculates the refrigerant circulation weight flow rate using the refrigerant circulation volume flow rate calculated in the claim 13 and the suction refrigerant density calculated in the claim 14. It is characterized by that.

  According to a sixteenth aspect of the present invention, in the fifteenth aspect, the control device uses the refrigerant circulation weight flow calculated in the fifteenth aspect and the compressor input enthalpy calculated from the suction refrigerant enthalpy and the discharge refrigerant enthalpy, Special feature is computing.

  According to a seventeenth aspect of the present invention, in the sixteenth aspect, the control device is configured to calculate the actual rotational speed of the compressor calculated from the compressor consumption power calculated in the sixteenth aspect and the engine speed detected by the engine speed means. And calculating the current compressor driving torque.

  According to an eighteenth aspect of the present invention, in any one of the first to seventeenth aspects, the pressure reducing means is provided with an internal release mechanism.

  According to a nineteenth aspect of the present invention, in any one of the first to eighteenth aspects, the pressure control by the control device, the ΔPecv characteristic of the differential pressure control type control valve, the ΔPecvmax characteristic, and the internal relief mechanism of the pressure reducing means, It is characterized by preventing an increase in high pressure.

  A twentieth aspect of the present invention is the control device according to any one of the first to nineteenth aspects, wherein the control device includes a suction pressure detected by the suction pressure detection means, a suction refrigerant temperature detected by the suction refrigerant temperature detection means, The compressor inlet superheat degree is calculated using the refrigerant density relational expression of the carbon dioxide Mollier diagram.

  According to a twenty-first aspect of the present invention, in the twentieth aspect, the control device calculates the high pressure calculated in the tenth aspect, the compressor inlet superheat degree calculated in the twentieth aspect, and the refrigerant density relational expression of the carbon dioxide Mollier diagram. And calculating a compressor discharge temperature, and controlling a duty value applied to the differential pressure control type control valve so that the compressor discharge temperature falls within a range of a set allowable allowable discharge temperature.

  According to the vehicle air conditioner of the present invention, the compressor driving torque is calculated using the detected values of the pressure sensor and the temperature sensor provided on the compressor suction side, and the discharged refrigerant capacity of the compressor is controlled according to the compressor driving torque. Therefore, the compressor driving torque can be appropriately controlled without using a high pressure sensor.

  Hereinafter, as a best mode for carrying out a vehicle air conditioner according to the present invention, an embodiment of a vehicle air conditioner using a refrigeration cycle using carbon dioxide gas as a refrigerant will be described.

  FIG. 1 is a system configuration diagram of a vehicle air conditioner according to a first embodiment.

  The vehicle air conditioner in the present embodiment includes a refrigeration cycle 101 that circulates carbon dioxide gas serving as a refrigerant and exchanges heat between the refrigerant and air.

  In the refrigeration cycle 101, an external capacity variable compressor 102, a gas cooler 103, an internal heat exchanger 104, an evaporator 105, and an accumulator 106 are connected by piping, and a refrigerant to which kinetic energy is given by the external capacity variable compressor 102 is between them. It is configured to circulate.

  An external capacity variable compressor (hereinafter referred to as a compressor) 102 is disposed outside a vehicle compartment such as an engine room, and compresses the low-pressure gaseous refrigerant sucked from the internal heat exchanger 104 to the gas cooler 103 as a high-pressure gaseous refrigerant. Discharge. The compressor 102 is driven by transmitting the power of the engine 107 through the belt 108. The compressor 102 is of a swash plate type, and the inclination of the swash plate is controlled by an electric signal from the outside.

  Therefore, the compressor 102 includes a differential pressure control valve (hereinafter referred to as ECV) 102a such as an electromagnetic valve that can be controlled by an electric signal from the outside. For example, when an electromagnetic valve communicating with the high pressure side is used as the ECV 102a, the inside of the crankcase and the low pressure side communicate with each other through a passage having a predetermined opening degree so that the pressure in the crankcase escapes to the low pressure side. It has become. Therefore, by controlling the pressure in the crankcase by introducing and shutting off the high pressure side pressure by adjusting the opening of this solenoid valve, the balance of the pressure applied to the piston is changed, and the inclination of the swash plate is changed. Thus, the discharge refrigerant capacity of the compressor 102 can be controlled. At this time, a duty signal having a predetermined duty ratio (duty value) is given to the solenoid valve as an external electrical signal from the control device 109 that controls the operation of the entire system. The valve opening time of the electromagnetic valve is determined according to the duty ratio at this time, and the discharge refrigerant capacity from the compressor 102 is set accordingly.

