EP3922926B1 - Procédé de régulation d'un processus de dégivrage d'un évaporateur d'une installation de réfrigération à compression et installation de réfrigération à compression - Google Patents

Procédé de régulation d'un processus de dégivrage d'un évaporateur d'une installation de réfrigération à compression et installation de réfrigération à compression Download PDF

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Publication number
EP3922926B1
EP3922926B1 EP21177570.5A EP21177570A EP3922926B1 EP 3922926 B1 EP3922926 B1 EP 3922926B1 EP 21177570 A EP21177570 A EP 21177570A EP 3922926 B1 EP3922926 B1 EP 3922926B1
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EP
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Prior art keywords
temperature
evaporator
refrigerant
dew
compressor
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EP21177570.5A
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German (de)
English (en)
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EP3922926A1 (fr
Inventor
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
    • F25B47/00—Arrangements for preventing or removing deposits or corrosion, not provided for in another subclass
    • F25B47/006—Arrangements for preventing or removing deposits or corrosion, not provided for in another subclass for preventing frost
    • 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
    • F25B47/00—Arrangements for preventing or removing deposits or corrosion, not provided for in another subclass
    • F25B47/02—Defrosting cycles
    • 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
    • F25B2500/00—Problems to be solved
    • F25B2500/15—Hunting, i.e. oscillation of controlled refrigeration variables reaching undesirable values
    • 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
    • F25B2500/19—Calculation of parameters
    • 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
    • F25B2700/19—Pressures
    • F25B2700/197—Pressures of the evaporator
    • 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
    • F25B2700/2106—Temperatures of fresh outdoor air
    • 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 present invention relates to a method and a correspondingly designed device for controlling a defrosting process of an evaporator of a compression refrigeration system, in particular a heat pump.
  • a compression refrigeration system includes as components an evaporator, a throttle element such as an expansion valve, a pressure increasing element such as a compressor and a condenser.
  • a throttle element such as an expansion valve
  • a pressure increasing element such as a compressor
  • condenser a condenser
  • a defrosting requirement detection must initiate defrosting to remove the ice layer when the maximum permissible ice layer thickness is exceeded. It is important that the defrosting is initiated in good time and lasts as short as possible. This increases the efficiency of the heat pump or compression refrigeration system and thus improves the coefficient of performance.
  • document DE 10 2012 208 819 A1 discloses a method for the control and regulation of the evaporation process in an air-loaded evaporator of a heat pump or cooling system, in which the process state that can be described using known process variables, characteristics and parameters is evaluated as to whether there is icing on the air-side evaporator surface by taking the position of the evaporator surface into account Working point i, given by the value of the virtual heat output in the virtual MSS characteristic curve and the stability limit of the evaporation process is recognized from the inclination angle of the time function of the temperature.
  • document EP 1 355 207 A1 refers to a method for operating a compression refrigeration system, in particular for cooling rooms with a refrigeration circuit comprising an evaporator, a compressor, a condenser and an expansion valve, in which the cooling capacity and the operating point of the evaporator are determined by comparing the setpoint and the actual value of suitable controlled variables the degree of opening of the expansion valve and the power, in particular the speed of the compressor, are regulated as manipulated variables, with the coupling of the controlled variables being at least approximately compensated for by decoupling elements with previously determined decoupling functions in order to improve the control when determining the manipulated values from the respective deviation of the controlled variables from their setpoint becomes.
  • Methods for controlling a defrosting process of an evaporator in a compression refrigeration system are available, for example DE 10 2005 054 101 A1 known.
  • the refrigeration circuit controller tries to reduce the evaporation temperature to such an extent that the energy is still transferred even with a reduced temperature difference in the air. It must also evaporate approximately 1 K lower for every K increase in the temperature difference in the air in order to continue to ensure overheating.
  • This reaction of the controller via an increase in the temperature difference of the heat source medium due to the volume flow is used to detect the need for defrosting.
  • a reference size calculation can be carried out in a defrosted state, initially or after defrosting or under conditions in which it is assumed that there will be no ice accretion.
  • the reference value calculation provides the temperature difference between the outside temperature entering the evaporator and the controlled evaporation temperature of the refrigerant.
  • the controller will reduce the evaporation temperature in order to achieve superheating, for example by reducing the degree of opening of the throttle element, whereby the evaporation temperature drops. This evaporation temperature could then be evaluated against the reference and an evaporation temperature reduction could be determined.
  • the evaporator If the reduction in the evaporation temperature exceeds a (factory) set threshold for evaporation temperature reduction of, for example, 2 K, the evaporator is sufficiently iced over that the need for defrosting is detected and the defrosting process is initiated.
  • a (factory) set threshold for evaporation temperature reduction of, for example, 2 K
  • a method for controlling a defrosting process of an evaporator of a compression refrigeration system wherein the compression refrigeration system has: a refrigeration circuit with refrigerant, an evaporator which has a fan and is designed to transfer heat from air to the refrigerant, a compressor, a control unit for Adjustment of a desired cooling capacity, which is defined as heat transfer in the evaporator.
  • the method has the following steps: a) measuring an evaporator outlet pressure, b) determining a dew temperature based on the evaporator outlet pressure, c) correcting the determined dew temperature by compensating for interference, d) determining a difference between the dew temperature and a dew temperature reference value, e) introduction a defrosting process if the difference exceeds a temperature limit.
  • the step of correcting the determined dew temperature includes compensating for interference using at least one of the influencing variables low pressure, cooling capacity, evaporator outlet overheating and/or fan performance.
  • the dew temperature is formed from a difference between a first dew temperature and an outside temperature of the outside air flowing through the evaporator, the first dew temperature being calculated from the evaporator outlet pressure and/or the outside temperature being measured.
