Method for operating a compression refrigeration system and compression refrigeration system

Abstract

The present invention relates to a method for operating a compression refrigeration system (200) and an associated compression refrigeration system (200). The method comprises the following steps: calculating a temperature difference between the dew point temperature and the boiling point temperature, called the temperature glide, of the refrigerant at the current operating point of the compression refrigeration system based on a measured low pressure (LP) of the refrigerant; calculating a temperature difference limit value relative to the calculated temperature glide of the measured low pressure (LP) of the refrigerant at the current operating point; evaluating the current evaporator outlet superheat for compliance with the temperature difference limit value; and providing a setpoint for superheat, defined as a difference between the temperature of the refrigerant and the dew point temperature, at the inlet to the compressor (210), called the compressor inlet superheat.Assessing whether the current compressor inlet superheat exceeds the target value, and detecting a refrigerant shortage if both results of the assessment steps are continuously met for a specified period.

Description

[0001] The invention relates to a method for operating a compression refrigeration system and an associated compression refrigeration system comprising a refrigerant, an evaporator, a compressor, a condenser, a throttling device and a control unit.

[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 a control unit, which for example has a regulator, are known.

[0003] The refrigerant is forced through the condenser, where it releases heat to a heating medium in a heat sink system. Internal heat is transferred in an optional internal heat exchanger, for example in the form of a recuperator, between the refrigerant flowing under high pressure from the condenser to the expansion valve and the refrigerant flowing under low pressure from the evaporator to the compressor.

[0004] The refrigerant is then guided in a high-pressure flow direction to an expansion valve controlled by the regulator, where the refrigerant is reduced from high pressure to low pressure according to a set control value. The refrigerant, now at low pressure, evaporates in the evaporator, absorbing heat from the surrounding environment.

[0005] From DE 101 59 892 A1, it is known to use a recuperator in a refrigeration machine, particularly in a heat pump, to increase the heating capacity 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 capacity 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.

[0006] 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, an throttling device, an evaporator, and a 4-2-way valve unit for switching between a first (heating) and a second (cooling) operating mode. The flow direction of the refrigerant in the refrigerant circuit can be reversed such that the first heat exchanger serves to condense the refrigerant in the first operating mode and to evaporate it in the second operating mode, and the second heat exchanger serves to evaporate the refrigerant in the first operating mode and to condense it in the second operating mode, wherein the first heat exchanger in the refrigerant circuit is configured to operate as a counterflow heat exchanger in both heating and cooling modes.

[0007] The control system of the compression refrigeration system must meet various requirements; for example, a high coefficient of performance (COP) is required to ensure the most energy-efficient operation possible. However, it is also crucial that the operating limits of the components are adhered to.

[0008] The coefficient of performance (COP) depends significantly on the correct amount of refrigerant circulating through the refrigeration circuit. Excess refrigerant can be collected in a refrigerant receiver, ensuring the correct amount circulates. If too little refrigerant is available to circulate, the efficiency of the compression refrigeration system will be poor. Such a refrigerant shortage can have various causes, each requiring different solutions.

[0009] Against this background, it is an object of the present invention to propose a method for operating the aforementioned compression refrigeration system that can reliably detect a refrigerant shortage situation and its cause.

[0010] The problem is solved by the method features of claim 1 and the device features of claim 6.

[0011] Accordingly, a method for operating a compression refrigeration system is proposed, comprising a refrigerant exhibiting a temperature glide, an evaporator, a compressor, a condenser, an throttling device, an internal heat exchanger (the internal heat exchanger being configured to transfer thermal energy from the refrigerant before it enters the throttling device to the refrigerant before it enters the compressor), and a control unit (the control unit being configured a) to detect superheating of the refrigerant after it exits the evaporator and to detect superheating of the refrigerant upon entry into the compressor, and b) to control the throttling device, the control of which is based on a control variable derived from both superheatings, and c) to detect a refrigerant shortage, at least in the condenser).

[0012] The procedure involves calculating a temperature difference between the dew point temperature and the boiling point temperature, called temperature glide, of the refrigerant at the current operating point of the compression refrigeration system based on a measured low pressure of the refrigerant.

[0013] The temperature glide depends on the current pressure to which it is referenced. According to the invention, the relevant pressure is in the low-pressure path of the refrigerant. Here, the temperature glide can be determined using a characteristic curve known for the refrigerant used. Refrigerants exhibiting a temperature glide are typically mixtures of two or more refrigerant components. In this case, a correction of the known characteristic curve due to deviations in the mixing ratio is particularly preferred.

[0014] The procedure also includes calculating a temperature difference limit value relative to the calculated temperature glide of the measured low pressure of the refrigerant at the current operating point.

[0015] The procedure also includes an assessment of the current evaporator outlet superheat to determine whether the temperature difference limit has been undercut.

[0016] For example, and this is not a limiting factor, the temperature glide in typical operating ranges for typical refrigerants, such as R454C, can be in the range of 5 K to 15 K, particularly between 7 and 8 K. The refrigerant exits the evaporator with an evaporator outlet superheat of -2 or -2.5 K, meaning the refrigerant temperature at the evaporator outlet is lower than the temperature at which evaporation is complete. The temperature difference limit can then determine that the condition for a refrigerant shortage is met when the refrigerant exits the evaporator at a temperature lower than -4 or -5 K below the temperature at which evaporation is complete.

[0017] The process further includes providing a setpoint for superheat, defined as the difference between the refrigerant temperature and the dew point temperature, at the compressor inlet, called compressor inlet superheat.

[0018] The procedure also includes an assessment of the current compressor inlet superheat to determine whether it exceeds the target value for compressor inlet superheat,

[0019] The procedure ultimately includes the detection of a refrigerant shortage situation if both results of the evaluation steps are continuously met for a specified period of time.

[0020] The method according to the invention therefore makes it possible to reliably detect an existing refrigerant shortage situation in compression refrigeration systems with an internal heat exchanger based on two superheat values ​​of the refrigerant in the low-pressure path, namely at the exit from the evaporator and at the entry into the compressor.

[0021] The specified time is preferably at least one minute, and particularly preferably 10 minutes. This gives the control loop the opportunity to return stably to the desired state, which requires a certain amount of time.

[0022] Only after a stabilization period is it worthwhile to assess a refrigerant shortage. If a refrigerant shortage actually exists, it is caused, for example, by small leaks over days, months, or years; no critical damage occurs, but the coefficient of performance (COP) decreases.

[0023] Preferably, the refrigerant has a temperature glide, wherein the refrigerant is in particular R454C or consists thereof, and wherein the compression refrigeration system in particular includes an internal heat exchanger for transferring thermal energy from the refrigerant before it enters the throttling device to the refrigerant before it enters the compressor. This is particularly relevant because high-performance internal heat exchangers are regularly used in refrigeration circuits with R454C.

[0024] Preferably, the method further comprises the following step: conducting an attempt at refrigerant recirculation into the condenser by temporarily increasing the setpoint of the compressor speed and / or temporarily increasing the setpoint of the evaporator outlet superheat.