  The gas cooler 103 is disposed outside the passenger compartment, and dissipates heat of the high-temperature and high-pressure gaseous refrigerant discharged from the compressor 102 to the outside air. The gas cooler 103 is driven by a blowing means such as an electric fan, for example, so that outside air is blown. Then, heat is exchanged between the high-temperature and high-pressure gaseous refrigerant passing through the gas cooler 103 and the outside air to be blown, thereby radiating the heat of the high-temperature and high-pressure gaseous refrigerant to the outside air.

  The internal heat exchanger 104 exchanges heat between the high-temperature and high-pressure gaseous refrigerant radiated by the gas cooler 103 and the low-temperature and low-pressure gaseous refrigerant evaporated by the evaporator 105 described later.

  The orifice 110 decompresses (expands) the high-temperature and high-pressure gaseous refrigerant output from the internal heat exchanger 104 to form a low-temperature and low-pressure mist refrigerant.

  The evaporator 105 is disposed in the air conditioning duct 111 and absorbs heat of the conditioned air generated by the air conditioning fan 112 into the low-temperature and low-pressure mist refrigerant supplied from the internal heat exchanger 104 through the orifice 110. The refrigerant supplied to the evaporator 105 as a low-temperature and low-pressure mist refrigerant at the orifice 110 takes the heat of the conditioned air flowing through the air-conditioning duct 111 and evaporates when passing through the evaporator 105. The conditioned air absorbed by the refrigerant in the evaporator 105 is dehumidified to become a cooling air and blown out from the air outlet 111a into the vehicle interior.

  The accumulator 106 separates the refrigerant discharged from the evaporator 105 into gas and liquid. Among these, the liquid refrigerant is stored, and the gaseous refrigerant separated from the liquid refrigerant is sent to the internal heat exchanger 104.

  A compressor suction temperature sensor 113 and a compressor suction pressure sensor 114 are disposed on the low pressure side (suction side) of the compressor 102. The compressor suction temperature sensor 113 is a sensor that detects the compressor suction temperature Ts of the refrigerant on the suction side of the compressor 102. The compressor suction pressure sensor 114 is a sensor that detects a refrigerant suction pressure (hereinafter referred to as a low pressure) Ps on the suction side of the compressor 102. Instead of the compressor suction pressure sensor 114, a blowing temperature sensor 115 may be provided on the downstream side of the evaporator 105. The refrigerant temperature, refrigerant pressure, etc. detected by these sensors are supplied to the control device 109.

  The control device 109 is constituted by a microcomputer including a CPU, a ROM, and a RAM. The control device 109 detects values from various sensors such as the compressor intake temperature sensor 113 and the compressor intake pressure sensor 114, and a rotation speed detection sensor 116 provided in the engine 107. Based on the engine speed detected in step 1, the calculation processing described later is performed, the duty ratio given to the ECV 102a is calculated, and the ECV 102a is controlled.

  Next, control of the refrigeration cycle in Example 1 will be described. First, the compressor driving torque calculation will be described.

  FIG. 2 is a flowchart illustrating a processing procedure when the compressor driving torque of the compressor 102 is calculated in the control device 109 according to the first embodiment. In FIG. 2, it is assumed that the processes in steps S101 to S106 are executed in parallel for each column.

  In step S101, the engine speed is input from the rotation speed detection sensor 116, and in step S102, the actual compressor speed Nc is calculated from the engine speed and the pulley ratio.

  In step S 103, the low pressure Ps is input from the compressor suction pressure sensor 114 (or the low pressure is calculated from the blowing temperature inputted from the blowing temperature sensor 115), and the compressor suction temperature Ts is inputted from the compressor suction temperature sensor 113. In step S104, the refrigerant density relationship described in relation (1), relation (2), and relation (3) described later is used in the carbon dioxide Mollier diagram shown in FIG. 3 using the low pressure Ps and the compressor suction temperature Ts. Used to calculate the suction refrigerant density.

  The control device 109 stores a function expression (saturation line, isothermal line, specific volume line, isentropic line) of the carbon dioxide Mollier diagram shown in FIG. 3 as a program. The control device 109 executes the following calculation using this program and input data.

  Relationship (1): The evaporation temperature of the evaporator is calculated based on the low pressure Ps of the compressor.

  Relationship (2): The low pressure Ps of the compressor is calculated based on the evaporation temperature of the evaporator.

  Relationship (3): The compressor inlet superheat degree SH is calculated based on the compressor suction temperature Ts and the low pressure Ps. At the same time, the suction refrigerant density is calculated from the isovolumetric line.

  Relationship (4): The above relationship (3) is calculated, and the high pressure difference Pd of the compressor is calculated by predicting the high and low differential pressure from the ECV characteristics.

  Relation (5): When the relation (4) and the adiabatic efficiency of the compressor are stored as a program, the compressor discharge temperature Td is calculated from the isothermal line. That is, in FIG. 3, since the slope of the thick line A is calculated by the compressor heat insulation efficiency, the intersection of the high pressure Pd and the thick line A is the intersection of the compressor high pressure Pd and the compressor discharge temperature Td. Therefore, the compressor discharge temperature Td can be calculated from this intersection and the isothermal line.

