METHOD FOR OPERATING A COMPRESSION REFRIGERATING SYSTEM AND COMPRESSION REFRIGERATING SYSTEM

Description

[0001] 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 entering the throttle element to the refrigerant before entering the compressor, and a control unit for detecting an overheating of the refrigerant upon entering the compressor, wherein the overheating is defined as a difference between a dew point temperature and a temperature of the refrigerant, and for controlling the throttle element based on the overheating,

[0002] Such compression refrigeration systems, for example in the form of heat pumps, with a vapor compression system in which a gaseous refrigerant is compressed from a low pressure to a high pressure by a compressor controlled by the control unit, which for example has a regulator, are known.

[0003] The refrigerant is forced through the condenser, where it transfers 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 at high pressure from the condenser to the expansion valve and the refrigerant flowing at low pressure from the evaporator to the compressor.

[0004] The refrigerant is then directed in a high-pressure flow direction to a regulator-controlled expansion valve, where the refrigerant is expanded from high pressure to low pressure depending on a set value. The low-pressure refrigerant evaporates in the evaporator by absorbing source heat.

[0005] Document EP 1 026 459 A1 discloses a vapor-compression type refrigeration system comprising a first refrigerant temperature detector mounted on a pipe for detecting the temperature of the refrigerant on the outlet side of an evaporator. A second refrigerant temperature detector is mounted on a pipe for detecting the temperature of the refrigerant on the inlet side of a compressor. A switching controller is connected to the first and second refrigerant temperature detectors and a superheat control valve, and changes / selects one of the refrigerant temperature detection value signals from the first and second refrigerant temperature detectors according to predetermined conditions. A superheat control valve adjusts the flow rate of the refrigerant flowing into the evaporator so that the refrigerant superheat on the inlet side of the compressor reaches a predetermined value.

[0006] Document WO 2020 / 071299 A1 discloses a refrigeration cycle device for which an injection pipe and an economical heat exchanger are provided for a main refrigerant circuit. Furthermore, the refrigeration cycle device is provided with a sub-refrigerant circuit including a sub-use-side heat exchanger. In the refrigeration cycle device, the sub-use-side heat exchanger serves as an evaporator for a sub-refrigerant to cool a main refrigerant that has been cooled in the economical heat exchanger or as a heat radiator for the sub-refrigerant to heat the main refrigerant that has been cooled in the economical heat exchanger.

[0007] Document EP 2 000 751 A2 discloses a refrigerant air conditioning system comprising a compressor, a room heat exchanger, a first pressure reducing device and an outdoor heat exchanger connected in a circuit for supplying hot heat from the room heat exchanger.The refrigerant air conditioning system comprises i) a first internal heat exchanger for exchanging refrigerant heat between the room heat exchanger and the first pressure reducing device with refrigerant heat between the outdoor heat exchanger and the compressor, ii) an injection circuit for bypassing a portion of the refrigerant between the room heat exchanger and the first pressure reducing device to inject it into a compression chamber within the compressor, iii) a pressure reducing device for injection provided along the injection circuit, and iv) a second internal heat exchanger for exchanging heat of the refrigerant whose pressure has been reduced by the pressure reducing device for injection with heat of the refrigerant between the room heat exchanger and the first pressure reducing device.

[0008] EP 1 965 160 A2 provides a method for controlling a refrigeration system comprising a refrigerant, an evaporator, a compressor, a condenser, and an electronically controllable throttle device. The refrigerant's superheat control is performed at the evaporator outlet. A special function control is activated when the refrigeration system is in certain operating modes. The special function control intervenes in the refrigerant's superheat control at the evaporator outlet. The special function control is terminated when the refrigeration system returns to normal operation.

[0009] EP 2 526 353 B1 discloses a method for controlling and regulating the superheat temperature of a refrigerant in an evaporation device in the circuit of a system of a heat pump or cooling system, which is characterized in that a virtual energy flow balance is formed for determining the setpoint for the controlled variable, wherein the virtual energy flow balance is formed in such a way that the virtual energy flow balance behaves approximately proportionally to the real energy flow balance at all operating points of the heat pump or cooling system by determining a virtual heat output of the evaporation device which is functionally dependent on the known variables of superheat temperature, specific heat capacity of the refrigerant and specific enthalpy in the saturated steam state, stroke of the expansion valve and a correction factor which is formed for each operating point of the expansion valve.

[0010] DE 101 59 892 A1 discloses the use of a recuperator in a refrigeration machine, particularly in a heat pump, to increase the heating output in a structurally simple manner at low outside temperatures. For this purpose, the recuperator is dimensioned such that, at low evaporation temperatures, it transfers at least approximately 15% of the heat pump's heating output from the liquid refrigerant to the gaseous refrigerant. An injection valve injects liquid refrigerant into the compressor, so that the final compression temperature remains below 120 °C.

[0011] A heat pump system with a refrigerant circuit is known from DE 10 2005 061 480 B3. It is equipped with a compressor, a first heat exchanger, a throttle element, an evaporator, and a 4-2-way valve unit for switching between a first (heating) and a second operating mode (cooling). The flow direction of the refrigerant in the refrigerant circuit can be switched such that the first heat exchanger serves to condense the refrigerant in the first operating mode and to evaporate the refrigerant in the second operating mode, and the second heat exchanger serves to evaporate the refrigerant in the first operating mode and to condense the refrigerant in the second operating mode. The first heat exchanger in the refrigerant circuit is connected such that it operates as a counterflow heat exchanger in both the heating and cooling operating modes.

[0012] In a refrigeration circuit with a refrigerant, an evaporator, a compressor, a condenser and a throttle device, when the refrigerant evaporates, energy from the environment, for example outside air in air / water heat pumps or brine in brine / water heat pumps, is transferred to the refrigerant at a comparatively low temperature level, the refrigerant is compressed in the compressor with the help of electrical energy and then when the refrigerant condenses, 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.

[0013] During normal heating operation of the refrigeration circuit, it is known to regulate the throttle valve in such a way that a target superheat of the refrigerant is achieved at the compressor inlet, which ensures the highest possible efficiency of the refrigeration circuit, thus maximizing its efficiency. Target superheats between 5 and 40 K are typically used for this purpose. For refrigeration circuits with an internal heat exchanger, a maximum efficiency at superheats of around 30 to 40 K is optimal for operating points with high pressure differences.

[0014] Due to the physical properties of gases, see, for example, the general gas equation, an increase in gas pressure, for example, as a result of compression of the gaseous refrigerant in the compressor, is accompanied by an increase in temperature. Furthermore, the gas temperature before compression also influences the gas temperature after compression; approximately, an increase in gas temperature before compression is also accompanied by an increase in gas temperature after compression.

[0015] It must now be taken into account that, in order to protect the components of the compressor, hot gas temperatures at the compressor outlet must not be exceeded, whereby these temperatures can be defined absolutely, ie applicable to all operating conditions, or relatively, ie depending on the operating condition.

