EP3922931B1 - Installation de réfrigération à compression et procédé de fonctionnement de celle-ci - Google Patents

Installation de réfrigération à compression et procédé de fonctionnement de celle-ci Download PDF

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EP3922931B1
EP3922931B1 EP21177574.7A EP21177574A EP3922931B1 EP 3922931 B1 EP3922931 B1 EP 3922931B1 EP 21177574 A EP21177574 A EP 21177574A EP 3922931 B1 EP3922931 B1 EP 3922931B1
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EP
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Prior art keywords
refrigerant
compressor
evaporator
superheat
value
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German (de)
English (en)
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EP3922931A1 (fr
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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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    • 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
    • F25B2339/00—Details of evaporators; Details of condensers
    • F25B2339/04—Details of condensers
    • F25B2339/047—Water-cooled condensers
    • 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
    • F25B2700/21—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, an internal heat exchanger, a throttle device 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 means of the control unit, which for example has a regulator, are known.
  • the refrigerant is driven through the condenser, where it gives off heat to a heating medium located in a heat sink system.
  • Internal heat is transferred in an internal heat exchanger, for example in the form of a recuperator, between the refrigerant flowing under high pressure from the condenser to the expansion valve and the refrigerant flowing under low pressure from the evaporator to the compressor.
  • the refrigerant is then guided in a high-pressure flow direction to an expansion valve controlled by the regulator, in which the refrigerant is expanded from high pressure to low pressure depending on a control value.
  • the refrigerant at low pressure evaporates in the evaporator by absorbing source heat.
  • EP 1 014 013 A1 discloses a vapor compression type 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 serving to adjust the flow rate of a liquid phase portion of the refrigerant supplied to the evaporator in accordance with a control signal to maintain a degree of superheat of a gas phase portion of the refrigerant supplied to a compressor.
  • Thermistors detect temperatures of the gas phase portion and an evaporated portion of the 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 the refrigerant flowing 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 refrigerant condition value to supply 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 the refrigerant on the outlet side of an evaporator.
  • a second refrigerant temperature detector is attached to a pipe for detecting the temperature of the refrigerant on the inlet side of a compressor.
  • a switching controller is connected to the first and second refrigerant temperature detectors and a superheat control valve, and changes/selects one of the refrigerant temperature detection value signals from the first and second refrigerant 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.
  • the refrigeration circuit comprises a refrigerant circuit including a compressor, a condenser, an electronically controlled expansion valve and an evaporator, an internal heat exchanger, a pressure sensor that detects the pressure of a refrigerant between the condenser and the internal heat exchanger, a first temperature sensor that detects the temperature of the evaporator or a second temperature sensor that has a measuring point between the evaporator and the internal heat exchanger, and a valve opening controller, wherein the valve opening controller uses as the superheat degree the value of a calculated superheat degree that is calculated based on parameters such as the detection value of the pressure sensor and the detection value of the first temperature sensor or the second temperature sensor.
  • recuperator in a refrigeration machine, especially in a heat pump, to increase the heating output in a structurally simple manner at low outside temperatures.
  • the recuperator is dimensioned in such a way that at low evaporation temperatures it can at least 15% of the heat pump's heating capacity is transferred from the liquid refrigerant to the gaseous refrigerant.
  • An injection valve injects liquid refrigerant into the compressor so that the final compression temperature remains below 120 °C.
  • a heat pump system with a refrigerant circuit is made of EN 10 2005 061 480 B3 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 in such a way that the first heat exchanger serves to liquefy the refrigerant in the first operating mode and to evaporate the refrigerant in the second operating mode, and the second heat exchanger serves to evaporate the refrigerant in the first operating mode and to liquefy the refrigerant in the second operating mode, wherein the first heat exchanger in the refrigerant circuit is connected in such a way that it works as a counterflow heat exchanger in the two operating modes heating and cooling.
  • the control of the compression refrigeration system must meet various requirements, for example, the coefficient of performance must be as high as possible in order to allow the most energy-efficient operation possible. However, it is also crucial that the operating limits of the components are adhered to.
  • the compressor is particularly important, as it compresses the gaseous refrigerant from low pressure (LP) to high pressure (HD). It is known that compressors that allow operation at different speeds can be used to optimize performance. In addition to other operating limits, the manufacturers of such variable-speed compressors stipulate, for example, that both upper and lower limits for the low pressure (LP) and high pressure (HD) must be observed depending on the speed. If the limits are exceeded on either the low pressure side or the high pressure side, the compressor must be switched off in the worst case scenario.
