Method for operating a compression refrigeration system and compression refrigeration system

Abstract

The present invention relates to a method for controlling a compression refrigeration system (200) and an associated compression refrigeration system (200).The procedure comprises determining, for the current compressor speed, a maximum permissible high pressure (HPmax), a minimum permissible high pressure (HPmin), a maximum permissible low pressure (LPmax), and a minimum permissible low pressure (LPmin); determining the distances of the current high pressure to the minimum and maximum permissible high pressure, as well as the distances of the current low pressure to the maximum and minimum permissible low pressure; determining the current speed range class to obtain the permissible counteracting actions; determining, for each permissible counteracting action, a change factor of the actuator belonging to the counteracting action as a function of at least one of the determined distances; and controlling the compression refrigeration system (200) based on the determined change factors of the actuators.

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 expanded 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] In this context, the compressor deserves particular mention, as it compresses the gaseous refrigerant from low pressure (LP) to high pressure (HP). It is known to use compressors capable of operating at different speeds to optimize performance. Manufacturers of such variable-speed compressors specify, among other operating limits, that both upper and lower limits for low pressure (LP) and high pressure (HP) must be observed, depending on the speed. If these limits are exceeded on either the low-pressure or high-pressure side, the compressor must, in the worst case, be shut down.

[0009] Against this background, it is an object of the present invention to propose a control method for the aforementioned compression refrigeration system, with which an exceedance of the operating limits of the components, in particular the compressor, can be reliably prevented and the resulting losses in performance or coefficient of performance can be limited to an absolutely necessary level.

[0010] The problem is solved by the method features of claim 1 and the device features of claim 9.

[0011] Accordingly, a method for controlling a compression refrigeration system comprising a refrigerant, an evaporator, a compressor, a condenser, an throttling device, and a control unit is proposed, comprising: a) detecting the high and low pressure of the refrigerant in the compression refrigeration system; b) detecting the speed of the compressor; c) providing speed range classes of the compressor, each speed range class defining, for a speed range of the compressor, which counteracting actions are permissible from a list of possible actions to maintain the permissible operating range of components of the compression refrigeration system, in particular the compressor; and d) controlling at least one actuator of the compression refrigeration system, in particular at least the throttling device and / or the compressor, based on the detected high pressure, the detected low pressure, and the detected superheat.

[0012] Each countermeasure preferably relates to an assigned process value range limit and an assigned actuator.

[0013] The counteracting action preferably influences the control of the assigned actuator when approaching the assigned process value range limit.

[0014] The process value range limits preferably refer to low pressure, high pressure, temperature and / or superheating.

[0015] The process includes the following steps: - Determine, for the current speed of the compressor, a maximum permissible high pressure, a minimum permissible high pressure, a maximum permissible low pressure and a minimum permissible low pressure; - Determining the distances between the current high pressure and the minimum and maximum permissible high pressure, as well as the distances between the current low pressure and the maximum and minimum permissible low pressure; - Determining the current speed range class to obtain the permissible counter-control actions; - Determine, for each permissible counteracting action, a change factor of the actor belonging to the counteracting action depending on at least one of the determined distances; - Rules of the compression refrigeration system based on the specific change factors of the actuators.

[0016] Preferably, the distances are determined as differences. Preferably, the change factors of the actuators for controlling the compression refrigeration system are combined with the actuator control signals that are being triggered.

[0017] The control system according to the invention therefore selects the permissible or appropriate actions depending on the current speed range of the compressor. Depending on the distance of the operating point from the compressor's operating limits, change factors for the individual actions are then determined. This ensures that the compression refrigeration system is operated safely within its operating limits.

[0018] Preferably, the maximum permissible high pressure, the minimum permissible high pressure, the maximum permissible low pressure, and the minimum permissible low pressure are determined using a compressor speed characteristic map, in particular by directly reading or interpolating tabulated values. For this purpose, tables or functions based on values ​​provided by compressor manufacturers may be used, for example.

[0019] Particularly preferred is the speed characteristic map dependent on the refrigerant used. According to the invention, it has been found that particularly good results of the method are achieved with sensitive refrigeration circuits. In the control engineering sense, sensitive refrigeration circuits are, for example, a result of using refrigerants that exhibit a temperature glide, such as R454C, which is particularly preferred.

[0020] Preferably, the method includes a step of combining the change factors of the same actuator. Thus, if several actions are related to the same actuator, a combined change factor is obtained from the individual change factors of these actions, which can then be easily transferred to or processed by the control system of the compression refrigeration system.

[0021] Preferably, the combining step includes selecting the change factor of the respective actuator that produces the greatest change. This ensures that the change is implemented to a sufficient degree even when multiple actions affect the same actuator are performed. Preferably, in this case, all actions affecting the same actuator are of the same direction, i.e., either increasing or decreasing.