[0025] In this preferred design, an attempt is made to resolve the identified refrigerant shortage. A distinction must be made between two fundamentally different mechanisms that lead to a refrigerant shortage in the condenser. In the first mechanism, leaks, leakage, etc., result in insufficient refrigerant in the system. This situation can only be remedied by manual intervention, i.e., refilling the refrigeration circuit with the required amount of refrigerant and sealing the circuit.

[0026] The second scenario of refrigerant shortage in the condenser is caused by refrigerant condensing within components of the refrigeration circuit. At certain operating points, refrigerant can condense, for example, at low positions in the evaporator, and due to low flow velocities and gravity, it cannot escape from the evaporator along with the gaseous refrigerant. In other words, the refrigerant remains trapped within the component and is unavailable for the cycle. It has been found that the risk of refrigerant accumulation in components of the refrigeration circuit increases, particularly with large evaporators and operating points with low power requirements, such as low compressor speeds. In such refrigerant shortage situations, shifting the operating point can initiate successful refrigerant recirculation.

[0027] Preferably, the method further comprises the following step: generating an error state upon unsuccessful attempt at refrigerant recovery, in particular upon multiple unsuccessful attempts at refrigerant recovery within a defined time period.

[0028] If the attempt to recover the refrigerant is unsuccessful, a leak and an insufficient amount of refrigerant are highly likely. Only manual intervention, which, as described above, involves refilling the refrigerant and sealing the refrigeration circuit, will then be suitable to restore optimized efficiency.

[0029] The fault status can preferably be displayed on the housing of the compression refrigeration system, for example, on a display. Alternatively or additionally, the fault status, along with further diagnostic information, can be transmitted to a server and / or displayed in an app. The diagnostic information can assist with repairs.

[0030] In one embodiment, a fixed time duration and a fixed value relative to the temperature glide can be set, indicating that significantly insufficient energy is being transferred in the evaporator, suggesting refrigerant displacement. At various operating points, a regular superheat value is assumed, which can also be negative, particularly based on known, model-based feedforward characteristics. Based on this regular superheat value, it can be determined according to the invention that a refrigerant shortage limit is always reached when a certain threshold, the temperature difference limit, for example 1 K or 2 K, particularly preferably 1.5 K, is reached by a distance from this limit.

[0031] Preferably, the temperature difference limit value is calculated relative to the calculated temperature glide of the measured low pressure of the refrigerant, depending on the current operating point of the compression refrigeration system.

[0032] Preferably, the temperature difference limit value is set proportionally to the difference between the heat sink temperature and the heat source temperature relative to the calculated temperature glide.

[0033] In this embodiment, the temperature difference limit can therefore also be corrected using a model, thus enabling a more precise diagnosis of refrigerant shortage that is dependent on the operating situation. Superheat reduction into negative ranges that occurs during regular operation—which must be prevented at the latest when the refrigerant enters the compressor—is made possible by the use of a recuperator and comes very close to the ranges in which refrigerant shortage is detected.

[0034] For this reason, it is advantageous to make the temperature difference limit dependent not only on the glide, which is a physical property of the refrigerant, but also on the evaporator's operating point, i.e., the temperature profile of the media temperatures. If there is a small driving temperature difference, only a small negative superheat is expected, so the temperature difference limit can be set close to the small negative superheat value. Therefore, for a large driving temperature difference, for example, at high supply temperatures and / or low heat source temperatures, it is advantageous to implement a larger temperature difference limit.

[0035] The object of the invention is further achieved by a compression refrigeration system comprising a refrigerant exhibiting a temperature glide, an evaporator, a compressor, a condenser, an expansion valve, an internal heat exchanger (the internal heat exchanger being designed to transfer thermal energy from the refrigerant in the high-pressure path before entering the expansion valve to the refrigerant in the low-pressure path before entering the compressor), and a control unit. The control unit is configured a) to detect superheating of the refrigerant after exiting the evaporator and to detect superheating of the refrigerant upon entering the compressor, b) to control the expansion valve using a control variable based on both superheating levels, and c) to detect a refrigerant shortage, at least in the condenser.

[0036] The control unit is further designed to: calculate a temperature difference between the dew point temperature and the boiling point temperature, called the temperature glide, of the refrigerant at the current operating point of the compression refrigeration system based on a measured low pressure of the refrigerant; calculate a temperature difference limit value relative to the calculated temperature glide of the measured low pressure of the refrigerant at the current operating point; assess whether the current evaporator outlet superheat falls below the temperature difference limit value; provide a setpoint for superheat, defined as the difference between the refrigerant temperature and the dew point temperature, at the compressor inlet, called compressor inlet superheat; assess whether the current compressor inlet superheat exceeds the setpoint; and detect a refrigerant shortage.if both results of the evaluation steps are continuously met for a specified period of time.

[0037] The problem is further solved according to the invention by a heat pump with a compression refrigeration system according to the invention.

[0038] Preferred embodiments and associated advantages of the method according to the invention are also applicable to the compression refrigeration system and the heat pump according to the invention.

[0039] The compression refrigeration system according to the invention is suitable regardless of the type of heat pump, for example air / water, brine / water heat pumps, and regardless of the location of installation.

[0040] Further advantages and preferred configurations are described in more detail below with reference to the accompanying figures. These show: Fig. 1 heat pump 100 with one vapor compression circuit 200 Fig. 2 log p / h - Diagram of the vapor compression process with recuperator 250 Fig. 3 the log p / h diagram of the Fig. 2 without refrigerant shortage and Fig. 4. A log p / h diagram showing refrigerant shortage.

[0041] Fig. Figure 1 shows a schematic and exemplary heat pump. 100 The heat pump 100 It essentially consists of a vapor compression system forming a compression refrigeration system. 200 . Fig. Figure 2 shows an exemplary log p / h diagram of the vapor compression process, which is present in the vapor compression system. 200 The ongoing process is depicted with the process variables shown. The steam compression system 200 contains the following components: • A compressor 210 for compressing the superheated refrigerant, • a liquefier 220 , with a refrigerant-side condenser inlet 221and a condenser outlet 222 for the transfer of heat energy Q H from the steam compression system 200 to a heating medium of a heating system 400 , with a heating medium inlet 401 , a heating medium leak 402 and a heating medium pump 410 , to a building heating system or a hot water heating system, • advantageously a refrigerant collector 260 , which is used as a refrigerant reservoir to compensate for varying refrigerant quantity requirements depending on operating conditions, • a throttling device designed as an expansion valve 230 to expand the refrigerant, • an evaporator 240 , with an evaporator inlet 241 , for the transmission of source energy Q Q from a heat source system 300 , with a heat source inlet 320 and a heat source outlet 310 , wherein the heat source system300 in particular a brine system which can store thermal energy Q Q absorbs heat from the ground or an air system which absorbs heat energy Q Q absorbs from the ambient air and transfers it to the vapor compression system 200 emits or any other heat source • a recuperator as an example of an optional internal heat exchanger 250 , which is designed to dissipate internal heat energy Q i between the liquefier 220 to the expansion valve 230 flowing refrigerant onto the surface from the evaporator 240 to the compressor 210 to transfer flowing refrigerants and • a refrigerant, in particular a refrigerant mixture of at least two substances or two refrigerants which flows in one direction S HD and S ND through the vapor compression circuit 200 flows, whereby in the vapor compression cycle 200 Refrigerant vapor through the compressor210 is brought to a high pressure (HD) and becomes a liquefier. 220 is guided, whereby a high-pressure path with the high-pressure flow direction S HD from the compressor 210 up to the expansion valve 230 is formed. After the expansion valve 230 up to the compressor 210 is a low-pressure path with a low-pressure flow direction S ND of the refrigerant in which the evaporator 240 lies.