  In step S105, the duty ratio of the ECV 102a is input. In step S106, the high / low differential pressure ΔPcomp is calculated from the high / low differential pressure relationship in the ECV characteristic diagram shown in FIG.

  In step S107, the high pressure Pd is calculated from the suction refrigerant density obtained in step S104, the low pressure Ps input in step S103, and the high / low differential pressure ΔPcomp obtained in step S106 (Pd = ΔPcomp + Ps). In step S108, the compressor discharge temperature Td is calculated from the suction refrigerant density obtained in step S104 and the compressor adiabatic efficiency described in relation (5) in FIG.

  In step S109, the suction refrigerant enthalpy Icompin is calculated from the suction refrigerant density obtained in step S104. Further, the discharge refrigerant enthalpy Icompout is calculated from the high pressure Pd obtained in step S107 and the compressor discharge temperature Td obtained in step S108. Then, the compressor charging enthalpy ΔIcomp is calculated from the suction refrigerant enthalpy Icompin and the discharged refrigerant enthalpy Icompout (ΔIcomp = Icompout−Icompin).

  In step S110, the refrigerant circulation volume flow rate Grv of the compressor 102 is calculated from the compressor charging enthalpy ΔIcomp obtained in step S109 and the pressure-flow rate characteristic of the orifice volume flow rate pressure loss characteristic diagram shown in FIG. Then, the refrigerant circulation weight flow rate Grw is calculated from the refrigerant circulation volume flow rate Grv and the suction refrigerant density obtained in step S104 (Grw = Grv × suction density).

  In step S111, the compressor power consumption Power is calculated from the refrigerant circulation weight flow rate Grw obtained in step S110 and the compressor charging enthalpy ΔIcomp obtained in step S109 (Power = Grw × ΔIcomp). Then, the compressor driving torque (value) Torque is calculated from the actual compressor speed Nc and the compressor power consumption Power calculated in step S102 (Torque = Power ÷ (Nc · 1.027)).

  Next, torque control of the compressor 102 will be described.

  FIG. 6 is a flowchart showing a processing procedure when the control device 109 performs torque control of the compressor 102 using the calculated compressor driving torque.

  In step S201, the compressor torque request value Trq to the compressor 102 is input, and in step S202, the current compressor driving torque Tcur (= Torque) of the current compressor 102 is calculated (according to the flowchart of FIG. 2). In step S203, a torque difference Δtorque between the compressor torque request value Trq and the current compressor drive torque Tcur is calculated (Δtorque = Trq−Tcur). In step S204, it is determined whether or not the torque difference Δtorque is within the set allowable width X (%). If not, the current state is maintained and the process returns to step S201. If the allowable width X (%) is exceeded, step S205 is performed to reversely calculate the compressor torque request value Trq and the process of the flowchart for calculating the compressor driving torque in FIG. . Subsequently, in step S206, the duty ratio is calculated from the required elevation differential pressure ΔPcomp and the elevation differential pressure relationship in the ECV characteristic diagram shown in FIG. In step S207, the instruction value having the duty ratio is supplied to the ECV 102a.

  Therefore, according to this embodiment, the compressor driving torque can be calculated without using the high pressure sensor. Then, by controlling the ECV 102a with an instruction value having a duty ratio calculated based on this value, the compressor driving torque of the compressor 102 can be controlled to become the compressor torque request value.

  Next, the superheat degree control of the compressor 102 will be described.

  FIG. 7 is a flowchart showing a processing procedure when the control device 109 performs compressor inlet superheat (SH) control.

  In step S301, the low pressure Ps input from the compressor suction pressure sensor 114 (or the low pressure calculated from the blowout temperature input from the blowout temperature sensor 115) and the compressor suction temperature Ts input from the compressor suction temperature sensor 113 are used. 3, the current compressor inlet superheat degree SHcur is calculated according to the refrigerant density relationship described in the above relationship (3).

  In step S302, it is determined whether or not the compressor inlet superheat degree SHcur obtained in step S301 is within the set allowable superheat degree SHlim. If not exceeded, the current state is maintained and the process returns to step S301.

  In step S302, if the compressor inlet superheat degree SHcur exceeds the set allowable superheat degree SHlim, in step S303, the duty ratio instruction value given to the ECV 102a is decreased by a predetermined amount. Hereinafter, the compressor inlet superheat degree SHcur can be appropriately controlled by lowering the instruction value of the duty ratio stepwise until the compressor inlet superheat degree SHcur falls within the set allowable superheat degree SHlim.

  Next, the discharge temperature control of the compressor 102 will be described.

  FIG. 8 is a flowchart showing a processing procedure when the controller 109 performs compressor discharge temperature control.