[0022] Preferably, because the factors are preferably applied multiplicatively, a combination of similar and opposite actions is not processed as a sum, but for example as follows: first, the maximum is formed from all increasing factors (factors greater than 1) and the minimum from all reducing factors (factors less than 1) is formed, and these partial results are multiplied.

[0023] Preferably, the control unit is designed to control the throttling device, the compressor, a heat source-side fan or a brine pump, and a heat sink-side actuator such as a circulation pump.

[0024] Preferably, the step of controlling the compression refrigeration system based on the change factors includes multiplying the actuator control variables by the change factors. This simplifies the determination of the change factors, as they can be calculated as relative changes independent of the actual actuator control variable.

[0025] Preferably, the list of possible permissible actions for several of the actuators controllable by the control unit includes increasing and decreasing the respective actuation level of the actuator.

[0026] Particularly preferably, the list includes at least one, preferably several, and in particular all of the following possible permissible actions: a) when approaching the minimum permissible low pressure: compressor speed increase, compressor speed reduction, fan speed increase, opening degree increase; b) when approaching the maximum permissible low pressure: compressor speed increase, compressor speed reduction, opening degree reduction; c) when approaching the maximum permissible high pressure: compressor speed increase, compressor speed reduction, fan speed increase, in particular in the case of heat pumps with an internal heat exchanger, pump speed reduction, pump speed increase; d) when approaching the minimum permissible high pressure: pump speed reduction, in particular in the case of heat pumps with an internal heat exchanger, pump speed increase.

[0027] The permissible actions therefore concern an actuator, e.g. compressor or throttling device, a direction of change, i.e. increase or reduction, and a cause, e.g. approaching minimum or maximum permissible values ​​for high or low pressure.

[0028] In addition, further actions based on other parameters may be provided, such as when approaching a maximum of the allowed compressor torque: compressor speed increase, compressor speed reduction or opening degree reduction, or a compressor speed reduction when approaching a temperature maximum of the inverter.

[0029] Depending on the speed range class in which the current speed of the compressor is located, some of these actions are now enabled, i.e., considered and calculated for the control, while other actions are not enabled and are therefore disregarded in the control.

[0030] For example, if an increase in compressor speed is enabled in one speed range class to approach the maximum permissible low pressure, a reduction in compressor speed may be considered when operating in a different speed range class while still approaching the maximum permissible low pressure in the same way.

[0031] Preferably, the change factor is determined as a function, in particular as a first-order linear function, of at least one process value with respect to its distance up to a distance threshold.

[0032] Preferably, the change factor is limited to a value range between 0.5 and 1 for reductions or between 1 and 2 for increases. This allows for sufficient corrections while avoiding excessively strong control interventions and thus instability of the system.

[0033] Preferably, the control step takes into account further operating conditions, in particular a maximum pressure ratio between high and low pressure and / or a minimum pressure differential between high and low pressure. Such conditions may, for example, also be specified by the compressor manufacturer for the entire operating range or for specific speed ranges.

[0034] According to a further aspect, the invention relates to a compression refrigeration system comprising a refrigerant, an evaporator, a compressor, a condenser, an expansion valve, and a control unit. The control unit is configured a) to detect the high and low pressures of the refrigerant in the compression refrigeration system, b) to detect the speed of the compressor, c) to provide speed range classes for the compressor, each speed range class defining, for a given speed range of the compressor, which counteracting actions from a list of possible actions are permissible; and d) to control at least one actuator of the compression refrigeration system, in particular at least one expansion valve and / or the compressor, based on the detected high pressure, the detected low pressure, and the detected superheat. The control unit is further configured to: - Determine, for the current speed of the compressor, a maximum permissible high pressure, a minimum permissible high pressure, a maximum permissible low pressure and a minimum permissible low pressure; - Determining the distances between the current high pressure and the minimum and maximum permissible high pressure, as well as the distances between the current low pressure and the maximum and minimum permissible low pressure; - Determining the current speed range class to obtain the permissible counter-control actions; - Determine, for each permissible counteracting action, a change factor of the actor belonging to the counteracting action depending on at least one of the determined distances; - Rules of the compression refrigeration system based on the specific change factors of the actuators.

[0035] The two output signals for the controller, namely the control signals of the actuators and the additional change factors, are combined in a suitable way so that the control according to the invention is also based on the specific change factors.

[0036] Preferably, the refrigerant has a temperature glide, wherein the refrigerant is particularly R454C or consists thereof, and wherein the compression refrigeration system particularly 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 used in refrigeration circuits with R454C.

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

[0038] 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.

[0039] The figures show an example of implementation: Fig. 1 heat pump 100 with a vapor compression cycle 200 Fig. 2 log p / h - Diagram of the vapor compression process with recuperator 250 Fig. Figure 3 shows a compressor characteristic curve schematically and as an example. Fig. Figure 4 shows a schematic and exemplary flowchart of a process. Fig. Figure 5 shows a progression of the reduction, for example, of the compressor speed as a function of the low pressure value. Fig. Figure 6 shows a schematic and exemplary flowchart of a procedure for controlling a compression refrigeration system.