[0042] The actuators listed below are advantageously at least partially connected to the controller via a data connection. 510 , which can be connected via cable, radio or other technologies: compressor 210 , heating medium pump 410 , brine pump 330 , expansion valve 230 Compressor inlet temperature sensor 501 , low pressure sensor 502 , high-pressure sensor 503 Hot gas temperature sensor 504, Recuperator inlet temperature sensor 505 Recuperator outlet temperature sensor 506 and / or evaporator outlet temperature sensor 508 Additionally or alternatively, a [something] can be [something] in the Fig. 1. Evaporator inlet temperature sensor (not shown): the temperature at the evaporator inlet 241 determine.

[0043] In the Fig. The first example shown is the heat pump. 100 The example shown is a brine-to-water heat pump. Of course, similar considerations and advantages can be achieved with air-to-water heat pumps. In particular, with air-source heat pumps, a brine pump can be used instead of a brine circuit. 330 A fan / ventilator is positioned as a heat source.

[0044] The compressor 210 It is used to compress the superheated refrigerant from an inlet port. 211 at a compressor outlet pressure PVa at a compressor outlet temperature corresponding to the hot gas temperature at the compressor outlet 212The compressor 210 It typically contains a drive unit with an electric motor, a compression unit, and advantageously the electric motor can be operated at variable speeds. The compression unit can be designed as a rotary piston unit, scroll unit, or otherwise. At the compressor outlet 212 is the compressed superheated refrigerant at the compressor outlet pressure P Va at a higher pressure level, especially a high pressure (HD), than at the inlet port 211 with a compressor inlet pressure P Ve , in particular a low-pressure LP, at a compressed inlet temperature T VE , which determines the state of the refrigerant temperature at the inlet connection 211 describes the process of entering a compression chamber.

[0045] In the liquefier 220 The transfer of heat energy takes place Q H from the refrigerant of the vapor compression system 200 to a heating medium of the heat sink system400 . First, the liquefied 220 The refrigerant is deheated, whereby superheated refrigerant vapor transfers some of its heat energy to the heating medium of the heat sink system through a temperature reduction. 400 transfers.

[0046] After the refrigerant vapor has been deheated, this advantageously takes place in the condenser. 220 further heat transfer Q H This occurs through condensation of the refrigerant during the phase transition from the gaseous to the liquid phase. This process releases further heat. Q H from the refrigerant in the vapor compression system 200 to the heating medium of the heat sink system 400 transmitted.

[0047] The one in the liquefier 220 The adjusting high pressure (HD) of the refrigerant corresponds to the operation of the compressor. 210 approximately with a condensation pressure of the refrigerant at a heating medium temperature Tws in the heat sink system.

[0048] The heating medium, especially water, is circulated by means of a heating medium pump. 410 through the heat sink system 400 in a SW direction through the condenser 220 promoted, thereby the heat energy is Q H transferred from the refrigerant to the heating medium.

[0049] In the following collector 260 is produced from the liquefier 220 escaping refrigerant is stored, the amount of which depends on the operating point of the vapor compression circuit. 200 It should not be fed into the circulating refrigerant. It is taken from the condenser. 220 More refrigerant was injected than through the expansion valve. 230 As the collector is forwarded, it fills up. 260 , otherwise it will become emptier or emptied.

[0050] In the following recuperator 250 , which can also be referred to as an internal heat exchanger, internal heat energy Q i from the refrigerant under high pressure (HD), which is supplied by the condenser 220 to the expansion valve 230 in a high-pressure flow direction S HD flows, transferred to the refrigerant flowing under low pressure (LP), which travels from the evaporator to the compressor in a low-pressure flow direction. S ND The fluid flows and is transferred. This process involves transferring the fluid from the condenser to the expansion valve. 230 The flowing refrigerant is advantageously subcooled.

[0051] First, the refrigerant flows through an expansion valve inlet. 231 into the expansion valve. In the expansion valve 230 The refrigerant pressure is throttled from high pressure (HP) to low pressure (LP) by the refrigerant advantageously passing through a nozzle arrangement or throttle with an advantageously variable opening cross-section, wherein the low pressure is advantageously approximately the suction pressure of the compressor. 210This corresponds to a replacement for an expansion valve. 230 Any other pressure-reducing device can also be used. Pressure-reducing pipes, turbines, or other pressure-relieving devices are advantageous.

[0052] An opening degree of the expansion valve 230 is adjusted by an electric motor, usually designed as a stepper motor, which is controlled by the control unit or regulator. 500 It is controlled. The low pressure (LP) at the expansion valve outlet is controlled in this process. 232 of the refrigerant from the expansion valve 230 controlled in such a way that the resulting low pressure (LP) of the refrigerant during compressor operation 210 approximately with the evaporation pressure of the refrigerant and the heat source medium temperature T WQ This corresponds to the evaporation temperature of the refrigerant, which is advantageously a few Kelvin below the temperature of the heat source medium T. WQThey lie so that the temperature difference drives heat transfer.

[0053] In the evaporator, heat energy Qv is transferred from the heat source fluid of the heat source system. 300 , which is a brine system, a geothermal system for the use of heat energy Q Q from the ground, an air system for energy utilization Q Q from the ambient air or another heat source that provides the source energy Q Q to the vapor compression system 200 hands over.

[0054] The one in the evaporator 240 The incoming refrigerant is reduced as it flows through the evaporator. 240 through heat absorption Q Q its wet steam content and leaves the evaporator 240 Advantageously with a low wet steam content or also advantageously as a superheated gaseous refrigerant. The heat source medium is supplied by means of a brine pump. 330In the case of brine-to-water heat pumps or an outdoor air fan in the case of air-to-water heat pumps, through the heat source medium path of the evaporator. 240 promoted, whereby the heat energy is transferred to the heat source medium as it flows through the evaporator. Q Q is withdrawn.

[0055] In the recuperator 250 heat energy is Q i between the liquefier 220 to the expansion valve 230 flowing refrigerant onto the surface from the evaporator 240 to the compressor 210 flowing refrigerants are transferred, with the one from the evaporator 240 to the compressor 210 The flowing refrigerants, in particular, are further overheated.