  • LP low pressure
  • HD high pressure
  • the dew point temperature of the refrigerant is pressure-dependent and can be tabulated or calculated in the form of characteristic curves, called wet vapor characteristic curves.
  • a low pressure in the evaporator can be regulated, for example by setting the opening degree of a throttle device.
  • a control value is influenced in a commissioning phase of the vapor compression system depending on a control deviation of an evaporator outlet superheat, which is used to control the expansion valve.
  • the control value is further determined after the commissioning phase, during a run-in operating state of the vapor compression system, depending on a compressor inlet superheat and the expansion valve is controlled after the commissioning phase depending on the determined evaporator outlet superheat and the compressor inlet superheat.
  • a controlled variable for the evaporator outlet superheat is calculated.
  • a control deviation for the evaporator outlet superheat is calculated using a target value for the evaporator outlet superheat.
  • a controlled variable for the compressor inlet superheat is calculated.
  • a control deviation for the compressor inlet superheat is calculated using a target value for the compressor inlet superheat.
  • the control value R is calculated from a weighted influence of the control deviation for the evaporator outlet superheat and a weighted influence of the control deviation for the compressor inlet superheat.
  • the expansion valve is controlled using the control value.
  • the compressor inlet superheat as well as the evaporator outlet superheat can be included proportionally to control the superheating of the refrigerant.
  • This method enables a faster control reaction to operating point changes in the refrigeration circuit because the reaction of the evaporator outlet superheat to disturbances such as operating mode/operating point changes is up to 10 times faster than the reaction of the compressor inlet superheat, depending on the refrigeration circuit operating point.
  • control deviation of the evaporator outlet superheat is usually included in the control to a greater extent than the control deviation of the compressor inlet superheat in order to optimize the overall response time of the control.
  • one goal is to regulate the superheat at the evaporator outlet as low as possible and to shift the superheat of the refrigerant as completely as possible into the internal heat exchanger.
  • the setpoint for evaporator outlet superheat will always be set approximately close to 0 Kelvin relative to the dew line of the refrigerant.
  • a method for adaptive compensation of the tolerances that are individual for each heat pump but are usually systematic. This means that the tolerances are caused by, for example, component tolerances that are individual for each heat pump but generally do not change on a daily basis.
  • a solution is proposed to design the setpoint values for evaporator outlet superheat and setpoint for evaporator outlet superheat such that in the steady state of the superheat control both the control deviation for the evaporator outlet superheat and the control deviation for the compressor inlet superheat are equally zero.
  • the method for controlling the compression refrigeration system includes the following steps: determining a target value of the evaporator outlet superheat and a target value of the compressor inlet superheat, calculating a correction value based on a control deviation of the compressor inlet superheat from the target value of the compressor inlet superheat, correcting the target value of the evaporator outlet superheat with the calculated correction value, calculating a control value after a commissioning phase of the compression refrigeration system depending on the target value of the evaporator outlet superheat and the target value of the compressor inlet superheat, and controlling the expansion valve on the basis of the control value.
  • the method according to the invention can react to tolerances of the wet steam characteristic curve and control the compression refrigeration system more precisely.
  • Tolerances in the wet steam characteristic curve regularly arise, for example, due to a separation of refrigerant components when filling the compression refrigeration system in a case where a refrigerant composed of several refrigerants is used.
  • a refrigerant known as R454C which is usually a mixture of 21.5% R32 and 78.5% R1235yf, can be used.
  • compression refrigeration systems are filled with the refrigerant, for example from a provided refrigerant container
  • the gravity alone causes a separation during the filling process, which leads to a deviation in the composition of the refrigerant mixture in the compression refrigeration system from the composition in the refrigerant storage container.
  • different mixing ratios of the refrigerant components arise between the compression refrigeration systems.
  • the correction value is calculated proportional to a time integral of the control deviation of the compressor inlet superheat to the target value of the compressor inlet superheat.
  • the correction value is corrected in discrete time steps by a proportional part of the control deviation of the compressor inlet superheat.
  • the correction value is limited to a permissible range of values.
  • the target value of the evaporator outlet superheat is corrected by adding the calculated correction value.