  In step S401, the low pressure Ps input from the compressor suction pressure sensor 114 (or the low pressure calculated from the blowout temperature input from the blowout temperature sensor 115) and the compressor suction temperature Ts input from the compressor suction temperature sensor 113 are used. 3, the current compressor inlet superheat degree SHcur is calculated according to the refrigerant density relationship described in the above relationship (3).

  In step S402, the compressor inlet superheat degree SHcur obtained in step S401 and the high / low differential pressure ΔPcomp obtained in step S106 in FIG. 2 are used, and the current compressor is expressed by the refrigerant density relationship described in relation (5) in FIG. The discharge temperature Tdcur is calculated.

  In step S403, it is determined whether or not the compressor discharge temperature Tdcur is within the set allowable discharge temperature Tdlim. If not exceeded, the current state is maintained and the process returns to step S401.

  In step S403, if the compressor discharge temperature Tdcur exceeds the set allowable discharge temperature Tdlim, the duty ratio instruction value given to the ECV 102a is decreased by a predetermined amount in step S404. Hereinafter, the compressor discharge temperature Tdcur can be appropriately controlled by lowering the duty ratio instruction value stepwise until the compressor discharge temperature Tdcur falls within the set allowable discharge temperature Tdlim.

  Next, the evaporation temperature control of the evaporator 105 will be described.

  FIG. 9 is a flowchart showing a processing procedure in the case where the control device 109 performs the evaporation temperature control.

  In step S501, the low-pressure pressure Ps input from the compressor suction pressure sensor 114 (or the low-pressure pressure calculated from the blowing temperature input from the blowing temperature sensor 115) is used, and the refrigerant density relationship described in relation (1) above in FIG. To calculate the current evaporation pressure Pecur of the evaporator 105.

  In step S502, it is determined whether or not the evaporation pressure Pecur is within the set evaporation pressure Pelim. If it does not exceed, the current state is maintained and the process returns to step S501.

  In step S502, if the evaporation pressure Pecur exceeds the set evaporation pressure Pelim, the instruction value of the duty ratio to be given to the ECV 102a is decreased by a predetermined amount in step S503. Hereinafter, the evaporation temperature of the evaporator 105 can be controlled by adjusting the evaporation pressure Pecur by gradually decreasing the indicated value of the duty ratio until the evaporation pressure Pecur of the evaporator 105 falls within the set evaporation pressure Pelim.

  According to the present embodiment, by using the differential pressure type ECV 102a, the compressor suction temperature sensor 113, and the compressor suction pressure sensor 114, appropriate pressure such as high pressure, compressor consumption power, and compressor drive torque can be obtained without using a high pressure sensor. Pressure control can be performed.

  Further, even when an abnormality occurs in the communication cable or control circuit of the ECV 102a, it is possible to prevent a predetermined pressure difference abnormality from occurring due to the ΔPecv characteristic and the Pecvmax characteristic value shown in FIG.

  Further, by adding an internal relief mechanism (relief valve) to the orifice 110, safety can be further increased.

  Further, when the compressor inlet superheat degree SH is obtained by calculation and exceeds a predetermined compressor inlet superheat degree SH, the compressor inlet superheat degree SH can be controlled by adjusting the capacity of the compressor 102 by the ECV 102a. Oil return can be secured and the reliability of the compressor can be improved.

  When the compressor discharge temperature Td is obtained by calculation and exceeds the predetermined compressor discharge temperature Td, the capacity of the compressor 102 can be adjusted by the ECV 102a to control the compressor discharge temperature Td, so that the compressor discharge temperature sensor is unnecessary. In addition to improving the reliability of the compressor, the thermal reliability of the system components can also be improved.

  Incidentally, in Japanese Patent No. 2931668, there is no description of a compressor discharge temperature sensor, but the refrigeration cycle of carbon dioxide gas has a high discharge temperature due to the characteristics of its refrigerant properties. Therefore, it is necessary to provide a compressor discharge temperature sensor to monitor the discharge temperature and to control it so as not to deteriorate the performance. However, since the compressor discharge temperature sensor is not required in the configuration of the present embodiment, it is possible to improve the reliability of the compressor and the thermal reliability of the system components.

  Further, in the first embodiment, the evaporation pressure temperature Te of the evaporator 105 is obtained by calculation from the low pressure Ps of the compressor, and when the evaporator evaporation temperature Te falls below a predetermined evaporator evaporation temperature Te, the evaporator evaporation temperature is adjusted by adjusting the capacity of the compressor 102 using the ECV 102a. Te can be controlled. For this reason, for example, it is possible to control the evaporator evaporation temperature Te so as not to fall below 0 ° C., and it is possible to achieve both sufficient securing of cooling power and prevention of freezing of the evaporator.