[0056] This superheated refrigerant, which has a superheating temperature T Ke from the recuperator 250 The outlet becomes the refrigerant inlet connection. 211 of the compressor 210 guided.

[0057] The recuperator 250is in the vapor compression circuit 200 used to determine the overall efficiency as the quotient of heat output and output. Q H and absorbed electrical power P e to increase the drive of the compressor motor.

[0058] For this purpose, the refrigerant which is in the condenser is 220 thermal energy Q H releases at a temperature level on the heat sink side to the heating medium, in the high-pressure path of the recuperator 250 Further heat energy is generated through subcooling Q i withdrawn.

[0059] The internal energy state of the refrigerant upon entering the evaporator 240 is caused by this heat extraction Q i reduced, so that the refrigerant releases more heat energy at the same evaporation temperature level Q Q from the heat source 300 can record.

[0060] Subsequently, the refrigerant is released after exiting the evaporator. 242 from the evaporator240 , in the low-pressure path at low pressure ND and at a low-pressure temperature T Va corresponding to an evaporator outlet temperature in the recuperator 250 the heat energy extracted in the high-pressure path Q i The energy is supplied again. The supply of energy advantageously reduces the wet steam content to a state without wet steam, and then further energy supply leads to superheating.

[0061] Furthermore, the following are required to record the operating status of the steam compression system. 200 Advantageously the following sensors are arranged, which are particularly useful for safeguarding and optimizing the operating conditions of the steam compression system. 200 In particular, model-based feedforward control is implemented when operating conditions change.

[0062] On the one hand, the process values ​​recorded by sensors advantageously provide safeguards regarding permissible operating ranges of the components, such as the compressor in particular. 210 On the other hand, model-based feedforwards are made based on the sensor data, in particular for the speed of the compressor. 210 and / or a valve opening degree of the expansion valve, so that the controller only needs to make smaller corrections to compensate for a control deviation that is nevertheless smaller due to the feedforward control: • A high-pressure sensor 503 advantageous for measuring 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 advantageous for measuring 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 liquefier inlet 221 , • an indoor temperature sensor 506 advantageous for measuring the indoor 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 also advantageously referred to as the "recuperator outlet temperature high-pressure path" and • A recuperator internal temperature sensor is advantageous 505 The recuperator internal temperature sensor 505 advantageously captures 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 is advantageously FA from the recuperator internal temperature sensor 505 measured.

[0063] The following sensors are particularly advantageous for carrying out the method according to the invention: • A low-pressure sensor 502 for measuring the low pressure (LP) 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 to measure the evaporator outlet temperature Tva of the refrigerant at the evaporator outlet 242 or between the evaporator outlet 242 and the low-pressure side entry 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 measure 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 .

[0064] The process parameter that has a significant influence on the overall efficiency of the steam compression circuit 200 as the quotient between that of the vapor compression circuit 200 transferred heating power Q H to one from the compressor 210 absorbed electrical power P e The problem is the overheating of the refrigerant at the compressor inlet. 211 To maintain permissible compressor operating conditions, it is advantageous to adhere to restrictions regarding the allowed superheat range of the refrigerant at the compressor inlet. Excessively low superheats particularly compromise the lubricating properties of the machine oil, while excessive superheats result in excessively high hot gas temperatures.

[0065] Superheating 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.

[0066] According to the invention, the compressor inlet superheat is preferably controlled in such a way that no condensate forms on components of the refrigeration circuit, particularly in the section between the refrigerant outlet of the recuperator, due to the water vapor content contained in the ambient air falling below the dew point. 252 and compressor inlet 211 fails. The refrigeration circuit section between evaporator outlet 242 and recuperator inlet 251 Although it is usually colder because it is typically only a short pipe section, it offers better insulation compared to the section between the refrigerant outlet of the recuperator. 252 and compressor inlet 211possible. For example, it is located at the compressor inlet. 211 The refrigerant separator, which needs protection, is located at the compressor. It's difficult to enclose it completely, so the temperature needs to be kept high enough to prevent condensation. Condensation problems don't typically occur on the high-pressure side. This also applies to the passage between the high-pressure side recuperator outlet. 252 and entry into the expansion valve 231 Cools regularly depending on the operating point under ideal heat transfer conditions in the recuperator 250 to the temperature level of the refrigerant at the evaporator outlet 242However, since this passage is also typically short and can be very well insulated, this section is also generally not problematic. It should be noted, however, that the method according to the invention can fundamentally prevent condensate from dropping throughout the entire heat pump circuit.

[0067] 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 at approximately -10°C, at the brine outlet 310 The temperature is approximately -13°C and 5°C at the compressor inlet, which affects overheating. 15K .

[0068] Many systems benefit from room temperature and humidity sensors, which allow for precise determination of the condensation conditions in the air. For example, at 21°C and 60% relative humidity, the condensation temperature is around 13°C. Under these conditions, no condensation occurs as long as the pipe temperature is above 13°C, plus a buffer if necessary (e.g., 1K).

[0069] Sticking with this non-exhaustive numerical example, we now assume that a superheat of 15K is achieved at a compressor inlet temperature of 5°C. This temperature is below the 13°C that is determined to be the condensation temperature of the water vapor present in the ambient air under the current ambient conditions. Therefore, condensation occurs. If the compressor inlet temperature is to be at least 14°C, i.e., the condensation temperature plus a buffer, the superheat must be increased by 9K, i.e., a superheat of 24K must be maintained.

[0070] Limit values, especially for overheating, define the permissible overheating range of the components at the compressor inlet, depending on the operating point. 211 fixed. However, dependencies also exist between the compressor inlet superheat dT. ÜE and the overall efficiency of the vapor compression circuit 200or also between compressor inlet superheat dT ÜE and a stability S of a control value R is advantageous in the control of compressor inlet superheating.

[0071] To take all these requirements into account, the following are advantageous depending on the operating point of the steam compression circuit. 200 , the heat source medium temperature, the heating medium temperature, the compressor power P e and target values ​​Z or the target value Z for a calculation of the compressor inlet superheat dT ÜE used. Alternatively or additionally, the refrigeration circuit parameters dependent on the operating point, such as heat source medium temperature, heating medium temperature, compressor power P, can be used. e and parameterizable coefficients, i.e. adapted to the behavior of the respective refrigeration circuit components, a calculation of the target value Z as a setpoint for the compressor inlet superheat dT ÜEto be carried out. In the simplest case, the target value for the compressor inlet superheat dT is ÜE It is constant regardless of all operating conditions, e.g., 10 Kelvin. With a more complex adjustment, it is determined as a function of an operating point parameter, e.g., the compressor power P. e It varies, or with even more complex adjustments, it varies as a function of several operating point parameters.