  • the refrigerant has a temperature glide, wherein the refrigerant in particular comprises or consists of R454C, and wherein the compression refrigeration system in particular contains an internal heat exchanger for transferring thermal energy of the refrigerant before entering the throttle element to the refrigerant before entering the compressor.
  • the refrigerant in particular comprises or consists of R454C
  • the compression refrigeration system in particular contains an internal heat exchanger for transferring thermal energy of the refrigerant before entering the throttle element to the refrigerant before entering the compressor.
  • the object is further achieved according to the invention by a compression refrigeration system and 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 location of installation.
  • a data connection 510 which can be made by 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.
  • a Fig.1 Evaporator inlet temperature sensor (not shown) determines the temperature at the evaporator inlet 241.
  • the heat pump 100 is shown as a brine heat pump.
  • a fan/ventilator 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 variable speed.
  • the compression unit can be designed as a rotary piston unit, scroll unit or otherwise.
  • the compressed superheated refrigerant is at the compressor outlet pressure P Va at a higher pressure level, 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 compressed inlet temperature T VE , which describes the state of the refrigerant temperature at the inlet connection 211 when entering a compression chamber.
  • the transfer of thermal energy Q H from the refrigerant of the vapor compression system 200 to a heating medium of the heat sink system 400 takes place.
  • the refrigerant is deheated in the condenser 220, with superheated refrigerant vapor transferring part of its thermal energy to the heating medium of the heat sink system 400 by reducing its temperature.
  • a further heat transfer Q H advantageously takes place in the condenser 220 through 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 established in the condenser 220 during operation of the compressor 210 corresponds approximately to a condensation pressure of the refrigerant at a heating medium temperature Tws in the heat sink system.
  • the heating medium in particular water, is pumped by means of a heating medium pump 410 through the heat sink system 400 in a direction SW through the condenser 220, whereby the thermal energy Q H is transferred from the coolant to the heating medium.
  • the subsequent collector 260 stores coolant exiting from the condenser 220, which should not be fed into the circulating coolant depending on the operating point of the vapor compression circuit 200. If more coolant is fed in from the condenser 220 than is passed on through the expansion valve 230, the collector 260 fills up, otherwise it becomes emptier or emptied.
  • 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 the refrigerant flowing under the low pressure ND, which flows from the evaporator to the compressor in a low pressure flow direction S ND flows.
  • 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, whereby the low pressure advantageously corresponds approximately to a suction pressure of the compressor 210.
  • any other desired pressure reduction device can also be used. Pressure reduction pipes, turbines or other expansion devices are advantageous.
  • the degree of opening of the expansion valve 230 is set by an electric motor, which is usually designed as a stepper 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 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 at the heat source medium temperature T WQ .
  • the evaporation temperature of the refrigerant will advantageously be a few Kelvin below the heat source medium temperature T WQ so that the temperature difference drives heat transfer.
  • a transfer of evaporation heat energy Qv takes place 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 releases the source energy Q Q to the vapor compression system 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 releases the source energy Q Q to the vapor compression system 200.
  • the coolant flowing into the evaporator 240 reduces its wet vapor content as it flows through the evaporator 240 by absorbing heat Q Q and leaves the evaporator 240 advantageously with a low wet vapor content or advantageously also as a superheated gaseous coolant.
  • 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, whereby the heat energy Q Q is extracted from the heat source medium as it flows through the evaporator.
  • thermal 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, wherein the refrigerant flowing from the evaporator 240 to the compressor 210 in particular continues to superheat.
  • This superheated refrigerant which exits the recuperator 250 with a superheat temperature T Ke , is led to the refrigerant inlet connection 211 of the compressor 210.
  • the recuperator 250 is used in the vapor compression circuit 200 to increase the overall efficiency as a quotient of the delivered heating power Q H and the absorbed electrical power P e to drive the compressor motor.
  • the internal energy state of the refrigerant upon entering the evaporator 240 is reduced by this heat removal 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 then fed back to the refrigerant after the evaporator outlet 242 from the evaporator 240 in the low-pressure path at low pressure ND and at a low-pressure temperature corresponding to an evaporator outlet temperature T Va in the recuperator 250.
  • the supply of energy advantageously causes a reduction in the wet steam portion to a state without a wet steam portion and then superheating occurs through further energy supply.