[0040] 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 , which contains the following components: • A compressor 210 for compressing the superheated refrigerant, • a liquefier 220 , with a refrigerant-side condenser inlet 221 and 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 system 300 in particular a brine system which can store thermal energy Q Q absorbs heat energy Q from the ground or an air system which extracts 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 a direction S HD and SND through the vapor compression circuit 200 flows, whereby in the vapor compression cycle 200 Refrigerant vapor through the compressor 210 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 210A low-pressure path with a low-pressure flow direction (SND) of the refrigerant is formed, in which the evaporator 240 lies.

[0041] 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) measures the temperature at the evaporator inlet. 241 determine.

[0042] 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.

[0043] The compressor 210 It is used to compress the superheated refrigerant from an inlet port. 211 to a compressor outlet pressure P Va at a compressor outlet temperature corresponding to the hot gas temperature at the compressor outlet 212 The compressor 210 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 212is 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 ND, 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.

[0044] In the liquefier 220 The transfer of heat energy Q takes place H from the refrigerant of the vapor compression system 200 to a heating medium of the heat sink system 400 . 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.

[0045] After the refrigerant vapor has been deheated, this advantageously takes place in the condenser. 220 another heat transfer Q H 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.

[0046] 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.

[0047] 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 Q Htransferred from the refrigerant to the heating medium.

[0048] 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.

[0049] 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 HDThe flow is transferred to the refrigerant flowing under low pressure (LP), which flows from the evaporator to the compressor in a low-pressure flow direction (SND). This transfer occurs from the condenser to the expansion valve. 230 The flowing refrigerant is advantageously subcooled.

[0050] 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. 210 This 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.

[0051] 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. WQ They lie so that the temperature difference drives heat transfer.

[0052] 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 utilizing energy 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.

[0053] The one in the evaporator 240 The incoming refrigerant is reduced as it flows through the evaporator. 240 by 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. 330 In 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 Q is transferred to the heat source medium as it flows through the evaporator. Q is withdrawn.

[0054] In the recuperator 250 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 flowing refrigerants are transferred, with the one from the evaporator 240 to the compressor 210 The flowing refrigerants, in particular, are further overheated.

[0055] 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.

[0056] The recuperator 250 is in the vapor compression circuit 200 used to determine the overall efficiency as the quotient of heat output Q H and absorbed electrical power P e to increase the drive of the compressor motor.

[0057] For this purpose, the refrigerant which is in the condenser is 220 Heat 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 Additional heat energy Q through subcooling i withdrawn.

[0058] 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 Q at the same evaporation temperature level Q from the heat source 300 can record.

[0059] Subsequently, the refrigerant is released after exiting the evaporator. 242 from the evaporator 240 , in the low-pressure path at low pressure ND and at a low-pressure temperature corresponding to an evaporator outlet temperature T Va in the recuperator 250 the heat energy Q extracted in the high-pressure path iThe energy is then 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.

[0060] 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.

[0061] 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. 210and / 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 le of the refrigerant between the high-pressure side internal recuperator outlet 252 of the refrigerant from the recuperator250 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.

[0062] 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 for measuring 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 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 .

[0063] The process parameter that has a significant influence on the overall efficiency of the steam compression circuit 200as 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.

[0064] 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.

[0065] According to the invention, the compressor inlet superheat is preferably controlled in such a way that no condensate forms due to the dew point being undershot in the surrounding environment. 5 Water vapor content of the air used in the refrigeration circuit, particularly in the section between the refrigerant outlet of the recuperator, affects components of the refrigeration circuit. 252 and compressor inlet 211 fails. The refrigeration circuit section between the evaporator outlet 242 and recuperator entry 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 211 possible. For example, it is located at the compressor inlet. 211The refrigerant separator, located at the compressor, needs protection. It's difficult to enclose it completely, so the temperature must be kept high enough to prevent condensation. Condensation is not a problem on the high-pressure side under normal circumstances. 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 242 However, 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.

[0066] 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 .

[0067] 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 of 1K if necessary.

[0068] Sticking with this non-exhaustive numerical example, we now assume that a superheat of 15 K 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 the buffer, the superheat must be increased by 9 K, i.e., a superheat of 24 K must be maintained.

[0069] 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 steam 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.

[0070] 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.

[0071] A deviation in the compressor inlet superheat dT is detected. ÜE and a deviation in the control of the evaporator outlet superheat dT ÜA They are 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

[0072] 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.

[0073] 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.

[0074] In the evaporator 250 Source heat energy Q is added to the refrigerant 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.

[0075] In the refrigerant high-pressure flow direction S HD subsequent recuperator 250 Heat energy Q is transferred to the refrigerant 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.