[0072] A deviation in the compressor inlet superheat dT is detected. ÜE and a deviation in the control of the evaporator outlet superheat dT ÜA combined and weighted together, resulting in the controller 500 A total control deviation is calculated, which is used to control the steam compression circuit. 200 is fed in. The control deviations from the compressor inlet superheat dT become advantageously more precise. ÜE and evaporator outlet superheat dT ÜAformed by calculating the differences between the respective measured values ​​and target values. • Control deviation of the compressor inlet superheat dT ÜE = Measured value compressor inlet superheat - Target value compressor inlet superheat Z TÜE • Control deviation of the evaporator outlet superheat dT ÜA = Measured value evaporator outlet superheat - Target value evaporator outlet superheat Z TÜA

[0073] Then, the weighted influence of the compressor inlet superheat deviation dT is advantageously applied. ÜE and the weighted influence of the control deviation of the evaporator outlet superheat dT ÜA in the controller 500 the total control deviation is calculated, which is used to control the steam compression circuit 200 is fed in.

[0074] In the vapor compression circuit 200 After expansion, the refrigerant passes through the expansion valve.230 two sequentially arranged heat exchangers, the evaporator 240 and the recuperator 250 in which the refrigerant heat energy Q Q and Q i is supplied.

[0075] In the evaporator 250 Source heat energy is applied to the refrigerant Q Q from the heat source system 300 supplied. The temperature level of the supplied source heat. Q Q is at a temperature level of the heat source, in particular such as the ground or the outside air.

[0076] In the high-pressure flow direction of the refrigerant S HD subsequent recuperator 250 Heat energy is transferred to the refrigerant Q i after leaving the condenser 220 extracted. The temperature level of the refrigerant at the condenser outlet is approximately equal to the return temperature of the heating medium.

[0077] This evaporator circuit 240 with the recuperator250 The series connection has a decisive influence on the transfer function of the control loop for the control of compressor inlet superheat dT. ÜE .

[0078] The control value R is advantageously the weighted relationship of the control deviation of the compressor inlet superheat dT. ÜE with the control deviation of the evaporator outlet superheat.

[0079] Actuator operating condition variables with an influence on the control value R, in particular the compressor inlet superheat dT ÜE , are in the relevant steam compression circuit 200 the compressor speed and / or the opening degree of the expansion valve 230 , which also advantageously determines the low pressure LP and the evaporation temperature level.

[0080] Actuators have a particularly advantageous influence on the control value R, especially on the weighted relationship between the control deviation of the compressor inlet superheat and the control deviation of the evaporator outlet superheat. In the relevant vapor compression circuit 200 are in particular the compressor 210 by varying the compressor speed and the expansion valve 230 These actuators influence the degree of opening. These two actuators affect the low pressure (LP) and the evaporation temperature level.

[0081] Not all influences are desirable. For example, changing the compressor speed to regulate the desired heating output without further compensatory changes to the opening degree of the expansion valve changes the control value R into undesirable ranges, so that a model-based supported change in the opening degree of the expansion valve, accompanying the change in compressor speed, is advantageous, and possibly even necessary, to regulate R.

[0082] Advantageous in the vapor compression circuit 200 the compressor speed is set so that the steam compression circuit 200 heating power transferred to the heating medium QH The required target value Z corresponds to this. To comply with this requirement, the compressor speed must be adjusted to control the compressor inlet superheat dT. ÜE advantageously subordinate or not appropriate.

[0083] The degree of opening of the expansion valve is advantageous. 230 as a control parameter for the regulation of the compressor inlet superheat dT ÜE used. The influence of the opening degree of the expansion valve 230 on the compressor inlet superheat dT ÜE It takes place as follows: The expansion valve 230 acts as a nozzle with an electrically adjustable nozzle cross-section, in which a needle-shaped nozzle needle is typically driven into a nozzle seat via a thread using a stepper motor.

[0084] The refrigerant flow rate through the expansion valve is at the expansion valve inlet when operating with liquid refrigerant. 231 approximately proportional to the square root of the pressure difference between the expansion valve inlet 231and outlet 232 multiplied by a current relative value of the nozzle cross-section or opening degree and advantageously one of the refrigerant and a geometry of the expansion valve 230 dependent constant.

[0085] 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 at the entry into the expansion valve is also 230 The degree of opening of the expansion valve, which can be assumed to be constant, is influenced by this. 230 The only relevant factor is the low pressure (LP), i.e., the outlet pressure from the expansion valve. 230 .

[0086] Will the opening degree of the expansion valve be adjusted? 230 This reduces the amount of refrigerant that passes through the expansion valve at constant high pressure (HP) and initially constant low pressure (LP). 230 Since the compressor 210but continues to initially convey the same refrigerant mass flow, in the high-pressure flow direction S HD through the expansion valve 230 Less refrigerant was supplied than by the compressor. 210 is sucked away.

[0087] Since refrigerant vapor is a compressible medium, the low pressure (LP) on the low-pressure side of the vapor compression circuit then decreases. 200 As the low pressure (LP) decreases, the mass flow rate of refrigerant through the compressor decreases approximately proportionally. 210 , 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 LP is established, at which the expansion valve 230 The supplied refrigerant mass flow is equal to that from the compressor. 210 discharged refrigerant mass flow is.

[0088] Will the opening degree of the expansion valve be adjusted?230 By increasing the pressure, more refrigerant passes through the expansion valve at constant high pressure (HP) and initially constant low pressure (LP). 230 Since the compressor 210 However, if the same refrigerant mass flow continues to be delivered initially, the low-pressure side (LP) of the refrigeration circuit is bypassed by the expansion valve. 230 More refrigerant was supplied than by the compressor. 210 is extracted. Since refrigerant vapor is a compressible medium, the low pressure (LP) on the low-pressure side of the vapor compression circuit increases. 200 As the low pressure (LP) increases, the mass flow rate of the compressor also increases. 210 approximately proportional, since its delivery rate can be described approximately as volume / time, and a correspondingly increased low pressure (LP) is established, at which the expansion valve 230 The supplied refrigerant mass flow is equal to that from the compressor. 210Discharged refrigerant mass flow is.

[0089] The low-pressure side (LP) in turn significantly influences the heat transfer between the heat source medium and the refrigerant in the evaporator. 240 The heat flow Q Q from the heat source system 300 is transferred between the heat source medium and the refrigerant at different temperatures, whereby the heat flow Q Q 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 is.

[0090] The heat transfer resistance between the heat source media path of the evaporator and the refrigerant path of the evaporator is in a respective vapor compression circuit. 200 It can be assumed to be approximately constant. Therefore, the magnitude of the heat transfer power in the evaporator is 240significantly dependent on the integral of the temperature differences of all surface elements of the heat transfer layer.

[0091] To ensure a sufficient level of heat energy Q Q from the heat source system 300 In order to transfer the heat to the refrigerant, it must be ensured that the temperature of the heat source medium is maintained in as many surface elements as possible of the heat exchanger's transfer layer, in this case the evaporator. 240 , is greater than the temperature of the refrigerant at the respective surface element.