  • the following sensors are advantageously arranged to detect the operating state of the vapor compression system 200, with which a model-based feedforward control is implemented, in particular to safeguard and optimize the operating conditions of the vapor compression system 200, in particular in the event of operating state changes.
  • the process variable which 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 an electrical power P e absorbed by the compressor 210 is the superheat of the refrigerant at the compressor inlet 211.
  • the superheat describes the temperature difference between the measured compressor inlet temperature T KE of the refrigerant and the evaporation temperature of the refrigerant at saturated vapor.
  • the compressor inlet superheat is controlled in such a way that no condensate is formed on components of the refrigeration circuit due to the water vapor content in the ambient air falling below the dew point, particularly 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 superheat is 15K.
  • room temperature sensors and room humidity sensors are advantageous, as they enable the condensation conditions of the air to be determined precisely. For example, at 21°C and 60% relative humidity, the condensation temperature is in the range of 13°C. Under these conditions, as long as the pipe temperature is above 13°C plus a buffer if necessary, e.g. 1K, no condensation takes place.
  • the numerical example which is of course not restrictive, is used to achieve a superheat of 15K at a compressor inlet temperature of 5°C. This temperature is below the 13°C that is determined for the current ambient conditions as the condensation temperature of the water vapor in the ambient air. Condensation therefore takes place. If the compressor inlet temperature is to be at least 14°C, i.e. condensation temperature plus buffer, the superheat must be increased by 9K, i.e. a superheat of 24K must be maintained.
  • Limit values in particular for superheating, determine the permissible superheating range of the components at the compressor inlet 211 depending on the operating point. Furthermore, there are also dependencies between the compressor inlet superheat dT ÜE and the overall efficiency of the vapor compression circuit 200 or between the compressor inlet superheat dTü ⁇ and a stability S of a control value R advantageous when regulating the compressor inlet superheat.
  • the heat source medium temperature, the heating medium temperature, the compressor output P e and target values Z or the target value Z for calculating the compressor inlet superheat dTü ⁇ are advantageously used.
  • the target value Z can be calculated as a default value for the compressor inlet superheat from the refrigeration circuit measured variables that depend on the operating point, such as the heat source medium temperature, the heating medium temperature, the compressor output P e and parameterizable coefficients that are adapted to the behavior of the respective refrigeration circuit components.
  • dTü ⁇ the target value for the compressor inlet superheat dTü ⁇ is constant, e.g. 10 Kelvin, regardless of all operating conditions. In a more complex adaptation, it is varied as a function of an operating point variable, e.g. the compressor power P e , or in an even more complex adaptation, it varies as a function of several operating point variables.
  • the total control deviation is 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 into the control of the vapor compression circuit 200.
  • the refrigerant after being released through the expansion valve 230, 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 the ground or the outside air.
  • thermal energy Q i is extracted from the refrigerant after it leaves the condenser 220.
  • the temperature level of the refrigerant at the outlet of the condenser is approximately equal to 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.
  • Actuator operating state variables with an influence on the control value R, in particular the compressor inlet superheat dTü ⁇ , are the compressor speed and/or the degree of opening of the expansion valve 230 in the relevant vapor compression circuit 200, which also advantageously determines the low pressure ND and the evaporation temperature level.
  • Actuators have a particularly advantageous influence on the control value R, in particular on the weighted link between the control deviation of the compressor inlet superheat and the control deviation of the evaporator outlet superheat.
  • 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 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 to regulate R, which accompanies the change in compressor speed, is advantageous, and may even be necessary.
  • the compressor speed in the vapor compression circuit 200 is set such that the heating power QH transferred 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 subordinate 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 is as follows:
  • the expansion valve 230 acts as a nozzle with an electric motor-adjustable nozzle cross-section, in which a needle-shaped nozzle needle is usually threaded into a nozzle seat by means of a stepper 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, 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 drops.
  • the mass flow of refrigerant through the compressor 210 drops approximately proportionally, since its delivery capacity can be approximately described as volume / time, 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. Since the compressor 210 continues to initially deliver the same refrigerant mass flow, more refrigerant is supplied to the low pressure side ND of the refrigeration circuit through the expansion valve 230 than is sucked off by the compressor 210. Since the refrigerant vapor is a compressible medium, the low pressure ND on the low pressure side of the vapor compression circuit 200 increases.