  • the dew temperature is filtered, in particular with a low-pass filter and particularly preferably filtered with a first-order low-pass filter.
  • the dew temperature reference value is the maximum dew temperature averaged over a certain period of time.
  • the defrosting process preferably comprises the following steps: e) determining a second difference between the evaporator output pressure and a switch-off pressure f) terminating the defrosting process if the second difference falls below a pressure limit.
  • a compression refrigeration system comprising: a refrigeration circuit with refrigerant, an evaporator which has a fan and is designed to transfer heat from air to the refrigerant, a compressor, a control unit for regulating a desired refrigeration output, which is used as heat transfer in the Evaporator is defined, wherein the control unit is designed to: measure an evaporator output pressure, determine a dew temperature based on the evaporator output pressure, correct the determined dew temperature by compensating for interference, determine a difference between the dew temperature and a dew temperature reference value, initiate a defrosting process if the difference is one Temperature limit exceeds.
  • the step of correcting the determined dew temperature includes the compensation of disruptive influences with the help of at least one of the influencing variables low pressure, cooling capacity, evaporator outlet overheating and/or fan performance.
  • the refrigerant preferably has a temperature glide and is in particular a mixture of R32 and R1234yf and particularly preferably R454C.
  • the task is further solved by a heat pump with a compression refrigeration system according to the invention.
  • the actuators listed below are advantageously at least partially connected to the controller via a data connection 510, which can be done 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, one in the Fig. 1 evaporator inlet temperature sensor, not shown, determine the temperature at the evaporator inlet 241.
  • the heat pump 100 is shown as a brine heat pump.
  • a fan/fan is arranged as a heat source instead of the brine circuit with brine pump 330.
  • the compressor 210 serves to compress the superheated refrigerant from an inlet connection 211 to a compressor outlet pressure P Va at a compressor outlet temperature corresponding to the hot gas temperature at the compressor outlet 212.
  • the compressor 210 usually contains a drive unit with an electric motor, a compression unit and advantageously the electric motor can be operated at a variable speed .
  • the compression unit can be designed as a rolling piston unit, scroll unit or otherwise.
  • the compressed superheated refrigerant is at a higher pressure level at the compressor outlet pressure P Va , in particular a high pressure HD, than at the inlet connection 211 with a compressor inlet pressure P Ve , in particular a low pressure ND, at a compressor inlet temperature T VE , which indicates the state of the refrigerant temperature Entry port 211 describes when entering a compression chamber.
  • thermal energy Q H is transferred from the refrigerant of the vapor compression circuit 200 to a heating medium of the heat sink system 400.
  • the refrigerant is deheated in the liquefier 220, with superheated refrigerant vapor transferring part of its heat energy to the heating medium of the heat sink system 400 through a temperature reduction .
  • 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 Tws in the heat sink system during operation of the compressor 210.
  • 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, thereby transferring the thermal energy Q H from the refrigerant to the heating medium.
  • refrigerant emerging 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 through the expansion valve 230, the collector 260 fills, otherwise it empties or empties.
  • recuperator 250 which can also be referred to as an internal heat exchanger
  • internal heat energy Q i is transferred from the refrigerant under the high pressure HD, which flows from the condenser 220 to the expansion valve 230 in a high-pressure flow direction S HD , to that flowing under the low pressure ND Transfer refrigerant, which flows from the evaporator to the compressor in a low-pressure flow direction S ND .
  • the refrigerant flowing from the condenser to the expansion valve 230 is advantageously subcooled.
  • the refrigerant flows into the expansion valve through an expansion valve inlet 231.
  • the refrigerant pressure is throttled from the high pressure HD to the low pressure ND by the refrigerant advantageously passing through a nozzle arrangement or throttle with an advantageously variable opening cross section, 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.
  • An opening degree of the expansion valve 230 is adjusted by an electric motor, which is usually designed as a stepper motor, which is controlled by the control unit or regulation 500.
  • the low pressure ND at the expansion valve outlet 232 of the refrigerant from the expansion valve 230 is controlled so that the resulting low pressure ND of the refrigerant 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 Qv 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 uses the source energy Q delivers Q to the vapor compression circuit 200.
  • 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 uses the source energy Q delivers Q to the vapor compression circuit 200.
  • the refrigerant flowing into the evaporator 240 reduces its wet steam content as it flows through the evaporator 240 by absorbing heat Q Q and advantageously leaves the evaporator 240 with a small wet steam content or advantageously also as 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 in the case of brine - water heat pumps or an outside air fan in the case of air / water heat pumps, the heat energy Q Q being removed from the heat source medium as it flows through the evaporator.
  • heat energy Q i is transferred between the refrigerant flowing from the condenser 220 to the expansion valve 230 to the refrigerant flowing from the evaporator 240 to the compressor 210, the refrigerant flowing from the evaporator 240 to the compressor 210 in particular being further overheated.
  • This superheated refrigerant which exits the recuperator 250 at a superheat temperature T Ke , is directed 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 a quotient of the output heating power Q H and the electrical power P e consumed to drive the compressor motor.
  • the refrigerant which in the condenser 220 releases heat energy Q H to the heating medium at a temperature level on the heat sink side, is removed from the refrigerant in the high-pressure path of the recuperator 250 by subcooling.
  • 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 heat energy Q i extracted in the high pressure path is fed back to the refrigerant in the low pressure path at low pressure ND and at a low pressure temperature corresponding to an evaporator outlet temperature T Va at the entry into the recuperator 250.
  • the supply of energy advantageously results in a reduction of the wet steam proportion to a state without wet steam proportion. Overheating is ensured by additional energy supply.