[0016] Against this background, it was an object of the present invention to provide a compression refrigeration system and a method for controlling such a compression refrigeration system, which enable an efficient operation while ensuring protection of the components.

[0017] The problem is solved by the method features of claim 1 and the device features of claim 7.

[0018] Accordingly, a method is proposed for controlling a compression refrigeration system with a refrigeration circuit, with a refrigerant, an evaporator, a compressor, a condenser, a throttle device, an internal heat exchanger for transferring thermal energy of the refrigerant before entering the throttle device to the refrigerant before entering the compressor, and a control unit a) for detecting an overheating of the refrigerant upon entering the compressor, wherein the overheating is defined as a difference between a dew point temperature and a temperature of the refrigerant, and b) for controlling the throttle device based on the overheating.The method comprises the following steps: determining a first target superheat of the refrigerant upon entering the compressor, wherein the first target superheat maximizes the efficiency of the compression refrigeration system as a function of an operating point of the compression refrigeration system, determining a second target superheat of the refrigerant upon entering the compressor based on a maximum permissible hot gas temperature at the outlet of the compressor, and controlling the throttle device based on the lower value of the first target superheat and the second target superheat.

[0019] By determining two target superheats, with the control system using the lower of the target superheats to regulate the compression refrigeration system, efficiency-maximizing control of the compression refrigeration system is possible while maintaining the maximum hot gas temperature at the compressor outlet. The control system is easy to implement, as only the setpoint for the superheat control needs to be adjusted if necessary, namely if the first target superheat would result in an excessively high hot gas temperature.

[0020] Preferably, the second target superheat is determined based on a refrigeration model that establishes a relationship between the superheat at the inlet of the compressor and the hot gas temperature at the outlet of the compressor.

[0021] The second target superheat can, for example, be calculated continuously, particularly based on measured values and / or parameters of the compression refrigeration system, or it can be provided for all operating points in the form of a table or similar. This allows the data processing effort during operation to be kept to a minimum.

[0022] Preferably, the refrigeration model contains at least the following input variables: low pressure, high pressure, compressor speed, or the following input variables: dew point temperature in low pressure, boiling point temperature in high pressure, compressor speed.

[0023] Preferably, the refrigeration model comprises a linear quadratic function of the boiling point temperature in high pressure and the dew point temperature in low pressure.

[0024] Preferably, the maximum permissible hot gas temperature at the outlet of the compressor is defined as a temperature which is below a hot gas temperature limit specified for the compressor, in particular specified by the manufacturer.

[0025] Preferably, the throttle body is also controlled based on a currently measured hot gas temperature.

[0026] It has been shown that the model-based calculation can be erroneous with regard to the time behavior, tolerances in the acquisition and processing of the included process values, component tolerances (compressor, refrigerant) and / or ambient conditions, e.g. machine room temperature, so that a correction of this calculation based on the acquisition and inclusion of the actual hot gas temperature is helpful.

[0027] In another aspect, the object is achieved by a compression refrigeration system having a refrigeration circuit with a refrigerant, an evaporator, a compressor, a condenser, a throttle element, an internal heat exchanger for transferring thermal energy of the refrigerant before entering the throttle element to the refrigerant before entering the compressor, and a control unit a) for detecting an overheating of the refrigerant upon entering the compressor, wherein the overheating is defined as a difference between a dew point temperature and a temperature of the refrigerant, and b) for controlling the throttle element based on the overheating, wherein the control unit is designed to: determine a first target superheat of the refrigerant upon entering the compressor, wherein the first target superheat maximizes the efficiency of the compression refrigeration system as a function of an operating point of the compression refrigeration system,Determining a second target superheat of the refrigerant entering the compressor based on a maximum permissible hot gas temperature at the compressor outlet, and controlling the throttle device based on the lower of the first target superheat and the second target superheat.

[0028] 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 preferred embodiments of the method is also advantageously possible.

[0029] Preferably, the refrigerant has a temperature glide, wherein the refrigerant in particular comprises or consists of R454C or contains components such as R32 or R1234yf.

[0030] According to a further aspect, a heat pump, in particular a heat pump installed inside a building, with a compression refrigeration system according to the invention is proposed.

[0031] The figures show an example of implementation: Fig. 1Heat pump 100 with a vapor compression cycle 200 Fig. 2log p / h - diagram of the vapor compression process with recuperator 250

[0032] Fig. 1 shows schematically and exemplarily a heat pump 100. The heat pump 100 essentially consists of a vapor compression system 200 forming a compression refrigeration system, which contains the following components: A compressor 210 for compressing the superheated refrigerant, a condenser 220, with a refrigerant-side condenser inlet 221 and a condenser outlet 222 for transferring heat energy QH from the vapor compression system 200 to a heating medium of a heating system 400, with a heating medium inlet 401, a heating medium outlet 402 and a heating medium pump 410, for a building heating system or a system for hot water heating, advantageously a refrigerant collector 260, which is used as a refrigerant reservoir to compensate for different refrigerant quantity requirements depending on the operating conditions, a throttle element 230 designed as an expansion valve for expanding the refrigerant, an evaporator 240, with an evaporator inlet 241, for transferring source energy QQ from a heat source system 300, with a heat source inlet 320 and a heat source outlet 310, wherein the heat source system 300 can in particular be a brine system,which absorbs thermal energy QQ from the ground or an air system which absorbs thermal energy QQ from the ambient air and transfers it to the vapor compression system 200 or any other heat source, a recuperator as an example of an internal heat exchanger 250, which is intended to transfer internal thermal energy Q i 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 and a refrigerant, in particular a refrigerant mixture of at least two substances or two refrigerants which flows in a flow direction S HD and S ND through the vapor compression circuit 200, wherein in the vapor compression circuit 200 refrigerant vapor is brought to a high pressure HD by the compressor 210 and is led to a condenser 220,wherein a high-pressure path with the high-pressure flow direction S HD is formed from the compressor 210 to the expansion valve 230. After the expansion valve 230 to the compressor 210, a low-pressure path with a low-pressure flow direction S ND of the refrigerant is formed, in which the evaporator 240 is located.

[0033] The actuators listed below are advantageously at least partially connected to the controller via 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. Additionally or alternatively, a Fig. 1An evaporator inlet temperature sensor (not shown) determines the temperature at the evaporator inlet 241.

[0034] In the Fig. 1 In the example shown, heat pump 100 is shown as a brine heat pump. Of course, similar considerations and advantages can be achieved with air-to-water heat pumps. In particular, with air-to-water heat pumps, a fan / ventilator is arranged as the heat source instead of the brine circuit with brine pump 330.

