EP3922925A1 - Compression cooling system and method for operating a compression cooling system - Google Patents

Abstract

The present invention relates to a method for regulating a compression refrigeration system (200) and an associated compression refrigeration system (200). The method comprises the following steps: determining a first target superheating of the refrigerant when entering the compressor, the first target superheating maximizing the efficiency of the compression refrigeration system as a function of an operating point of the compression refrigeration system, determining a second target superheating of the refrigerant when entering the compressor on the basis of a maximum permissible hot gas temperature at the outlet of the compressor, and regulating the throttle element (230) on the basis of the lower value from the first setpoint superheat and the second setpoint superheat.

Description

The invention relates to a method for operating a compression refrigeration system and an associated compression refrigeration system with a refrigerant, an evaporator, a compressor, a condenser, a throttle element, an internal heat exchanger for transferring thermal energy of the refrigerant before it enters the throttle element to the refrigerant before it enters the Compressor, and a control unit for detecting overheating of the refrigerant when it enters the compressor, the overheating being defined as a difference between a dew point temperature and a temperature of the refrigerant, and for regulating the throttle element based on the overheating,

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

The refrigerant is driven through the condenser, in which it gives off heat to a heating medium located in a heat sink system. Internal heat is transferred in an internal heat exchanger, for example in the form of a recuperator, between the refrigerant flowing under high pressure from the condenser to the expansion valve and the refrigerant flowing from the evaporator to the compressor under low pressure.

The refrigerant is guided further in a high-pressure flow direction to an expansion valve controlled by the regulator, in which the refrigerant is expanded from high pressure to low pressure as a function of a control value. The refrigerant at the low pressure evaporates in the evaporator when it absorbs source heat.

Out DE 101 59 892 A1 It is known to use a recuperator in a refrigeration machine, in particular in a heat pump, which is intended to increase the heating power in a structurally simple manner at low outside temperatures. For this purpose, the recuperator is dimensioned in such a way that at low evaporation temperatures it transfers at least around 15% of the heat output of the heat pump from the liquid refrigerant to the gaseous refrigerant. An injection valve injects liquid refrigerant into the compressor so that the compression end temperature remains below 120 ° C.

A heat pump system with a refrigerant circuit is off DE 10 2005 061 480 B3 known. It is equipped with a compressor, a first heat exchanger, a throttle element, an evaporator and a 4-2-way valve unit for switching between a first (heating) and a second operating mode (cooling). A direction of flow of the refrigerant in the refrigerant circuit can be switched so that the first heat exchanger is used to liquefy the refrigerant in the first operating mode and to evaporate the refrigerant in the second operating mode, and the second heat exchanger in the first operating mode to evaporate the refrigerant and in the second operating mode is used to liquefy the refrigerant, the first heat exchanger in the refrigerant circuit being connected in such a way that it works as a countercurrent heat exchanger in the two operating modes heating and cooling.

In a refrigeration circuit with a refrigerant, an evaporator, a compressor, a condenser and a throttle device, energy from the environment, for example outside air in air / water heat pumps or brine in brine / water heat pumps, is generated at a comparatively low temperature level when the refrigerant evaporates the refrigerant is transferred, the refrigerant is compressed in the compressor with the aid of electrical energy and then is when the refrigerant is liquefied, energy is transferred to the working medium of a heat sink circuit, for example a heating circuit and / or a hot water charging circuit, at a comparatively high temperature level.

It is known to regulate the throttle element in normal heating operation of the refrigeration circuit in such a way that a target overheating of the refrigerant occurs at the inlet of the compressor, which ensures the greatest possible efficiency of the refrigeration circuit, i.e. maximizes the efficiency of the refrigeration circuit. Normally, target overheating between 5 and 40 K is used for this. For cooling circuits with internal heat exchangers, an efficiency maximum at overheating of 30 to 40 K is optimal for operating points with high pressure differences.

Due to the physical properties of gases, cf., for example, the general gas equation, an increase in pressure of the gas, for example as a result of the compression of the gaseous refrigerant in the compressor, is accompanied by an increase in temperature. Furthermore, the gas temperature before compression also has an influence on the gas temperature after compression; an increase in gas temperature before compression is approximately also associated with an increase in gas temperature after compression.

It should now be taken into account that, in order to protect the components of the compressor, hot gas temperatures at the outlet of the compressor must not be exceeded, whereby these temperatures can be defined as absolute, i.e. applicable to all operating states, or relatively, i.e. depending on the operating state.

Against this background, it was an object of the present invention to specify a compression refrigeration system and a method for regulating such a compression refrigeration system, which enable operation with optimum efficiency while ensuring protection of the components.