  • the following sensors are advantageously arranged, with which a model-based pilot control is implemented, in particular to safeguard and optimize the operating conditions of the vapor compression system 200, particularly in the event of changes in the operating state.
  • the process variable that has a significant influence on the overall efficiency of the vapor compression circuit 200 as a quotient between the heating power Q H transmitted by the vapor compression circuit 200 and that received by the compressor 210 electrical power P e is the overheating of the refrigerant at the compressor inlet 211.
  • Superheat describes the temperature difference between the recorded compressor inlet temperature T KE of the refrigerant and the evaporation temperature of the refrigerant with saturated vapor.
  • the compressor inlet overheating is not regulated according to the invention in such a way that no condensate occurs on components of the refrigeration circuit due to the water vapor content in the ambient air falling below the dew point, in particular in the section between the refrigerant outlet of the recuperator 252 and the compressor inlet 211.
  • the refrigeration circuit section between the evaporator outlet 242 and the recuperator inlet 251 is usually colder, because this 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 that needs to be protected is located at the location of the compressor inlet 211 on the compressor.
  • the heat source medium temperature, the heating medium temperature, the compressor power P e and target values Z or the target value Z are not used according to the invention, depending on the operating point of the vapor compression circuit 200, for a calculation of the compressor inlet superheat dTü ⁇ .
  • a calculation of the target value Z can be carried out as a default value for the compressor inlet overheating dTü ⁇ from the refrigeration circuit measurement 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 refrigeration circuit components.
  • the target value for the compressor inlet superheat dTü ⁇ is constant, eg 10 Kelvin, regardless of all operating conditions. In a more complex adjustment it is varied as a function of an operating point size, for example the compressor power P e or in an even more complex adjustment it varies as a function of several operating point sizes.
  • the total control deviation is then advantageously calculated from the weighted influence of the control deviation of the compressor inlet superheat dTü ⁇ and the weighted influence of the control deviation of the evaporator outlet superheat dT ÜA in the controller 500, which is fed in to control the vapor compression circuit 200.
  • the refrigerant passes through two sequentially arranged heat exchangers, the evaporator 240 and the recuperator 250, in which thermal 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 supplied source heat 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 removed from the refrigerant after it leaves the condenser 220.
  • the temperature level of the refrigerant at the outlet of the condenser is approximately at the same level as the return temperature of the heating medium.
  • the control value R is advantageously the weighted combination of the control deviation of the compressor inlet superheat dTü ⁇ with the control deviation of the evaporator outlet superheat.
  • Actuators have a particularly advantageous influence on the control value R, in particular on the weighted connection of the control deviation of the compressor inlet overheating with the control deviation of the evaporator outlet overheating.
  • such actuators are in particular the compressor 210 by varying the compressor speed and the expansion valve 230 by influencing the degree of opening. These two actuators influence the low pressure LP and the evaporation temperature level.
  • a change in the compressor speed to regulate the desired heating output without further compensatory changes in the degree of opening of the expansion valve changes the control value R into undesirable ranges, so that a model-based, supported change in the degree of opening of the expansion valve associated with the change in compressor speed is advantageous and possibly even necessary for regulating R.
  • the compressor speed is advantageously set in the vapor compression circuit 200 so that the heating power QH transmitted from the vapor compression circuit 200 to the heating medium corresponds to the requested target value Z.
  • influencing the compressor speed to control the compressor inlet superheat dT ÜE is advantageously subordinated or not appropriate.
  • the degree of opening of the expansion valve 230 is advantageously used as a control value for controlling the compressor inlet superheat dTü ⁇ .
  • the influence of the degree of opening of the expansion valve 230 on the compressor inlet superheat dT ÜE occurs as follows:
  • the expansion valve 230 acts as a nozzle with an electromotively adjustable nozzle cross section, in which a needle-shaped nozzle needle is usually threaded into a nozzle seat using a stepper motor.
  • the refrigerant throughput 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 a constant dependent on the refrigerant and a geometry of the expansion valve 230.
  • the degree of opening of the expansion valve 230 significantly influences only 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 decreases.
  • the mass flow of refrigerant through the compressor 210 decreases approximately proportionally, since its delivery rate is approximately described as volume / time can, due in particular to the piston strokes, and a correspondingly reduced low pressure value ND is established, at which the refrigerant mass flow supplied through the expansion valve 230 is equal to the refrigerant mass flow discharged by the compressor 210.
  • the degree of opening of the expansion valve 230 is increased, more refrigerant passes through the expansion valve 230 at constant high pressure HD and initially constant low pressure ND. However, since the compressor 210 initially continues to deliver the same refrigerant mass flow, the low pressure side ND of the refrigeration circuit is supplied with 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 delivery rate of the compressor 210 increases approximately proportionally, since its delivery rate can be approximately described as volume / time, and it A correspondingly increased low pressure ND occurs, at which the refrigerant mass flow supplied through the expansion valve 230 is equal to the refrigerant mass flow discharged by the compressor 210.
  • the low pressure ND in turn significantly influences the heat transfer between the heat source medium and the 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 media path of the evaporator and the coolant path of the evaporator can be assumed to be approximately constant in a respective vapor compression circuit 200. Therefore, the size of the heat transfer performance in the evaporator 240 is largely dependent on the integral of the temperature differences of all surface elements of the heat transfer layer.
  • a refrigerant temperature is established which is a function of the low pressure ND of the refrigerant due to the saturation vapor characteristic curve as a material property of the refrigerant.
  • the thermal energy Q Q which is transferred from the heat source system to the refrigerant flowing through the evaporator 240, influences the physical state 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 at the exit from the evaporator 240 depending on the manipulated variable "degree of opening of the expansion valve 230".
  • control system behavior of the "isolated" controlled system "evaporator 240" results in a controlled system behavior with moderate steepness.
