EP3922932B1 - Procédé de fonctionnement d'une installation de réfrigération à compression et installation de réfrigération à compression - Google Patents

Procédé de fonctionnement d'une installation de réfrigération à compression et installation de réfrigération à compression Download PDF

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Publication number
EP3922932B1
EP3922932B1 EP21177576.2A EP21177576A EP3922932B1 EP 3922932 B1 EP3922932 B1 EP 3922932B1 EP 21177576 A EP21177576 A EP 21177576A EP 3922932 B1 EP3922932 B1 EP 3922932B1
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EP
European Patent Office
Prior art keywords
refrigerant
temperature
compressor
superheating
evaporator
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EP21177576.2A
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German (de)
English (en)
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EP3922932A1 (fr
Inventor
Silvia Hildebrandt
Martin Herrs
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Stiebel Eltron GmbH and Co KG
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Stiebel Eltron GmbH and Co KG
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Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B49/00—Arrangement or mounting of control or safety devices
    • F25B49/02—Arrangement or mounting of control or safety devices for compression type machines, plants or systems
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B40/00—Subcoolers, desuperheaters or superheaters
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2500/00—Problems to be solved
    • F25B2500/22—Preventing, detecting or repairing leaks of refrigeration fluids
    • F25B2500/222—Detecting refrigerant leaks
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2600/00—Control issues
    • F25B2600/21—Refrigerant outlet evaporator temperature
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2600/00—Control issues
    • F25B2600/25—Control of valves
    • F25B2600/2513—Expansion valves
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2700/00—Sensing or detecting of parameters; Sensors therefor
    • F25B2700/19—Pressures
    • F25B2700/193—Pressures of the compressor
    • F25B2700/1933—Suction pressures
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2700/00—Sensing or detecting of parameters; Sensors therefor
    • F25B2700/21—Temperatures
    • F25B2700/2115—Temperatures of a compressor or the drive means therefor
    • F25B2700/21151—Temperatures of a compressor or the drive means therefor at the suction side of the compressor
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2700/00—Sensing or detecting of parameters; Sensors therefor
    • F25B2700/21—Temperatures
    • F25B2700/2117—Temperatures of an evaporator
    • F25B2700/21175—Temperatures of an evaporator of the refrigerant at the outlet of the evaporator

Definitions

  • the invention relates to a method for operating a compression refrigeration system and an associated compression refrigeration system with a refrigerant, an evaporator, a compressor, a condenser, a throttle element and a control unit.
  • Such compression refrigeration systems for example in the form of heat pumps, with a vapor compression system in which a gaseous refrigerant is compressed from a low pressure to a high pressure by a compressor controlled by the control unit, which has a regulator, for example, are known.
  • the refrigerant is driven through the condenser, in which it gives off thermal heat to a heating medium located in a heat sink system.
  • Internal heat is transferred in an optional internal heat exchanger, for example in the form of a recuperator, between the high-pressure refrigerant flowing from the condenser to the expansion valve and the low-pressure refrigerant flowing from the evaporator to the compressor.
  • the refrigerant is further conducted in a high-pressure flow direction to an expansion valve controlled by the regulator, in which the refrigerant is expanded from the high pressure to the low pressure depending on a control value.
  • the refrigerant at the low pressure evaporates in the evaporator when absorbing heat from the source.
  • WO 2019/124409 A1 relates to a refrigeration cycle device capable of performing a refrigeration cycle using a refrigerant with low global warming potential (GWP: global warming potential), the refrigeration cycle device comprising a refrigeration cycle and a refrigerant included in the refrigeration cycle, the refrigeration cycle including a compressor , a capacitor, a Includes decompression section and an evaporator and the refrigerant at least
  • GWP global warming potential
  • document WO 2019/124409 A1 discloses a method for operating a compression refrigeration system according to the preamble of claim 1 and a compression refrigeration system according to the preamble of claim 6.
  • EP 1 014 013 A1 discloses a vapor compression refrigeration cycle, wherein in a vapor compression type refrigeration cycle using carbon dioxide as a refrigerant, a superheat control valve is connected between an evaporator and an internal heat exchanger, the superheat control valve for adjusting the flow rate of a liquid-phase portion of the refrigerant supplied to the evaporator in accordance with a control signal for maintaining a superheat level of a gas phase portion of refrigerant supplied to a compressor.
  • Thermistors sense temperatures of the gas phase portion and a vaporized portion of refrigerant flowing out of the evaporator to generate first and second temperature signals, respectively.
  • a temperature/pressure sensor detects a condition of a radiated portion of refrigerant that has flowed out of a radiator to generate a refrigerant condition value.
  • a selector selects as a selected signal one of the first and second temperature signals in accordance with the coolant condition value to provide the selected signal as the control signal to the superheat control valve.
  • document EP 1 026 459 A1 discloses a vapor compression type refrigeration system including a first refrigerant temperature detector attached to a pipe for detecting the temperature of refrigerant on the side of an outlet of an evaporator.
  • a second refrigerant temperature detector is attached to a pipe to detect the temperature of refrigerant on an inlet side of a compressor.
  • a switching controller is connected to the first and second coolant temperature detectors and an overheat control valve and changes/selects one of the coolant temperature detection value signals from the first and second coolant temperature detectors according to predetermined conditions.
  • a superheat control valve adjusts the flow rate of the refrigerant flowing into the evaporator so that the refrigerant superheat on the inlet side of the compressor reaches a predetermined value.
  • Out of DE 101 59 892 A1 is known in a refrigerating machine, in particular in a heat pump, to use a recuperator, with which the heat output is to be increased in a structurally simple manner at low outside temperatures.
  • the recuperator is dimensioned in such a way that at low evaporation temperatures it at least approximately 15% of the heating capacity of the heat pump is transferred from the liquid refrigerant to the gaseous refrigerant.
  • An injection valve injects liquid refrigerant into the compressor so that the discharge temperature remains below 120 °C.
