EP3922931A1 - Procédé de fonctionnement d'une installation de réfrigération à compression - Google Patents

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

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Publication number
EP3922931A1
EP3922931A1 EP21177574.7A EP21177574A EP3922931A1 EP 3922931 A1 EP3922931 A1 EP 3922931A1 EP 21177574 A EP21177574 A EP 21177574A EP 3922931 A1 EP3922931 A1 EP 3922931A1
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EP
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Prior art keywords
refrigerant
compressor
value
compressor inlet
evaporator
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EP21177574.7A
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German (de)
English (en)
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EP3922931B1 (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
    • 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 element and a control unit.
  • Such compression refrigeration systems for example in the form of heat pumps, with a vapor compression system in which a gaseous refrigerant is compressed from a low pressure to a high pressure by a compressor controlled by means of the control unit, which for example has a regulator, are known.
  • the refrigerant is driven through the condenser, in which 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 from the evaporator to the compressor under low pressure.
  • the refrigerant is guided further 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 as a function of a control value.
  • the refrigerant at the low pressure evaporates in the evaporator when it absorbs source heat.
  • recuperator in a refrigeration machine, in particular in a heat pump, which is intended to increase the heating power in a structurally simple manner at low outside temperatures.
  • the recuperator is dimensioned in such a way that at low evaporation temperatures it transfers at least around 15% of the heat output of the heat pump from the liquid refrigerant to the gaseous refrigerant.
  • An injection valve injects liquid refrigerant into the compressor so that the compression end temperature remains below 120 ° C.
  • a heat pump system with a refrigerant circuit is off DE 10 2005 061 480 B3 known. It is equipped with a compressor, a first heat exchanger, a throttle element, an evaporator and a 4-2-way valve unit for switching between a first (heating) and a second operating mode (cooling).
  • a direction of flow of the refrigerant in the refrigerant circuit can be switched so that the first heat exchanger is used to liquefy the refrigerant in the first operating mode and to evaporate the refrigerant in the second operating mode, and the second heat exchanger in the first operating mode to evaporate the refrigerant and in the second operating mode is used to liquefy the refrigerant, the first heat exchanger in the refrigerant circuit being connected in such a way that it works as a countercurrent heat exchanger in the two operating modes heating and cooling.
  • the regulation of the compression refrigeration system has to meet various requirements, for example it is required that a coefficient of performance is as high as possible in order to allow the most energy-efficient operation possible. But it is also crucial that the operating limits of the components are observed.
  • the compressor which, as stated, compresses the gaseous refrigerant from low pressure (LP) to high pressure (HP). It is known to provide compressors to optimize performance, which allow operation at different speeds. 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 the high pressure (HP) must be adhered to, depending on the speed. If the limits are exceeded either on the low pressure side or on the high pressure side, the operation of the compressor must be switched off in the worst case.
  • the dew point temperature of the refrigerant is pressure-dependent and can be tabulated or calculated in the form of characteristics, called wet steam characteristics.
  • a low pressure in the evaporator can therefore be regulated, for example, by adjusting the degree of opening of a throttle element.
  • a control value is influenced as a function of a control deviation of an evaporator outlet overheating, with which the expansion valve is controlled.
  • the control value is still determined after the commissioning phase, during a retracted operating state of the vapor compression system, depending on compressor inlet overheating and the expansion valve is controlled after the commissioning phase depending on the determined evaporator outlet overheating and the compressor inlet overheating.
  • a controlled variable for the superheating of the evaporator outlet is calculated. With a target value for evaporator outlet superheating, a control deviation for evaporator outlet superheating is calculated. A controlled variable for compressor inlet superheating is calculated. With A control deviation for compressor inlet superheating is calculated for a target value for compressor inlet superheating.
  • the control value R is calculated from a weighted influence of the control deviation evaporator outlet overheating and a weighted influence of the control deviation compressor inlet overheating.
  • the expansion valve is regulated with the control value.
  • the compressor inlet superheating and the evaporator outlet can be proportionally included to regulate the overheating of the refrigerant.
  • This method enables a faster control-related reaction to changes in the operating point of the refrigeration circuit, because the reaction of the evaporator outlet overheating to disturbance variables such as operating mode / operating point changes, depending on the refrigeration circuit operating point, is up to 10 times faster than the reaction of the compressor inlet overheating.