  • the mass flow rate of the compressor 210 increases approximately proportionally, since its delivery rate can be approximately described as volume / time, and a correspondingly increased low pressure ND is set in, 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, whereby the heat flow Q Q is dependent 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 is assumed to be approximately constant in a respective vapor compression circuit 200. Therefore, the magnitude 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.
  • the temperature of the heat source medium in as many surface elements of the transfer layer of the heat exchanger, here the evaporator 240, as possible is greater than the temperature of the coolant at the respective surface element.
  • 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.
  • an evaporation pressure can indirectly control the evaporation temperature of the refrigerant as it flows through the recuperator 250.
  • the thermal energy Q Q which is transferred from the heat source system to the refrigerant flowing through the evaporator 240, influences the state of aggregation of the refrigerant.
  • recuperator 250 For complete evaporation, additional energy is supplied in the recuperator 250 to superheat the refrigerant beyond the state of saturated vapor.
  • a corresponding refrigerant state is set at the outlet from the evaporator 240 under given operating conditions of the vapor compression circuit 200 as a function of the manipulated variable “opening degree of expansion valve 230”.
  • control system behavior of the "isolated" control system "evaporator 240" has a moderate slope.
  • the control system behavior is particularly characterized by the control system output value of the evaporator outlet superheat as a function of the control system input value of the expansion valve opening degree.
  • a refrigerant in particular a refrigerant mixture
  • a refrigerant mixture which has a "temperature glide", in particular R454C is advantageously used.
  • a superheat change of less than 1 K is usually set at the outlet of the refrigerant from the evaporator.
  • a delay in the corresponding change in the state of the refrigerant when exiting the evaporator outlet 242 occurs, and a total time constant Z total is advantageously in the range of 30 seconds to about 5 minutes, depending on the operating point.
  • the refrigerant After flowing through the evaporator 240, the refrigerant enters the low pressure path of the recuperator 250 at low pressure ND.
  • the state of aggregation of the refrigerant when flowing into the recuperator 250 is, in a normal operating case, therefore advantageously either saturated vapor with a low vapor content between 0 and 20% or, particularly advantageously, already superheated refrigerant.
  • a refrigerant temperature is established which is a function of the refrigerant pressure due to the saturation steam characteristic of the refrigerant.
  • the refrigerant temperature will assume a maximum value which corresponds to the inlet temperature of the heat source medium.
  • the value preferably corresponds to the inlet temperature of the refrigerant in the high-pressure path of the recuperator 250, i.e. the temperature of the refrigerant after exiting the condenser 220.
  • 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, has an effect on the state of aggregation of the refrigerant on the low-pressure side.
  • the wet vapor portion of the refrigerant flowing through the recuperator 250 on the low-pressure side at low pressure ND decreases when heat is transferred to the refrigerant and after complete evaporation, the refrigerant is advantageously overheated.
  • a significantly higher heat transfer is set in the evaporator 240 than in the recuperator 250, since a significantly greater amount of energy is to be extracted from the environment by means of the evaporator 240 than is transferred within the refrigeration circuit in the recuperator 250 alone.
  • the driving temperature difference in the recuperator is, for example, between 20 K and 60 K, while in the evaporator it is only between 3 K and 10 K.
  • the exchange surface of the evaporator 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 total after changing the manipulated variable "expansion valve opening degree".
  • the time behavior of the recuperator 250 can advantageously be taken into account as the total recuperator time constant Z total 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 compressor inlet superheat dTü ⁇ and evaporator outlet superheat dT ÜA is advantageously carried out by calculating the total control deviation, which is fed into the controller 500 for controlling the vapor compression circuit 200, in particular by means of a weighted combination of the control deviation of the compressor superheat and the control deviation of the evaporator outlet superheat dT ÜA .
  • Step 1 First, the process variables compressor inlet superheat dT ÜE are advantageously measured as the main control variable and the evaporator outlet superheat dT ÜA are advantageously measured as the auxiliary variable in a first process step.
  • the refrigerant temperature is recorded by means of temperature sensors 501, 508 at the respective superheat measuring point, in particular at the evaporator outlet 242 and/or at the compressor inlet 211.
  • the temperature difference of the refrigerant at the respective measuring point and the evaporation temperature is then calculated and this temperature difference value then corresponds to the respective superheat of the refrigerant at the measuring point.