  • a heat pump system with a refrigerant circuit is off DE 10 2005 061 480 B3 known. It is equipped with a compressor, a first heat exchanger, a throttle element, an evaporator and a 4-2-way valve unit for switching between a first (heating) and a second operating mode (cooling).
  • a flow direction of the refrigerant in the refrigerant circuit can be switched such that the first heat exchanger is used in the first operating mode to liquefy the refrigerant and in the second operating mode to evaporate the refrigerant, and the second heat exchanger in the first operating mode to evaporate the refrigerant and is used in the second mode of operation for condensing the refrigerant, the first heat exchanger in the refrigerant circuit being connected in such a way that it operates as a counterflow heat exchanger in the two operating modes of heating and cooling.
  • the control of the compression refrigeration system must meet various requirements, for example, it is required that a coefficient of performance is as high as possible in order to allow the most energy-efficient operation possible. However, it is also important that the operating limits of the components are observed.
  • the COP is largely dependent on the right amount of refrigerant circulating through the refrigeration cycle. Excess refrigerant can be collected in a refrigerant receiver, for example, so that the right amount of refrigerant circulates. If there is not enough refrigerant that can circulate through the refrigeration circuit, the efficiency of the compression refrigeration system is poor. Such a lack of refrigerant can have various causes that require different responses.
  • a method for operating a compression refrigeration system is proposed with a refrigerant that has a temperature glide, an evaporator, a compressor, a condenser, a throttling element, an internal heat exchanger, with the internal heat exchanger for transferring thermal energy of the refrigerant before it enters the throttling element the refrigerant is formed before entering the compressor, and a control unit, wherein the control unit is formed a) for detecting overheating of the refrigerant after exiting the evaporator and for detecting overheating of the refrigerant upon entry into the compressor and b) for controlling the Throttle element is formed, the regulation of the throttle element using a control variable based on both overheating and c) for the detection of a lack of refrigerant is formed at least in the condenser.
  • the method includes calculating a temperature difference between the dew point temperature and the boiling point temperature, called temperature glide, of the refrigerant at the current operating point of the compression refrigeration system on the basis of a detected low pressure of the refrigerant.
  • the temperature glide depends on the current pressure to which it is related. According to the invention, that in the low-pressure path of the refrigerant is relevant.
  • the temperature glide can be determined using a characteristic curve known for the refrigerant used. Refrigerants that exhibit a temperature glide are usually mixtures of two or more refrigerant components. In this case, a correction of the known characteristic due to deviations in the mixing ratio is particularly preferred.
  • the method further includes calculating a temperature difference threshold relative to the calculated temperature glide of the sensed low pressure of the refrigerant at the current operating point.
  • the method also includes an evaluation of the current evaporator outlet overheating for falling below the temperature difference limit value.
  • the temperature glide in typical working areas for typical refrigerants such as R454C can be in a range from 5K to 15K, in particular between 7 and 8K.
  • the refrigerant leaves the evaporator with an evaporator leaving superheat of -2 or -2.5 K, that is, the temperature of the refrigerant at the evaporator outlet is lower than the temperature at which evaporation is complete.
  • the temperature difference limit value can now specify that the prerequisite for the existence of a lack of refrigerant is met if the refrigerant exits the evaporator at a lower temperature than -4 or -5 K after the end of evaporation.
  • the method further includes providing a target value of a superheat, which is defined as a difference between the temperature of the refrigerant and the dew point temperature, upon entry into the compressor, called the compressor entry superheat.
  • a target value of a superheat which is defined as a difference between the temperature of the refrigerant and the dew point temperature
  • the method also includes evaluating the current superheating of the compressor inlet to determine whether it exceeds the target value for superheating of the compressor inlet,
  • the method includes recognizing a lack of refrigerant situation if both results of the steps of evaluating are continuously fulfilled for a specified time.
  • the method according to the invention therefore makes it possible to reliably detect an existing refrigerant shortage situation in compression refrigeration systems with an internal heat exchanger based on two overheating values of the refrigerant in the low-pressure path, namely when it exits the evaporator and when it enters the compressor.
  • the set time is preferably at least one minute, more preferably 10 minutes. This gives the control loop the opportunity to stably return to the desired state, which takes a certain amount of time.
  • the refrigerant preferably has a temperature glide, the refrigerant particularly having or consisting of R454C, and the compression refrigeration system containing in particular an internal heat exchanger for transferring heat energy from the refrigerant before it enters the throttle element to the refrigerant before it enters the compressor.
  • R454C refrigerant particularly having or consisting of R454C
  • the compression refrigeration system containing in particular an internal heat exchanger for transferring heat energy from the refrigerant before it enters the throttle element to the refrigerant before it enters the compressor. This is particularly relevant because powerful internal heat exchangers are regularly used in refrigeration circuits with R454C.
  • the method preferably also has the following step: attempting to return refrigerant to the condenser by increasing the setpoint value for the compressor speed for a limited time and/or increasing the setpoint value for the evaporator outlet overheating for a limited time.
  • an attempt is made to resolve the detected lack of refrigerant.
  • the first mechanism there is insufficient refrigerant in the system due to leaks, leaks, etc. This situation can only be remedied by manual intervention, ie refilling the refrigeration circuit with the required amount of refrigerant and sealing the refrigeration circuit.
  • the second situation of a lack of refrigerant in the condenser is caused by refrigerant that has condensed out in the components of the refrigeration cycle.
  • refrigerant can fail, for example at low-lying positions in the evaporator, and cannot escape from the evaporator together with the gaseous refrigerant due to low flow speeds and gravitation. In other words, the refrigerant remains trapped in the component and is not available for the cycle. It has been found that the risk of refrigerant accumulating in components of the refrigeration cycle increases, particularly in the case of evaporators with a large surface area and operating points with low performance requirements, for example low compressor speeds. In such situations where there is a shortage of refrigerant, there is the possibility of triggering a successful return of the refrigerant by shifting the operating point.