  • control deviation of the evaporator outlet overheating is proportionally more included in the control than the control deviation of the compressor inlet overheating in order to optimize the overall reaction time of the control.
  • one goal is to regulate the lowest possible overheating at the evaporator outlet and to shift the overheating of the refrigerant as completely as possible into the internal heat exchanger.
  • a method for adaptive compensation of the tolerances, which are individual, but mostly systematically, for each heat pump. What is meant is that the tolerances are caused by component tolerances, for example, which are individual for each heat pump, but generally do not change on a daily basis.
  • a solution is proposed to design the setpoints for evaporator outlet overheating and setpoint for evaporator outlet overheating in such a way that when the overheating control is in a steady state, both the control deviation for the evaporator output overheating and the control deviation for the compressor inlet overheating are equally zero.
  • the method for regulating the compression refrigeration system contains the following steps: determining a target value for the evaporator outlet superheating and a target value for the compressor inlet superheating, calculating a correction value based on a control deviation of the compressor inlet superheating from the target value for the compressor inlet superheating, correcting the target value the evaporator outlet superheat with the calculated correction value, calculation of 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 regulation of the expansion valve based on the control value.
  • the method according to the invention can react to tolerances of the wet steam characteristic and regulate the compression refrigeration system more precisely.
  • Tolerances of the wet steam characteristic arise, for example, regularly as a result of the segregation of refrigerant components when the compression refrigeration system is filled in a case in which 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 force of gravity during the filling process already results in segregation, which leads to a discrepancy between the composition of the refrigerant mixture in the compression refrigeration system and the composition in the refrigerant storage container.
  • different mixing ratios of the refrigerant components arise between the compression refrigeration systems.
  • the correction value is preferably calculated proportionally to a time integral of the control deviation of the compressor inlet overheating from the target value of the compressor inlet overheating.
  • the correction value is preferably corrected in discrete time steps by a proportional portion of the control deviation of the compressor inlet overheating.
  • the correction value is preferably limited to a permissible range of values.
  • the target value for the superheating of the evaporator outlet is preferably corrected by adding the calculated correction value.
  • Another advantageous method is to perform the addition "inversely” to correct the target value of the evaporator outlet overheating, the actual value being subtracted to form the control deviations from the target value.
  • the control deviation is advantageously calculated by calculating or forming a difference between an “actual value” and a “setpoint value”.
  • the superheating of the evaporator outlet is weighted with a parameter in relation to the superheating of the compressor in order to determine the control value.
  • the control deviations of an evaporator outlet overheating or the compressor inlet overheating from a setpoint value are particularly advantageously weighted.
  • the evaporator outlet overheating is weighted against a weighting of the compressor inlet overheating in order to determine the controlled variable by means of a parameter.
  • the control deviations of an evaporator outlet overheating or a compressor inlet overheating are particularly advantageously weighted with a setpoint value.
  • the evaporator outlet overheating is determined from a measured evaporator outlet temperature value, which is measured with an evaporator outlet temperature sensor, and from a low pressure, which is measured with a low pressure sensor.
  • the compressor inlet overheating is determined from a compressor inlet temperature, measured with a compressor inlet temperature sensor, and a low pressure, which is measured with a low pressure sensor.
  • a first control deviation of the compressor inlet superheat is calculated in the controller with a second control deviation of the evaporator outlet superheat to a total control deviation and the total control deviation is used to set the expansion valve.
  • the method is advantageously carried out in such a way that the parameter is gradually changed, preferably after the commissioning phase, in particular a weighting of the compressor inlet overheating begins with an influence of less than 20% compared to the evaporator outlet overheating and the weighting is then taken into account when determining the control value, changing the weighting, in particular increasing it is, in particular up to a predetermined target weighting of the parameter.
  • the weighting coefficient of the compressor inlet overheating is particularly advantageously set to a value between 0% and 20%, which is small compared to the operating phase, and after the commissioning phase it rises in a ramp over time to a target value for the operating phase.
  • the parameter value is advantageously controlled in such a way that the weighting of the compressor inlet overheating for determining the controlled variable has an influence of 0% or close to 0%, in particular below 20%. This means that the evaporator outlet overheating is included exclusively or primarily in the controlled variable.
  • the parameter P is controlled in such a way that the weighting of the controlled variable of the compressor inlet overheating to determine the controlled variable is increased up to a target value and in the subsequent operating phase the weighting of the compressor inlet overheating reaches the target value.