  • the starting values 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 offset in a second step to form assigned control deviations with respective assigned setpoints:
  • the setpoint for the compressor inlet superheat dTü ⁇ is advantageously varied in the range between approx. 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 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 the evaporator superheat in the steady state control case 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 precalculated on a model-based basis depending on an operating mode or an operating point, depending on the evaporation temperature, the condensation temperature, the compressor output, a setpoint for the compressor inlet superheat dTü ⁇ at the evaporator outlet 242 and/or component properties and can be adaptively corrected.
  • the control deviation of the compressor inlet superheat dT ÜE is then calculated by subtracting the setpoint value of the compressor inlet superheat dTü ⁇ from the process value of the compressor inlet superheat dT ÜE .
  • the control deviation of the evaporator outlet superheat dT ÜA is then calculated by subtracting the setpoint value of the evaporator outlet superheat dT ÜA from the process value of the evaporator outlet superheat dT ÜA .
  • Step 3 In a third process step, the control deviation of the compressor inlet superheat dT ÜE and the control deviation of the evaporator outlet superheat dT ÜA are advantageously combined to form a total control deviation superheat.
  • 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 proportional combination of the individual control deviations and can, in extreme cases, result in the exclusive inclusion of only 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 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 opening degree and/or the compressor 210 with adjustable compressor speed, such that in the regulated case a control deviation of the superheat is set equal to approximately 0 Kelvin.
  • a P, I, PI, PID controller can be used, whereby the control components are advantageously dynamically adapted to the respective operating mode and the operating point.
  • the recuperator 250 between the refrigerant path for subcooled refrigerant after exiting the heat sink-side heat exchanger, here the condenser 220, and the refrigerant path after exiting the heat source-side heat exchanger, here the evaporator 210, and the compressor inlet 211, the compressor inlet superheat T ÜE as well as the evaporator outlet superheat T ÜA can be included proportionally to regulate the superheating of the refrigerant.
  • this is implemented by correcting the setpoint value for the evaporator outlet superheat.
  • the control deviation of the compressor inlet superheat i.e. the difference between the setpoint superheat of the compressor inlet and the actual superheat of the compressor inlet, is used as the input variable for the correction.
  • a compensation variable for the temperature difference adaptation is calculated by using a time function to make the value of the adaptation time constant take on the inverse value of the control deviation of the compressor inlet superheat.
  • the adaptation time constant together with the time function, determines the duration within which the temperature difference adaptation follows the control deviation.
  • Other forms of filtering rapid changes in the control deviation such as low-pass filters, are also conceivable.
  • Fig.3 shows schematically and as an example three wet steam characteristic curves 1010, 1020 and 1030 for different mixing ratios of refrigerant components, in this example of R32 and R1234yf.
  • the wet steam characteristic curve 1020 corresponds in this example to the wet steam characteristic curve of R454C.
  • the proportion of R32 is reduced compared to R454C, while it is increased for the wet steam characteristic curve 1030.
  • the three wet steam characteristic curves 1010, 1020 and 1030 are therefore typical mixture values obtained from filling a compression refrigeration system from a container with R454C.
  • the dew point temperature as a function of the pressure for lower proportions of R32 see wet steam characteristic curve 1010
  • the dew point temperature for higher proportions of R32 see wet steam characteristic curve 1020, 1030.
  • these tolerances of the wet steam characteristic curves can be compensated by correcting and including the correction value.
  • the temperature difference adaptation can be calculated in two consecutive steps.
  • an unlimited new value of the temperature difference adaptation evaporator output superheat setpoint (unlimited) is calculated using the value of the temperature difference correction evaporator output superheat setpoint calculated in the last loop run.
  • this value is limited to the parameterizable range limit and then processed further as a newly calculated process value.
  • the temperature difference adaptation evaporator output superheat setpoint is preferably adapted in such a way that it would assume the value of the control deviation of the superheat at the compressor inlet in a specified time, called the adaptation time constant evaporator output superheat setpoint.
  • the range of the temperature difference correction evaporator outlet superheat setpoint is limited to the range set with an adaptation range parameter by limiting the previously calculated value to the range +/- adaptation range parameter (in Kelvin).
  • Fig.4 shows schematically and exemplarily a curve 2010 of the control deviation of the compressor inlet superheat and a curve 2020 of the correction of the evaporator outlet superheat setpoint resulting from the control deviation over time in minutes on the horizontal axis.