[0092] Is the state of matter of the refrigerant when it flows through the evaporator 240When the vapor is saturated, a refrigerant temperature is established which, according to the saturation vapor characteristic (a material property of the refrigerant), is a function of the refrigerant's low pressure (LP). Therefore, by controlling the low pressure (LP) or the evaporation pressure, the evaporation temperature of the refrigerant as it flows through the recuperator can be indirectly controlled. 250 steer.

[0093] The heat energy Q Q , which from the heat source system to the evaporator 240 The transfer of the refrigerant flowing through the system causes a change in the state of matter of the refrigerant.

[0094] The wet vapor fraction in saturated refrigerant vapor decreases at constant low pressure during heat transfer to the refrigerant. With incomplete evaporation, the wet vapor fraction, and thus also the internal energy state of the refrigerant upon exiting the heat exchanger, is a function of: • Wet steam content at the entrance to the evaporator 240 , • Refrigerant mass flow, • Transferred heat output Q Q , and from a • Enthalpy difference in the wet vapor region at the respective low pressure LP, which the refrigerant exhibits as an associated function of the pressure.

[0095] For complete evaporation, an additional energy supply is provided in the recuperator. 250 , in order to superheat the refrigerant beyond the state of saturated vapor.

[0096] The method is used under given operating conditions of the steam compression circuit. 200 Depending on the control variable “opening degree of expansion valve 230”, a corresponding refrigerant state is achieved upon exit from the evaporator. 240 set.

[0097] In steady state, the control loop behavior of the “isolated” control loop “evaporator 240” is moderately steep.

[0098] The control loop behavior is characterized in particular by the control loop output value of the evaporator outlet superheat as a function of the control loop input value of the expansion valve opening degree.

[0099] Advantageously, a refrigerant, in particular a refrigerant mixture, is used which exhibits a "temperature glide," especially R454C. Advantageously, with a refrigerant mixture exhibiting a temperature glide, a relative change in the opening degree of the expansion valve actuator of 1% rel. typically results in a superheat change of approximately less than 1 K at the refrigerant outlet from the evaporator.

[0100] This state can advantageously also be achieved by influencing at least one or more of the following various time constants through control engineering; these ultimately determine the process variable refrigerant superheat at the evaporator outlet. 242 influence: • A first time constant advantageously delays the mechanical change in the opening degree of the expansion valve. 230 by limiting the travel speed through the controller 500 The control value R is reduced in this first time constant Z in the travel speed by a braking value. The braking value can, for example, be the controller cycle time in which a travel step of the expansion valve takes place. 230 is controlled, include. • A second time constant acts through the controller 500Advantageously specified for a delayed adjustment of a corresponding low pressure when the opening degree of the expansion valve changes. 230 due to the compressibility of the refrigerant vapor at low pressure (LP) 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 is advantageously obtained from delayed changes in the state of matter of the refrigerant during changes in evaporation temperature. • A fifth time constant is advantageously obtained from the transport of the refrigerant through the evaporator. 240 with a finite flow velocity.

[0101] Therefore, after changing the control variable "opening degree of the expansion valve 230", it is advantageous to delay the corresponding refrigerant state change at the exit from the evaporator outlet. 242 The total time constant Zges is advantageously in the range of 30 seconds to about 5 minutes, depending on the operating point.

[0102] After flowing through the evaporator 240 The refrigerant enters the low-pressure path of the recuperator at low pressure (ND). 250 a.

[0103] The state of matter of the refrigerant when flowing into the recuperator 250 In a typical operating scenario, it is advantageous to use either saturated vapor with a low vapor content between 0 and 20% or, in particular, already superheated refrigerant.

[0104] With advantageously saturated vapor, a refrigerant temperature is established which, according to the refrigerant's saturation vapor characteristic, is a function of the refrigerant pressure. Upon the entry of superheated refrigerant, the refrigerant temperature will at most reach a value corresponding to the inlet temperature of the heat source medium. In this case, this value preferably corresponds to the inlet temperature of the refrigerant entering the high-pressure path of the recuperator. 250 , that is, the temperature of the refrigerant after it exits the condenser 220 .

[0105] To transfer a sufficient amount of heat energy from the refrigerant of the high-pressure side refrigerant path to the refrigerant of the low-pressure side refrigerant path in the recuperator 250To enable the transfer of heat, it must be ensured that the temperature of the refrigerant in the high-pressure side refrigerant path reaches high pressure (HD) in as many surface elements of the recuperator's transfer layer as possible. 250 greater than the temperature of the refrigerant in the low-pressure side refrigerant path at low pressure LP at the respective surface element.

[0106] The corresponding temperatures of the heating system 400 of the steam compression system 200 are higher in a heating scenario than the corresponding temperatures of the heat source, such as the ground or the outside air.

[0107] The heat energy Q i , which transfers from the refrigerant at high pressure (HD) of the high-pressure refrigerant path to the refrigerant at low pressure in the low-pressure refrigerant path of the recuperator 250The transfer of this process causes a change in the state of matter of the refrigerant on the low-pressure side. The wet vapor component of the recuperator 250 On the low-pressure side, when the refrigerant flows through the low-pressure side (LP), the temperature decreases during heat transfer to the refrigerant, and after complete evaporation, superheating of the refrigerant advantageously occurs.

[0108] The internal energy state of the refrigerant upon exiting the low-pressure side path of the recuperator is advantageously influenced by one or more of the following factors. It should be noted that the change in energy state is based solely on physical dependencies, with the controller influencing the control of the actuators, which in turn naturally affects physical quantities such as the refrigerant mass flow rate. • Wet steam content upon entry into the recuperator 250 , • Refrigerant mass flow, • transferred heat output Q i , which advantageously allows for control depending on the temperature difference between the temperature of the refrigerant at high pressure (HP) in the high-pressure refrigerant path and the temperature of the refrigerant in the low-pressure refrigerant path at low pressure (LP), and / or • an enthalpy difference in the wet steam region at the respective low pressure (LP).

[0109] This advantageously results in the following: depending on the given operating conditions of the steam compression circuit 200 as well as, depending on the control variable "opening degree of expansion valve 230", a corresponding refrigerant state at the outlet. 252 from the recuperator 250 It sets the low pressure (ND).

[0110] In steady state, the gradient of the "isolated" control loop at the low pressure (LP) of the refrigerant in the low-pressure-side path of the recuperator is determined. 250a control loop behavior with a high steepness, with an approximately constant internal energy state of the refrigerant at the inlet 251 in the low-pressure side LP path of the recuperator 250 A change in the relative opening degree of the expansion valve of 1% results in a change in superheat at the refrigerant outlet from the evaporator. 230 It is advantageous to set the temperature to approximately 10 K or even above 10 K.

[0111] opposite the recuperator 250 This results in a significantly higher heat transfer in the evaporator. 240 between the source medium and the refrigerant in the evaporator 240 .