[0076] This evaporator circuit 240with the recuperator 250 The series connection has a decisive influence on the transfer function of the control loop for the control of compressor inlet superheat dT. ÜE .

[0077] 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.

[0078] 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.

[0079] 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.

[0080] 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 alters 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 may even be necessary, to regulate R.

[0081] Advantageous in the vapor compression circuit 200 the compressor speed is set so that the steam compression circuit 200 The heating power QH transferred to the heating medium corresponds to the requested target value Z. To comply with this requirement, the compressor speed is adjusted to control the compressor inlet superheat dT. ÜE advantageously subordinate or not appropriate.

[0082] The degree of opening of the expansion valve is advantageous. 230 as a control parameter for 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.

[0083] 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 a refrigerant and expansion valve geometry 230 dependent constant.

[0084] 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 .

[0085] 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.

[0086] 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.

[0087] Will the opening degree of the expansion valve be adjusted?230 As the pressure increases, 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.

[0088] 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 Heat 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.

[0089] 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.

[0090] To obtain a sufficient amount 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 of the heat exchanger's transfer layer as possible, in this case the evaporator. 240 , is greater than the temperature of the refrigerant at the respective surface element.

[0091] 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.

[0092] 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.

[0093] 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 a material constant as a function of pressure.

[0094] 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.

[0095] 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.

[0096] In steady state, the gradient of the "isolated" control loop "evaporator" is determined with regard to the control loop gradient. 240“A control loop behavior with moderate steepness. 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.”

[0097] 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.

[0098] 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.

[0099] Therefore, after changing the control variable "opening degree of the expansion valve 230", it is advantageous to delay the corresponding change in the refrigerant state 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.

[0100] After flowing through the evaporator 240 The refrigerant enters the low-pressure path of the recuperator at low pressure (ND). 250 a.

[0101] 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.

[0102] 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 .

[0103] 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.

[0104] 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.

[0105] 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.

[0106] 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 power 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).

[0107] 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).

[0108] 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.250 a 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.

[0109] 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 .

[0110] 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 .

[0111] 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.

[0112] 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”.

[0113] After changing the control variable "opening degree of expansion valve" 230 “This results in a further delay in the corresponding change in the state of the refrigerant due to the time behavior of the recuperator.” 250 at the exit from the low-pressure side refrigerant path of the recuperator 250 a.

[0114] 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.

[0115] 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.

[0116] 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 1

[0117] First, the process variables compressor inlet superheat dT are determined. 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.

[0118] 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.

[0119] 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.

[0120] Input variables of the calculation in step 1 are then the compressor inlet superheat dT ÜE and the evaporator outlet superheat dT ÜA . Step 2

[0121] The process variables compressor inlet superheat dT ÜE and evaporator outlet superheat dT ÜA In a second step, these deviations are advantageously offset against each other to form the corresponding target values: The target value for the compressor inlet superheat dT ÜE 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.

[0122] The target value for the evaporator outlet superheat dT ÜA at the evaporator outlet 242 The temperature is then varied depending on the refrigeration circuit operating mode and the refrigeration circuit operating point so that the evaporator superheat in the steady-state normal case 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.

[0123] 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.

[0124] 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. Step 3

[0125] 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.

[0126] The combination is achieved in particular by means of a weighted addition of the individual rule deviations.

[0127] 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.

[0128] 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 Uoverheating = (weighting influence * control deviation compressor inlet Uoverheating) + ((1 - weighting influence) * control deviation evaporator outlet Uoverheating)

[0129] 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 dT compared to the evaporator outlet superheat ÜA Larger signal delays and steeper track gradients prevented oscillations. Step 4:

[0130] In a fourth process step, the calculated total control deviation of the 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.

[0131] 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.

[0132] Fig. Figure 3 shows a compressor characteristic curve schematically and as an example. 800, in which low pressure is plotted on the horizontal axis and high pressure on the vertical axis. Different permissible limits can be seen for both high pressure (HP) and low pressure (LP) for each specific compressor speed.

[0133] Shown are pairs of values ​​from ND and HD, which define the operating range limits for each data point. The permissible operating range for the speed value corresponding to a specific data point lies within a curve connecting the value pairs.

[0134] A business area 802 , which corresponds to an exemplary speed value of a maximum of 20 Hz, as well as an operating range 804 , which corresponds to an exemplary speed value of 120 Hz, are described in detail below.

[0135] The relevant value pairs from ND and HD are for the operating range 802 with reference mark 803 , for the operational area 804with reference mark 805 shown. It can be seen that for the lower speed value of the operating range 802 a permissible maximum high pressure of approximately 25 bar is present, which is significantly lower than the permissible maximum high pressure of approximately 25 bar for the operating range 804 The deviations on the low-pressure side are also significant; for example, a maximum permissible low pressure of over 15 bar applies to the operating range. 802 a maximum permissible ND of just under 8 bar for the operating range 804 opposite.