  • a control deviation that has deviated in the opposite direction.
  • the value of the adaptation increases gradually to counteract the control deviation.
  • the slope of the adaptation 2020 in the time range 2050 is greater than the slope in the time range 2030, since the slope is preferably selected to be proportional to the value of the control deviation.
  • the correction remains constant in a time range 2070, since the control deviation in this time range is zero.

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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)
  • Heat-Pump Type And Storage Water Heaters (AREA)
  • Compression-Type Refrigeration Machines With Reversible Cycles (AREA)

Claims (8)

  1. Procédé de régulation d'une installation de réfrigération à compression (200), en particulier une pompe à chaleur (100), dans laquelle
    - un fluide frigorigène à l'état gazeux est comprimé d'une basse pression (ND) à une haute pression (HD) par un compresseur (210) commandé par un régulateur (500),
    - le fluide frigorigène traverse un condenseur (220) dans lequel il transfère de la chaleur de chauffage (Qh) à un fluide de chauffage se trouvant dans un système dissipateur de chaleur (400),
    - une chaleur interne (Qi) du fluide frigorigène est échangée dans un récupérateur (250) entre le fluide frigorigène circulant à haute pression (HD) depuis le condenseur (220) vers le détendeur (230) et le fluide frigorigène circulant à basse pression (ND) depuis l'évaporateur (240) vers le compresseur (210),
    - le fluide frigorigène est guidé dans une direction d'écoulement à haute pression (SHD) jusqu'à un détendeur (230) commandé par le régulateur (500) et dans lequel le fluide frigorigène est détendu de la haute pression (HD) à la basse pression (ND) de manière régulée en fonction d'une valeur de régulation (R), sachant que
    - le fluide frigorigène à basse pression ND s'évapore dans l'évaporateur (240) en absorbant de la chaleur source QQ,
    comportant les étapes de procédé :
    - détermination d'une valeur cible (ZÜA) de la surchauffe de sortie d'évaporateur (TÜA) et d'une valeur cible (ZÜE) de la surchauffe d'entrée de compresseur (TÜE),
    - calcul d'une valeur de correction en fonction d'un écart de régulation entre la surchauffe d'entrée de compresseur (TÜE) et la valeur cible (ZÜE) de la surchauffe d'entrée de compresseur (TÜE),
    - correction de la valeur cible (ZÜA) de la surchauffe de sortie d'évaporateur (TÜA) avec la valeur de correction calculée,
    - calcul d'une valeur de régulation (R) après une phase de mise en service de l'installation de réfrigération à compression (200), en fonction de la valeur cible (ZÜA) de la surchauffe de sortie d'évaporateur (TÜA) et de la valeur cible (ZÜE) de la surchauffe d'entrée de compresseur (TÜE), et
    - régulation du détendeur (230) en fonction de la valeur de régulation (R).
  2. Procédé selon la revendication 1, dans lequel la valeur de correction est calculée proportionnellement à une intégrale dans le temps de l'écart de régulation entre la surchauffe d'entrée de compresseur (TÜE) et la valeur cible (ZÜE) de la surchauffe d'entrée de compresseur (TÜE).
  3. Procédé selon la revendication 2, dans lequel la valeur de correction est corrigée en mettant en oeuvre une part proportionnelle de l'écart de régulation de la surchauffe d'entrée de compresseur (TÜE) par pas de temps discrets.
  4. Procédé selon l'une des revendications précédentes, dans lequel la valeur de correction est limitée à une plage de valeurs autorisées.
  5. Procédé selon l'une des revendications précédentes, dans lequel la valeur cible (ZÜA) de la surchauffe de sortie d'évaporateur (TÜA) est corrigée en ajoutant la valeur de correction calculée.
