EP3922925A1 - Procédé de fonctionnement d'une installation de réfrigération à compression et installation de réfrigération à compression - Google Patents

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

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Publication number
EP3922925A1
EP3922925A1 EP21177578.8A EP21177578A EP3922925A1 EP 3922925 A1 EP3922925 A1 EP 3922925A1 EP 21177578 A EP21177578 A EP 21177578A EP 3922925 A1 EP3922925 A1 EP 3922925A1
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EP
European Patent Office
Prior art keywords
refrigerant
compressor
temperature
overheating
hot gas
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Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
Pending
Application number
EP21177578.8A
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German (de)
English (en)
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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Publication date
Application filed by Stiebel Eltron GmbH and Co KG filed Critical Stiebel Eltron GmbH and Co KG
Publication of EP3922925A1 publication Critical patent/EP3922925A1/fr
Pending legal-status Critical Current

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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
    • 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
    • F25B49/00—Arrangement or mounting of control or safety devices
    • F25B49/02—Arrangement or mounting of control or safety devices for compression type machines, plants or systems
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2500/00—Problems to be solved
    • F25B2500/19—Calculation of parameters
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2500/00—Problems to be solved
    • F25B2500/28—Means for preventing liquid refrigerant entering into 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
    • F25B2600/00—Control issues
    • F25B2600/25—Control of valves
    • F25B2600/2513—Expansion valves
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2700/00—Sensing or detecting of parameters; Sensors therefor
    • F25B2700/19—Pressures
    • F25B2700/193—Pressures of the compressor
    • F25B2700/1933—Suction pressures
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2700/00—Sensing or detecting of parameters; Sensors therefor
    • F25B2700/21—Temperatures
    • F25B2700/2115—Temperatures of a compressor or the drive means therefor
    • F25B2700/21151—Temperatures of a compressor or the drive means therefor at the suction side of the compressor