  • the controlled system behavior is characterized in particular by the controlled system output value of the evaporator outlet superheat as a function of the controlled system input value of the expansion valve opening degree.
  • a refrigerant is advantageously used, in particular a refrigerant mixture which has a "temperature glide", in particular R454C is advantageously used.
  • the refrigerant After flowing through the evaporator 240, the refrigerant enters the low-pressure path of the recuperator 250 at low pressure ND.
  • the physical state of the refrigerant when it flows into the recuperator 250 is, in a normal operating case, advantageously either saturated steam with a low vapor content between 0 to 20% or, in particular, advantageously already overheated refrigerant.
  • a refrigerant temperature is established which is a function of the refrigerant pressure due to the saturation vapor characteristic curve of the refrigerant.
  • the refrigerant temperature will assume a maximum value that corresponds to the inlet temperature of the heat source medium.
  • the size preferably corresponds to the inlet temperature of the refrigerant into the high-pressure path of the recuperator 250, i.e. the temperature of the refrigerant after it leaves the condenser 220.
  • the temperature of the refrigerant of the high-pressure-side refrigerant path is greater than high pressure HD in as many surface elements of the transfer layer of the recuperator 250 as possible is the temperature of the refrigerant of the low-pressure side refrigerant path at low pressure ND on the respective surface element.
  • the corresponding temperatures of the heating system 400 of the vapor compression circuit 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, influences the physical state of the refrigerant on the low-pressure side.
  • the wet steam portion of the refrigerant flowing through the recuperator 250 on the low pressure side at low pressure LP decreases as heat is transferred to the refrigerant and after complete evaporation, the refrigerant advantageously overheats.
  • the low pressure ND of the refrigerant in the low-pressure side path of the recuperator 250 results in a controlled system behavior with a high steepness, with an approximately constant internal energy state of the refrigerant at the entry 251 into the low-pressure side ND path of the recuperator 250.
  • a particularly relative change in the degree of opening of the expansion valve of 1% results in a change in superheat at the outlet of the refrigerant from the evaporator 230 of advantageously approximately 10 K or even over 10 K.
  • the driving temperature difference can, for example, be between 20 and 60K in the recuperator, while in the evaporator it is only between 3 and 10K.
  • the exchanger surface of the evaporator for example, is designed to be approximately 5 to 20 times larger than that of the recuperator 250.
  • the low-pressure side refrigerant path of the recuperator 250 is fed from the evaporator outlet 242 of the evaporator 240.
  • the internal energy state of the refrigerant is already delayed by at least two time constants Z, Z 11 , Z 12 , Z 13 , Z 14 , Z 15 , Z tot after changing the manipulated variable “opening degree of expansion valve”.
  • the time behavior of the recuperator 250 can advantageously be taken into account as the overall recuperator time constant Z tot depending on the respective operating point of the vapor compression circuit in the range between approximately 1 minute and 30 minutes.
  • a weighted combination of the compressor inlet superheat dTü ⁇ and the evaporator outlet superheat dT ÜA is advantageously carried out, in particular by means of a weighted combination of the control deviation of the compressor overheating and the control deviation of the evaporator outlet superheat dT ÜA , the total control deviation is calculated, which is fed into the controller 500 for controlling the vapor compression circuit 200.
  • Step 1 First, the process variables compressor inlet superheat dT ÜE are advantageously recorded as the main controlled variable and the evaporator outlet superheating dT ÜA is advantageously recorded as an auxiliary variable in a first process step.
  • the refrigerant temperature is detected 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 output variables for the calculation in step 1 are then the compressor inlet superheat dTü ⁇ and the evaporator outlet superheat dT ÜA .
  • Step 2 The process variables compressor inlet superheat dTü ⁇ and evaporator outlet superheat dT ÜA are advantageously calculated in a second step to form assigned control deviations with assigned setpoints:
  • the setpoint for the compressor inlet superheat dTü ⁇ is advantageously varied in the range between approx. 5 K to 20 K to ensure the permissible compressor operating range and the highest possible efficiency of the refrigeration circuit.
  • the setpoint for the evaporator outlet superheat dT ÜA at the evaporator outlet 242 is then varied depending on the refrigeration circuit operating mode and the refrigeration circuit operating point so that, in the steady state, this approximately corresponds to the resulting process value of the evaporator outlet superheat dT ÜA .
  • This setpoint for the evaporator outlet superheat dT ÜA can be pre-calculated and adaptively corrected based on a model depending on an operating mode or an operating point depending on the evaporation temperature, the condensation temperature, the compressor output, a setpoint of the compressor inlet superheat dTü ⁇ at the compressor inlet 211 and/or component properties.
  • the control deviation of the compressor inlet overheating dT ÜE is then calculated by subtracting the setpoint of the compressor inlet overheating dTü ⁇ from the process value of the compressor inlet overheating dT ÜE .
  • the control deviation of the evaporator outlet superheat dT ÜA is then calculated by subtracting the setpoint of the evaporator outlet superheat dT ÜA from the process value of the evaporator outlet superheat dT ÜA .
  • Step 3 In a third method step, the control deviation of the compressor inlet superheat dTü ⁇ and the control deviation of the evaporator outlet superheat dT ÜA are advantageously combined to form a total control deviation overheating.
  • the combination is carried out 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 only one individual control deviation, but usually the weighted inclusion of both individual control deviations.