We solved the problem by the method features of claim 1 and the device features of claim 7.

Accordingly, a method for regulating a compression refrigeration system with a refrigeration circuit, with a refrigerant, an evaporator, a compressor, a Condenser, a throttle element, an internal heat exchanger for the transfer of thermal energy of the refrigerant before entry into the throttle element to the refrigerant before entry into the compressor, and a control unit a) for detecting overheating of the refrigerant when entering the compressor, the overheating as a The difference between a dew point temperature and a temperature of the refrigerant is defined, and b) for regulating the throttle element based on the overheating is proposed. The method comprises the following steps: determining a first target superheating of the refrigerant when entering the compressor, the first target superheating maximizing the efficiency of the compression refrigeration system as a function of an operating point of the compression refrigeration system, determining a second target superheating of the refrigerant when entering the compressor on the basis of a maximum permissible hot gas temperature at the outlet of the compressor, and regulating the throttle element on the basis of the lower value from the first setpoint superheat and the second setpoint superheat.

By determining two target superheats, with the control system using the lower of the target superheats to control the compression refrigeration system, an efficiency-maximized control of the compression refrigeration system is made possible while maintaining the maximum hot gas temperature at the compressor outlet. The control is easy to implement here, since only the setpoint for the overheating control needs to be adjusted if necessary, namely if the first setpoint overheating would result in too high a hot gas temperature.

The second setpoint overheating is preferably determined based on a refrigeration model that establishes a relationship between the overheating at the inlet of the compressor and the hot gas temperature at the outlet of the compressor.

The second target overheating can, for example, be continuously calculated, in particular on the basis of measured values and / or parameters of the compression refrigeration system, or also be provided for all operating points in the form of a table or the like. As a result, the data processing effort can be kept low during operation.

The refrigeration model preferably contains at least the following input variables: low pressure, high pressure, speed of the compressor, or the following input variables: dew point temperature in low pressure, boiling point temperature in high pressure, speed of the compressor.

The refrigeration model preferably comprises a linear function of quadratic order of the boiling point temperature in high pressure and the dew point temperature in low pressure.

The maximum permissible hot gas temperature at the outlet of the compressor is preferably defined as a temperature which is below a hot gas temperature limit defined for the compressor, in particular defined by the manufacturer.

The throttle element is preferably also regulated based on a currently measured hot gas temperature.

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

In another aspect, the task is achieved by a compression refrigeration system with a refrigeration circuit with a refrigerant, an evaporator, a compressor, a condenser, a throttle element, an internal heat exchanger for the transfer of thermal energy of the refrigerant before it enters the throttle element to the refrigerant before it enters the Compressor, and a control unit a) for detecting overheating of the refrigerant upon entry into the compressor, the overheating being defined as a difference between a dew point temperature and a temperature of the refrigerant, and b) for regulating the throttle element based on the overheating released, the The control unit is designed to: determine a first setpoint overheating of the refrigerant when it enters the compressor, the first setpoint overheating as a function of an operating point of the compression refrigeration system maximizes the efficiency of the compression refrigeration system, determining a second target superheating of the refrigerant when entering the compressor on the basis of a maximum permissible hot gas temperature at the outlet of the compressor, and regulating the throttle element based on the lower value from the first target superheating and the second target Overheating.

The compression refrigeration system according to the invention enables the same advantages to be achieved as the method according to the invention. A combination with all of the preferred embodiments of the method is also advantageously possible.

The refrigerant preferably has a temperature glide, the refrigerant in particular having or consisting of R454C or containing components such as R32 or R1234yf.

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

Fig. 1

Wärmepumpe 100 mit einem Dampfkompressionskreislauf 200

Fig. 2

log p / h - Diagramm des Dampfkompressionsprozesses mit Rekuperator 250

The figures show an exemplary embodiment:

Fig. 1

Heat pump 100 with a vapor compression circuit 200

Fig. 2

log p / h - diagram of the vapor compression process with recuperator 250

• Einen Verdichter 210 zum Verdichten des überhitzten Kältemittels,
• einen Verflüssiger 220, mit einem kältemittelseitigem Verflüssigereintritt 221 und einem Verflüssigeraustritt 222 zur Übertragung von Wärmeenergie Q_H aus dem Dampfkompressionssystem 200 an ein Heizmedium eines Heizsystems 400, mit einem Heizmediumeintritt 401, einem Heizmediumaustritt 402 und einer Heizmediumpumpe 410, zu einer Gebäudeheizung oder ein System zur Warmwassererhitzung,
• vorteilhaft einen Kältemittelsammler 260, welcher als Kältemittelreservoir zum Ausgleich von betriebsbedingungsabhängig unterschiedlich hohen Kältemittelmengenbedarfen verwendet wird,
• ein als Expansionsventil ausgebildetes Drosselorgan 230 zum Expandieren des Kältemittels,
• einen Verdampfer 240, mit einem Verdampfereinlass 241, zur Übertragung von Quellenenergie Q_Q aus einem Wärmequellensystem 300, mit einem Wärmequelleinlass 320 und einem Wärmequellauslass 310, wobei das Wärmequellsystem 300 insbesondere ein Solesystem sein kann, welches Wärmeenergie Q_Q aus dem Erdreich aufnimmt oder ein Luftsystem, welches Wärmeenergie Q_Q aus der Umgebungsluft aufnimmt und an das Dampfkompressionssystem 200 abgibt oder eine beliebige andere Wärmequelle,
• einen Rekuperator als Beispiel eines internen Wärmeübertragers 250, welcher dazu bestimmt ist, innere Wärmeenergie Q_i zwischen dem vom Verflüssiger 220 zum Expansionsventil 230 strömenden Kältemittel auf das vom Verdampfer 240 zum Verdichter 210 strömende Kältemittel zu übertragen und
• ein Kältemittel, insbesondere ein Kältemittelgemisch aus wenigsten zwei Stoffen oder zwei Kältemitteln welches in einer Strömungsrichtung S_HD und S_ND durch den Dampfkompressionskreis 200 strömt, wobei im Dampfkompressionskreislauf 200 Kältemitteldampf durch den Verdichter 210 auf einen Hochdruck HD gebracht wird und zu einem Verflüssiger 220 geführt ist, wobei ein Hochdruckpfad mit der Hochdruckströmungsrichtung S_HD vom Verdichter 210 bis zum Expansionsventil 230 gebildet ist. Nach dem Expansionsventil 230 bis zum Verdichter 210 ist ein Niederdruckpfad mit einer Niederdruckströmungsrichtung S_ND des Kältemittels gebildet, in dem der Verdampfer 240 liegt.

Fig. 1 shows schematically and by way of example a heat pump 100. The heat pump 100 essentially consists of a vapor compression system 200 which forms a compression refrigeration system and which contains the following components:

• A compressor 210 for compressing the superheated refrigerant,
• a condenser 220, with a refrigerant-side condenser inlet 221 and a condenser outlet 222 for transferring thermal energy Q _H from the vapor compression system 200 to a heating medium Heating system 400, with a heating medium inlet 401, a heating medium outlet 402 and a heating medium pump 410, for heating a building or a system for warm water heating,
• Advantageously, a refrigerant collector 260, which is used as a refrigerant reservoir to compensate for different refrigerant quantities depending on the operating conditions,
• a throttle element 230 designed as an expansion valve for expanding the refrigerant,
• an evaporator 240, 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 Wärmequellauslass 310, wherein the heat source system 300 may be in particular a brine system which absorbs heat energy Q _Q from the ground or an air system , which absorbs thermal energy Q _Q from the ambient air and emits it to the vapor compression system 200 or any other heat source,
• a recuperator as an example of an internal heat exchanger 250, which is intended to transfer internal thermal energy Q _i between the refrigerant flowing from the condenser 220 to the expansion valve 230 to the refrigerant flowing from the evaporator 240 to the compressor 210 and
• a refrigerant, in particular a refrigerant mixture of at least two substances or two refrigerants which flows in a flow direction S _HD and S _ND through the vapor compression circuit 200, wherein in the vapor compression circuit 200 refrigerant vapor is brought to a high pressure HD by the compressor 210 and is led to a condenser 220 , wherein a high pressure path with the high pressure flow direction S _HD from the compressor 210 to the expansion valve 230 is formed. After the expansion valve 230 up to the compressor 210, a low-pressure path with a low-pressure flow direction S _{ND of} the refrigerant, in which the evaporator 240 is located, is formed.

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

In the in Fig. 1 The example shown, the heat pump 100 is shown as a brine heat pump. Of course, analogous considerations and advantages can be achieved with air / water heat pumps. In particular in the case of air heat pumps, a fan / fan is arranged as a heat source instead of the brine circuit with brine pump 330.