Definitions

  • the invention relates to a method for operating a compression refrigeration system and an associated compression refrigeration system with a refrigerant, an evaporator, a compressor, a condenser, a throttle element, 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, and a control unit for detecting overheating of the refrigerant when it enters the compressor, the overheating being defined as a difference between a dew point temperature and a temperature of the refrigerant, and for regulating the throttle element based on the overheating,
  • 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.
  • a refrigeration circuit with a refrigerant, an evaporator, a compressor, a condenser and a throttle device
  • energy from the environment for example outside air in air / water heat pumps or brine in brine / water heat pumps
  • the refrigerant is compressed in the compressor with the aid of electrical energy and then is when the refrigerant is liquefied
  • energy is transferred to the working medium of a heat sink circuit, for example a heating circuit and / or a hot water charging circuit, at a comparatively high temperature level.
  • an increase in pressure of the gas for example as a result of the compression of the gaseous refrigerant in the compressor, is accompanied by an increase in temperature.
  • the gas temperature before compression also has an influence on the gas temperature after compression; an increase in gas temperature before compression is approximately also associated with an increase in gas temperature after compression.
  • hot gas temperatures at the outlet of the compressor must not be exceeded, whereby these temperatures can be defined as absolute, i.e. applicable to all operating states, or relatively, i.e. depending on the operating state.
  • a method for regulating a compression refrigeration system with a refrigeration circuit with a refrigerant, an evaporator, a compressor, a Condenser, a throttle element, an internal heat exchanger for the transfer of thermal energy of the refrigerant before entry into the throttle element to the refrigerant before entry into the compressor, and a control unit a) for detecting overheating of the refrigerant when entering the compressor, the overheating as a The difference between a dew point temperature and a temperature of the refrigerant is defined, and b) for regulating the throttle element based on the overheating is proposed.
  • the method comprises the following steps: determining a first target superheating of the refrigerant when entering the compressor, the first target superheating maximizing the efficiency of the compression refrigeration system as a function of an operating point of the compression refrigeration system, determining a second target superheating of the refrigerant when entering the compressor on the basis of a maximum permissible hot gas temperature at the outlet of the compressor, and regulating the throttle element on the basis of the lower value from the first setpoint superheat and the second setpoint superheat.
  • the second setpoint overheating is preferably determined based on a refrigeration model that establishes a relationship between the overheating at the inlet of the compressor and the hot gas temperature at the outlet of the compressor.
  • the second target overheating can, for example, be continuously calculated, in particular on the basis of measured values and / or parameters of the compression refrigeration system, or also be provided for all operating points in the form of a table or the like. As a result, the data processing effort can be kept low during operation.
  • the refrigeration model preferably contains at least the following input variables: low pressure, high pressure, speed of the compressor, or the following input variables: dew point temperature in low pressure, boiling point temperature in high pressure, speed of the compressor.
  • the refrigeration model preferably comprises a linear function of quadratic order of the boiling point temperature in high pressure and the dew point temperature in low pressure.
  • the maximum permissible hot gas temperature at the outlet of the compressor is preferably defined as a temperature which is below a hot gas temperature limit defined for the compressor, in particular defined by the manufacturer.
  • the throttle element is preferably also regulated based on a currently measured hot gas temperature.
  • model-based calculation can be incorrect with regard to the time behavior, with regard to tolerances in the acquisition and processing of the included process values, component tolerances (compressor, refrigerant) and / or environmental conditions, e.g. engine room temperature, so that a correction of this calculation based on a recording and inclusion of the actual hot gas temperature is helpful.
  • the task is achieved by a compression refrigeration system with a refrigeration circuit with a refrigerant, an evaporator, a compressor, a condenser, a throttle element, an internal heat exchanger for the transfer of thermal energy of the refrigerant before it enters the throttle element to the refrigerant before it enters the Compressor, and a control unit a) for detecting overheating of the refrigerant upon entry into the compressor, the overheating being defined as a difference between a dew point temperature and a temperature of the refrigerant, and b) for regulating the throttle element based on the overheating released, the The control unit is designed to: determine a first setpoint overheating of the refrigerant when it enters the compressor, the first setpoint overheating as a function of an operating point of the compression refrigeration system maximizes the efficiency of the compression refrigeration system, determining a second target superheating of the refrigerant when entering the compressor on the basis of a maximum permissible hot gas temperature at the
  • the compression refrigeration system according to the invention enables the same advantages to be achieved as the method according to the invention.
  • a combination with all of the preferred embodiments of the method is also advantageously possible.
  • the refrigerant preferably has a temperature glide, the refrigerant in particular having or consisting of R454C or containing components such as R32 or R1234yf.
  • a heat pump in particular a heat pump installed inside a building, with a compression refrigeration system according to the invention is proposed.
  • 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 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 T Va at the compressor outlet 212.
  • the compressor 210 usually contains a drive unit with an electric motor, a compression unit and the electric motor can advantageously 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 compressor inlet temperature T VE , which indicates the state of the refrigerant at the inlet connection 211 describes 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 from the high pressure HP to the low pressure LP by the refrigerant advantageously being a nozzle arrangement or throttle with an advantageously variable Opening cross-section happens, 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 that from the evaporator 240 Transferring refrigerant flowing to the compressor 210, the refrigerant flowing from the evaporator 240 to the compressor 210 in particular 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 in the low-pressure path at low pressure ND and at a low pressure temperature corresponding to an evaporator outlet temperature T Va at the inlet to the recuperator 250 is supplied with the heat energy Q i extracted in the high-pressure path.
  • the supply of energy has the advantageous effect of reducing the proportion of wet steam to a state without a proportion of wet steam. Overheating is ensured by additional energy supply.
  • 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 to safeguard and optimize the operating conditions of the vapor compression system 200, in particular 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 compressor inlet overheating 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. For example, sits at the point of the compressor inlet 211 on the compressor is the refrigerant separator that is to be protected.
  • room temperature sensors and room humidity sensors are advantageous, since they enable precise determination of the condensation conditions of the air, for example at 21 ° C and 60% rel. Moisture the condensation temperature in the range of 13 ° C. Under these conditions, no condensation takes place as long as the pipe temperature is above 13 ° C plus, if necessary, a buffer, e.g. 1K.
  • 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.
  • Actuators or 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 in the relevant vapor compression circuit 200 Expansion valve 230, which also advantageously determines the low pressure LP 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 power without further compensatory changes in the degree of opening of the expansion valve advantageously 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 fed as 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 rises.
  • the mass flow rate of the compressor 210 increases approximately proportionally, since its rate can be approximately described as volume / time, and it 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.
  • a control of the low pressure LP or also an evaporation pressure indirectly control a control of 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 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”.
  • a controlled system behavior with moderate steepness results with regard to a control path steepness of the "isolated" controlled system "evaporator 240".
  • the 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 relative change in the degree of opening of the expansion valve actuator of 1% rel. at the outlet of the refrigerant from the evaporator usually set with a superheating change of advantageously approximately less than 1 K.
  • the refrigerant After flowing through the evaporator 240, the refrigerant enters the low-pressure path of the recuperator 250 at low pressure LP.
  • the state of aggregation of the refrigerant when flowing into the recuperator 250 is in a normal operating case, i.e. 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.
  • the "isolated" controlled system at low pressure ND of the refrigerant in the low-pressure side path of the recuperator 250 results in a control system behavior with high steepness, with an approximately constant internal energy state of the refrigerant upon entry 251 into the low-pressure side LP path of the recuperator 250.
  • a particularly relative change in the degree of opening of the expansion valve of 1% there is a change in superheating at the outlet of the refrigerant from the recuperator 250 of approximately 10 K or even more than 10 K.
  • 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 driving temperature difference in the recuperator can be between 20 and 60 K, while it is only between 3 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
  • the compressor inlet overheating dT ÜE is advantageously used as the main control variable and the corresponding signal flows and signal processing takes place in particular in the following process steps:
  • the process variables compressor inlet overheating dT ÜE are advantageously measured 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 .
  • the process variables compressor inlet superheating dT ÜE and evaporator outlet superheating dT ÜA are advantageously offset in a second step to form 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 corresponds approximately to the process value of the evaporator outlet overheating dT ÜA.
  • This setpoint for the evaporator outlet superheat 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 compressor inlet 211 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.
  • 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 - 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 control deviations and, in extreme cases, can only be included exclusively an individual control deviation, but usually the weighted inclusion of both individual control deviations.
  • 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 the If a control deviation of the overheating is equal to about 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 compressor To protect components (mechanical components, refrigeration machine oil), the compressor must be protected from excessively high refrigerant gas temperatures when exiting the compression chamber; the compressor manufacturer can also ensure compliance with an absolute (applicable to all operating states) or relative (applicable depending on operating states) prescribe maximum hot gas temperature (e.g. 120 ° C).
  • a hot gas temperature limit value relevant to the control technology (which is below the hot gas temperature limit specified by the compressor manufacturer) is specified (e.g. 110 ° C), which is used as a limit for the controller activities that limit the hot gas temperature.
  • thermodynamic calculation is made for the current Operating point of the refrigeration circuit calculates the overheating of the refrigerant at the compressor inlet, known as compressor inlet overheating, when operating with the hot gas temperature limit value relevant to the control technology.
  • Influencing variables in this model-based calculation can be low pressure, high pressure, (compressor speed) or dew point temperature ND, boiling point temperature HD, (compressor speed).
  • the compressor speed does not play a role in the calculation, but the dependency of the compression losses of the compressor as a function of the compressor speed can have an influence on the thermodynamic behavior.
  • a parameterizable approximation of the compressor behavior with regard to the hot gas temperature is preferably used in order to estimate a temperature difference between the hot gas temperature and the superheat at the compressor inlet at the respective operating point of the compressor based on the process data for the evaporation temperature and the condensation temperature.
  • the corresponding compressor inlet superheat calculated on the basis of an estimated hot gas temperature is managed as a process variable, setpoint superheat compressor inlet V-ND, estimated as a maximum .
  • the current hot gas temperature can also be included in the calculation, since the model-based calculation can be incorrect with regard to the time behavior, with regard to tolerances in the acquisition and processing of the included process values, component tolerances (compressor, refrigerant), ambient conditions, e.g. machine room temperature, so that a correction of this calculation based on a detection and inclusion of the actual hot gas temperature is helpful.
  • component tolerances compressor, refrigerant
  • ambient conditions e.g. machine room temperature
  • Another alternative embodiment is the maximum integrated in the formulas - formation of the overheating setpoint calculated for limiting the hot gas temperature and a lower limit provided for limiting the value range of this overheating setpoint, here the value 0 Kelvin is implemented as an example.
  • Such a limitation can be advantageous because the result of the model-based calculation of the superheating setpoint calculated for the limitation of the hot gas temperature can also result in negative superheating setpoints which, if regulated to these values, would cause undesired wet steam suction of the compressor. This is avoided by limiting the overheating to a minimum, which can also be interpreted in the (slightly) negative value range if necessary; Priority is subordinate) that the hot gas temperature is not limited exactly to the desired limit value, but may exceed it to a small extent.
  • the compressor inlet superheating setpoint designed for optimum efficiency and the superheating setpoint calculated for limiting the hot gas temperature are combined in such a way that, in the event of a required hot gas temperature limitation, an overheating setpoint that is optimized in terms of efficiency can be reduced:
  • Setpoint U ⁇ Overheating compressor inlet V - ND minimum Setpoint U ⁇ Overheating compressor inlet V - LP efficiency setpoint U ⁇ Overheating compressor inlet V - ND maximum
  • Compressor inlet superheat setpoint V-LP 10 K.