  Further, there are the following effects as the incidental effects of the first embodiment.

  By controlling the duty ratio of the ECV 102a, the torque of the compressor 102 is instantaneously reduced and acceleration is improved. Further, the torque reduction allowance can be freely controlled by the duty ratio. Specifically, the acceleration performance is dramatically improved by instantaneously reducing the torque of the compressor 102 during acceleration from low-speed driving. On the other hand, when the engine speed decreases, the torque of the compressor 102 is instantaneously decreased and recovered slowly, thereby preventing engine stall (or engine stall) of low-rotation characters near the idle.

  In a refrigeration cycle using carbon dioxide gas, which is supercritical on the high-pressure side, as a refrigerant, there is a correlation between the lift work of the compressor (differential pressure between high pressure and low pressure) and the generated torque. That is, by controlling the lift head work of the compressor with the pressure control valve, the target generated torque can be obtained, so an engine that requires precise control, such as a lean burn engine or a direct injection engine, or a light engine Cooperative control with an engine such as an automobile that generates less engine torque is also possible.

  FIG. 10 is a system configuration diagram of the vehicle air conditioner according to the second embodiment. The same parts as those in FIG. 1 are denoted by the same reference numerals.

  The vehicle air conditioner in this embodiment has a compressor discharge temperature sensor 117 disposed on the high pressure side (discharge side) of the compressor 102 instead of the compressor intake temperature sensor 113, and the other configurations are the same as those in FIG. is there. In the present embodiment, the intake temperature (and the degree of superheat and the intake refrigerant density) is calculated by calculating back the adiabatic efficiency from the compressor discharge temperature.

  FIG. 11 is a flowchart illustrating a processing procedure when the compressor driving torque of the compressor 102 is calculated in the control device 109 according to the second embodiment. In FIG. 11, the processing in steps S601 to S611 is executed in parallel for each column.

  In step S601, the engine speed is input from the rotation speed detection sensor 116, and in step S602, the actual compressor speed Nc is calculated from the engine speed and the pulley ratio.

  In step S603, the duty ratio of the ECV 102a is input. In step S604, the high / low differential pressure ΔPcomp is calculated from the high / low differential pressure relationship in the ECV characteristic diagram shown in FIG.

  In step S605, the low pressure Ps is input from the compressor suction pressure sensor 114 (or the low pressure is calculated from the blowing temperature input from the blowing temperature sensor 115). In step S606, the high / low differential pressure ΔPcomp obtained in step S604 and the step are calculated. The (discharge) high pressure Pd is calculated from the low pressure Ps input in S605 (Pd = ΔPcomp + Ps).

  In step S607, the compressor discharge temperature Td is input from the compressor discharge temperature sensor 117. In step S608, the high pressure Pd obtained in step S606, the compressor discharge temperature Td input in step S607, and the refrigerant density relationship shown in FIG. Is used to calculate the high-pressure refrigerant density. This high-pressure refrigerant density is a physical quantity determined by (discharge) high-pressure pressure and compressor discharge temperature, and can be calculated by a relational expression. That is, the isovolume line shown in FIG. 3 is an inverse function of the high-pressure refrigerant density, and if the specific volume is obtained from the compressor discharge temperature and the high-pressure pressure, the inverse high-function refrigerant density can be obtained.

  In step S609, the compressor intake temperature Ts is calculated using the high-pressure refrigerant density obtained in step S608 and the compressor adiabatic efficiency described in relation (5) in FIG. In step S610, the compressor suction temperature Ts obtained in step S609, the low pressure Ps input in step S605, and the refrigerant density relationship described in relation (1), relation (2), and relation (3) in FIG. Is used to calculate the suction refrigerant density.

  In step S611, an intake refrigerant enthalpy Icompin is calculated from the intake refrigerant density obtained in step S610. Further, the discharge refrigerant enthalpy Icompout is calculated from the high pressure Pd obtained in step S606 and the compressor discharge temperature Td input in step S607. Then, the compressor charging enthalpy ΔIcomp is calculated from the suction refrigerant enthalpy Icompin and the discharged refrigerant enthalpy Icompout (ΔIcomp = Icompout−Icompin).

  In step S612, the refrigerant circulation volume flow rate Grv of the compressor 102 is calculated from the compressor charging enthalpy ΔIcomp obtained in step S611 and the pressure-flow rate characteristic of the orifice volume flow rate pressure loss characteristic diagram shown in FIG. Then, the refrigerant circulation weight flow rate Grw is calculated from the refrigerant circulation volume flow rate Grv and the suction refrigerant density obtained in step S610 (Grw = Grv × suction density).

  In step S613, the compressor power consumption Power is calculated from the refrigerant circulation weight flow rate Grw obtained in step S612 and the compressor charging enthalpy ΔIcomp obtained in step S611 (Power = Grw × ΔIcomp). Then, the compressor driving torque (value) Torque is calculated from the actual compressor speed Nc obtained in step S602 and the compressor power consumption Power (Torque = Power ÷ (Nc · 1.027)).