[0035] The compressor 210 serves to compress the superheated refrigerant from an inlet connection 211 to a compressor outlet pressure P Va at a compressor outlet temperature T Va at the compressor outlet 212. The compressor 210 usually contains a drive unit with an electric motor, a compression unit and advantageously the electric motor can be operated at a variable speed. The compression unit can be designed as a rolling piston unit, scroll unit or otherwise. At the compressor outlet 212 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 describes the state of the refrigerant at the inlet connection 211 when entering a compression chamber.

[0036] In the condenser 220, the transfer of heat energy QH from the refrigerant of the vapor compression system 200 to a heating medium of the heat sink system 400 takes place. First, the desuperheating of the refrigerant takes place in the condenser 220, whereby superheated refrigerant vapor transfers part of its heat energy to the heating medium of the heat sink system 400 by reducing its temperature.

[0037] After the refrigerant vapor has been deheated, further heat transfer QH advantageously occurs in the condenser 220 through condensation of the refrigerant during the phase transition from the gas phase of the refrigerant to the liquid phase of the refrigerant. In this process, further heat QH is transferred from the refrigerant in the vapor compression system 200 to the heating medium of the heat sink system 400.

[0038] The high pressure HD of the refrigerant that occurs in the condenser 220 during operation of the compressor 210 corresponds approximately to a condensation pressure of the refrigerant at a heating medium temperature T WS in the heat sink system.

[0039] 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, whereby the heat energy QH is transferred from the refrigerant to the heating medium.

[0040] The downstream receiver 260 stores refrigerant exiting the condenser 220, 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 from the condenser 220 than is passed through the expansion valve 230, the receiver 260 fills; otherwise, it empties or empties.

[0041] In the subsequent recuperator 250, which can also be referred to as an internal heat exchanger, internal heat energy Q i is transferred from the high-pressure refrigerant (HD), which flows from the condenser 220 to the expansion valve 230 in a high-pressure flow direction S HD, to the low-pressure refrigerant (ND), which flows from the evaporator to the compressor in a low-pressure flow direction S ND. In this process, the refrigerant flowing from the condenser to the expansion valve 230 is advantageously subcooled.

[0042] First, the refrigerant flows into the expansion valve through an expansion valve inlet 231. In the expansion valve 230, the refrigerant pressure is throttled from the high pressure (HP) to the low pressure (ND) by the refrigerant advantageously passing through a nozzle arrangement or throttle with an advantageously variable opening cross-section, whereby the low pressure advantageously corresponds approximately to the suction pressure of the compressor 210. Instead of an expansion valve 230, any other desired pressure-reducing device can also be used. Pressure-reducing pipes, turbines, or other expansion devices are advantageous.

[0043] The degree of opening of the expansion valve 230 is adjusted by an electric motor, typically embodied as a stepper motor, which is controlled by the control unit or regulator 500. The low pressure ND at the expansion valve outlet 232 of the refrigerant from the expansion valve 230 is controlled such that the resulting low pressure ND of the refrigerant during operation of the compressor 210 approximately corresponds to the evaporation pressure of the refrigerant at the heat source temperature T WQ . Advantageously, the evaporation temperature of the refrigerant will be a few Kelvin below the heat source temperature T WQ so that the temperature difference drives heat transfer.

[0044] In the evaporator, a transfer of evaporation heat energy QV takes place from the heat source fluid of the heat source system 300, which may be a brine system, a geothermal system for utilizing heat energy QQ from the ground, an air system for utilizing energy QQ from the ambient air, or another heat source that transfers the source energy QQ to the vapor compression system 200.

[0045] The refrigerant flowing into the evaporator 240 reduces its wet vapor content by absorbing heat QQ as it flows through the evaporator 240 and advantageously leaves the evaporator 240 with a low wet vapor content or advantageously also as a superheated gaseous refrigerant. The heat source medium is conveyed through the heat source medium path of the evaporator 240 by means of a brine pump 330 in brine-water heat pumps or an outside air fan in air-to-water heat pumps, whereby the heat energy QQ is extracted from the heat source medium as it flows through the evaporator.

[0046] In the recuperator 250, heat energy Q i is transferred between the refrigerant flowing from the condenser 220 to the expansion valve 230 and the refrigerant flowing from the evaporator 240 to the compressor 210, wherein the refrigerant flowing from the evaporator 240 to the compressor 210 is further superheated.

[0047] This superheated refrigerant, which exits the recuperator 250 with a superheat temperature T Ke, is led to the refrigerant inlet connection 211 of the compressor 210.

[0048] The recuperator 250 is used in the vapor compression circuit 200 to increase the overall efficiency as a quotient of the delivered heating power QH and the absorbed electrical power P e to drive the compressor motor.

[0049] For this purpose, further heat energy Q i is extracted from the refrigerant, which releases heat energy QH to the heating medium at a heat sink temperature level in the condenser 220, by subcooling in the high pressure path of the recuperator 250.

[0050] The internal energy state of the refrigerant upon entering the evaporator 240 is reduced by this heat removal Q i , so that the refrigerant can absorb more heat energy QQ from the heat source 300 at the same evaporation temperature level.

[0051] Subsequently, after the evaporator outlet 242 from the evaporator 240, the heat energy Q i extracted in the high-pressure path is reintroduced into 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. The addition of energy advantageously reduces the wet vapor content to a state without any wet vapor content. Superheating is ensured by further energy input.

[0052] Furthermore, the following sensors are advantageously arranged to detect the operating state of the vapor compression system 200, with which a model-based pre-control 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.

[0053] On the one hand, the process values recorded by sensors advantageously provide safeguards with regard to permissible operating ranges of the components, such as in particular the compressor 210. On the other hand, model-based pre-controls, in particular a speed of the compressor 210 and / or a valve opening degree of the expansion valve, are carried out based on the sensor data, so that the controller only has to carry out minor corrections to compensate for a control deviation that is nevertheless smaller due to the pre-control: A high-pressure sensor 503 advantageously for detecting the high pressure HD of the refrigerant at the compressor outlet 212 or between the compressor outlet 212 and the expansion valve inlet 231, a hot gas temperature sensor 504 advantageously for detecting a hot gas temperature T HG of the refrigerant at the compressor outlet 212, or in the refrigeration circuit section between the compressor outlet 212 and the condenser inlet 221, an internal temperature sensor 506 advantageously for detecting the internal temperature T Ie of the refrigerant between the high-pressure side internal recuperator outlet 252 of the refrigerant from the recuperator 250 and the expansion valve inlet 231. The internal temperature is advantageously also referred to as "recuperator outlet temperature high pressure path" and advantageously a recuperator internal temperature sensor 505.The recuperator internal temperature sensor 505 advantageously detects the condenser outlet temperature T FA of the refrigerant in the flow direction at the condenser outlet or the high-pressure side recuperator inlet, and therefore the condenser outlet temperature T FA is advantageously measured by the recuperator internal temperature sensor 505.