  • the method preferably also has the following step: generation of an error state in the event of an unsuccessful attempt to return the refrigerant, in particular in the event of multiple unsuccessful attempts to return the refrigerant, in a specified time period.
  • the fault status can preferably be displayed on a housing of the compression refrigeration system, for example on a display.
  • the error status together with further diagnostic information can be transmitted to a server and/or displayed in an app.
  • the diagnostic information can assist in a repair.
  • a fixed amount of time and value relative to the temperature glide may be set that indicates that significantly under-energy is being transferred in the evaporator, indicating refrigerant migration.
  • a regular overheating value is assumed at various operating points, which can also be negative, in particular based on known, model-based pre-control characteristics. Based on this regular overheating value, it can be determined according to the invention that a limit for a lack of refrigerant is always reached when a distance therefrom reaches a certain threshold value, the temperature difference limit value, for example 1 K or 2 K, particularly preferably 1.5 K.
  • the temperature difference limit value is preferably calculated relative to the calculated temperature glide of the detected low pressure of the refrigerant as a function of the current operating point of the compression refrigeration system.
  • the temperature difference limit is set relative to the calculated temperature glide proportional to a difference between heat sink temperature and heat source temperature.
  • the temperature difference limit value can therefore also be corrected based on a model, with which the diagnosis of a lack of refrigerant can be determined even more precisely as a function of the operating situation.
  • Overheating reduction in negative ranges that takes place in regular operation - which must be ruled out at the latest when entering the compressor - is possible through the use of a recuperator and comes close to areas in which a lack of refrigerant is detected.
  • the temperature difference limit value dependent on the operating point of the evaporator, i.e. the temperature constellation of the medium temperatures, in addition to the sliding, which is a physical property of the refrigerant. If there is a small driving temperature difference, only a small negative superheat is expected, so the temperature difference limit value is close to the small negative overheating value can be used. For a large driving temperature difference, for example in the case of high flow temperatures and/or low heat source temperatures, it is therefore advantageous to implement a larger temperature difference limit value
  • a compression refrigeration system with a refrigerant that has a temperature glide, an evaporator, a compressor, a condenser, a throttle element, an internal heat exchanger, the internal heat exchanger for transferring heat energy of the refrigerant in the high-pressure path before it enters the throttle element is formed on the refrigerant in the low-pressure path before entering the compressor, and solved a control unit.
  • the control unit is designed a) to detect overheating of the refrigerant after exiting the evaporator and to detect overheating of the refrigerant upon entry into the compressor b) to regulate the throttle element using a controlled variable based on both overheatings and c) to detect a lack of refrigerant at least in the condenser.
  • the control unit is also designed to: Calculate a temperature difference between the dew point temperature and the boiling point temperature, known as the temperature glide, of the refrigerant at the current operating point of the compression refrigeration system on the basis of a detected low pressure of the refrigerant, Calculate a temperature difference limit value relative to the calculated temperature glide of the detected low pressure of the refrigerant in the current operating point, evaluation of the current evaporator outlet overheating for falling below the temperature difference limit value, provision of a target value for overheating, which is defined as a difference between the temperature of the refrigerant and the dew point temperature, upon entry into the compressor, called compressor inlet overheating, evaluation of the current compressor inlet overheating for exceeding the target value compressor inlet overheating, recognizing a refrigerant shortage situation if both results of the steps of evaluating are continuously fulfilled for a specified time.
  • the object is also achieved according to the invention by a heat pump with a compression refrigeration system according to the invention.
  • the compression refrigeration system according to the invention is suitable regardless of the type of heat pump, for example air/water, brine/water heat pumps, and regardless of the place of installation.
  • a data connection 510 which can be via cable, radio or other technologies: compressor 210, heating medium pump 410, brine pump 330, expansion valve 230, compressor inlet temperature sensor 501, low pressure sensor 502, high pressure sensor 503, hot gas temperature sensor 504, recuperator inlet temperature sensor 505, recuperator outlet temperature sensor 506 and/or evaporator outlet temperature sensor 508. Additionally or alternatively, a 1 not shown evaporator inlet temperature sensor determine the temperature at the evaporator inlet 241.
  • the heat pump 100 is shown as a brine heat pump.
  • a ventilator/fan is arranged as the heat source instead of the brine circuit with brine pump 330 .
  • Compressor 210 is used to compress the superheated refrigerant from an inlet connection 211 to a compressor outlet pressure PVa at a compressor outlet temperature corresponding to the hot gas temperature at compressor outlet 212.
  • Compressor 210 usually contains a drive unit with an electric motor, a compression unit and the electric motor can advantageously be operated at variable speeds.
  • the compression unit can be designed as a rolling piston unit, scroll unit or otherwise.
  • the compressed superheated refrigerant at the compressor outlet pressure P Va is at a higher pressure level, in particular a high pressure HD, than at the inlet port 211 with a compressor inlet pressure P Ve , in particular a low pressure ND, at a compressed inlet temperature T VE , which determines the state of the refrigerant temperature at the Describes entry port 211 upon entry into a compression chamber.
  • thermal energy Q H is transferred from the refrigerant of vapor compression system 200 to a heating medium of heat sink system 400.
  • the refrigerant is heated in condenser 220, with superheated refrigerant vapor transferring part of its thermal energy to the heating medium of heat sink system 400 by reducing the temperature .
  • a further heat transfer Q H advantageously takes place in the condenser 220 by condensation of the refrigerant during the phase transition from the gas phase of the refrigerant to the liquid phase of the refrigerant.
  • further heat Q H is transferred from the refrigerant from the vapor compression system 200 to the heating medium of the heat sink system 400 .
  • the high pressure HD of the refrigerant that occurs in the condenser 220 corresponds approximately to a condensation pressure of the refrigerant at a heating medium temperature T WS in the heat sink system when the compressor 210 is in operation.