The compressor 210 is used to compress the superheated refrigerant from an inlet connection 211 to a compressor _outlet pressure P Va at a compressor outlet _temperature T Va at the compressor outlet 212. The compressor 210 usually contains a drive unit with an electric motor, a compression unit and the electric motor can advantageously be operated at variable speed. The compression unit can be designed as a rolling piston unit, scrolling unit or otherwise. At the compressor outlet 212, the compressed superheated refrigerant at the compressor _{outlet pressure P Va is} at a higher pressure level, in particular a high pressure HD, than at the inlet connection 211 with a compressor inlet pressure P _Ve , in particular a low pressure ND, at a compressor _{inlet temperature T VE} , which indicates the state of the refrigerant at the inlet connection 211 describes when entering a compression chamber.

In the condenser 220, thermal energy Q _{H is transferred} from the refrigerant of the vapor compression system 200 to a heating medium of the heat sink system 400. First, the refrigerant is de-heated in the liquefier 220, with superheated refrigerant vapor transferring part of its thermal energy to the heating medium of the heat sink system 400 by reducing the temperature .

After the refrigerant vapor has been de-heated, a further heat transfer Q _H advantageously takes place in the condenser 220 by condensation of the refrigerant during the phase transition from the gas phase of the refrigerant to the liquid phase of the refrigerant. In the process, further heat Q _{H is transferred} from the refrigerant from the vapor compression system 200 to the heating medium of the heat sink system 400.

The high pressure HD of the refrigerant established in the condenser 220 corresponds approximately to a condensation pressure of the refrigerant at a heating medium temperature Tws in the heat sink system when the compressor 210 is in operation.

The heating medium, in particular water, is conveyed by means of a heating medium pump 410 through the heat sink system 400 in a direction SW through the condenser 220, while the thermal energy Q _{H is transferred} from the refrigerant to the heating medium.

In the following collector 260, refrigerant emerging from the condenser 220 is stored, which, depending on the operating point of the vapor compression circuit 200, should not be fed into the circulating refrigerant. If more refrigerant is fed in from the condenser 220 than is passed on through the expansion valve 230, the collector 260 fills, otherwise it is emptied or emptied.

In the following recuperator 250, which can also be referred to as an internal heat exchanger, internal thermal energy Q _i from the refrigerant under high pressure HD, which flows from condenser 220 to expansion valve 230 in a high pressure flow direction S _HD , is transferred to the refrigerant flowing under low pressure LP Transfer refrigerant, which flows from the evaporator to the compressor in a low-pressure _{flow direction S ND} , transferred. The refrigerant flowing from the condenser to the expansion valve 230 is advantageously supercooled.

First, the refrigerant flows into the expansion valve through an expansion valve inlet 231. In the expansion valve 230, the refrigerant pressure is throttled from the high pressure HP to the low pressure LP by the refrigerant advantageously being a nozzle arrangement or throttle with an advantageously variable Opening cross-section happens, the low pressure advantageously corresponding approximately to a suction pressure of the compressor 210. Instead of an expansion valve 230, any other pressure reducing device can also be used. Pressure reducing pipes, turbines or other expansion devices are advantageous.

An opening degree of the expansion valve 230 is set by an electric motor, which is usually designed as a stepping motor, which is controlled by the control unit or regulation 500. The low pressure ND at the expansion valve outlet 232 of the refrigerant from the expansion valve 230 is controlled in such a way that the resulting low pressure ND of the refrigerant during operation of the compressor 210 corresponds _{approximately to the evaporation pressure of the refrigerant with the heat source medium temperature T WQ.} The evaporation temperature of the refrigerant will advantageously be a few Kelvin below the heat source _medium temperature T WQ so that the temperature difference drives heat transfer.

In the evaporator there is a transfer of evaporation heat energy Qv from the heat source fluid of the heat source system 300, which can be a brine system, a geothermal system for using heat energy Q _Q from the ground, an air system for using energy Q _Q from the ambient air or another heat source that uses the source energy Q _Q delivers to vapor compression system 200.

The refrigerant flowing into the evaporator 240 reduces its wet steam portion when flowing through the evaporator 240 by absorbing heat Q _Q and leaves the evaporator 240 advantageously with a low wet steam portion or advantageously also as superheated gaseous refrigerant. The heat source medium is conveyed through the heat source media path of the evaporator 240 by means of a brine pump 330 in the case of brine - water heat pumps or an outside air fan in the case of air / water heat pumps, the thermal energy Q _{Q being} withdrawn from the heat source medium as it flows through the evaporator.

In the recuperator 250, thermal energy Q _{i is transferred} between the refrigerant flowing from the condenser 220 to the expansion valve 230 to that from the evaporator 240 Transferring refrigerant flowing to the compressor 210, the refrigerant flowing from the evaporator 240 to the compressor 210 in particular further overheating.

This superheated refrigerant, which _{exits the recuperator 250 at an overheating temperature T Ke} , is passed to the refrigerant inlet connection 211 of the compressor 210.