  • the weighting influence is advantageously estimated as a value between 0 to 1, i.e. 0 to 100% and this value is based on the degree of inclusion of the control deviation of the compressor inlet overheating dTü ⁇ in the total control deviation, which results in the following dependency for the calculation of the total control deviation :
  • ⁇ Deviation from the rule Overheating Weighting influence * Deviation from the rule Compressor inlet overheating + 1 ⁇ Weighting influence * Deviation from the rule Evaporator outlet superheat
  • Step 4 In a fourth method step, the calculated total control deviation of the superheat 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, in such a way that In the regulated case, a control deviation of the superheating is as close as possible to approximately 0 Kelvin.
  • a P, I, PI, PID controller can be used, with the control components being advantageously dynamically adapted to the respective operating mode and the operating point.
  • the evaporation temperature of the refrigerant in the evaporator in heating mode depends, among other things, on the following process variables: outside temperature, degree of icing of the evaporator and the resulting reduced air mass flow, cooling capacity of the heat pump and evaporator outlet overheating.
  • the higher the cooling capacity of the heat pump, the lower the evaporation temperature is at a constant outside temperature and constant degree of icing.
  • the air mass flow reduced by the icing is the parameter which should be measured and evaluated by the defrosting requirement recognition according to the invention, as independently as possible from the influences of the other process variables. If the outside temperature, degree of icing and cooling capacity are approximately constant, a variation of the evaporator outlet superheat, e.g. by changing the degree of expansion valve opening from less than one to a few percent, means a proportional change in the evaporation temperature, which is also synonymously referred to as the dew point temperature.
  • the dew point temperature that would occur with a parameterized reference cooling output is calculated.
  • a parameterizable linear relationship between the change in cooling capacity and the change in evaporation temperature is calculated.
  • This compensation is based on the refrigeration connection that, with otherwise unchanged operating conditions of the refrigeration circuit, i.e. that heat source temperatures, heat sink temperatures and cooling capacity are approximately constant, a varied evaporator outlet superheat has an influence on the temperature difference between the heat source temperature and the evaporation temperature: An increase in superheating caused by disturbance variables and thus also in the control deviation of superheating is generally accompanied by an increase in the temperature difference between the heat source temperature and the evaporation temperature.
  • a reduction in superheating caused by disturbance variables and thus also in the control deviation of superheating is generally accompanied by a reduction in the temperature difference between the heat source temperature and the evaporation temperature
  • the relationship between a change in superheat and the change in the temperature difference between the heat source temperature and the evaporation temperature is generally non-linear; if the superheat is sufficiently high, the ratio of the change in the temperature difference between the heat source temperature and the evaporation temperature to the change in the control deviation of the superheat is almost one, that is to say An increase in superheating by 1 Kelvin is accompanied by an increase in the temperature difference between the heat source temperature and the evaporation temperature by almost 1 Kelvin.
  • the steepness of the compensation and thus the ratio of the change in the temperature difference between the heat source temperature and the evaporation temperature to the change in an overheating correction is preferably adjustable using parameters.
  • the range of compensation is preferably limited to a range of values.
  • a third step the influence of the fan performance and the associated influence on the air mass flow through the evaporator on the resulting dew point temperature in low pressure is compensated.
  • This compensation is based on the physical connection that, with otherwise unchanged operating conditions of the refrigeration circuit, i.e. that the heat source inlet temperature into the heat pump, heat sink temperatures and cooling capacity are approximately constant, a varied fan performance influences the temperature difference between the heat source inlet temperature and the heat source inlet temperature and thus also influences the Temperature difference between heat source inlet temperature and evaporation temperature has: An increase in the air mass flow through the evaporator caused by an increase in fan power is generally accompanied by a reduction in the temperature difference between the heat source temperature and the evaporation temperature
  • a reduction in the air mass flow through the evaporator caused by fan power reduction is generally accompanied by an increase in the temperature difference between the heat source temperature and the evaporation temperature
  • the relationship between a change in the fan performance and thus the air mass flow through the evaporator and the change in the temperature difference between the heat source temperature and the evaporation temperature is generally non-linear, regardless of the operating point of the refrigeration circuit.
  • the gradient between a change in the fan power and thus the air mass flow through the evaporator and the change in the temperature difference between the heat source temperature and the evaporation temperature depends on the operating point of the refrigeration circuit.
  • a relative change in the air mass flow at higher cooling capacities causes a larger change in the temperature difference between the heat source temperature and the evaporation temperature than with smaller cooling capacities.
  • the refrigeration circuit controller would regulate an evaporation temperature just below the heat source medium outlet temperature from the evaporator in order to completely evaporate the refrigerant.
  • the heat source media outlet temperature in turn results from a) heat source media inlet temperature into the evaporator, b) specific heat capacity of the heat source medium, c) heat source media - mass flow and d) cooling power transferred in the evaporator.
  • the heat source media outlet temperature is reduced proportionally to a product of the change in the heat source media mass flow and the cooling capacity.
  • the dependency between heat source media mass flow and evaporation temperature is preferably determined by measuring by varying the fan power
  • Fig. 3 shows schematically and exemplary relationships between evaporation temperatures on the vertical axis and a relative fan speed, which is indicative of the fan performance, on the horizontal axis.
  • the different curves 3100, 3200, 3300 are determined for different compressor outputs, with higher compressor outputs being closer to the outside temperature 3400, which is in Fig.3 is constant. It can be seen that the evaporation temperature approaches the outside temperature as the fan power increases and the compressor power increases.
  • FIG. 4a and 4b show schematically and exemplary temperature curves of the media flows in the evaporator at a high media flow of 3400 in Fig. 4a and at low media flow 3500 in Fig. 4b .
  • the temperature of the heat source medium WQ decreases from an inlet temperature WQ_In to an outlet temperature WQ_Out as it passes through the evaporator by releasing heat to the refrigerant KM.