[0112] This is how it happens in the evaporator. 240 a significantly higher heat transfer than in the recuperator 250 adjusted, as the environment is controlled by means of an evaporator 240 a significantly greater amount of energy is to be extracted than is stored in the recuperator alone. 250to transfer heat within the refrigeration cycle. The driving temperature difference in the recuperator, for example, is between 20 K and 60 K, while in the evaporator it is only between 3 K and 10 K. In order to transfer the desired energy despite the different driving temperature differences, the heat exchanger surface area of ​​the evaporator is designed to be approximately 5 to 20 times larger than that of the recuperator. 250 .

[0113] This state is advantageously set using at least one of the following time constants Z: • With an eleventh time constant Z 11 A delay in the mechanical opening degree change of the expansion valve is advantageous. 230 by limiting the travel speed. • A twelfth time constant Z 12This has a beneficial effect on the delayed adjustment of a corresponding low pressure (LP) when the opening degree of the expansion valve changes. 230 due to the compressibility of the refrigerant vapor in the low-pressure path LP. • One 13 Time constant Z 13 is a thermal time constant of the evaporator's heat transfer layer. Therefore, a change in the evaporation pressure, and thus the evaporation temperature, causes a delayed temperature change in the heat transfer layer, which often contains several kilograms of metal, and in the refrigerant in the evaporator's low-pressure path. 240 . • One 14 Time constant Z 14 is advantageously determined or specified from delayed changes in the state of matter of the refrigerant during changes in evaporation temperature. • One 15 Time constant Z 15 This results advantageously from the transport of the refrigerant through the evaporator.240 with a finite flow velocity and is taken into account.

[0114] The low-pressure side refrigerant path of the recuperator 250 is extracted from the evaporator outlet 242 of the evaporator 240 supplied. The internal energy state of the refrigerant is also determined here by at least two time constants Z, Z. 11 , Z 12 , Z 13 , Z 14 , Z 15 , Total time is delayed after changing the control variable “opening degree of expansion valve”.

[0115] After changing the control variable "opening degree of expansion valve 230", a further delay in the corresponding refrigerant state change occurs due to the time behavior of the recuperator. 250 at the exit from the low-pressure side refrigerant path of the recuperator 250 a.

[0116] The time response of the recuperator 250The total recuperator time constant Zges can be advantageously considered depending on the respective operating point of the steam compression circuit in the range of approximately 1 minute to 30 minutes.

[0117] A weighted combination of compressor inlet superheat dT is advantageously used. ÜE and the evaporator outlet superheat dT ÜA , 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 in the controller 500 for regulating the steam compression circuit 200 is fed in.

[0118] The compressor inlet superheat dT ÜE is advantageously used as the main controlled variable and the corresponding signal flows and signal processing take place in particular in the following process steps: Step 1First, the process variables compressor inlet superheat dT are considered. ÜE advantageous as the main control variable and the evaporator outlet superheat dT ÜA Advantageously, it can be measured as an auxiliary variable in a first process step.

[0119] For this purpose, the evaporation temperature of the refrigerant at the respective measurement point is either • directly measured using a temperature sensor positioned to detect a temperature corresponding to the refrigerant temperature in the wet vapor region or • indirectly determined using a pressure sensor which detects the refrigerant pressure of the refrigerant evaporating in the wet steam region and then calculates the evaporation temperature from the refrigerant-specific dependence between pressure and temperature in the wet steam region.

[0120] Furthermore, the respective overheating measurement point, in particular the evaporator outlet, is used. 242 and / or at the compressor inlet 211 assigned temperatures of the refrigerant temperature using 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 corresponds to the respective superheat of the refrigerant at the measuring point.

[0121] Input variables of the calculation in step 1 Are the compressor inlet superheat dT then ÜE and the evaporator outlet superheat dT ÜA .

[0122] Step 2 The process variables compressor inlet superheat dT ÜE and evaporator outlet superheat dT ÜA In a second step, the associated deviations from the standard are advantageously offset against their respective target values ​​to form the corresponding deviations from the standard: The target value for the compressor inlet superheat dT ÜE at the evaporator outlet 242 The temperature 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.

[0123] The target value for the evaporator outlet superheat dT ÜA Depending on the refrigeration circuit operating mode and the refrigeration circuit operating point, the evaporator superheat is then varied so that, in the steady-state normal case, it is approximately equal to the established process value of the evaporator outlet superheat dT. ÜA This corresponds to the target value for the evaporator outlet superheat dT. ÜA can be model-based depending on an operating mode or operating point, the evaporation temperature, the condensation temperature, the compressor power, and a setpoint for the compressor inlet superheat dT. ÜE at the evaporator outlet242 and / or component properties are pre-calculated and adaptively corrected.

[0124] The deviation of the compressor inlet superheat dT will then be determined. ÜE calculated by taking the process value of the compressor inlet superheat dT ÜE the target value of the compressor inlet superheat dT ÜE is subtracted.

[0125] The deviation from the control parameters of the evaporator outlet superheat dT will then be determined. ÜA calculated by taking from the process value of the evaporator outlet superheat dT ÜA the target value of the evaporator outlet superheat dT ÜA is subtracted.

[0126] Step 3 In a third process step, the deviation of the compressor inlet superheat dT is measured. ÜE and the control deviation of the evaporator outlet superheat dT ÜA advantageously combined to prevent overall control deviation and overheating.

[0127] The combination is achieved in particular by means of a weighted addition of the individual rule deviations.

[0128] The weighting influence is a measure of the proportional combination of the individual rule deviations and, in extreme cases, can result in the exclusive inclusion of only one individual rule deviation, but usually it results in the weighted inclusion of both individual rule deviations.

[0129] Advantageously, the weighting influence is estimated as a value between 0 and 1, i.e., 0 to 100%, and this value is applied to the degree of inclusion of the control deviation of the compressor inlet superheat dT. ÜE included in the total control deviation, resulting in the following dependency for calculating the total control deviation: Total control deviation overheating = (Weighting influence * control deviation compressor inlet superheat) + ((1 - Weighting influence) * Control deviation evaporator outlet superheat)

[0130] The value of the weighting influence can be advantageously influenced by the operating mode and / or the operating point of the heat pump. 100 The following variations may occur depending on the circumstances: • During the changeover between operating modes = operation with the compressor switched off 210 and operating mode = operation with compressor switched on 210 During heating operation, due to the dynamic process value changes when starting up the steam compression system, 200 Advantageously, only the control deviation of the evaporator outlet superheat dT is initially addressed. ÜA included in the overall rule deviation, in particular the value of a weighting influence is initially = 0 or a value advantageously below 20%. • After a stabilization phase of the vapor compression system 200It is advantageous not to switch spontaneously to the weighting influence value designed for normal operation, but rather to design the transition as a ramp. In this case, it is advantageous for the weighting influence value to be increased from the initial value of 0, or a value particularly below 20%, to the intended target value in a ramp-like manner. This avoids, in particular, value discontinuities during a spontaneous switch and thus prevents control oscillations. • The target value of the weighting influence is advantageously adapted to the respective operating mode and operating point. Operating points characterized by increased oscillation tendency advantageously require a lower weighting of the control deviation of the compressor inlet superheat dT. ÜE , in particular, this addresses a control-technically critical signal behavior of the compressor inlet superheat dT ÜEdue to the evaporator outlet superheat dT ÜA Larger signal delays and steeper track gradients prevented oscillations.