  • the heating medium in particular water, is conveyed by means of a heating medium pump 410 through the heat sink system 400 in a direction SW through the condenser 220, the thermal energy Q H being transferred from the refrigerant to the heating medium.
  • exiting refrigerant from the condenser 220 is stored, which, depending on the operating point of the vapor compression circuit 200 , should not be fed into the circulating refrigerant. If more refrigerant is fed in from the condenser 220 than is passed on through the expansion valve 230, the accumulator 260 fills up, otherwise it becomes emptier or emptied.
  • downstream recuperator 250 which can also be referred to as an internal heat exchanger
  • internal heat energy Q i is transferred from the high-pressure HD refrigerant, which flows from the condenser 220 to the expansion valve 230 in a high-pressure flow direction S HD , to the low-pressure ND flow Transfer refrigerant flowing from the evaporator to the compressor in a low pressure flow direction S ND transferred.
  • the refrigerant flowing from the condenser to the expansion valve 230 is advantageously supercooled.
  • the refrigerant flows through an expansion valve inlet 231 into the expansion valve.
  • the refrigerant pressure is throttled from the high pressure HD to the low pressure ND, in that the refrigerant advantageously passes through a nozzle arrangement or throttle with an advantageously variable opening cross section, with the low pressure advantageously approximately corresponding to a suction pressure of the compressor 210.
  • any other pressure-reducing device can also be used. Pressure reducing pipes, turbines or other expansion devices are advantageous.
  • a degree of opening of the expansion valve 230 is set by an electric motor, which is usually implemented as a stepping motor, which is controlled by the control unit or regulator 500 .
  • the low pressure ND at the expansion valve outlet 232 of the refrigerant from the expansion valve 230 is controlled in such a way that the low pressure ND of the refrigerant that occurs during operation of the compressor 210 corresponds approximately to the evaporation pressure of the refrigerant with the heat source medium temperature T WQ .
  • the evaporation temperature of the refrigerant will be a few Kelvin below the heat source medium temperature T WQ so that the temperature difference drives heat transfer.
  • evaporation heat energy Q V is transferred from the heat source fluid of the heat source system 300, which can be a brine system, a geothermal system for using heat energy Q Q from the ground, an air system for using energy Q Q from the ambient air or another heat source that Source energy Q Q to the vapor compression system 200 outputs.
  • the heat source fluid of the heat source system 300 can be a brine system, a geothermal system for using heat energy Q Q from the ground, an air system for using energy Q Q from the ambient air or another heat source that Source energy Q Q to the vapor compression system 200 outputs.
  • the coolant flowing into the evaporator 240 reduces its proportion of wet vapor as it flows through the evaporator 240 by absorbing heat Q Q and leaves the evaporator 240 advantageously with a low proportion of wet vapor or advantageously also as a superheated gaseous refrigerant.
  • the heat source medium is conveyed through the heat source medium path of the evaporator 240 by means of a brine pump 330 for brine/water heat pumps or an outside air fan for air/water heat pumps, with the thermal energy Q Q being extracted from the heat source medium as it flows through the evaporator.
  • recuperator 250 thermal energy Q i is transferred between the refrigerant flowing from condenser 220 to expansion valve 230 to the refrigerant flowing from evaporator 240 to compressor 210, with the refrigerant flowing from evaporator 240 to compressor 210 in particular continuing to overheat.
  • This overheated refrigerant which exits the recuperator 250 at an overheating temperature T Ke , is conducted to the refrigerant inlet connection 211 of the compressor 210 .
  • the recuperator 250 is used in the vapor compression circuit 200 in order to increase the overall efficiency as the quotient of the heat output Q H delivered and the electrical power P e consumed to drive the compressor motor.
  • the coolant which in the condenser 220 gives off thermal energy Q H at a temperature level on the heat sink side to the heating medium, is withdrawn in the high-pressure path of the recuperator 250 by supercooling further thermal energy Q i .
  • the internal energy state of the refrigerant when it enters the evaporator 240 is reduced by this heat extraction Q i , so that the refrigerant can absorb more heat energy Q Q from the heat source 300 at the same evaporation temperature level.
  • the refrigerant is then returned to the refrigerant in the low-pressure path at low pressure ND and at a low-pressure temperature T Va corresponding to an evaporator outlet temperature in the recuperator 250 .
  • the supply of energy advantageously has the effect of reducing the proportion of wet steam to a state without a proportion of wet steam, and then overheating occurs as a result of further supply of energy.
  • the following sensors are advantageously arranged to detect the operating state of the vapor compression system 200, with which, in particular, for safeguarding and optimization of the operating conditions of the vapor compression system 200, in particular when there are changes in the operating state, a model-based pilot control is implemented.
  • the process variable that has a significant influence on the overall efficiency of the vapor compression circuit 200 as a quotient between the heat output Q H transmitted by the vapor compression circuit 200 and an electrical power P e consumed by the compressor 210 is the overheating of the refrigerant at the compressor inlet 211.
  • the overheating of the refrigerant at the compressor inlet 211 is advantageously observed, however, with restrictions regarding the permitted overheating range of the refrigerant at the compressor inlet.
  • Overheating describes the temperature difference between the recorded compressor inlet temperature T KE of the refrigerant and the evaporation temperature of the refrigerant with saturated steam.
  • the superheating of the compressor inlet is preferably controlled in such a way that no condensate forms on components of the refrigeration circuit, particularly in the section between the refrigerant outlet of the recuperator 252 and the compressor inlet 211, due to the water vapor content in the ambient air falling below the dew point.
  • the refrigeration cycle section between the evaporator outlet 242 and the recuperator inlet 251 is usually colder because it is typically only a short pipe section, better insulation is possible compared to the section between the refrigerant outlet of the recuperator 252 and the compressor inlet 211 .
  • the refrigerant separator which is to be protected, is located at the point of the compressor inlet 211 on the compressor.