  • the refrigerant preferably has a temperature glide, the refrigerant in particular having or consisting of R454C, and where the compression refrigeration system contains in particular an internal heat exchanger for transferring thermal energy of the refrigerant before it enters the throttle element to the refrigerant before it enters the compressor. This is particularly relevant because high-performance internal heat exchangers are used in refrigeration circuits with R454C.
  • the object is also 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 the installation.
  • the actuators listed below are advantageously at least partially connected to the controller via a data connection 510, which can be done 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. Additionally or alternatively, an in the Fig. 1 Evaporator inlet temperature sensor (not shown) determine the temperature at 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 is used 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 rolling piston unit, scrolling unit or otherwise.
  • the compressed, superheated refrigerant at the compressor outlet pressure P Va is at a higher pressure level, in particular a high pressure HD, than at the inlet connection 211 with a compressor inlet pressure P Ve , in particular a low pressure ND, at a compressed inlet temperature T VE , which indicates the state of the refrigerant temperature at Describes inlet port 211 when entering a compression chamber.
  • thermal energy Q H is transferred from the refrigerant of the vapor compression system 200 to a heating medium of the heat sink system 400.
  • the refrigerant is de-heated in the liquefier 220, with superheated refrigerant vapor transferring part of its thermal energy to the heating medium of the heat sink system 400 by reducing the temperature .
  • a further heat transfer Q H advantageously takes place in the condenser 220 by condensation of the refrigerant during the phase transition from the gas phase of the refrigerant to the liquid phase of the refrigerant.
  • further heat Q H is transferred from the refrigerant from the vapor compression system 200 to the heating medium of the heat sink system 400.
  • the high pressure HD of the refrigerant established in the condenser 220 corresponds approximately to a condensation pressure of the refrigerant at a heating medium temperature Tws in the heat sink system when the compressor 210 is in operation.
  • the heating medium in particular water, is conveyed by means of a heating medium pump 410 through the heat sink system 400 in a direction SW through the condenser 220, while the thermal energy Q H is transferred from the refrigerant to the heating medium.
  • collector 260 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 on through the expansion valve 230, the collector 260 fills, otherwise it is emptied or emptied.
  • recuperator 250 which can also be referred to as an internal heat exchanger
  • internal thermal energy Q i from the refrigerant under high pressure HD which flows from condenser 220 to expansion valve 230 in a high pressure flow direction S HD
  • Q i from the refrigerant under low pressure LP Transfer refrigerant, which flows from the evaporator to the compressor in a low-pressure flow direction S ND , transferred.
  • the refrigerant flowing from the condenser to the expansion valve 230 is advantageously supercooled.
  • the refrigerant flows into the expansion valve through an expansion valve inlet 231.
  • the refrigerant pressure is throttled in the expansion valve 230 High pressure HP to low pressure LP, in that the refrigerant advantageously passes through a nozzle arrangement or throttle with an advantageously variable opening cross section, the low pressure advantageously corresponding approximately 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 set by an electric motor, which is usually designed as a stepping 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 in such a way 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 advantageously be a few Kelvin below the heat source medium temperature T WQ so that the temperature difference drives heat transfer.
  • the evaporator there is a transfer of evaporation heat energy Qv 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 Q delivers to vapor compression system 200.
  • 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 Q delivers to vapor compression system 200.
  • the refrigerant flowing into the evaporator 240 reduces its wet steam portion when flowing through the evaporator 240 by absorbing heat Q Q and leaves the evaporator 240 advantageously with a low wet steam portion or advantageously also as superheated gaseous refrigerant.
  • the heat source medium is conveyed through the heat source media 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 thermal energy Q Q being withdrawn 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, the refrigerant flowing from the evaporator 240 to the compressor 210 particularly further overheating.
  • This superheated refrigerant which exits the recuperator 250 at an overheating temperature T Ke , is passed to the refrigerant inlet connection 211 of the compressor 210.
  • the recuperator 250 is used in the vapor compression circuit 200 in order to increase the overall efficiency as the quotient of the heat output Q H emitted and the electrical power P e consumed to drive the compressor motor.
  • thermal energy Q i is withdrawn from the refrigerant, which in the condenser 220 emits thermal energy Q H at a temperature level on the heat sink side, by subcooling in the high pressure path of the recuperator 250.