[0136] For speed values ​​that lie between the operating ranges shown as reference points, interpolation is possible. These ranges can be defined as tables or functions. For example, interpolation can be performed such that if the current compressor speed lies between two reference point speeds, a coefficient is calculated that describes the position of the current compressor speed relative to the flanking reference points. This coefficient represents the ratio of the influence of the respective high / low pressure value pairs assigned to the flanking reference point speeds on the calculation of the high / low pressure value pairs applicable to the current speed.

[0137] The characteristic map 800 It is dependent on the refrigerant and the compressor.

[0138] In addition, further boundary conditions, such as a maximum value of the HD / ND pressure ratio, as well as minimum values ​​for a pressure difference HD / ND for certain values ​​of the compressor speed or over the entire compressor speed, can be defined via the characteristic curve itself.

[0139] The permissible operating range with regard to the current compressor speed is determined based on the characteristic map. 800 calculated. For example, the speed support values ​​flanking the current speed are determined and linearly interpolated element by element.

[0140] If the compressor speed value is outside the range enclosed by the speed reference points, the field with the corresponding speed reference point that has the smallest difference to the compressor speed can be used for further calculations.

[0141] For the control system according to the invention, a minimum low-pressure limit (NDmin), a maximum low-pressure limit (NDmax), a minimum high-pressure limit (HDmin) and a maximum high-pressure limit (HDmax) are then determined for a specific compressor speed.

[0142] Other optional input parameters for the control system include the current inverter temperature, a limit for the maximum inverter temperature, a current compressor torque, and a limit for the maximum compressor torque.

[0143] In particular, limit values ​​for the permissible low pressure range are calculated for the calculation of the permissible range as a function of the current high pressure and the compressor speed: • Low-pressure limit min (NDmin) • Low-pressure limit max (NDmax)

[0144] Furthermore, limit values ​​for the permissible high-pressure range are calculated as a function of the current low pressure and the compressor speed: • Minimum high-pressure limit (HDmin) • Maximum high-pressure limit (HDmax)

[0145] To prevent these limits from being exceeded, a selection of countermeasures can be taken, depending on various other influencing factors: • Increase in compressor speed • Reduction of compressor speed • Increasing the fan speed or pump speed to circulate the heat transfer medium on the heat source side in brine / water heat pumps • Reducing the fan speed or pump speed to circulate the heat transfer medium on the heat source side in brine / water heat pumps • Increase in the water pump speed (for circulating the heat transfer medium used for heating / cooling / hot water charging) assigned to the heat pump • Reduction of the water pump speed (for circulating the heat transfer medium used for heating / cooling / hot water charging) assigned to the heat pump • Increasing the opening degree of the throttling device, for example a main expansion valve (to reduce superheating at the compressor inlet) • Reduction of the throttle opening degree (to increase superheat at the compressor inlet / reduction of low pressure) • Possibly further possibilities for reducing the hot gas temperature by manipulating the opening degree of the expansion valve for intermediate injection. The selection of countermeasures taken depends on • the respective limit (NDmin, NDmax, HDmin, HDmax) • the speed range in which the compressor is currently operating • Setting the assigned parameter for action configuration. According to the invention, speed ranges are first parameterized within which the measures counteracting an exceedance of the range are the same. These so-called speed range classes thus describe which counteracting actions from the list of possible actions are permissible.

[0146] For each speed range class, for example, the speed that represents the upper limit for the transition to the next range is set via a parameter.

[0147] The first speed range class has a lower limit of 0 compressor speed. The permissible total speed range of the compressor can be divided into 2, 3 or a larger number of speed range classes.

[0148] The division into speed range classes is preferably based on the operating range limits applicable to the respective compressor type when operating with the respective refrigerant. These requirements are defined, for example, by the compressor manufacturer for the respective compressor type in conjunction with the refrigerant. The controller developer then derives the speed range classes from these requirements in such a way that the effort and benefit are in the best possible ratio. To achieve the best possible control precision for all criteria, one could certainly define more than five compressor speed ranges individually, but limiting oneself to three or four ranges comes quite close to the optimum without excessively increasing the complexity and effort of the definition.

[0149] Using parameters for the action range setting, it is possible to determine at what distance to the respective limits (NDmin, NDmax, HDmax, current max, moment max) counteracting measures are initiated.

[0150] For the various limit ranges (ND min, ND max, HD max, current max, torque max), the respective assigned range size can preferably be set separately via parameter.

[0151] Furthermore, parameters for the amplification setting of the counteracting measures can be used to determine which change in the triggering process value (e.g., high pressure enters the limit range) causes which change in the counteracting measure (e.g., a factor for a change in rotational speed is calculated).