  • the refrigerant is initially evaporates, which corresponds to the constant/horizontal range of the temperature curve between the temperature at the inlet KM_In and the temperature at the outlet KM_Out before overheating occurs, i.e. the area to which the curve 3410 or 3420 of the refrigerant temperature increases.
  • the evaporation temperature level and the non-linearity of the relationship between a) the difference between the outside air temperature and the evaporation temperature and b) the cooling capacity are preferably included in the calculation of the power-compensated dew point temperature in low pressure.
  • the inclusions are preferably carried out using a correction parameter for the evaporation temperature level and the nonlinearity, which are particularly preferably included in the calculation as an exponent.
  • the gradient difference temperature divided by the cooling capacity has a dependence on the heat source temperature.
  • a heat source temperature-dependent evaporation pressure can play a role here.
  • the suction gas density and the flow rate of the refrigerant change in proportion to the evaporation pressure, which in turn influence the heat transfer between the refrigerant and the heat exchanger.
  • This dependency is included via the correction parameter for the evaporation temperature level, particularly preferably as an exponent of the refrigerant pressure in the low-pressure path of the refrigeration circuit, which is multiplicatively included in the calculation of the power-compensated dew point temperature.
  • a shorter residence time means lower energy transfer based on the driving temperature difference. In order to achieve the desired energy transfer, in particular to achieve the desired degree of evaporation, a disproportionately higher driving temperature difference is required. Furthermore, the Reynold number, which depends on the average flow velocity, can influence the heat transfer.
  • non-linearity of cooling capacity are particularly preferably compensated for by a factor which is an exponential factor of the cooling capacity of the compressor to the power of a correction parameter, the factor also preferably being included multiplicatively in the calculation of the capacity-compensated dew point temperature.
  • Fig. 5 and 6 show schematically and as an example a temperature difference between the outside air temperature TA and the evaporation temperature T0, called the difference temperature, on the vertical axis as a function of the cooling capacity on the horizontal axis.
  • measured curves 4010, 4020, 4030 are shown for four different compressor speeds at different operating points of the compression refrigeration system.
  • the different operating points are preferably determined by different outside temperatures and by different heat sink flow temperatures, for example heating flow or hot water flow temperatures.
  • Fig. 5 the corrections of the dependencies "evaporation temperature level” and “non-linearity cooling capacity” are not included, whereas Fig. 6 shows the curves with corrected, calculated dew point temperature or evaporation temperature T0.
  • the associated calculated, power-compensated curves 4012, 4022, 4032 in Fig. 5 are correspondingly linear, while the curves 4014, 4024, 4034 in Fig. 6 the exponential corrections “evaporation temperature level” and “non-linearity cooling capacity” are taken into account.
  • the power-compensated dew point temperature is filtered for further processing, for example low-pass filtered and preferably filtered with a first-order low pass.
  • the defrost detection can be deactivated in the first time, for example an adjustable value such as 10 minutes, of a heating cycle and only activated after the first time has elapsed.
  • the outside temperature-compensated dew point temperature changes in the outside temperature are included in the calculation.
  • the outside temperature is preferably filtered, particularly preferably filtered with a low pass, and then further processed.
  • a dew point temperature reference is important for the method according to the invention.
  • the dew point temperature reference is only calculated if defrosting requirement recognition is enabled.
  • the dew point temperature reference is initialized when the device is switched on or after successful defrosting. Defrosts are considered successful if they are ended by the regular defrost end detection, in this case, for example, high pressure above limit pressure.
  • Defrosts are considered unsuccessful if they are terminated by other criteria, such as condenser temperature below limit, compressor shutdown due to a lockout, or maximum defrost time exceeded. If the defrost is not successful, the value of the dew point temperature reference is continued.
  • the dew point temperature reference is assigned a value that is below any possible value during operation, e.g. - 100 °C.
  • the dew point temperature reference is updated by setting it to a maximum of the dew point temperature and the dew point temperature reference.
  • the dew point temperature is preferably the dew point temperature corrected, filtered and/or outside temperature compensated according to the invention.
  • the need for defrosting is recognized when the dew point temperature plus a temperature difference parameter to trigger defrosting, for example 2 K, is smaller than the dew point temperature reference.
  • the dew point temperature is preferably the dew point temperature corrected, filtered and/or outside temperature compensated according to the invention.
  • a settling process of the refrigeration circuit process values relevant for the defrost trigger detection takes place.
  • the refrigeration circuit element to which the evaporator outlet temperature sensor is coupled is subjected to hot gas temperature; the outside air temperature sensor is also heated by the evaporator heated in defrosting mode.
  • both temperature values adjust over time to the process temperatures actually present in the refrigeration circuit. Only after this adjustment period is it possible to precisely evaluate the evaporator icing based on the process values, in particular the calculated dew point temperature.
  • the temperature difference value for triggering defrosting as a criterion for detecting icing of the evaporator with a time-dependent tolerance, which causes the temperature difference value to increase over time during the transient process of the process temperatures.
  • a time-dependent tolerance which causes the temperature difference value to increase over time during the transient process of the process temperatures.
  • the course of the tolerance as a time function follows a 1/x characteristic curve, so that the tolerance decreases in inverse proportion to the current time and approaches the temperature difference parameter.
  • defrosting is triggered depending on the operating situation of the refrigeration circuit, always at an evaporation temperature that is reduced in proportion to the cooling capacity in relation to the uniced state of the evaporator.
  • This parametrically adjustable, operating state-dependent reduction in the evaporation temperature triggers defrosting at a relative degree of icing, which is relatively independent of the operating state of the refrigeration circuit and, regardless of the operating state, little different defrosting cycle times under the same ambient conditions.
  • the proportional share of the cooling capacity is included in a parameterizable exponent, which increases successively for the curves 5010, 5020, 5030 and 5040.