  • the temperature here should be kept high enough that nothing condenses.
  • the problem of condensation occurs on the high-pressure side usually not on.
  • the passage between the recuperator outlet 252 on the high-pressure side and the entry into the expansion valve 231 also regularly cools down to the temperature level of the refrigerant at the evaporator outlet 242 depending on the operating point under ideal heat transfer conditions in the recuperator 250 .
  • this passage is also typically short and can be isolated very well, this section is usually not problematic either.
  • the method according to the invention can in principle prevent a condensate drop over the entire circuit of the heat pump.
  • a room temperature sensor and a room humidity sensor which enable an exact determination of the condensing conditions of the air, e.g. is 21°C and 60% rel. Humidity the condensation temperature in the range of 13°C. Under these conditions, no condensation takes place as long as the pipe temperature is above 13°C plus a buffer if necessary, e.g. 1K.
  • Limit values in particular for overheating, determine the permissible overheating range of the components at the compressor inlet 211 depending on the operating point.
  • the compressor inlet overheating dTü ⁇ and the overall efficiency of the vapor compression circuit 200 or between the compressor inlet overheating dTü ⁇ and a stability S of a control value R, which is advantageous when correcting the compressor inlet overheating.
  • the heat source medium temperature, the heating medium temperature, the compressor output P e and target values Z or the target value Z used for a calculation of the compressor inlet overheating dT ÜE can be calculated as the default value for the compressor inlet overheating dT ÜE from the cooling circuit measured variables that depend on the operating point, such as heat source medium temperature, heating medium temperature, compressor output P e and parameterizable coefficients, i.e. coefficients that are adapted to the behavior of the respective cooling circuit components.
  • the target value for the compressor inlet overheating dT ÜE is constant regardless of all operating conditions, eg 10 Kelvin.
  • it is varied as a function of an operating point variable, for example the compressor output P e , or in the case of an even more complex adaptation, it varies as a function of a number of operating point variables.
  • the total control deviation, which is fed in to control the vapor compression circuit 200, is then advantageously calculated from the weighted influence of the control deviation of the compressor inlet overheating dT ÜE and the weighted influence of the control deviation of the evaporator outlet overheating DT ÜA in the controller 500.
  • the refrigerant passes through two sequentially arranged heat exchangers, the evaporator 240 and the recuperator 250, in which heat energy Q Q and Q i is supplied to the refrigerant.
  • source heat energy Q Q from the heat source system 300 is supplied to the refrigerant.
  • the temperature level of the source heat supplied Q Q is at a temperature level of the heat source, in particular such as the ground or the outside air.
  • thermal energy Q i is withdrawn from the refrigerant after leaving the condenser 220 .
  • the temperature level of the refrigerant at the outlet of the condenser is approximately the same as the return flow temperature of the heating medium.
  • the control value R is advantageously the weighted combination of the control deviation of the compressor inlet overheating dT ÜE with the control deviation of the evaporator outlet overheating.
  • Actuator operating state variables with an influence on the control value R, in particular the compressor inlet overheating dT ÜE , are the compressor speed and/or the degree of opening of the expansion valve 230 in the vapor compression circuit 200 in question, with which the low pressure ND and the evaporation temperature level are also advantageously determined.
  • Actuators have a particularly advantageous influence on the control value R, in particular on the weighted linkage of the control deviation of the compressor inlet overheating with the control deviation of the evaporator outlet overheating.
  • the compressor 210 by varying the compressor speed and the expansion valve 230 by influencing the degree of opening are such actuators. These two actuators influence the low pressure ND and the evaporation temperature level.
  • a change in the compressor speed to regulate the desired heating output without further compensatory changes in the opening degree of the expansion valve advantageously changes the control value R into undesirable ranges, so that a model-based, supported change in the opening degree of the expansion valve associated with the compressor speed change is advantageous for regulating R, and may even be necessary.
  • the compressor speed is set in the vapor compression circuit 200 in such a way that the heat output QH transmitted from the vapor compression circuit 200 to the heating medium corresponds to the requested target value Z.
  • influencing the compressor speed for controlling the compressor inlet overheating dT ÜE is advantageously secondary or not appropriate.
  • the degree of opening of the expansion valve 230 is advantageously used as a control value for controlling the superheating of the compressor inlet dT ÜE .
  • the influence of the degree of opening of the expansion valve 230 on the compressor inlet overheating dT ÜE is as follows:
  • the expansion valve 230 acts as a nozzle with a nozzle cross-section that can be adjusted by an electric motor, in which a needle-shaped nozzle needle is usually threaded into a nozzle seat by means of a stepping motor.
  • the refrigerant flow rate through the expansion valve is approximately proportional to the square root of the pressure difference between the expansion valve inlet 231 and outlet 232 multiplied by a current relative value of the nozzle cross section or degree of opening and advantageously one of the refrigerant - and a geometry of the expansion valve 230 dependent constant.
  • the degree of opening of the expansion valve 230 only significantly influences the low pressure ND, i.e. the outlet pressure from the expansion valve 230.
  • the low pressure ND on the low-pressure side of the vapor compression circuit 200 then falls Compressor 210, since its capacity can be approximately described as volume / time, due in particular to the piston strokes, and a correspondingly reduced low-pressure value ND is set, at which the refrigerant mass flow supplied through expansion valve 230 is equal to the refrigerant mass flow discharged from compressor 210.
  • the degree of opening of the expansion valve 230 is increased, more refrigerant passes through the expansion valve 230 at a constant high pressure HD and initially a constant low pressure LP.
  • the compressor 210 initially continues to deliver the same refrigerant mass flow, the low-pressure side ND of the refrigeration circuit receives more refrigerant through the expansion valve 230 supplied than is sucked out by the compressor 210 . Since the refrigerant vapor is a compressible medium, the low pressure ND increases on the low-pressure side of the vapor compression circuit 200.