[0152] The parameters are preferably defined such that the parameter value is set to the factor for the corresponding counteracting measure, which is to be controlled directly at the boundary (of the working area). For example, a factor of 0.5 can be set for a reduction and a factor of 2 for an increase.

[0153] Outside the action range, until the point where a counteracting action enters the action range, the factor is 1 and is then preferably interpolated linearly within the action range up to the value of the parameter (for the gain setting) at the boundary (of the working range). Entry into the action range can occur at a distance threshold to the respective boundary (of the working range); the interpolated factor is preferably referred to as the change factor.

[0154] The determination of change factors is described schematically and exemplarily with reference to Fig. 4 and Fig. 5 explains the lower limit of low pressure NDmin.

[0155] In Fig. 4 will be a course 900 The increase in compressor speed, for example, is determined as a change factor between 2 and 1, depending on the value of the low pressure (ND). For a low pressure (ND) near the lower limit (NDMin), the change factor is 2. In this example, it decreases linearly up to a low pressure (ND) value that corresponds to the lower limit (NDMin) plus a threshold value (NDmin_range). For low pressure values ​​greater than (NDmin + NDmin_range), the change factor is 1, meaning no change occurs. The change factor at the NDMin limit can also be between 1 and 2, or greater than 2, depending on the specific application.

[0156] In Fig. 5 will be a course 950The reduction of, for example, the compressor speed is determined as a change factor between 0.5 and 1, depending on the value of the low pressure (LP). Analogously to Fig. 4. At the limit NDmin, the largest change in absolute value, i.e., in the case of reduction, the smallest change factor, is achieved, which increases to 1, i.e., no change, up to the low pressure value NDmin+NDmin_range.

[0157] Similarly, similar behaviors of the change factors can be determined for different boundaries and actors.

[0158] Subsequently, those factors which, although due to a counter-control for different limits, e.g. NDmin, HDmax, but • in the same direction of action • operate the same actuator as a counteracting measure, combined with each other, preferably aligned.

[0159] For example, all factors that would cause a reduction in compressor speed are balanced in such a way that the factor that would cause the greatest reduction in speed is processed further.

[0160] Therefore, the minimum of all factors prevails as the balanced value for a compressor speed reduction.

[0161] For example, all factors that would cause an increase in compressor speed are balanced in such a way that the factor that would cause the greatest increase in speed is processed further.

[0162] Therefore, the maximum of all factors prevails here as the balanced value for increasing the compressor speed.

[0163] As a next step, all the balanced factors, which relate to the same actor but produce opposing effects, are calculated pairwise. This calculation is performed as a multiplication, so that the effects of the opposing factors are linked in such a way that, in the resulting direction of effect, the factor that differs relatively more from the value 1 dominates.

[0164] If, for example, the factor for compressor speed reduction is 0.8 and the factor for compressor speed increase is 1.4, then a compressor speed change by a factor of 1.12 is calculated.

[0165] Furthermore, limit values ​​of other process variables can also be integrated into the range control according to the invention, such as • Maximum torque • Compressor current max • Inverter temperature max

[0166] For selecting the actions to be considered, i.e., for determining the action configuration, the operating mode is preferably taken into account in addition to the speed range class. A distinction is made between heating and cooling modes, because countermeasures specifically adapted to the operating mode are required to maintain the permissible operating range of the compressor.

[0167] For example, increasing the fan speed during cooling operation is useful when the high pressure approaches the maximum permissible limit.

[0168] Increasing the fan speed would cause the heat energy of the finned / tube heat exchanger, which acts as a condenser in cooling mode, to be better released to the ambient air, thereby reducing the boiling point of the refrigerant in the exchanger and lowering the high pressure.

[0169] In heating mode, however, increasing the fan speed would slightly improve the heat transfer between the evaporator and the outside air, increasing the cooling capacity and thus the heating capacity and therefore the high pressure (slightly), which would not be useful under these conditions (high pressure near the upper limit).

[0170] Thus, different possible actions and different speed range classes result in very different outcomes regarding the approval of each action. The list of possible actions preferably includes several, and in particular all, of the following actions to be approved: • NDmin compressor speed increase • NDmin compressor speed reduction • NDmin fan speed increase / heat source pump speed increase • NDmin Throttle Organ Opening Degree Increase • NDmax compressor speed increase • NDmax compressor speed reduction • NDmax throttle element opening degree reduction • HDmax compressor speed increase • HDmax compressor speed reduction • HDmax fan speed increase • HDmax pump speed increase • HDmin pump speed reduction • Maximum torque compressor speed increase • Maximum torque, compressor speed reduction • Maximum torque, throttle valve opening degree reduction • Inverter temperature max, compressor speed reduction Example 1