  • the exponent is equal to 0 so that complete independence from the cooling capacity can be set for calculating the temperature difference for triggering the defrost.
  • the exponent of the performance correction is equal to 1, so that a complete cooling capacity dependence (proportionality) is set for the calculation of the temperature difference for triggering the defrost.
  • the exponents of the performance correction are between 0 and 1, so that there is a gradual cooling performance dependence (proportionality) between the temperature difference for defrost detection and the cooling performance.
  • the choice of the exponent and thus the influence of the cooling capacity is preferably made depending on the specific compression refrigeration system.
  • Fig. 7a There is also a minimum and a maximum value for the temperature difference for triggering the defrost.
  • the temperature difference is limited to at least 2 K and in an area 5055 the temperature difference of the curve 5040 is limited to a maximum of 4 K.
  • Fig. 7b equals to Fig. 7a , whereby in an area of low cooling output 5060 threshold values for the temperature difference of less than 2 K are also permitted.
  • the filtered dew point temperature is compared with the dew point temperature reference.
  • the need for defrosting is preferably recognized if the following condition is continuously met for longer than a predetermined period of time, for example one minute a) the sum of the (corrected, filtered and/or outside temperature compensated) dew point temperature and the specific temperature difference defrost trigger is smaller than the dew point temperature reference and b) the (corrected, filtered and/or outside temperature compensated) dew point temperature is smaller than a parameter for enabling defrost detection.
  • the defrosting process is completed when the high pressure of the high-pressure sensor is greater than a parameterized defrost end limit pressure.
  • a defrost time program is preferably superimposed, which monitors a parameterizable minimum defrost cycle depending on the outside temperature.
  • the evaporator can become so heavily icy even before the defrost is triggered according to the time program that the low-pressure switch is triggered. If the low pressure switch responds several times within a set period of time, this can cause the heat pump to shut down, which should be avoided. The response of the low pressure monitor can therefore directly trigger the defrosting process to be carried out.
  • the triggering of a refrigeration circuit data-based defrost is subject to a high degree of influence by static disturbance variables such as a) inaccuracy in the calculation of the cooling capacity compensation b) inaccuracy in the compensation of the fan power influence c) inaccuracy in the compensation of the evaporation temperature influence d) inaccuracy in the compensation of the overheating influence; as well as a high degree of influence by dynamic processes such as e) Compressor start with oscillation of the superheat controller / settling of the process temperatures f) Compressor speed changes g) Operating point changes when switching from heating mode to hot water charging mode
  • a disruptive influence can cause incorrect calculation of the defrost reference - differential temperature as well as the filtered differential temperature between the outside temperature and the dew point temperature, which leads to premature defrosting (without ice formation).
  • the blocking of a defrost triggering is implemented in a period of, for example, 10 minutes after the compressor starts / the end of defrosting.
  • an additional defrost lock can be parameterized, which is based on the period of time that is calculated for a defrost to be triggered according to the time program.
  • the defrost triggering according to the time program is a higher-level safety function that occurs in the event of failure, in particular a faulty non-triggering despite a sufficient amount of ice, a refrigeration circuit-based defrost detection according to the invention forces defrosting.
  • the running time of a timed defrost trigger is set parametrically so long that a regular defrost usually takes place before a timed defrost is triggered.
  • the model-based calculation of the defrost triggering according to the time program is based approximately on assumed realistic operating conditions of the heat pump, this calculation can also be used for a model-based suppression of premature defrosts due to interference.
  • a time period relative to the defrost triggering period is calculated, in which icing below a level for defrost triggering is assumed. If a refrigeration circuit data-based defrost request occurs within this period of time, defrost triggering is suppressed.
  • the compressor speed is preferably limited to a parameterizable maximum value before switching to defrosting mode.
  • the time period used to reduce the compressor speed before initiating a defrost is set by a lead time parameter.
  • the refrigerant filling level in the tube/fin heat exchanger is high in defrosting mode, the refrigerant only condenses in small areas of the heat exchanger, which are then heated accordingly, while the remaining areas in which there is liquefied refrigerant are hardly heated, so that the ice deposit only forms there is insufficiently melted.
  • the refrigerant outlet temperature from the tube/fin heat exchanger in defrosting mode helps to detect complete ice melting, as this can only reach temperature values greater than 0 °C - assuming there is a uniform flow - when there is no ice at all.
  • the link between pressure-based defrost end detection and temperature-based defrost end detection occurs when a corresponding temperature sensor is activated via sensor configuration as an "and" link. If one of the triggering methods is undesirable, the corresponding parameter must be set to a value which is equivalent to masking the corresponding method.
  • the sensor is configured or enabled to record the evaporator inlet temperature and the error status for this sensor value is inactive, then inclusion of the evaporator inlet temperature in the calculation is enabled and the following applies: If the operating mode is "Defrost mode" and the high pressure exceeds the limit pressure for the end of defrost and the current defrost duration is greater than the value set with the process variable Minimum defrost duration and the evaporator inlet temperature is greater than the Limit temperature defrost end parameter, then the defrost operation is ended.
  • the sensor for detecting the evaporator inlet temperature is not configured (enabled) or the error status for this sensor value is active, then inclusion of the evaporator inlet temperature in the calculation is not enabled and the following applies: If the operating mode is "defrost mode" and If the high pressure exceeds the limit pressure for the end of defrosting and the current defrosting duration is greater than the value set with the process variable Minimum defrosting duration, then the defrosting operation is ended.