  • the mass flow capacity of the compressor 210 increases approximately proportionally, since its capacity can be described approximately as volume / time, and it a correspondingly increased low pressure ND sets in, at which the refrigerant mass flow supplied through the expansion valve 230 is equal to the refrigerant mass flow discharged from the compressor 210 .
  • the low pressure LP in turn significantly influences the heat transfer between the heat source medium and refrigerant in the evaporator 240.
  • the heat flow Q Q from the heat source system 300 is transferred between the heat source medium and the refrigerant at different temperatures, with the heat flow Q Q depending on the temperature difference between the heat source medium and the refrigerant and the heat transfer resistance of a heat transfer layer of the evaporator 240 .
  • the heat transfer resistance between the heat source medium path of the evaporator and the refrigerant path of the evaporator can be assumed to be approximately constant in a respective vapor compression circuit 200 .
  • the size of the heat transfer performance in the evaporator 240 is therefore largely dependent on the integral of the temperature differences of all surface elements of the heat transfer layer.
  • a refrigerant temperature is established which, due to the saturated vapor characteristic curve as a material property of the refrigerant, is a function of the low pressure ND of the refrigerant.
  • the evaporation temperature of the refrigerant as it flows through the recuperator 250 can be controlled indirectly.
  • the thermal energy Q Q which is transferred from the heat source system to the refrigerant flowing through the evaporator 240, causes an influencing of the state of aggregation of the refrigerant.
  • recuperator 250 For complete evaporation, additional energy is supplied in the recuperator 250 in order to superheat the refrigerant beyond the state of saturated vapor.
  • a corresponding refrigerant state is set when it exits the evaporator 240 as a function of the manipulated variable “degree of opening of the expansion valve 230”.
  • the controlled system behavior is characterized in particular by the controlled system output value for the evaporator outlet overheating as a function of the controlled system input value for the degree of opening of the expansion valve.
  • a refrigerant is used, in particular a refrigerant mixture as the refrigerant, which has a "temperature glide", in particular R454C is advantageously used.
  • a refrigerant mixture with a temperature glide with a relative change in the degree of opening of the expansion valve actuator of 1% rel. at the exit of the refrigerant from the evaporator, a superheating change of about less than 1 K is usually set.
  • the refrigerant After flowing through the evaporator 240 , the refrigerant enters the low-pressure path of the recuperator 250 at low pressure LP.
  • the state of aggregation of the coolant when it flows into the recuperator 250 is, in a normal operating case, ie advantageously either saturated steam with a low steam content of between 0 and 20% or, in particular, also advantageously already overheated coolant.
  • a refrigerant temperature sets in which is a function of the refrigerant pressure due to the saturation vapor characteristic of the refrigerant.
  • the refrigerant temperature will at most assume a size that corresponds to the inlet temperature of the heat source medium.
  • the variable preferably corresponds to the inlet temperature of the refrigerant in the high-pressure path of recuperator 250, i.e. the temperature of the refrigerant after it has exited condenser 220.
  • recuperator 250 In order to be able to transfer a sufficient amount of thermal energy from the refrigerant of the high-pressure side refrigerant path to the refrigerant of the low-pressure side refrigerant path in recuperator 250, it must be ensured that the temperature of the refrigerant of the high-pressure side refrigerant path at high pressure HD is greater than in as many surface elements of the transfer layer of recuperator 250 as possible is the temperature of the refrigerant of the low-pressure side refrigerant path at low pressure ND at the respective surface element.
  • the corresponding temperatures of the heating system 400 of the vapor compression system 200 are higher than the corresponding temperatures of the heat source such as the ground or the outside air.
  • the thermal energy Q i which is transferred from the refrigerant at high pressure HD of the high-pressure-side refrigerant path to the refrigerant at low pressure in the low-pressure-side refrigerant path of the recuperator 250, causes an influencing of the state of aggregation refrigerant on the low-pressure side.
  • the proportion of wet vapor in the refrigerant flowing through the recuperator 250 on the low-pressure side at low pressure LP decreases when heat is transferred to the refrigerant, and after complete evaporation, the refrigerant advantageously overheats.
  • recuperator 250 there is advantageously a significantly higher heat transfer in the evaporator 240 between the source medium and the refrigerant in the evaporator 240.
  • a significantly higher heat transfer is set in the evaporator 240 than in the recuperator 250, since the environment is to be extracted by means of the evaporator 240 significantly more energy than it is to transfer only in the recuperator 250 within the refrigeration cycle.
  • the driving temperature difference is, for example, between 20 K and 60 K in the recuperator, while it is only between 3 K and 10 K in the evaporator.
  • the exchanger surface of the evaporator is designed to be approx. 5 to 20 times larger than that of the recuperator 250.
  • the refrigerant path of the recuperator 250 on the low-pressure side is fed from the evaporator outlet 242 of the evaporator 240 .
  • the internal energy state of the refrigerant is also delayed here by at least two time constants Z, Z11 , Z12 , Z13 , Z14 , Z15 , Ztot after changing the manipulated variable "degree of opening of the expansion valve".
  • the behavior over time of the recuperator 250 can advantageously be taken into account as the total recuperator time constant Z ges depending on the respective operating point of the vapor compression circuit in the range between approximately 1 minute and 30 minutes.
  • Step 1 First, the process variables compressor inlet overheating dTü ⁇ are advantageously recorded as the main control variable and the evaporator outlet overheating dT ÜA is advantageously recorded as an auxiliary variable in a first method step.
  • the temperatures of the refrigerant temperature are recorded by means of temperature sensors 501, 508.
  • the temperature difference of the refrigerant at the respective measuring point and the evaporation temperature are then calculated and this temperature difference value then corresponds to the respective overheating of the refrigerant at the measuring point.
  • the starting variables of the calculation in step 1 are then the compressor inlet overheating dT ÜE and the evaporator outlet overheating DT ÜA .