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  • General Engineering & Computer Science (AREA)
  • Air Conditioning Control Device (AREA)
  • Heat-Pump Type And Storage Water Heaters (AREA)
EP21177578.8A 2020-06-09 2021-06-03 Procédé de fonctionnement d'une installation de réfrigération à compression et installation de réfrigération à compression Pending EP3922925A1 (fr)

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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
EP2000751A2 (fr) * 2006-03-27 2008-12-10 Mitsubishi Electric Corporation Dispositif de climatisation frigorifique
WO2020071299A1 (fr) * 2018-10-02 2020-04-09 ダイキン工業株式会社 Dispositif à cycle frigorifique

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DE4303533A1 (de) 1993-02-06 1994-08-11 Stiebel Eltron Gmbh & Co Kg Verfahren zur Begrenzung der Heißgastemperatur in einem Kältemittelkreislauf und Expansionsventil
US7665321B2 (en) 2002-12-11 2010-02-23 Bms-Energietechnik Ag Evaporation process control used in refrigeration
DE102007010645B4 (de) 2007-03-02 2020-12-17 Stiebel Eltron Gmbh & Co. Kg Verfahren zum Steuern einer Kompressionskälteanlage und eine Kompressionskälteanlage
CN103717985B (zh) 2009-12-18 2016-08-03 开利公司 运输制冷系统和用于运输制冷系统以解决动态条件的方法

Patent Citations (5)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
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
EP2000751A2 (fr) * 2006-03-27 2008-12-10 Mitsubishi Electric Corporation Dispositif de climatisation frigorifique
WO2020071299A1 (fr) * 2018-10-02 2020-04-09 ダイキン工業株式会社 Dispositif à cycle frigorifique

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