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  • Engineering & Computer Science (AREA)
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  • General Engineering & Computer Science (AREA)
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Claims (9)

  1. Procédé de régulation d'un processus de dégivrage d'un évaporateur (240) d'une installation de réfrigération à compression (100), l'installation de réfrigération à compression (100) présentant :
    - un circuit frigorifique (200) avec du fluide frigorigène,
    - un évaporateur (240) qui présente un ventilateur et qui est conçu pour la transmission de la chaleur de l'air au fluide frigorigène,
    - un compresseur (210),
    - une unité de régulation (500) destinée au réglage d'une puissance frigorifique souhaitée qui est définie comme transfert de chaleur dans l'évaporateur,
    le procédé présentant les étapes suivantes consistant à :
    a) mesurer une pression de sortie d'évaporateur,
    b) déterminer une température de condensation sur la base de la pression de sortie d'évaporateur,
    c) corriger la température de condensation déterminée par compensation d'influences perturbatrices,
    d) déterminer une différence entre la température de condensation et une valeur de référence de température de condensation,
    e) engager une opération de dégivrage si la différence dépasse une valeur limite de température,
    caractérisé en ce que l'étape de correction de la température de condensation déterminée comporte la compensation d'influences perturbatrices à l'aide d'au moins une des grandeurs d'influence basse pression (BP), puissance frigorifique, surchauffe de sortie d'évaporateur et/ou puissance de ventilation.
  2. Procédé selon la revendication 1, dans lequel la température de condensation est formée à partir d'une différence entre une première température de condensation et une température extérieure de l'air extérieur circulant dans l'évaporateur (240), la première température de condensation étant calculée à partir de la pression de sortie d'évaporateur et/ou la température extérieure étant mesurée.
  3. Procédé selon l'une des revendications précédentes, dans lequel la température de condensation est filtrée, en particulier filtrée par un filtre passe-bas et de manière particulièrement préférée par un filtre passe-bas du premier ordre.
  4. Procédé selon l'une des revendications précédentes, dans lequel la valeur de référence de température de condensation est la moyenne de température de condensation maximale sur une période déterminée.
  5. Procédé selon l'une des revendications précédentes, dans lequel l'étape de la correction de la température de condensation déterminée comporte au moins une, de préférence plusieurs et de manière particulièrement préférée toutes les étapes suivantes consistant à :
    - corriger la température de condensation au moyen d'un rapport paramétrable, en particulier linéaire, entre une modification de la puissance frigorifique et une modification de la température d'évaporation, en particulier après le calcul d'une température du point de rosée obtenue à une puissance frigorifique de référence paramétrable,
    - compenser l'influence de la surchauffe de sortie d'évaporateur sur la température du point de rosée obtenue en basse pression, en particulier compenser un rapport non linéaire entre une modification de la surchauffe et la modification de la différence de température entre la température de source de chaleur et la température de condensation,
    - compenser l'influence de la vitesse relative du ventilateur sur la température du point de rosée obtenue à basse pression et/ou
    - compenser une non-linéarité de la puissance frigorifique par un facteur exponentiel de la puissance frigorifique du compresseur avec comme exposant un paramètre de correction.
  6. Procédé selon l'une des revendications précédentes, dans lequel le processus de dégivrage comprend les étapes suivantes consistant à :
    e) déterminer une seconde différence entre la pression de sortie d'évaporateur et une pression de coupure
    f) terminer le processus de dégivrage si la seconde différence passe en dessous d'une valeur limite de pression.
  7. Installation de réfrigération à compression (100), comportant :
    - un circuit frigorifique avec du fluide frigorigène,
    - un évaporateur (240) qui présente un ventilateur et qui est conçu pour la transmission de la chaleur de l'air au fluide frigorigène,
    - un compresseur (210),
    - une unité de régulation (500) destinée au réglage d'une puissance frigorifique souhaitée qui est définie comme transfert de chaleur dans l'évaporateur (240),
    l'unité de régulation étant en outre conçue pour :
    a) mesurer une pression de sortie d'évaporateur,
    b) déterminer une température de condensation sur la base de la pression de sortie d'évaporateur,
    c) corriger la température de condensation déterminée par compensation d'influences perturbatrices,
    d) déterminer une différence entre la température de condensation et une valeur de référence de température de condensation,
    e) engager une opération de dégivrage si la différence dépasse une valeur limite de température,
    caractérisée en ce que l'étape de correction de la température de condensation déterminée comporte la compensation d'influences perturbatrices à l'aide d'au moins une des grandeurs d'influence basse pression, puissance frigorifique, surchauffe de sortie d'évaporateur et/ou puissance de ventilation.
  8. Installation de réfrigération à compression selon la revendication 7, dans laquelle le fluide frigorigène présente un glissement de température, est en particulier un mélange de R32 et de R1234yf et est de manière particulièrement préférée du R454C.
  9. Pompe à chaleur air-eau dotée d'une installation de réfrigération à compression selon la revendication 7 ou 8.
EP21177570.5A 2020-06-09 2021-06-03 Procédé de régulation d'un processus de dégivrage d'un évaporateur d'une installation de réfrigération à compression et installation de réfrigération à compression Active EP3922926B1 (fr)

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EP1355207A1 (fr) 2002-04-16 2003-10-22 Otto Egelhof GmbH & Co. Procédé de fonctionnement pour un système frigorifique à compression et système frigorifique à compression
EP1775533B1 (fr) * 2005-10-13 2018-03-28 STIEBEL ELTRON GmbH & Co. KG Procédé pour faire fonctionner un système frigorifique à compression
DE102005054101B3 (de) 2005-11-12 2007-03-01 Voith Turbo Gmbh & Co. Kg Abwärtsfördernde Förderanlage
DE102012208819B4 (de) 2012-05-25 2018-08-16 Honeywell Technologies Sarl Verfahren für die steuerung und regelung von kälteanlagen und wärmepumpen mit luftbeaufschlagtem verdampfer

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