  In this embodiment, the torque control of the compressor 102, the compressor inlet superheat (SH) control, and the evaporation temperature control are the same as those in FIGS. 6, 7, and 9 of the first embodiment.

  According to the present embodiment, by using the differential pressure type ECV 102a, the compressor suction pressure sensor 114, and the compressor discharge temperature sensor 117, it is possible to appropriately set the high pressure, compressor consumption power, compressor driving torque, etc. without using the high pressure sensor. Pressure control can be performed.

  Particularly in the second embodiment, since the discharge temperature is directly monitored by the compressor discharge temperature sensor, the discharge temperature can be accurately monitored and accurately controlled. Therefore, not only the reliability of the compressor is improved, but also the thermal reliability of the system components can be improved.

  FIG. 12 is a system configuration diagram of the vehicle air conditioner according to the third embodiment. 1 and 10 are denoted by the same reference numerals.

  The vehicle air conditioner according to this embodiment includes a compressor suction temperature sensor 113 and a compressor suction pressure sensor 114 on the low pressure side (suction side) of the compressor 102, and a compressor discharge temperature on the high pressure side (discharge side) of the compressor 102. A sensor 117 is arranged. Other configurations are the same as those in FIGS. FIG. 13 is a flowchart illustrating a processing procedure when the compressor driving torque of the compressor 102 is calculated in the control device 109 according to the third embodiment. In FIG. 13, it is assumed that the processing in steps S701 to S707 is executed in parallel for each column.

  In step S701, the engine speed is input from the rotation speed detection sensor 116, and in step S702, the actual compressor speed Nc is calculated from the engine speed and the pulley ratio.

  In step S 703, the low pressure Ps is input from the compressor suction pressure sensor 114 (or the low pressure is calculated from the blowing temperature input from the blowing temperature sensor 115), and the compressor suction temperature Ts is input from the compressor suction temperature sensor 113. In step S704, the low-pressure pressure Ps and the compressor suction temperature Ts are used, and the refrigerant density relationship described in relation (1), relation (2), and relation (3) is used in the carbon dioxide Mollier diagram shown in FIG. To calculate the suction refrigerant density.

  In step S705, the duty ratio of the ECV 102a is input. In step S706, the high / low differential pressure ΔPcomp is calculated from the high / low differential pressure relationship in the ECV characteristic diagram shown in FIG.

  In step S707, the compressor discharge temperature Td is input from the compressor discharge temperature sensor 117.

  In step S708, the high pressure Pd is calculated from the suction refrigerant density obtained in step S704, the low pressure Ps input in step 703, and the high / low differential pressure ΔPcomp obtained in step S706 (Pd = ΔPcomp + Ps).

  In step S709, an intake refrigerant enthalpy Icompin is calculated from the intake refrigerant density obtained in step S704. Further, the discharge refrigerant enthalpy Icompout is calculated from the high pressure Pd obtained in step S708 and the compressor discharge temperature Td input in step S707108. Then, the compressor input enthalpy ΔIcomp is calculated from the suction refrigerant enthalpy Icompin and the discharged refrigerant enthalpy Icompout (ΔIcomp = Icompout−Icompin).

  In step S710, the refrigerant circulation volume flow rate Grv of the compressor 102 is calculated from the compressor charging enthalpy ΔIcomp obtained in step S709 and the pressure-flow rate characteristic of the orifice volume flow rate pressure loss characteristic diagram shown in FIG. Then, the refrigerant circulation weight flow rate Grw is calculated from the refrigerant circulation volume flow rate Grv and the suction refrigerant density obtained in step S104 (Grw = Grv × suction density).

  In step S711, the compressor power consumption Power is calculated from the refrigerant circulation weight flow rate Grw obtained in step S710 and the compressor charging enthalpy ΔIcomp obtained in step S709 (Power = Grw × ΔIcomp). Then, the compressor driving torque (value) Torque is calculated from the actual compressor speed Nc and the compressor power consumption Power calculated in step S702 (Torque = Power ÷ (Nc · 1.027)).

  In this embodiment, the torque control of the compressor 102, the compressor inlet superheat (SH) control, and the evaporation temperature control are the same as those in FIGS. 6, 7, and 9 of the first embodiment.

  According to the present embodiment, by using the differential pressure type ECV 102a, the compressor suction temperature sensor 113, the compressor suction pressure sensor 114, and the compressor discharge temperature sensor 117, high pressure, compressor consumption power, Appropriate pressure control such as compressor driving torque can be performed.