[0054] The following sensors are particularly advantageous for carrying out the method according to the invention: A low-pressure sensor 502 for detecting the low pressure ND of the refrigerant at the compressor inlet 211, or between the expansion valve 230 and the compressor inlet 211, an evaporator outlet temperature sensor 508 for detecting the evaporator outlet temperature T Va of the refrigerant at the evaporator outlet 242 or between the evaporator outlet 242 and the low-pressure side inlet of the refrigerant into the recuperator inlet 251 of the recuperator 250 and a low-pressure temperature sensor 501 advantageously measures a compressor inlet temperature or is advantageously used to detect the refrigerant low-pressure temperature T ND or advantageously a compressor inlet temperature T KE at the compressor inlet 211, or between the low-pressure side recuperator outlet 252 of the refrigerant from the recuperator 250 and the compressor inlet 211.

[0055] The process variable that significantly influences the overall efficiency of the vapor compression circuit 200, as the quotient between the heat output QH transferred by the vapor compression circuit 200 and the electrical power P e absorbed by the compressor 210, is the superheat of the refrigerant at the compressor inlet 211. However, to maintain permissible compressor operating conditions, it is advantageous to maintain restrictions regarding the permissible superheat range of the refrigerant at the compressor inlet. Excessively low superheats, in particular, endanger the lubricating properties of the machine oil, while excessive superheats, in particular, result in excessively high hot gas temperatures.

[0056] The superheat describes the temperature difference between the measured compressor inlet temperature T KE of the refrigerant and the evaporation temperature of the refrigerant at saturated vapor.

[0057] Preferably, the compressor inlet superheat is controlled such that no condensate forms on components of the refrigeration circuit due to the water vapor content of the ambient air falling below the dew point, particularly in the section between the refrigerant outlet of the recuperator 252 and the compressor inlet 211. Although the refrigeration circuit section between the evaporator outlet 242 and the recuperator inlet 251 is usually colder because it is typically only a short pipe section, better insulation is possible compared to the section between the refrigerant outlet of the recuperator 252 and the compressor inlet 211. For example, the refrigerant separator, which is to be protected, is located at the location of the compressor inlet 211 on the compressor. This is difficult to enclose, so the temperature here should be kept high enough that no condensation occurs. The problem of condensation does not generally occur on the high-pressure side.The passage between the high-pressure recuperator outlet 252 and the inlet of the expansion valve 231 also regularly cools down to the temperature level of the refrigerant at the evaporator outlet 242, depending on the operating point under ideal heat transfer conditions in the recuperator 250. However, since this passage is also typically short and can be very well insulated, this section is generally not problematic. However, it should be noted that the process can fundamentally prevent condensate drop throughout the entire heat pump circuit.

[0058] If - for the purpose of a numerical example - an evaporation temperature level of approximately -10°C is assumed and the temperature at the brine inlet 330 is approximately -10°C, at the brine outlet 310 approximately -13°C and at the compressor inlet 5°C, the superheat would be 15K.

[0059] Room temperature sensors and room humidity sensors are advantageous for many systems, allowing for precise determination of the air's condensation conditions. For example, at 21°C and 60% relative humidity, the condensation temperature is in the range of 13°C. Under these conditions, no condensation occurs as long as the pipe temperature is above 13°C plus a buffer, e.g., 1°C.

[0060] Based on this numerical example, which is of course not limiting, a superheat of 15 K is achieved at a compressor inlet temperature of 5°C. This temperature is below the 13°C specified as the condensation temperature of the water vapor in the ambient air for the current ambient conditions. Condensation therefore occurs. If the compressor inlet temperature is to be at least 14°C, i.e., the condensation temperature plus the buffer, the superheat must be increased by 9 K, i.e., a superheat of 24 K must be maintained.

[0061] Limit values, in particular for superheating, determine the permissible superheating range of the components at the compressor inlet 211 depending on the operating point. Furthermore, there are also dependencies between the compressor inlet superheat dT ÜE and the overall efficiency of the vapor compression circuit 200 or between the compressor inlet superheat dT ÜE and a stability S of a control value R advantageous for controlling the compressor inlet superheat.

[0062] To take all of these requirements into account, the heat source medium temperature, the heating medium temperature, the compressor capacity P e and target values Z or the target value Z for calculating the compressor inlet superheat dT ÜE are advantageously used depending on the operating point of the vapor compression circuit 200. Alternatively or additionally, the target value Z can be calculated as a default value for the compressor inlet superheat dT ÜE from the refrigeration circuit measured variables that are dependent on the operating point, such as the heat source medium temperature, the heating medium temperature, the compressor capacity P e and parameterizable coefficients that are adapted to the behavior of the respective refrigeration circuit components. In the simplest case, the target value for the compressor inlet superheat dT ÜE is constant regardless of all operating conditions, e.g., 10 Kelvin. With a more complex adaptation, it is calculated as a function of an operating point variable, e.g.,the compressor power P e varies or, with even more complex adaptation, it varies as a function of several operating point variables.

[0063] A control deviation of the compressor inlet superheat dT ÜE and a control deviation of the evaporator outlet superheat dT ÜA are combined in a weighted manner, from which a total control deviation is calculated in the controller 500, which is fed into the control of the vapor compression circuit 200. Advantageously, the control deviations of the compressor inlet superheat dT ÜE and evaporator outlet superheat dT ÜA are first calculated more precisely by calculating the differences between the respective measured values and target values. Control deviation of the compressor inlet superheat dT ÜE = measured value of the compressor inlet superheat - target value of the compressor inlet superheat Z TÜE Control deviation of the evaporator outlet superheat dT ÜA = measured value of the evaporator outlet superheat - target value of the evaporator outlet superheat Z TÜA

[0064] Then, the total control deviation is advantageously calculated from the weighted influence of the control deviation of the compressor inlet superheat dT ÜE and the weighted influence of the control deviation of the evaporator outlet superheat dT ÜA in the controller 500, which is fed in to control the vapor compression circuit 200.

[0065] In the vapor compression circuit 200, after being expanded through the expansion valve 230, the refrigerant passes through two sequentially arranged heat exchangers, the evaporator 240 and the recuperator 250, in which heat energy QQ and Q i is added to the refrigerant.

[0066] In the evaporator 250, source heat energy QQ from the heat source system 300 is supplied to the refrigerant. The temperature level of the supplied source heat QQ is at a temperature level of the heat source, in particular such as the ground or the outside air.

[0067] In the recuperator 250 following in the high-pressure refrigerant flow direction S HD, thermal energy Q i is extracted from the refrigerant after leaving the condenser 220. The temperature level of the refrigerant at the condenser outlet is approximately equal to the return temperature of the heating medium.

[0068] This connection of the evaporator 240 with the recuperator 250 in series has a decisive influence on the transfer function of the control system for the control of the compressor inlet superheat dT ÜE .

[0069] The control value R is advantageously the weighted combination of the control deviation of the compressor inlet superheat dT ÜE with the control deviation of the evaporator outlet superheat.

[0070] Actuators or operating state variables with an influence on the control value R, in particular the compressor inlet superheat 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.