  • 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 thermal energy Q Q from the heat source 300 at the same evaporation temperature level.
  • the refrigerant is supplied with the heat energy Q i extracted in the high pressure path again 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 has the advantageous effect of reducing the proportion of wet steam to a state without a proportion of wet steam and then overheating occurs through further supply of energy.
  • the following sensors are advantageously arranged to detect the operating state of the vapor compression system 200, with which a model-based precontrol is implemented, in particular for safeguarding and optimizing the operating conditions of the vapor compression system 200, particularly in the event of changes in the operating state.
  • the process variable which has a significant influence on the overall efficiency of the vapor compression circuit 200 as the quotient between the heating power Q H transferred by the vapor compression circuit 200 to the electrical power P e consumed by the compressor 210, is the overheating of the refrigerant at the compressor inlet 211 Compressor operating conditions, however, restrictions with regard to the permitted overheating range of the refrigerant at the compressor inlet are advantageously observed. Overheating that is too low endangers the lubricating properties of the machine oil in particular, while overheating that is too high particularly results in a hot gas temperature that is too high.
  • the overheating describes the temperature difference between the recorded compressor inlet temperature T KE of the refrigerant and the evaporation temperature of the refrigerant in the case of saturated steam.
  • the superheating of the compressor inlet is preferably regulated in such a way that no condensate precipitates due to the water vapor content in the ambient air falling below the dew point in components of the refrigeration circuit, particularly in the section between the refrigerant outlet of the recuperator 252 and the compressor inlet 211.
  • the refrigeration circuit section between evaporator outlet 242 and recuperator inlet 251 is usually colder, because it is typically only a short pipe section, better insulation is possible compared to the section between the refrigerant outlet of recuperator 252 and compressor inlet 211.
  • the refrigerant separator that is to be protected is located at the location of the compressor inlet 211 on the compressor.
  • Limit values in particular for overheating, define the permissible overheating range of the components at the compressor inlet 211 as a function of the operating point. Furthermore, there are also dependencies between the compressor inlet overheating dT ÜE and the overall efficiency of the vapor compression circuit 200 or between the compressor inlet overheating dT ÜE and a stability S of a control value R, which is advantageous when regulating the compressor inlet overheating.
  • the heat source medium temperature, the heating medium temperature, the compressor power P e and target values Z or the target value Z are used to calculate the compressor inlet superheat dT ÜE.
  • a calculation of the target value Z as a default value for the compressor inlet superheating dT ÜE can be carried out from the refrigeration circuit measurement variables that are dependent on the operating point, such as heat source medium temperature, heating medium temperature, compressor power P e and parameterizable coefficients that are adapted to the behavior of the respective refrigeration circuit components.
  • the target value for the compressor inlet superheat dT ÜE is constant regardless of all operating conditions, eg 10 Kelvin. In the case of a more complex adaptation it is varied as a function of an operating point variable, for example the compressor power P e , or in the case of an even more complex adaptation it varies as a function of several operating point variables.
  • the total control deviation is calculated, which is fed in to regulate the vapor compression circuit 200.
  • the refrigerant passes two sequentially arranged heat exchangers, the evaporator 240 and the recuperator 250, in which the refrigerant is supplied with thermal energy Q Q and Q i.
  • the refrigerant is supplied with source heat energy Q Q from the heat source system 300.
  • the temperature level of the supplied source heat Q Q is at a temperature level of the heat source, in particular such as that of the ground or the outside air.
  • thermal energy Q i is withdrawn from the refrigerant after it has left the condenser 220.
  • the temperature level of the refrigerant at the outlet of the condenser is approximately at the level of the return temperature of the heating medium.
  • the control value R is advantageously the weighted link between the control deviation of the compressor inlet superheat dT ÜE and 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 overheating dT ÜE, 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 linkage of the control deviation of the compressor inlet overheating with the control deviation of the evaporator outlet overheating.
  • the compressor 210 by varying the compressor speed and the expansion valve 230 by influencing the degree of opening are such actuators. These two actuators influence the low pressure 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 undesired ranges, so that a model-based, supported change in the degree of opening of the expansion valve to regulate R is advantageous, if necessary, even necessary.
  • the compressor speed is advantageously set in the vapor compression circuit 200 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 superheating 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 regulating the superheating of the compressor inlet dT ÜE.