[0171] NDmin compressor speed increase is enabled in the speed range up to 92 Hz. • At compressor speeds between 20 Hz and 45 Hz, an increase in compressor speed across the entire high-pressure range results in a desired reduction of the low-pressure minimum limit. • At compressor speeds between 46 Hz and 75 Hz, an increase in compressor speed does not result in a reduction of the low pressure across the entire high pressure range. - min - limit. Here, the selected parameterization is ineffective or even disadvantageous. However, activation is still necessary to increase the compressor speed sufficiently (possibly by a factor of 2 to 90 Hz) at low outside temperatures (-15°C) and high supply temperatures (>55°C) so that the operating point of the compression refrigeration system is again within the permissible range. • At compressor speeds between 76 Hz and 92 Hz, an increase in compressor speed results in a desired reduction of the ND - min limit, at least at high pressures above 27 bar. Example 2

[0172] NDmin compressor speed reduction is enabled in the speed range between 93 Hz and 120 Hz. • At compressor speeds between 93 Hz and 120 Hz, reducing the compressor speed does not produce the desired reduction of the minimum low pressure (ND) limit in almost the entire high-pressure range. However, reducing the compressor speed reduces the cooling capacity approximately proportionally, thus decreasing the temperature difference between the evaporation temperature and the ambient temperature. This is accompanied by an increase in low pressure and therefore increases the distance to the minimum low pressure limit, which is a desired effect. Example 3

[0173] NDmax compressor speed reduction is enabled in the speed range between 76 Hz and 120 Hz. • At compressor speeds between 76 Hz and 120 Hz, a reduction in compressor speed results in a desired increase in the LP - max - limit in almost the entire high-pressure range. Example 4

[0174] HDmax compressor speed increase is enabled in the speed range up to 45 Hz. • At compressor speeds between 20 Hz and 45 Hz, an increase in compressor speed across the entire low-pressure range results in a desired increase in the HD max limit. Example 5

[0175] HDmax compressor speed reduction is enabled in the speed range between 93 Hz and 120 Hz. • At compressor speeds between 93 Hz and 120 Hz, a reduction in compressor speed results in a desired increase in the HD max limit in almost the entire low-pressure range.

[0176] If a process value parameterized as relevant enters the action range of a counter-control measure influencing compressor speed (for example, the high pressure HD approaches the high pressure limit HDmax), then a factor of the compressor speed change is varied unfiltered as a counter-control measure according to the desired relative speed correction.

[0177] However, if a calculated modified high-pressure limit HDmax changes after varying the compressor speed, this in turn can cause a change in the value of the compressor speed change.

[0178] Under unfavorable circumstances, this self-reinforcing overall system reaction may lead to a tendency to oscillate.

[0179] Preferably, to dampen the tendency to vibrate, a limitation of the rate of change of the compressor speed change is provided, wherein the maximum possible rate of change, called compressor speed ramp, is adjustable, in particular in % / sec.

[0180] Setting the compressor speed ramp to 1% / sec would mean that the compressor speed change factor would only change by a maximum of 1% / sec, which in turn would correspond to a maximum compressor speed change rate of 1% / sec resulting from the method according to the invention.

[0181] Fig.Figure 6 shows a schematic and exemplary flowchart of a process. 1000 for controlling a compression refrigeration system 200 .

[0182] In one step 1010 For the current speed of the compressor, a maximum permissible high pressure (HDmax), a minimum permissible high pressure (HDmin), a maximum permissible low pressure (NDmax) and a minimum permissible low pressure (NDmin) are determined.

[0183] In one step 1020 will be based on the limit values ​​from step 1010 Distances between the current high pressure and the minimum and maximum permissible high pressure, as well as distances between the current low pressure and the maximum and minimum permissible low pressure, are determined.

[0184] In one step 1030The current speed range class is determined from the multitude of speed range classes to obtain the permissible counteracting actions. The permissible counteracting actions can optionally also depend on the operating mode, for example, heating and cooling.

[0185] In one step 1040 will be, for each permissible countermeasure from step 1030 , a change factor of the actor belonging to the counteracting action depending on at least one of the in step 1020 calculated at specific distances.

[0186] In one step 1050 The compression refrigeration system will then be ( 200 ) based on the specific change factors of the actors.