[0071] Actuators have a particularly advantageous influence on the control value R, in particular on the weighted link between the control deviation of the compressor inlet superheat and the control deviation of the evaporator outlet superheat. In the vapor compression circuit 200 in question, such actuators are, in particular, compressor 210, which varies the compressor speed, and expansion valve 230, which influences its opening degree. These two actuators influence the low pressure ND and the evaporation temperature level.

[0072] Not all influences are desirable here. For example, a change in the compressor speed to adjust the desired heating output without further compensatory changes to the expansion valve opening degree can change the control value R into undesirable ranges. Therefore, a model-based, supported change in the expansion valve opening degree associated with the change in the compressor speed is advantageous, and may even be necessary, to adjust R.

[0073] Advantageously, the compressor speed in the vapor compression circuit 200 is adjusted such that the heating power QH transferred from the vapor compression circuit 200 to the heating medium corresponds to the requested target value Z. To comply with this requirement, influencing the compressor speed to control the compressor inlet superheat dT ÜE is advantageously subordinate or not appropriate.

[0074] Advantageously, the degree of opening of expansion valve 230 is used as a control value for controlling the compressor inlet superheat dT ÜE. The influence of the degree of opening of expansion valve 230 on the compressor inlet superheat dT ÜE is as follows: Expansion valve 230 acts as a nozzle with an electric motor-adjustable nozzle cross-section, in which a needle-shaped nozzle needle is typically threaded into a nozzle seat by means of a stepper motor.

[0075] When operating with liquid refrigerant at the expansion valve inlet 231, the refrigerant flow rate through the expansion valve is approximately proportional to the square root of the pressure difference between the expansion valve inlet 231 and outlet 232 multiplied by a current relative value of the nozzle cross-section or degree of opening and advantageously a constant dependent on the refrigerant and a geometry of the expansion valve 230.

[0076] Since, at an operating point with a compressor speed assumed to be constant and a heating medium temperature TWS assumed to be constant, the corresponding high pressure HD of the refrigerant upon entry into the expansion valve 230 can also be assumed to be constant, the degree of opening of the expansion valve 230 significantly influences only the low pressure ND, i.e. the outlet pressure from the expansion valve 230.

[0077] If the degree of opening of the expansion valve 230 is reduced, less refrigerant passes through the expansion valve 230 at constant high pressure HD and initially constant low pressure ND. However, since the compressor 210 continues to initially deliver the same refrigerant mass flow, less refrigerant is supplied through the expansion valve 230 in the high pressure flow direction S HD than is sucked off by the compressor 210.

[0078] Since refrigerant vapor is a compressible medium, the low pressure ND on the low-pressure side of the vapor compression circuit 200 then drops. As the low pressure ND drops, the mass flow of refrigerant through the compressor 210 drops approximately proportionally, since its delivery rate can be approximately described as volume / time, due in particular to the piston strokes, and a correspondingly reduced low pressure value ND is established, at which the refrigerant mass flow supplied through the expansion valve 230 is equal to the refrigerant mass flow discharged by the compressor 210.

[0079] If the degree of opening of the expansion valve 230 is increased, more refrigerant passes through the expansion valve 230 at constant high pressure HD and initially constant low pressure ND. However, since the compressor 210 continues to initially deliver the same refrigerant mass flow, more refrigerant is supplied to the low pressure side ND of the refrigeration circuit through the expansion valve 230 than is extracted by the compressor 210. Since the refrigerant vapor is a compressible medium, the low pressure ND on the low pressure side of the vapor compression circuit 200 increases. As the low pressure ND increases, the mass flow rate of the compressor 210 increases approximately proportionally, since its delivery rate can be approximately described as volume / time, and a correspondingly increased low pressure ND is established, at which the refrigerant mass flow supplied through the expansion valve 230 is equal to the refrigerant mass flow discharged by the compressor 210.

[0080] 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 QQ from the heat source system 300 is transferred between the heat source medium and the refrigerant at different temperatures, whereby the heat flow QQ is dependent on the temperature difference between the heat source medium and the refrigerant and the heat transfer resistance of a heat transfer layer of the evaporator 240.

[0081] The heat transfer resistance between the heat source medium path of the evaporator and the refrigerant path of the evaporator is assumed to be approximately constant in a given vapor compression circuit 200. Therefore, the magnitude of the heat transfer performance in the evaporator 240 depends significantly on the integral of the temperature differences of all surface elements of the heat transfer layer.

[0082] In order to be able to transfer a sufficient amount of heat energy QQ from the heat source system 300 to the refrigerant, it must be ensured that the temperature of the heat source medium in as many surface elements of the transfer layer of the heat exchanger, here the evaporator 240, as possible is greater than the temperature of the refrigerant at the respective surface element.

[0083] If the refrigerant's state of aggregation is saturated vapor as it flows through the evaporator 240, a refrigerant temperature is established which, due to the saturation vapor characteristic curve as a material property of the refrigerant, is a function of the low pressure ND of the refrigerant. Thus, by controlling the low pressure ND or an evaporation pressure, the evaporation temperature of the refrigerant as it flows through the recuperator 250 can be indirectly controlled.

[0084] The heat energy QQ , which is transferred from the heat source system to the refrigerant flowing through the evaporator 240, influences the state of aggregation of the refrigerant.

[0085] The wet vapor fraction in the saturated refrigerant vapor decreases at constant low pressure during heat transfer to the refrigerant. In the case of incomplete evaporation, the wet vapor fraction and thus the internal energy state of the refrigerant upon exiting the heat exchanger is a function of: Wet vapor fraction upon entry into the evaporator 240, refrigerant mass flow, transferred heat output QQ , and an enthalpy difference in the wet vapor region at the respective low pressure ND, which the refrigerant has as an associated function of the pressure.

[0086] For complete evaporation, additional energy is supplied in the recuperator 250 to superheat the refrigerant beyond the saturated vapor state.

[0087] With the method, a corresponding refrigerant state is set at the outlet from the evaporator 240 under given operating conditions of the vapor compression circuit 200 as a function of the manipulated variable "degree of opening of the expansion valve 230".

[0088] In the steady state, the control system slope of the "isolated" control system "Evaporator 240" exhibits a moderate control system behavior. The control system behavior is particularly characterized by the control system output value of the evaporator outlet superheat as a function of the control system input value of the expansion valve opening degree.

[0089] Advantageously, a refrigerant, in particular a refrigerant mixture, is used that exhibits a temperature glide; in particular, R454C is advantageously used. Advantageously, with a refrigerant mixture with a temperature glide, a relative opening degree change of the expansion valve actuator of 1% rel. at the refrigerant outlet from the evaporator is usually set with a superheat change of advantageously less than 1 K.