  6. Installation de réfrigération à compression (200), en particulier pompe à chaleur (100), comportant
    - un circuit frigorifique comportant du fluide frigorigène,
    - un compresseur (210),
    - un condenseur (220),
    - un évaporateur (240),
    - un détendeur (230),
    - un échangeur de chaleur interne (250), dit récupérateur, et
    - un régulateur (500),
    dans laquelle, lors du fonctionnement de l'installation de réfrigération à compression (200),
    - du fluide frigorigène à l'état gazeux est comprimé d'une basse pression (ND) à une haute pression (HD) par le compresseur (210) commandé par le régulateur (500),
    - le fluide frigorigène traverse le condenseur (220), le condenseur (220) étant conçu de telle manière que le fluide frigorigène transfère, dans le condenseur (220), une chaleur de chauffage (Qh) à un fluide de chauffage se trouvant dans un système dissipateur de chaleur (400),
    - une chaleur interne (Qi) du fluide frigorigène est échangée dans un récupérateur (250) entre le fluide frigorigène circulant à haute pression (HD) depuis le condenseur (220) vers le détendeur (230) et le fluide frigorigène circulant à basse pression (ND) depuis l'évaporateur (240) vers le compresseur (210),
    - le fluide frigorigène est guidé dans une direction d'écoulement à haute pression (SHD) jusqu'à un détendeur (230) commandé par le régulateur (500) et dans lequel le fluide frigorigène est détendu de la haute pression (HD) à la basse pression (ND) de manière régulée en fonction d'une valeur de régulation (R), sachant que
    - le fluide frigorigène à basse pression ND s'évapore dans l'évaporateur (240) en absorbant de la chaleur source QQ,
    le régulateur (500) étant conçu pour :
    - déterminer une valeur cible (ZÜA) de la surchauffe de sortie d'évaporateur (TÜA) et une valeur cible (ZÜE) de la surchauffe d'entrée de compresseur (TÜE),
    - calculer une valeur de correction en fonction d'un écart de régulation entre la surchauffe d'entrée de compresseur (TÜE) et la valeur cible (ZÜE) de la surchauffe d'entrée de compresseur (TÜE),
    - corriger la valeur cible (ZÜA) de la surchauffe de sortie d'évaporateur (TÜA) avec la valeur de correction calculée,
    - calculer une valeur de régulation (R) après une phase de mise en service de l'installation de réfrigération à compression (200), en fonction de la valeur cible (ZÜA) de la surchauffe de sortie d'évaporateur (TÜA) et de la valeur cible (ZÜE) de la surchauffe d'entrée de compresseur (TÜE), et
    - réguler le détendeur (230) en fonction de la valeur de régulation (R).
  7. Installation de réfrigération à compression (200) selon la revendication 6, dans laquelle le fluide frigorigène contient un mélange de plusieurs composants de fluide frigorigène.
  8. Installation de réfrigération à compression (200) selon la revendication 7, dans laquelle un mélange contenant du R32 et du R1234yf, en particulier du R454C, est utilisé comme fluide frigorigène.
EP21177574.7A 2020-06-09 2021-06-03 Installation de réfrigération à compression et procédé de fonctionnement de celle-ci Active EP3922931B1 (fr)

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DE3442169A1 (de) 1984-11-17 1986-05-28 Süddeutsche Kühlerfabrik Julius Fr. Behr GmbH & Co KG, 7000 Stuttgart Verfahren zum regeln eines kaeltekreisprozesses fuer eine waermepumpe oder eine kaeltemaschine und eine waermepumpe oder kaeltemaschine hierzu
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
DE19925744A1 (de) 1999-06-05 2000-12-07 Mannesmann Vdo Ag Elektrisch angetriebenes Kompressionskältesystem mit überkritischem Prozeßverlauf
FR2815397B1 (fr) 2000-10-12 2004-06-25 Valeo Climatisation Dispositif de climatisation de vehicule utilisant un cycle supercritique
DE10157461A1 (de) 2001-11-20 2003-05-28 Daimler Chrysler Ag Verfahren zum Betrieb eines Kältemittelkreislaufs und Verfahren zum Betrieb eines Kraftfahrzeugantriebsmotors
DE10159892B4 (de) 2001-12-06 2006-08-24 Stiebel Eltron Gmbh & Co. Kg Kältemaschine mit einem Rekuperator
DE102005061480B3 (de) 2005-12-22 2007-04-05 Stiebel Eltron Gmbh & Co. Kg Wärmepumpenanlage
JP4948374B2 (ja) 2007-11-30 2012-06-06 三菱電機株式会社 冷凍サイクル装置
JP2010032159A (ja) 2008-07-30 2010-02-12 Denso Corp 冷凍サイクル装置
JP2017088137A (ja) * 2015-11-17 2017-05-25 株式会社ヴァレオジャパン 車両用空調装置の冷凍サイクル及びこれを搭載した車両

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