  • Step 2 The process variables compressor inlet overheating dTü ⁇ and evaporator outlet overheating DT ÜA are advantageously offset in a second step to form associated control deviations with respectively associated setpoint values:
  • the target value for the compressor inlet overheating dT ÜE at the evaporator outlet 242 is advantageously varied in the range between approximately 5 K and 20 K to ensure the permissible compressor operating range and the highest possible efficiency of the refrigeration circuit.
  • the setpoint for the evaporator outlet overheating DT ÜA is then varied depending on the refrigeration circuit operating mode and the refrigeration circuit working point so that the evaporator overheating in the settled control case roughly corresponds to the process value of the evaporator outlet overheating dT ÜA that is set.
  • This target value for the evaporator outlet overheating DT ÜA can be precalculated and adaptively corrected based on an operating mode or an operating point depending on the evaporation temperature, the condensation temperature, the compressor capacity, a target value for the compressor inlet overheating dT ÜE at the evaporator outlet 242 and/or component properties.
  • the control deviation of the compressor inlet overheating dTü ⁇ is then calculated by subtracting the setpoint value of the compressor inlet overheating dT ÜE from the process value of the compressor inlet overheating dTü ⁇ .
  • the control deviation of the evaporator outlet overheating dT ÜA is then calculated by subtracting the setpoint of the evaporator outlet overheating DT ÜA from the process value of the evaporator outlet overheating dT ÜA .
  • Step 3 In a third method step, the control deviation of the compressor inlet overheating dTü ⁇ and the control deviation of the evaporator outlet overheating dT ÜA are advantageously combined to form a total control deviation overheating.
  • the combination takes place in particular by means of a weighted addition of the individual control deviations.
  • the weighting influence is a measure of the proportionate combination of the individual control deviations and, in extreme cases, can result in the exclusive inclusion of just one individual control deviation, but usually the weighted inclusion of both individual control deviations.
  • Step 4 In a fourth method step, the calculated total control deviation of the overheating is then processed in the controller 500, which controls the corresponding actuators of the refrigeration circuit, in particular the expansion valve 230 with the adjustable degree of opening and/or the compressor 210 with the adjustable compressor speed, so that in the regulated case, a control deviation of the overheating is set to about 0 Kelvin if possible.
  • a P, I, PI, PID controller can be used, with the control components advantageously being dynamically adapted to the respective operating mode and the operating point.
  • a step 810 the refrigerant in the compressor 210 is compressed from low pressure ND to high pressure HD, with the temperature likewise increasing from the evaporator outlet temperature T KE to the hot gas temperature T HG .
  • the heat output Q H is first released in step 820 , before the refrigerant is then liquefied in step 830 .
  • the liquefaction is preferably completely completed in the liquefier 220 .
  • the liquid refrigerant is supercooled in step 850 before it is expanded in step 860 in the expansion valve 230 to the low pressure ND.
  • step 870 the liquid and expanded refrigerant is almost completely evaporated in the evaporator 240 before it is finally overheated in the recuperator 250 in step 890 .
  • the superheated refrigerant is then compressed again in compressor 210 in step 810 so that the cycle can be repeated.
  • the first difference occurs in step 830.
  • the liquefaction is not completed in the liquefier 220, but the still gaseous refrigerant enters the recuperator 250 and is liquefied in a step 840 in the recuperator 250.
  • the energy is transferred internally from the high pressure to the low pressure path, reducing the COP.
  • step 870 the vaporization in the vaporizer 240 is not completed. A significant part of the evaporation takes place in a step 880 in the recuperator 250 .

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  • Engineering & Computer Science (AREA)
  • Physics & Mathematics (AREA)
  • Mechanical Engineering (AREA)
  • Thermal Sciences (AREA)
  • General Engineering & Computer Science (AREA)
  • Air Conditioning Control Device (AREA)

Claims (7)

  1. Procédé d'exploitation d'une installation de réfrigération à compression (200) comportant
    - un fluide frigorigène qui présente un glissement de température,
    - un évaporateur (240),
    - un compresseur (210),
    - un condenseur (220),
    - un organe d'étranglement (230),
    - un échangeur de chaleur (250) interne, l'échangeur de chaleur interne étant conçu pour transmettre de l'énergie thermique du fluide frigorigène situé dans le chemin haute pression (HD), avant son entrée dans l'organe d'étranglement (230), au fluide frigorigène situé dans le chemin basse pression (ND) avant son entrée dans le compresseur (210) et
    - une unité de commande (500), l'unité de commande étant conçue pour
    a) détecter une surchauffe du fluide frigorigène après sa sortie de l'évaporateur (240) ainsi que pour détecter une surchauffe du fluide frigorigène à son entrée dans le compresseur (210) et
    b) réguler l'organe d'étranglement (230) à l'aide d'une grandeur de régulation reposant sur les deux surchauffes,
    caractérisée en ce que l'unité de commande est en outre conçue pour
    c) détecter un manque de fluide frigorigène au moins dans le condenseur (220), la régulation comprenant les étapes suivantes :
    - calcul d'un écart de température entre la température du point de rosée et la température du point d'ébullition, dit glissement de température, du fluide frigorigène au point de fonctionnement actuel de l'installation de réfrigération à compression compte tenu d'une basse pression (ND) détectée du fluide frigorigène,
    - calcul d'une valeur limite d'écart de température relativement au glissement de température calculé concernant la basse pression (ND) détectée du fluide frigorigène au point de fonctionnement actuel,
    - évaluation de la surchauffe actuelle à la sortie de l'évaporateur pour déterminer s'il y a passage en dessous de la valeur limite d'écart de température,
    - mise à disposition d'une valeur de consigne d'une surchauffe qui est définie comme écart entre la température du fluide frigorigène et la température du point de rosée, à l'entrée dans le compresseur (210), dit surchauffe à l'entrée du compresseur,
    - évaluation de la surchauffe actuelle à l'entrée du compresseur pour déterminer s'il y a dépassement de la valeur de consigne de la surchauffe à l'entrée du compresseur,
    - détection d'une situation de manque de fluide frigorigène quand les deux résultats des étapes d'évaluation s'appliquent sans interruption pendant une durée définie.