[0090] This condition is advantageously also achieved by controlling at least one or more of the following time constants, which ultimately influence the process variable refrigerant superheat at the evaporator outlet 242: A first time constant advantageously delays the mechanical opening degree change of the expansion valve 230 by limiting the travel speed by the controller 500. The control value R is reduced in the travel speed by a braking value in this first time constant Z. The braking value can, for example, comprise the controller cycle time in which a travel step of the expansion valve 230 is controlled. A second time constant, predetermined by the controller 500, advantageously delays the setting of a corresponding low pressure when the opening degree of the expansion valve 230 changes due to the compressibility of the refrigerant vapor at low pressure ND in the low-pressure path.A third time constant is advantageously a thermal time constant of the heat transfer layer of the evaporator 240, whereby a change in the evaporation pressure and thus the evaporation temperature results in a delayed temperature change of the heat transfer layer of the evaporator, which often contains several kilograms of metal, and of the heat source medium. A fourth time constant advantageously results from delayed changes in the state of aggregation of the refrigerant upon changes in the evaporation temperature. A fifth time constant advantageously results from the transport of the refrigerant through the evaporator 240 at a finite flow velocity.

[0091] Thus, after changing the manipulated variable "degree of opening of the expansion valve 230", a delay in the corresponding change in the refrigerant state at the outlet from the evaporator outlet 242 advantageously occurs and a total time constant Z ges is advantageously in the range of 30 seconds to about 5 minutes, depending on the operating point.

[0092] After flowing through the evaporator 240, the refrigerant enters the low pressure path of the recuperator 250 at low pressure ND.

[0093] The aggregate state of the refrigerant when flowing into the recuperator 250 in a normal operating case is therefore advantageously either saturated vapor with a low vapor content between 0 and 20% or, in particular, also advantageously already superheated refrigerant.

[0094] With advantageously saturated vapor, a refrigerant temperature is established which, due to the saturation vapor characteristic of the refrigerant, is a function of the refrigerant pressure. When superheated refrigerant enters, the refrigerant temperature will assume a maximum value corresponding to the inlet temperature of the heat source medium. In this case, the value preferably corresponds to the inlet temperature of the refrigerant in the high-pressure path of the recuperator 250, i.e., the temperature of the refrigerant after exiting the condenser 220.

[0095] 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 the 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 the recuperator 250 as possible is greater than the temperature of the refrigerant of the low-pressure side refrigerant path at low pressure ND at the respective surface element.

[0096] In a heating case, the corresponding temperatures of the heating system 400 of the vapor compression system 200 are higher than the corresponding temperatures of the heat source such as the ground or the outside air.

[0097] The thermal energy Q i , which is transferred from the refrigerant at high pressure HD in 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 state of the refrigerant on the low-pressure side. The wet vapor content of the refrigerant flowing through the recuperator 250 on the low-pressure side at low pressure ND decreases during heat transfer to the refrigerant, and after complete evaporation, the refrigerant is advantageously superheated.

[0098] The internal energy state of the refrigerant upon exiting the low-pressure path of the recuperator is advantageously influenced by one or more of the following factors. It should be noted that the energy state change is based exclusively on physical dependencies, with the controller influencing the control of the actuators, which then naturally also influences physical variables such as the refrigerant mass flow: Wet steam fraction upon entry into the recuperator 250, refrigerant mass flow, transferred heat output Q i , which is advantageously controlled depending on the temperature difference between the temperature of the refrigerant at high pressure HD in the high-pressure side refrigerant path and the temperature of the refrigerant in the low-pressure side refrigerant path at low pressure ND, and / or an enthalpy difference in the wet steam region at the respective low pressure ND.

[0099] This advantageously ensures that, depending on the given operating conditions of the vapor compression circuit 200 and depending on the manipulated variable "degree of opening of expansion valve 230", a corresponding refrigerant state is established at the outlet 252 from the recuperator 250 at low pressure ND.

[0100] In the steady state, the control path slope of the "isolated" control path at the low pressure ND of the refrigerant in the low-pressure side path of the recuperator 250 results in a control path behavior with a high slope, with an approximately constant internal energy state of the refrigerant at the inlet 251 into the low-pressure side ND path of the recuperator 250. With a relative opening degree change of the expansion valve of 1%, in particular, a superheat change at the outlet of the refrigerant from the recuperator 250 of advantageously about 10 K or even more than 10 K results.

[0101] Compared to the recuperator 250, a significantly higher heat transfer takes place in the evaporator 240 between the source medium and the refrigerant in the evaporator 240.

[0102] Thus, a significantly higher heat transfer occurs in the evaporator 240 than in the recuperator 250, which is necessary because a significantly greater amount of energy is to be extracted from the environment by means of the evaporator 240 than is transferred solely in the recuperator 250 within the refrigeration circuit. However, the driving temperature difference in the recuperator can, for example, be between 20 and 60 K, while in the evaporator it is only between 3 and 10 K. In order to be able to transfer the desired energies despite different driving temperature differences, the exchange surface of the evaporator, for example, is designed to be approximately 5 to 20 times larger than that of the recuperator 250.

[0103] The setting of this state is advantageously carried out using at least one of the following time constants Z: An eleventh time constant Z11 advantageously specifies a delay in the mechanical opening degree change of the expansion valve 230 by limiting a travel speed. A twelfth time constant Z12 advantageously has a delayed effect on the setting of a corresponding low pressure ND upon changes in the opening degree of the expansion valve 230 due to the compressibility of the refrigerant vapor in the low-pressure path ND. A thirteenth time constant Z13 is a thermal time constant of the heat transfer layer of the evaporator. Thus, a change in the evaporation pressure and thus the evaporation temperature causes a delayed temperature change of the heat transfer layer, which often contains several kilograms of metal, and of the refrigerant in the low-pressure path of the evaporator 240. A fourteenth time constant Z14 is advantageously determined or specified from delayed changes in the state of aggregation of the refrigerant upon changes in the evaporation temperature. A fifteenth time constant Z14Time constant Z 15 results advantageously from the transport of the refrigerant through the evaporator 240 with a finite flow velocity and is taken into account.

[0104] The low-pressure side refrigerant path of the recuperator 250 is fed from the evaporator outlet 242 of the evaporator 240. Here, too, the internal energy state of the refrigerant is already delayed by at least two time constants Z, Z 11 , Z 12 , Z 13 , Z 14 , Z 13 , Z ges after changing the manipulated variable "expansion valve opening degree."

[0105] After changing the control variable "Opening degree of expansion valve 230", a further delay of the corresponding refrigerant state change occurs due to the time behavior of the recuperator 250 when exiting the low-pressure side refrigerant path of the recuperator 250.

[0106] The time behavior of the recuperator 250 can advantageously be taken into account as the total recuperator time constant Z ges depending on the respective operating point of the vapor compression circuit in the range between approximately 1 minute and 30 minutes.