[0187] In one example, the original controller output signals, such as compressor speed, throttle opening degree, fan speed setpoint, and pump speed setpoint, from a power controller, overheat controller, fan controller, or pump controller, are combined with the respective actuator change factors as follows. The original controller output signals are also referred to as actuator control signals. Advantageously, the combination is multiplicative. Furthermore, as an advantageous embodiment when applying a factor other than 1, e.g.,The integral component of the respective controller of the original (from the power controller, overheating controller, fan controller, pump controller) controller output signals is set to the value before the application of a factor other than 1, in order to avoid the controllers working against each other in such a way that the control action of the operating range controller according to the invention, which generates the change factors, is canceled out at least after some time. 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 controlling a compression refrigeration system (200) with - a refrigerant, - an evaporator (240), - a compressor (210), - a liquefier (220), - a throttling organ (230) and - a control unit (500) a) for measuring a high pressure (HP) and a low pressure (LP) of the refrigerant in the compression refrigeration system (200), b) to detect the speed of the compressor (210), c) for providing compressor speed range classes, each speed range class defining for a compressor speed range which counteracting actions are permissible from a list of possible actions to maintain the permissible operating range of components of the compression refrigeration system (200), in particular the compressor (210); and d) for controlling at least one actuator of the compression refrigeration system, in particular at least the throttling device (230) and / or the compressor (210), based on the detected high pressure (HP), the detected low pressure (LP) and the detected superheat (dT). ÜE ) in particular by means of a control signal assigned to the respective actuator, the procedure comprising the following steps: - Determine, for the current speed of the compressor, a maximum permissible high pressure (HDmax), a minimum permissible high pressure (HDmin), a maximum permissible low pressure (NDmax) and a minimum permissible low pressure (NDmin); - Determining the distances between the current high pressure and the minimum and maximum permissible high pressure, as well as the distances between the current low pressure and the maximum and minimum permissible low pressure; - Determining the current speed range class to obtain the permissible counter-control actions; - Determine, for each permissible counteracting action, a change factor of the actuator's control signal belonging to the counteracting action, depending on at least one of the specified distances; - Rules of the compression refrigeration system (200) based on the determined change factors of the actuators. [2] Method according to claim 1, wherein the maximum permissible high pressure (HDmax), the minimum permissible high pressure (HDmin), the maximum permissible low pressure (NDmax) and the minimum permissible low pressure (NDmin) are determined by means of a speed characteristic map of the compressor, in particular by direct reading or interpolation of tabulated values. [3] Method according to one of the preceding claims, wherein the method includes a step of combining the change factors of the same actuator, wherein, in the case of change factors which would cause a change in the same direction, the step of combining includes in particular selecting the change factor of the respective actuator which causes the greatest change. [4] Method according to one of the preceding claims, wherein the method includes a step of combining the change factors of the same actuator, wherein, in the case of change factors which would cause an opposite change, the step of combining includes in particular multiplying the change factors. [5] Method according to one of the preceding claims, wherein the control unit is designed to control the throttling device (230), the compressor (210), a heat source-side fan or a brine pump, and a heat sink-side actuator such as a circulation pump. [6] Method according to one of the preceding claims, wherein the step of controlling the compression refrigeration system (200) based on the change factors comprises multiplying the actuator control variables by the change factors. [7] Method according to one of the preceding claims, wherein the list of possible permissible actions for several of the actuators controllable by the control unit includes increasing and decreasing the respective actuation level of the actuator. [8] Method according to one of the preceding claims, wherein the change factor is determined as a function, in particular as a first-order linear function, of at least one of the distances up to a distance threshold, wherein the change factor is in particular limited to a range of values ​​between 0.5 and 1 for reductions or between 1 and 2 for increases. [9] Method according to one of the preceding claims, wherein the step of regulating takes into account further boundary conditions of the operation, in particular a maximum pressure ratio between high pressure and low pressure and / or a minimum value for a pressure difference between high pressure and low pressure. [10] Compression refrigeration system (200) with - a refrigerant, - an evaporator (240), - a compressor (210), - a liquefier (220), - a throttling organ (230) and - a control unit (500) a) for measuring a high pressure (HP) and a low pressure (LP) of the refrigerant in the compression refrigeration system (200), b) to detect the speed of the compressor (210), c) to provide compressor speed range classes, each speed range class defining for a compressor speed range which counteracting actions from a list of possible actions are permitted; and d) for controlling at least one actuator of the compression refrigeration system, in particular at least the throttling device (230) and / or the compressor (210), based on the detected high pressure (HP), the detected low pressure (LP) and the detected superheat (dT). ÜE ), wherein the control unit (500) is further configured to: - Determine, for the current speed of the compressor, a maximum permissible high pressure (HDmax), a minimum permissible high pressure (HDmin), a maximum permissible low pressure (NDmax) and a minimum permissible low pressure (NDmin); - Determining the distances between the current high pressure and the minimum and maximum permissible high pressure, as well as the distances between the current low pressure and the maximum and minimum permissible low pressure; - Determining the current speed range class to obtain the permissible counter-control actions; - Determine, for each permissible counteracting action, a change factor of the actor belonging to the counteracting action depending on at least one of the determined distances; - Rules of the compression refrigeration system (200) based on the determined change factors of the actuators. [11] Compression refrigeration system according to claim 9, wherein the refrigerant has a temperature glide, wherein the refrigerant in particular comprises or consists of R454c, and wherein the compression refrigeration system (200) in particular includes an internal heat exchanger (250) for transferring heat energy of the refrigerant before entering the throttling device (230) to the refrigerant before entering the compressor (210). [12] Heat pump (100) with a compression refrigeration system (200) according to claim 9 or 10.

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