  2. Procédé selon la revendication 1, dans lequel le procédé comporte en outre l'étape suivante :
    - exécution d'une tentative de retour de fluide frigorigène dans le condenseur (220) par augmentation, de manière limitée dans le temps, de la valeur de consigne du régime de compresseur et/ou par augmentation, de manière limitée dans le temps, de la valeur de consigne de la surchauffe à la sortie de l'évaporateur.
  3. Procédé selon la revendication 2, dans lequel le procédé comporte en outre l'étape suivante :
    - génération d'un état de défaut lorsque la tentative de retour de fluide frigorigène est sans succès, en particulier lorsque la tentative de retour de fluide frigorigène est sans succès à plusieurs reprises, pendant un laps de temps défini.
  4. Procédé selon l'une quelconque des revendications précédentes, dans lequel le calcul de la valeur limite d'écart de température relativement au glissement de température calculé concernant la basse pression détectée du fluide frigorigène est effectué en fonction du point de fonctionnement actuel de l'installation de réfrigération à compression.
  5. Procédé selon la revendication 4, dans lequel la valeur limite d'écart de température relativement au glissement de température calculé est réglée proportionnellement à un écart entre la température de dissipateur de chaleur et la température de source de chaleur.
  6. Installation de réfrigération à compression (200) comportant
    - un fluide frigorigène qui présente un glissement de température,
    - un évaporateur (240),
    - un compresseur (210),
    - un condenseur (220),
    - un organe d'étranglement (230),
    - un échangeur de chaleur (250) interne, l'échangeur de chaleur interne étant conçu pour transmettre de l'énergie thermique du fluide frigorigène situé dans le chemin haute pression (HD), avant son entrée dans l'organe d'étranglement (230), au fluide frigorigène situé dans le chemin basse pression (ND) avant son entrée dans le compresseur (210) et
    - une unité de commande (500), l'unité de commande étant conçue pour
    a) détecter une surchauffe du fluide frigorigène après sa sortie de l'évaporateur (240) ainsi que pour détecter une surchauffe du fluide frigorigène à son entrée dans le compresseur (210) et
    b) réguler l'organe d'étranglement (230) à l'aide d'une grandeur de régulation reposant sur les deux surchauffes,
    caractérisée en ce que l'unité de commande est en outre conçue pour
    c) détecter un manque de fluide frigorigène au moins dans le condenseur (220), l'unité de commande étant en outre conçue pour :
    - calculer un écart de température entre la température du point de rosée et la température du point d'ébullition, dit glissement de température, du fluide frigorigène au point de fonctionnement actuel de l'installation de réfrigération à compression compte tenu d'une basse pression (ND) détectée du fluide frigorigène,
    - calculer une valeur limite d'écart de température relativement au glissement de température calculé concernant la basse pression (ND) détectée du fluide frigorigène au point de travail actuel,
    - évaluer la surchauffe actuelle à la sortie de l'évaporateur pour déterminer s'il y a passage en dessous de la valeur limite d'écart de température,
    - mettre à disposition une valeur de consigne d'une surchauffe qui est définie comme écart entre la température du fluide frigorigène et la température du point de rosée, à l'entrée dans le compresseur (210), dit surchauffe à l'entrée du compresseur,
    - évaluer la surchauffe actuelle à l'entrée du compresseur pour déterminer s'il y a dépassement de la valeur de consigne de la surchauffe à l'entrée du compresseur,
    - détecter une situation de manque de fluide frigorigène quand les deux résultats des étapes d'évaluation s'appliquent sans interruption pendant une durée définie.
  7. Pompe à chaleur comportant une installation de réfrigération à compression selon la revendication 6.
EP21177576.2A 2020-06-09 2021-06-03 Procédé de fonctionnement d'une installation de réfrigération à compression et installation de réfrigération à compression Active EP3922932B1 (fr)

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Publication number Priority date Publication date Assignee Title
JP4200532B2 (ja) 1997-12-25 2008-12-24 三菱電機株式会社 冷凍装置
DE19829335C2 (de) 1998-07-01 2000-06-08 Kki Klima-, Kaelte- Und Industrieanlagen Schmitt Kg Kälteanlage
JP2000179960A (ja) * 1998-12-18 2000-06-30 Sanden Corp 蒸気圧縮式冷凍サイクル
JP4202505B2 (ja) * 1999-01-11 2008-12-24 サンデン株式会社 蒸気圧縮式冷凍サイクル
DE10159892B4 (de) 2001-12-06 2006-08-24 Stiebel Eltron Gmbh & Co. Kg Kältemaschine mit einem Rekuperator
EP1775533B1 (fr) 2005-10-13 2018-03-28 STIEBEL ELTRON GmbH & Co. KG Procédé pour faire fonctionner un système frigorifique à compression
DE102005061480B3 (de) 2005-12-22 2007-04-05 Stiebel Eltron Gmbh & Co. Kg Wärmepumpenanlage
EP2669597B1 (fr) 2011-01-27 2017-05-17 Mitsubishi Electric Corporation Appareil de conditionnement d'air
US10001308B2 (en) 2011-12-22 2018-06-19 Mitsubishi Electric Corporation Refrigeration cycle device
DE102014221106A1 (de) 2014-10-17 2016-04-21 Bayerische Motoren Werke Aktiengesellschaft Verfahren zur Steuerung oder Regelung eines Fahrzeugklimaanlagen-Kältemittelkreislaufs
DE102015007564B4 (de) 2015-06-12 2023-11-23 Audi Ag Verfahren zum Betreiben einer Klimaanlage
KR102655619B1 (ko) * 2017-12-18 2024-04-09 다이킨 고교 가부시키가이샤 냉동 사이클 장치

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