  • the influence of the degree of opening of the expansion valve 230 on the compressor inlet superheat dT ÜE takes place as follows:
  • the expansion valve 230 acts as a nozzle with a nozzle cross-section that can be adjusted by an electric motor, in which a needle-shaped nozzle needle is usually threaded into a nozzle seat by means of a stepper motor.
  • the refrigerant throughput through the expansion valve is roughly proportional to the square root of the pressure difference between the expansion valve inlet 231 and outlet 232 multiplied by a current relative value of the nozzle cross-section or degree of opening and advantageously one of the refrigerant and a geometry of the expansion valve 230 dependent constant.
  • the degree of opening of the expansion valve 230 significantly influences only the low pressure ND, that is, the outlet pressure from the expansion valve 230.
  • the low pressure LP then falls on the low pressure side of the vapor compression circuit 200.
  • the mass flow of refrigerant through the compressor 210 decreases approximately proportionally, since its delivery rate is 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 by 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 a constant high pressure HP and initially still a constant low pressure LP.As the compressor 210 initially continues to convey the same refrigerant mass flow, the low pressure side LP of the refrigeration circuit becomes 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 LP increases on the low pressure side of the vapor compression circuit 200.
  • the mass flow rate of the compressor 210 increases roughly proportionally, since its delivery rate can approximately be described as volume / time, and a correspondingly increased low pressure ND is established, at which the refrigerant mass flow supplied by 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, 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 refrigerant path of the evaporator is to be assumed to be approximately constant in a respective vapor compression circuit 200.
  • the size of the heat transfer capacity in the evaporator 240 is therefore critically dependent on the integral of the temperature differences of all surface elements of the heat transfer layer.
  • a refrigerant temperature is established which, as a material property of the refrigerant, is a function of the low pressure ND of the refrigerant as a result of the saturation vapor characteristic.
  • control of the evaporation temperature of the refrigerant as it flows through the recuperator 250 can be controlled indirectly.
  • the thermal energy Q Q which is transferred from the heat source system to the refrigerant flowing through the evaporator 240, influences the state of the refrigerant.
  • a corresponding refrigerant state when exiting the evaporator 240 is set as a function of the manipulated variable “degree of opening of the expansion valve 230”.
  • control system behavior is characterized in particular by the control system output value of the evaporator outlet overheating as a function of the control system input value of the expansion valve opening degree.
  • a refrigerant is advantageously used, in particular a refrigerant mixture as refrigerant, which has a “temperature glide”, in particular R454C is advantageously used.
  • a refrigerant mixture with a temperature glide it is advantageous if the relative degree of opening of the actuator expansion valve changes by 1% rel.
  • a change in superheating of approximately less than 1 K is usually set at the outlet of the refrigerant from the evaporator.
  • the refrigerant After flowing through the evaporator 240, the refrigerant enters the low-pressure path of the recuperator 250 at low pressure LP.
  • the physical state of the refrigerant when flowing into the recuperator 250 is in a normal operating case, that is, advantageously either saturated steam with a low vapor content between 0 to 20% or, in particular, also advantageously already superheated refrigerant.
  • a refrigerant temperature is established which, due to the saturation vapor characteristic curve of the refrigerant, is a function of the refrigerant pressure.
  • the refrigerant temperature will at most assume a size which corresponds to the entry 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, that is to say the temperature of the refrigerant after it exits the condenser 220.
  • recuperator 250 In order to be able to transfer a sufficient amount of thermal energy from the refrigerant of the high-pressure-side refrigerant path to the refrigerant of the low-pressure-side refrigerant path in recuperator 250, it must be ensured that the temperature of the refrigerant of the high-pressure-side refrigerant path at high pressure HD in as many surface elements of the transfer layer of recuperator 250 as possible is greater than is the temperature of the refrigerant of the low-pressure side refrigerant path at low pressure LP on the respective surface element.
  • the corresponding temperatures of the heating system 400 of the vapor compression system 200 are higher in a heating case 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 proportion of the refrigerant flowing through the recuperator 250 on the low pressure side at low pressure LP decreases when heat is transferred to the refrigerant and, after complete evaporation, the refrigerant is advantageously overheated.
  • recuperator 250 there is advantageously a significantly higher heat transfer in the evaporator 240 between the source medium and the refrigerant in the evaporator 240.