[0107] A weighted combination of the compressor inlet superheat dT ÜE and the evaporator outlet superheat dT ÜA is advantageously carried out, in particular by means of a weighted combination of the control deviation of the compressor superheat and the control deviation of the evaporator outlet superheat dT ÜA the total control deviation is calculated, which is fed into the controller 500 for controlling the vapor compression circuit 200.

[0108] The compressor inlet superheat 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: Step 1

[0109] First, the process variables compressor inlet superheat dT ÜE are advantageously measured as the main control variable and the evaporator outlet superheat dT ÜA are advantageously measured as an auxiliary variable in a first process step.

[0110] For this purpose, an evaporation temperature of the refrigerant at the respective detection point is either directly determined by measurement, with a temperature sensor which is positioned in such a way that it detects a temperature corresponding to the refrigerant temperature in the wet steam area or indirectly determined by measurement, with a pressure sensor which detects a refrigerant pressure of the refrigerant evaporating in the wet steam area and the evaporation temperature is then calculated from the refrigerant-specific relationship between pressure and temperature in the wet steam area.

[0111] Furthermore, the refrigerant temperature is measured by means of temperature sensors 501, 508 at the respective superheat measuring point, in particular at the evaporator outlet 242 and / or at the compressor inlet 211. The temperature difference 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 superheat of the refrigerant at the measuring point.

[0112] The starting values for the calculation in step 1 are the compressor inlet superheat dT ÜE and the evaporator outlet superheat dT ÜA . Step 2

[0113] The process variables compressor inlet superheat dT ÜE and evaporator outlet superheat dT ÜA are advantageously calculated in a second step to form assigned control deviations with respectively assigned setpoints: The setpoint for the compressor inlet superheat dT ÜE is advantageously varied in the range between approximately 5 K and 20 K to ensure the permissible compressor operating range and the highest possible efficiency of the refrigeration circuit.

[0114] The setpoint for the evaporator outlet superheat dT ÜA at the evaporator outlet 242 is then varied depending on the refrigeration circuit operating mode and the refrigeration circuit operating point such that the evaporator superheat in the steady-state control case approximately corresponds to the resulting process value of the evaporator outlet superheat dT ÜA. This setpoint for the evaporator outlet superheat dT ÜA can be precalculated model-based depending on an operating mode or an operating point, depending on the evaporation temperature, the condensation temperature, the compressor capacity, a setpoint for the compressor inlet superheat dT ÜE at the compressor inlet 211 and / or component properties, and can be adaptively corrected.

[0115] The control deviation of the compressor inlet superheat dT ÜE is then calculated by subtracting the setpoint value of the compressor inlet superheat dT ÜE from the process value of the compressor inlet superheat dT ÜE.

[0116] The control deviation of the evaporator outlet superheat dT ÜA is then calculated by subtracting the setpoint of the evaporator outlet superheat dT ÜA from the process value of the evaporator outlet superheat dT ÜA. Step 3

[0117] In a third process step, the control deviation of the compressor inlet superheat dT ÜE and the control deviation of the evaporator outlet superheat dT ÜA are advantageously combined to form a total control deviation superheat.

[0118] The combination is carried out in particular by means of a weighted addition of the individual control deviations.

[0119] The weighting influence is a measure of the proportional combination of the individual control deviations and can, in extreme cases, result in the exclusive inclusion of only one individual control deviation, but usually the weighted inclusion of both individual control deviations.

[0120] It is advantageous to estimate the weighting influence as a value between 0 and 1, i.e. 0 to 100%, and this value is used to determine the degree to which the control deviation of the compressor inlet superheat dT ÜE is included in the total control deviation, which results in the following dependency for the calculation of the total control deviation:

[0121] The value of the weighting influence can advantageously be varied depending on the operating mode and / or the operating point of the heat pump 100: During the operating mode transition between operating mode = operation with the compressor 210 switched off and operating mode = operation with the compressor 210 switched on in heating mode, due to the dynamic changes in the process values when starting up the vapor compression system 200, it is advantageous to initially only include the control deviation of the evaporator outlet superheat dT ÜA in the overall control deviation. In particular, the value of a weighting influence is then initially = 0 or a value advantageously below 20%. After a stabilization phase of the vapor compression system 200, it is advantageous not to switch spontaneously to the value of the weighting influence designed for control operation, but rather to design the transition in a ramp-like manner. In this case, it is advantageous for the value of the weighting influence to be increased from the starting value = 0, or a value in particular below 20%, advantageously in a ramp-like manner to the intended target value.This particularly prevents value discontinuity during spontaneous switching, thus avoiding control oscillations. The target value of the weighting influence is advantageously adapted to the respective operating mode and the operating point. Operating points characterized by an increased tendency to oscillate advantageously require a lower weighting of the control deviation of the compressor inlet superheat dT ÜE . In particular, this avoids a tendency to oscillate due to the control-technically critical signal behavior of the compressor inlet superheat dT ÜE due to the longer signal delay and greater path steepness compared to the evaporator outlet superheat dT ÜA . Step 4:

[0122] In a fourth method step, the calculated total control deviation of the superheat is then processed in the controller 500, which controls the corresponding actuators of the refrigeration circuit, in particular the expansion valve 230 with the adjustable opening degree and / or the compressor 210 with adjustable compressor speed, in such a way that in the controlled case a control deviation of the superheat is set equal to approximately 0 Kelvin.

[0123] A P, I, PI, PID controller can be used, whereby the control components are advantageously adapted dynamically to the respective operating mode and the operating point.

[0124] To protect components (mechanical components, refrigeration oil), the compressor must be protected from excessively high refrigerant gas temperatures when leaving the compression chamber. The compressor manufacturer can specify compliance with an absolute (applicable for all operating conditions) or relative (applicable depending on operating conditions) maximum hot gas temperature (e.g. 120 °C).

[0125] In order to avoid an inadmissible exceedance of the hot gas temperature limit, a control-relevant hot gas temperature limit (which is below the hot gas temperature limit specified by the compressor manufacturer) is defined (e.g. 110 °C), which is used as a limit for the hot gas temperature-limiting controller activities.

[0126] Based on the control-relevant hot gas temperature limit, a (simplified) thermodynamic calculation is used to calculate for the current operating point of the refrigeration circuit which superheat of the refrigerant at the inlet of the compressor, called compressor inlet superheat, would result when operating with the control-relevant hot gas temperature limit.

[0127] 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).

[0128] From a purely thermodynamic perspective, the compressor speed plays no role in the calculation, but the dependence of the compressor's compression losses as a function of the compressor speed may have an influence on the thermodynamic behavior.

[0129] Preferably, a parameterizable approximation of the compressor behavior with respect to hot gas temperature is used to estimate a temperature difference between hot gas temperature and compressor inlet superheat that occurs at the respective operating point of the compressor on the basis of the process data evaporation temperature and condensation temperature.

[0130] Based on this calculation, it is then possible to calculate a corresponding compressor inlet superheat at a target hot gas temperature.