  In particular, in the third embodiment, by using sensor values from the compressor suction temperature sensor 113, the compressor suction pressure sensor 114, and the compressor discharge temperature sensor 117, some values can be obtained directly, so that the calculation is simplified. Can be Specifically, by using the compressor discharge temperature Td input from the compressor discharge temperature sensor 117, the calculation in step S108 shown in FIG.

Further, in this embodiment, by using the carbon dioxide Mollier diagram shown in FIG. 3, the compressor adiabatic efficiency can be calculated from the suction refrigerant density, the high pressure Pd, and the compressor discharge temperature Td. For this reason, it is possible to detect a cycle abnormality (refrigerant missing) and a compressor abnormality (failure or previous generation of sliding heat).

1 is a system configuration diagram of a vehicle air conditioner according to Embodiment 1. FIG. 3 is a flowchart illustrating a processing procedure when calculating a current compressor driving torque in the first embodiment. Carbon dioxide Mollier diagram and cycle control diagram of diagram. ECV characteristic diagram. Fig. 3 is a volume flow pressure loss characteristic diagram of an orifice. The flowchart which shows the process sequence in the case of performing the required torque control of a compressor using the present compressor drive torque in Examples 1-3. 3 is a flowchart illustrating a processing procedure when performing compressor inlet superheat (SH) control in the first embodiment. 3 is a flowchart illustrating a processing procedure when compressor discharge temperature control is performed in the first embodiment. The flowchart which shows the process sequence in the case of performing evaporation temperature control in Examples 1-3. The system block diagram of the vehicle air conditioner concerning Example 2. FIG. 9 is a flowchart illustrating a processing procedure when calculating a current compressor driving torque in the second embodiment. FIG. 6 is a system configuration diagram of a vehicle air conditioner according to a third embodiment. 9 is a flowchart showing a processing procedure when calculating a current compressor driving torque in the third embodiment.

Explanation of symbols

101 ... Refrigeration cycle 102 ... External capacity variable compressor 102a ... ECV
DESCRIPTION OF SYMBOLS 103 ... Gas cooler 104 ... Internal heat exchanger 105 ... Evaporator 106 ... Accumulator 107 ... Engine 109 ... Control apparatus 110 ... Orifice 111 ... Air conditioning duct 112 ... Air conditioning fan 113 ... Compressor suction temperature sensor 114 ... Compressor suction pressure sensor 115 ... Temperature sensor 116 ... Rotational speed detection sensor 117 ... Compressor discharge temperature sensor

Claims (21)