  • the evaporator 240 a significantly higher heat transfer is set in the evaporator 240 than in the recuperator 250, since a much greater amount of energy is to be extracted from the surroundings by means of the evaporator 240 than is only to be transferred in the recuperator 250 within the refrigeration circuit.
  • the driving temperature difference in the recuperator is between 20 K and 60 K, for example, while it is only between 3 K and 10 K in the evaporator.
  • the exchanger surface of the evaporator is, for example, 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 here by at least two time constants Z, Z 11 , Z 12 , Z 13 , Z 14 , Z 15 , Z tot after the manipulated variable "expansion valve opening degree" has been changed.
  • recuperator 250 After changing the manipulated variable “opening degree of expansion valve 230”, there is a further delay in the corresponding change in refrigerant state due to the time behavior of recuperator 250 when it exits the low-pressure side refrigerant path of recuperator 250.
  • the time behavior of the recuperator 250 can advantageously be taken into account as the total 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 superheating dT UE and the evaporator outlet superheat dT ÜA by dividing the total deviation is calculated in particular by means of a weighted combination of the control deviation of the compressor overheating and the deviation of the evaporator outlet superheat dT ÜA which is fed in the controller 500 for controlling the vapor compression cycle 200th
  • Step 1 First, the process variables compressor inlet overheating dT ÜE are advantageously recorded as the main control variable and the evaporator outlet overheating dT ÜA is advantageously recorded as an auxiliary variable in a first process step.
  • the temperatures of the refrigerant temperature assigned to the overheating measuring point are recorded by means of temperature sensors 501, 508.
  • the temperature difference between the refrigerant at the respective measuring point and the evaporation temperature is then calculated and this temperature difference value then corresponds to the respective overheating of the refrigerant at the measuring point.
  • the output variables of the calculation in step 1 are then the compressor inlet superheat dT ÜE and the evaporator outlet superheat dT ÜA .
  • Step 2 The process variables compressor inlet superheating dT ÜE and evaporator outlet superheating dT ÜA are advantageously offset in a second step to create assigned control deviations with the respectively assigned setpoints:
  • the setpoint value for the compressor inlet superheat dT ÜE is advantageously varied in the range between approx. 5 K to 20 K in order to ensure the permissible compressor operating range and the highest possible efficiency of the refrigeration circuit.
  • the setpoint for the evaporator outlet overheating dT ÜA 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 overheating in the steady normal case roughly corresponds to the process value of the evaporator outlet overheating dT ÜA .
  • This setpoint for the evaporator outlet superheating dT ÜA can be precalculated 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 for the compressor inlet superheat dT ÜE at the evaporator outlet 242 and / or on component properties.
  • the control deviation of the compressor inlet superheat dT ÜE is then calculated by subtracting the setpoint of the compressor inlet superheat dT ÜE from the process value of the compressor inlet superheat dT ÜE.
  • Step 3 In a third process step, the control deviation of the compressor inlet superheating dT ÜE and the control deviation of the evaporator outlet superheating dT ÜA are advantageously combined to form an overall control deviation for superheating.
  • the combination takes place in particular by means of a weighted addition of the individual control deviations.
  • the weighting influence is a measure of the proportional combination of the individual system deviations and, in extreme cases, can result in the exclusive inclusion of only one individual system deviation, but usually the weighted inclusion of both individual system deviations.
  • Step 4 In a fourth method step, the calculated overall control deviation of the overheating is then processed in the controller 500, which controls the corresponding actuators of the refrigeration circuit, in particular the expansion valve 230 with the adjustable degree of opening and / or the compressor 210 with adjustable compressor speed, so that In the regulated case, there is a control deviation of the overheating equal to approximately 0 Kelvin, if possible.
  • a P, I, PI, PID controller can be used, the control components being advantageously dynamically adapted to the respective operating mode and the operating point.
  • the recuperator 250 between the refrigerant path for supercooled 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, proportionally the compressor inlet superheat T ÜE and proportionally the evaporator outlet superheat T ÜA are included to regulate the overheating of the refrigerant.
  • this is implemented by the adaptation of the setpoint value for the evaporator outlet superheating that has been carried out.
  • the control deviation for superheating compressor inlet i.e. the difference between the target superheating at the compressor inlet and the actual superheating at the compressor inlet, is used as the input variable for the adaptation.