[0131] Step 4 In a fourth process step, the calculated total control deviation due to overheating is then entered into the controller. 500 processed, which includes 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, so that in the regulated case the control deviation of the superheat is set to approximately 0 Kelvin.

[0132] A P, I, PI, PID controller can be used, whereby the control components are advantageously dynamically adapted to the respective operating mode and the operating point.

[0133] Fig. Figure 3 shows the log p / h diagram of the Fig. 2, where instead of the in Fig. The two process variables shown describe the underlying processes. The process variables themselves are not shown for clarity.

[0134] In one step 810 The refrigerant is in the compressor 210 The air is compressed from low pressure (LP) to high pressure (HP), with the temperature also differing from the evaporator outlet temperature (T). KE to the hot gas temperature T HG increased.

[0135] In the liquefier 220 it occurs when the heating output is released Q H first, a deheating step 820 , before the refrigerant is then added in step 830 It is liquefied. The liquefaction takes place in the liquefier. 220 preferably completely finished.

[0136] During subsequent passage through the recuperator 250 The liquid refrigerant is added in step 850 chilled before it goes into step 860in the expansion valve 230 The low pressure (ND) is relaxed.

[0137] In step 870 The liquid and expanded refrigerant is in the evaporator 240 almost completely evaporates before entering the recuperator 250 finally in step 890 overheating.

[0138] The overheated refrigerant is then returned to step 810 in the compressor 210 compacted, so that the cycle can be repeated.

[0139] In contrast, a log p / h diagram of an operating point with refrigerant shortage is used in Fig. 4 shown, with the same process steps as in Fig. 3 are designated with the same reference numerals.

[0140] The first difference occurs in step 830 The liquefaction does not occur in the liquefier. 220 The refrigerant, still in its gaseous state, enters the recuperator. 250one and will be done in one step 840 in the recuperator 250 liquefies. Instead of releasing the heat energy to the heat source system, the energy is transferred internally from the high-pressure to the low-pressure path, which reduces the coefficient of performance.

[0141] Accordingly, in step 870 the evaporation not in the evaporator 240 completed. A significant portion of the evaporation takes place in one step. 880 only in the recuperator 250 instead of.

[0142] The larger, solid arrow is clearly visible, representing the amount of heat transfer in the recuperator. 250 The further the arrow lies within the area defined by the boiling point and dew point lines, the less efficiently the refrigeration cycle operates. According to the invention, this condition is detected and, if possible, rectified. QUOTES INCLUDED IN THE DESCRIPTION

[0000] This list of documents cited by the applicant was automatically generated and is included solely for the reader's convenience. The list is not part of the German patent or utility model application. The DPMA accepts no liability for any errors or omissions. Cited patent literature

[0000] DE 10159892 A1

[0005] DE 102005061480 B3

[0006]

Claims

[1] Method for operating a compression refrigeration system (200) with - a refrigerant that exhibits a temperature glide, - an evaporator (240), - a compressor (210), - a liquefier (220), - a thoracic organ (230), - an internal heat exchanger (250), wherein the internal heat exchanger is designed to transfer heat energy from the refrigerant in the high-pressure path (HP) before entering the throttling device (230) to the refrigerant in the low-pressure path (LP) before entering the compressor (210), and - a control unit (500), wherein the control unit is designed a) to detect superheating of the refrigerant after exiting the evaporator (240) and to detect superheating of the refrigerant upon entry into the compressor (210) and b) for controlling the throttling device (230) using a control variable based on both superheats and c) for the detection of a refrigerant shortage at least in the condenser (220), the method comprising the following steps: - Calculating a temperature difference between the dew point temperature and the boiling point temperature, called the temperature glide, of the refrigerant at the current operating point of the compression refrigeration system based on a measured low pressure (LP) of the refrigerant, - Calculating a temperature difference limit value relative to the calculated temperature glide of the measured low pressure (LP) of the refrigerant at the current operating point, - Assessing whether the current evaporator outlet superheat is below the temperature difference limit value, - Providing a setpoint for superheat, defined as the difference between the refrigerant temperature and the dew point temperature, at the compressor inlet (210), called compressor inlet superheat, - Evaluating the current compressor inlet superheat to determine if it exceeds the target value for compressor inlet superheat, - Detecting a refrigerant shortage situation when both results of the assessment steps are continuously met for a specified period of time. [2] The method of claim 1, wherein the method further comprises the following step: - Conducting a test of refrigerant recirculation into the condenser (220) by temporarily increasing the setpoint of the compressor speed and / or temporarily increasing the setpoint of the evaporator outlet superheat. [3] The method of claim 2, wherein the method further comprises the following step: - Generation of an error state upon unsuccessful attempt at refrigerant recovery, particularly upon multiple unsuccessful attempts at refrigerant recovery within a defined time period. [4] Method according to one of the preceding claims, wherein the calculation of the temperature difference limit value relative to the calculated temperature glide of the detected low pressure of the refrigerant is dependent on the current operating point of the compression refrigeration system. [5] Method according to claim 4, wherein the temperature difference limit value is set relative to the calculated temperature glide proportional to a difference between heat sink temperature and heat source temperature. [6] Compression refrigeration system (200) with - a refrigerant that exhibits a temperature glide, - an evaporator (240), - a compressor (210), - a liquefier (220), - a thoracic organ (230), - an internal heat exchanger (250), wherein the internal heat exchanger is designed to transfer heat energy from the refrigerant in the high-pressure path (HP) before entering the throttling device (230) to the refrigerant in the low-pressure path (LP) before entering the compressor (210), and - a control unit (500), wherein the control unit is designed a) to detect superheating of the refrigerant after exiting the evaporator (240) and to detect superheating of the refrigerant upon entry into the compressor (210) and b) for controlling the throttling device (230) using a control variable based on both superheats and c) for detecting a refrigerant shortage at least in the condenser (220), wherein the control unit is further configured to: - Calculating a temperature difference between the dew point temperature and the boiling point temperature, called the temperature glide, of the refrigerant at the current operating point of the compression refrigeration system based on a measured low pressure (LP) of the refrigerant, - Calculating a temperature difference limit value relative to the calculated temperature glide of the measured low pressure (LP) of the refrigerant at the current operating point, - Assessing whether the current evaporator outlet superheat is below the temperature difference limit value, - Providing a setpoint for superheat, defined as the difference between the refrigerant temperature and the dew point temperature, at the compressor inlet (210), called compressor inlet superheat, - Evaluating the current compressor inlet superheat to determine if it exceeds the target value for compressor inlet superheat, - Detecting a refrigerant shortage situation when both results of the assessment steps are continuously met for a specified period of time. [7] Heat pump with a compression refrigeration system according to claim 6.

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