A compressor (102) driven by the engine (107) and configured to control the discharge refrigerant capacity by an external electric signal, a differential pressure control type control valve (102a) for adjusting the capacity of the compressor (102), A radiator (103) that radiates the refrigerant compressed by the compressor (102), a decompression unit (110) that decompresses the refrigerant radiated by the radiator (103), and a decompression unit (110) An evaporator (105) for evaporating the refrigerant, an engine rotation speed means (116) for detecting the rotation speed of the engine (107), and an intake refrigerant temperature detection means (113) for detecting the intake refrigerant temperature of the compressor (102). ), Suction pressure detection means (114) for detecting the suction pressure of the compressor (102), and detection values from the detection means. There are a vehicle air-conditioning apparatus provided with a control device (109) for controlling the discharge refrigerant volume of the compressor (102),
The control device (109) includes an intake refrigerant temperature detected by the intake refrigerant temperature detection means (113), an intake pressure detected by the intake pressure detection means (114), and a differential pressure control type control valve ( The current compressor driving torque is calculated using the high-low pressure difference obtained from the input duty value of 102a), and the duty value given to the differential pressure control type control valve (102a) according to the current compressor driving torque and the torque request value The current compressor driving torque is controlled by adjusting the air conditioner. A vehicle air conditioner comprising discharge temperature detection means (117) for detecting the discharge temperature of the refrigerant supplied from the compressor (102) instead of the suction refrigerant temperature detection means (113),
The control device (109) includes the discharge temperature detected by the discharge temperature detecting means (117), the suction pressure detected by the suction pressure detecting means (114), and the differential pressure control type control valve (102a). The current compressor driving torque is calculated using the high pressure-low pressure pressure difference obtained from the input duty value, and the duty value given to the differential pressure control type control valve (102a) is adjusted according to the current compressor driving torque and the torque request value. The vehicle air conditioner according to claim 1, wherein the current compressor driving torque is controlled. In addition to the suction refrigerant temperature detection means (113) and the suction pressure detection means (114), the vehicle further includes a discharge temperature detection means (117) for detecting the discharge temperature of the refrigerant supplied from the compressor (102). Air conditioner for
The control device (109) includes the suction refrigerant temperature detected by the suction refrigerant temperature detection means (113), the suction pressure detected by the suction pressure detection means (114), and the discharge temperature detection means (117). The current compressor driving torque is calculated by calculating the current compressor driving torque using the high pressure-low pressure pressure difference obtained from the discharge temperature detected in step 5 and the input duty value of the differential pressure control type control valve (102a). The vehicle air conditioner according to claim 1, wherein the current compressor driving torque is controlled by adjusting a duty value given to the differential pressure control type control valve (102a) according to a required value. In place of the suction pressure detection means (114), a blowout temperature detection means (115) for detecting a blowout temperature downstream of the evaporator (105) is provided.
The said control apparatus (109) calculates the suction pressure of the said compressor (102) based on the blowing temperature detected by the said blowing temperature detection means (115). The vehicle air conditioner according to item.   In the torque control of the compressor (102), the control device (109) calculates a torque difference between the required compressor driving torque and the compressor driving torque obtained by the calculation, and the torque difference is set to a setting allowable width X (%). The vehicle air conditioner according to any one of claims 1 to 4, wherein a duty value applied to the differential pressure control type control valve (102a) is controlled so as to be within a range of (5).   The differential pressure control type control valve (102a) has a characteristic capable of controlling a high pressure-low pressure differential pressure with respect to an input duty value. When the input duty value is low, the differential pressure is small and the input duty value is high. The vehicle air conditioner according to any one of claims 1 to 5, wherein the differential pressure sometimes increases.   The vehicle air conditioner according to any one of claims 1 to 6, wherein the differential pressure control type control valve (102a) has a characteristic that a maximum control pressure is provided in a region of a specified duty value or more. .   8. The control device according to claim 1, wherein the control device calculates a high-low pressure differential pressure with respect to an input duty value of the differential pressure control type control valve. The vehicle air conditioner described.   The vehicle according to any one of claims 1 to 8, wherein the control device (109) calculates an evaporation temperature of the evaporator (105) based on a suction pressure of the compressor (102). Air conditioner.   The control device (109) calculates the compressor (102) from the high-low pressure differential pressure calculated based on the suction pressure of the compressor (102) and the input duty value of the differential pressure control type control valve (102a). The vehicle air conditioner according to any one of claims 1 to 9, wherein the high-pressure pressure is calculated.   11. The vehicle air conditioner according to claim 10, wherein the control device (109) controls the calculated high pressure so as not to exceed a predetermined normal maximum pressure.   When the calculated high pressure exceeds a predetermined maximum allowable pressure, the control device (109) uses the calculation of claim 5 to reduce the input duty value and temporarily control to the minimum differential pressure. The vehicle air conditioner according to claim 10.   The control device (109) calculates the refrigerant circulation volume flow rate of the compressor (102) using a correlation equation or a correlation map between the pressure difference between the pressure reduction means (110) and the refrigerant circulation volume flow rate. The vehicle air conditioner according to any one of claims 1 to 12.   The vehicle air conditioner according to any one of claims 1 to 13, wherein the control device (109) calculates an intake refrigerant density from an intake temperature and an intake pressure of the compressor (102).   The control device (109) calculates the refrigerant circulation weight flow rate using the refrigerant circulation volume flow rate calculated in claim 13 and the suction refrigerant density calculated in claim 14. The vehicle air conditioner described.   The control device (109) particularly calculates the compressor power consumption using the refrigerant circulation weight flow rate calculated in claim 15 and the compressor charging enthalpy calculated from the suction refrigerant enthalpy and the discharge refrigerant enthalpy. Item 15. The vehicle air conditioner according to Item 15.   The control device (109) calculates the actual rotational speed of the compressor (102) calculated from the compressor consumption power calculated in claim 16 and the rotational speed of the engine (107) detected by the engine rotational speed means (116). The vehicle air conditioner according to claim 16, wherein the current compressor driving torque is calculated using   The vehicle air conditioner according to any one of claims 1 to 17, wherein an internal release mechanism is provided in the decompression means (110).   The pressure control by the control device (109), the ΔPecv characteristic of the differential pressure control type control valve (102a), the ΔPecvmax characteristic, and the internal relief mechanism of the pressure reducing means (110) prevent an increase in high pressure. The vehicle air conditioner according to any one of claims 1 to 18.   The control device (109) includes an intake pressure detected by the intake pressure detection means (114), an intake refrigerant temperature detected by the intake refrigerant temperature detection means (113), and a refrigerant density relational expression in a carbon dioxide Mollier diagram. The air conditioner for a vehicle according to any one of claims 1 to 19, wherein a compressor inlet superheat degree is calculated using The control device (109) calculates the compressor discharge temperature using the high pressure calculated in claim 10, the compressor inlet superheat calculated in claim 20, and the refrigerant density relational expression of the carbon dioxide Mollier diagram. The vehicle air conditioner according to claim 20, wherein a duty value applied to the differential pressure control type control valve (102a) is controlled so that the compressor discharge temperature falls within a range of a set allowable allowable discharge temperature. .

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