  • a compensation variable temperature difference adaptation is calculated in that the value of the adaptation time constant assumes the inverse value of the control deviation overheating compressor inlet with the help of a time function.
  • 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 for example low-pass filters, are also conceivable.
  • Fig. 3 shows schematically and exemplarily three wet steam characteristics 1010, 1020 and 1030 for different mixing ratios of refrigerant components, in this example of R32 and R1234yf.
  • the wet steam curve 1020 corresponds to the wet steam curve of R454C.
  • the proportion of R32 is reduced compared to R454C, while it is increased for the wet steam curve 1030.
  • the three wet steam characteristics 1010, 1020 and 1030 are therefore typical mixture values that are obtained based on a container filled with R454C in a compression refrigeration system.
  • the temperature difference adaptation can be calculated in two successive steps. In a first step, taking into account the value of the temperature difference adaption evaporator output overheating target value calculated in the last loop pass, an unlimited new value of temperature difference adaption evaporator output overheating target value (unlimited) is calculated, in the second step this value is limited to the parameterizable range limit and then processed further as a newly calculated process value.
  • the temperature difference adaptation of the 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 within a specified time, called the adaptation time constant of the evaporator output superheat setpoint.
  • the range of the temperature difference adaptation evaporator output superheat setpoint is limited to the range set with an adaptation range parameter by limiting the previously calculated value to the range +/- adaptation range parameters (in Kelvin).
  • Fig. 4 shows schematically and by way of example a curve 2010 of the control deviation of the compressor inlet superheating and a curve 2020 of the adaptation of the evaporator outlet overheating setpoint over time in minutes on the horizontal axis as a result of the control deviation.
  • a control deviation in the opposite direction there is now a control deviation in the opposite direction.
  • the value of the adaptation increases successively in order 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)
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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Cited By (2)

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CN118729623A (zh) * 2024-07-29 2024-10-01 江苏拓米洛高端装备股份有限公司 蒸发器出口过热度的控制方法、装置、电子设备及存储介质
EP4579151A1 (fr) * 2023-12-20 2025-07-02 Stiebel Eltron GmbH & Co. KG Procédé et dispositif de régulation conçu selon ce procédé pour un cycle de réfrigération et système de réfrigération à compression

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EP1014013A1 (fr) * 1998-12-18 2000-06-28 Sanden Corporation Cycle frigorifique à compression de vapeur
EP1026459A1 (fr) * 1999-01-11 2000-08-09 Sanden Corporation Système frigorifique à compression de vapeur
DE10159892A1 (de) 2001-12-06 2003-06-26 Stiebel Eltron Gmbh & Co Kg Kältemaschine mit einem Rekuperator
DE102005061480B3 (de) 2005-12-22 2007-04-05 Stiebel Eltron Gmbh & Co. Kg Wärmepumpenanlage
DE112016005264T5 (de) * 2015-11-17 2018-08-02 Valeo Japan Co., Ltd. Kältekreislauf einer klimaanlage für fahrzeuge und damit ausgerüstetes fahrzeug

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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
JP4948374B2 (ja) 2007-11-30 2012-06-06 三菱電機株式会社 冷凍サイクル装置
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EP1014013A1 (fr) * 1998-12-18 2000-06-28 Sanden Corporation Cycle frigorifique à compression de vapeur
EP1026459A1 (fr) * 1999-01-11 2000-08-09 Sanden Corporation Système frigorifique à compression de vapeur
DE10159892A1 (de) 2001-12-06 2003-06-26 Stiebel Eltron Gmbh & Co Kg Kältemaschine mit einem Rekuperator
DE102005061480B3 (de) 2005-12-22 2007-04-05 Stiebel Eltron Gmbh & Co. Kg Wärmepumpenanlage
DE112016005264T5 (de) * 2015-11-17 2018-08-02 Valeo Japan Co., Ltd. Kältekreislauf einer klimaanlage für fahrzeuge und damit ausgerüstetes fahrzeug

Cited By (2)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
EP4579151A1 (fr) * 2023-12-20 2025-07-02 Stiebel Eltron GmbH & Co. KG Procédé et dispositif de régulation conçu selon ce procédé pour un cycle de réfrigération et système de réfrigération à compression
CN118729623A (zh) * 2024-07-29 2024-10-01 江苏拓米洛高端装备股份有限公司 蒸发器出口过热度的控制方法、装置、电子设备及存储介质

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