[0131] With the help of the Parameter Limitation Hot Gas Temperature Suction Gas Superheat V- ND allows a limitation of the hot gas temperature to be set, which is then controlled by a corresponding limitation of the compressor inlet superheat.

[0132] The corresponding compressor inlet superheat calculated on the basis of an estimated hot gas temperature is used as a process variable Setpoint superheat compressor inlet V-ND maximum estimated led.

[0133] A calculation of the maximum permissible Setpoint superheat compressor inlet V-ND maximum estimated based on the Parameter Limitation Hot Gas Temperature Suction Gas Superheat V-ND The set hot gas temperature limit is then determined using the Temperature difference hot gas temperature minus superheat compressor inlet V-ND estimated calculated as an auxiliary value as follows:

[0134] As a further refinement, the current hot gas temperature can also be included in the calculation, since the model-based calculation can be erroneous with regard to the time behavior, tolerances in the recording 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 recording and inclusion of the actual hot gas temperature is helpful.

[0135] The correction is made by calculating the difference between Parameter Limitation Hot Gas Temperature Suction Gas Superheat V-ND and Hot gas temperature, Is the Hot gas temperature larger than Parameter Limitation Hot Gas Temperature Suction Gas Superheat V-ND, i.e. the difference is negative, the pre-calculated Setpoint superheat compressor inlet V-ND maximum estimatedby the difference multiplied by Parameter P-factor Limitation of hot gas temperature Suction gas superheat V-ND reduced and as Setpoint superheat compressor inlet V-ND maximum further processed Is the Hot gas temperature less than Parameter Limitation Hot Gas Temperature Suction Gas Superheat V-ND, If the difference is positive, the pre-calculated Setpoint superheat compressor inlet V-ND maximum estimated by the difference multiplied by Parameter P-factor Bear. Hot gas temperature Suction gas superheat V-ND increased and as Setpoint superheat compressor inlet V-ND maximum further processing

[0136] The following applies:

[0137] A further alternative design is the maximum formation of the superheat setpoint calculated for limiting the hot gas temperature and a lower limit provided for limiting the value range of this superheat setpoint, which is integrated into the formulas; here the value 0 Kelvin is implemented as an example.

[0138] Such a limitation can be advantageous because the model-based calculation of the superheat setpoint calculated for limiting the hot gas temperature can also result in negative superheat setpoints, which, if controlled to these values, would then cause undesirable wet steam intake into the compressor. This is avoided by limiting the superheat to a minimum, which can also be designed in the (slightly) negative value range if necessary. In this case, it is accepted (because this is of secondary importance) that the hot gas temperature is not limited exactly to the desired limit, but may exceed it slightly.

[0139] In a final step, the compressor inlet superheat setpoint designed for optimum efficiency and the superheat setpoint calculated for the limitation of the hot gas temperature are combined in such a way that, in the event of a required hot gas temperature limitation, a superheat setpoint optimized in terms of efficiency can be reduced:

[0140] Calculated with exemplary values without hot gas temperature influence: Setpoint specification efficiency optimized: Setpoint superheat compressor inlet V-ND efficiency = 30 K

[0141] Temperature difference calculated from boiling point HD, dew point temperature ND hot gas temperature minus superheat compressor inlet V-ND estimated = 100 K.

[0142] With Parameter Limitation Hot Gas Temperature Suction Gas Superheat V-ND = 110°C calculated Setpoint superheat compressor inlet V-ND maximum estimated = 10 K

[0143] Calculated with the above calculation results Setpoint superheat compressor inlet V-ND = 10 K

[0144] The values are of course to be understood as examples; for actual applications, higher, lower and values adapted to the respective operating point are also conceivable.

Claims

1. A method for controlling a compression refrigeration system (200) comprising: - a refrigerant circuit comprising - a refrigerant, - an evaporator (240), - a compressor (210), - a condenser (220), - a throttle valve (230), - an internal heat exchanger (250) for transferring thermal energy of the refrigerant before entering the throttle valve (230) to the refrigerant before entering the compressor (210), and - a control unit (500) a) for detecting overheating (TÜE) of the refrigerant at the entry to the compressor, the overheating (TÜE) being defined as a difference between a dew point temperature and a temperature of the refrigerant, and b) for controlling the throttle valve (230) based on the overheating (TÜE), wherein the method includes the following steps: - determining a first target overheating of the refrigerant at the entry to the compressor, wherein the first target overheating maximises the efficiency of the compression refrigeration system as a function of an operating point of the compression refrigeration system, - determining a second target overheating of the refrigerant at the entry to the compressor based on a maximum allowable hot gas temperature at the outlet of the compressor, and - control of the throttle valve (230) based on the lower value from the first target overheating and the second target overheating.

2. Method according to claim 1, wherein the determination of the second target overheating is based on a refrigeration model which establishes a relationship between the overheating at the entry to the compressor and the hot gas temperature at the outlet of the compressor.

3. Method according to claim 2, wherein the refrigeration model contains at least the following input variables: Low pressure, high pressure, speed of the compressor, or the following input variables: Dew point temperature at low pressure, boiling point temperature at high pressure, speed of the compressor.

4. Method according to claim 2 or 3, wherein the refrigeration model comprises a linear function of quadratic order of the boiling point temperature at high pressure and the dew point temperature at low pressure.

5. Method according to any one of the preceding claims, wherein the maximum allowable hot gas temperature at the outlet of the compressor is defined as a temperature which is below a hot gas temperature limit defined for the compressor, in particular defined by the manufacturer.

6. Method according to any one of the preceding claims, wherein the control of the throttle valve (230) is further based on a currently measured hot gas temperature.

7. A compression refrigeration system (200) comprising - a refrigerant circuit comprising - a refrigerant, - an evaporator (240), - a compressor (210), - a condenser (220), - a throttle valve (230), - an internal heat exchanger (250) for transferring thermal energy of the refrigerant before entering the throttle valve (230) to the refrigerant before entering the compressor (210), and - a control unit (500) a) for detecting overheating (TÜE) of the refrigerant at the entry to the compressor, the overheating (TÜE) being defined as a difference between a dew point temperature and a temperature of the refrigerant, and b) for controlling the throttle valve (230) based on the overheating (TÜE), wherein the control unit (500) is configured for: - determining a first target overheating of the refrigerant at the entry to the compressor, wherein the first target overheating maximises the efficiency of the compression refrigeration system as a function of an operating point of the compression refrigeration system, - determining a second target overheating of the refrigerant at the entry to the compressor based on a maximum allowable hot gas temperature at the outlet of the compressor, and - control of the throttle valve (230) based on the lower value from the first target overheating and the second target overheating.

8. The compression refrigeration system according to claim 7, wherein the refrigerant has a temperature glide, wherein the refrigerant in particular comprises or consists of R454C.

9. A heat pump (100), in particular a heat pump (100) installed inside a building, comprising a compression refrigeration system (200) according to claim 7 or claim 8.