The recuperator 250 is used in the vapor compression circuit 200 in order to increase the overall efficiency as the quotient of the heat output Q _H emitted and the electrical power P _{e consumed} to drive the compressor motor.

_{For this purpose, further thermal energy Q i is} withdrawn from the refrigerant, which in the condenser 220 _{emits thermal energy Q H} at a temperature level on the heat sink side, by subcooling in the high pressure path of the recuperator 250.

The internal energy state of the refrigerant when it enters the evaporator 240 is reduced by this heat extraction Q _{i , so that the refrigerant can absorb more thermal energy Q Q} from the heat source 300 at the same evaporation temperature level.

Thereafter, after the evaporator outlet 242 from the evaporator 240, the refrigerant in the low-pressure path at low pressure ND and at a low pressure temperature corresponding to an evaporator outlet temperature T _Va at the inlet to the recuperator 250 is supplied with the _{heat energy Q i extracted in the high-pressure path.} The supply of energy has the advantageous effect of reducing the proportion of wet steam to a state without a proportion of wet steam. Overheating is ensured by additional energy supply.

Furthermore, the following sensors are advantageously arranged to detect the operating state of the vapor compression system 200, with which a model-based precontrol is implemented, in particular to safeguard and optimize the operating conditions of the vapor compression system 200, in particular in the event of changes in the operating state.

• Ein Hochdrucksensor 503 vorteilhaft zur Erfassung des Hochdrucks HD des Kältemittels am Verdichteraustritt 212 oder zwischen dem Verdichteraustritt 212 und dem Expansionsventileintritt 231,
• ein Heißgastemperatursensor 504 vorteilhaft zur Erfassung einer Heißgastemperatur T_HG des Kältemittels am Verdichteraustritt 212, oder im Kältekreisabschnitt zwischen dem Verdichteraustritt 212 und dem Verflüssigereintritt 221,
• ein Innentemperatursensor 506 vorteilhaft zur Erfassung der Innentemperatur T_le des Kältemittels zwischen dem hochdruckseitigem internen Rekuperatorauslass 252 des Kältemittels aus dem Rekuperator 250 und dem Expansionsventileitritt 231. Die Innentemperatur ist vorteilhaft auch als "Rekuperatoraustrittstemperatur Hochdruckpfad" benannt und
• vorteilhaft ein Rekuperatorinnentemperatursensor 505. Der Rekuperatorinnentemperatursensor 505 erfasst vorteilhaft Verflüssigeraustrittstemperatur T_FA des Kältemittels in der Strömungsrichtung am Verflüssigeraustritt oder dem hochdruckseitigen Rekuperatoreintritt und daher wird vorteilhaft die Verflüssigeraustrittstemperatur T_FA vom Rekuperatorinnentemperatursensor 505 gemessen.

On the one hand, with the help of the process values recorded by sensors, safeguards are advantageously made with regard to permissible working ranges of the components, such as in particular the compressor 210; Adjustment of a control deviation that is nevertheless smaller due to the feedforward control only needs to make minor corrections:

• A high pressure sensor 503 is advantageous for detecting the high pressure HD of the refrigerant at the compressor outlet 212 or between the compressor outlet 212 and the expansion valve inlet 231,
• A hot gas temperature sensor 504 is advantageous for detecting a hot gas temperature T _{HG of} the refrigerant at the compressor outlet 212, or in the refrigeration circuit section between the compressor outlet 212 and the condenser inlet 221,
• an internal temperature sensor 506 advantageously for detecting the internal temperature T _{le of} the refrigerant between the high-pressure side internal recuperator outlet 252 of the refrigerant from the recuperator 250 and the expansion valve entry 231. The internal temperature is advantageously also referred to as the "recuperator outlet temperature high pressure path" and
• advantageously a recuperator internal temperature sensor 505. The recuperator internal temperature sensor 505 advantageously detects the condenser outlet temperature T _{FA of} the refrigerant in the direction of flow at the condenser outlet or the high-pressure recuperator inlet and therefore the condenser outlet temperature T _FA is advantageously measured by the recuperator interior temperature sensor 505.

• Ein Niederdrucksensor 502 zur Erfassung des Niederdrucks ND des Kältemittels am Verdichtereintritt 211, oder zwischen dem Expansionsventil 230 und dem Verdichtereintritt 211,
• ein Verdampferaustrittstemperatursensor 508 zur Erfassung der Verdampferaustrittstemperatur T_Va des Kältemittels am Verdampferaustritt 242 oder zwischen dem Verdampferaustritt 242 und dem niederdruckseitigen Eintritt des Kältemittels in den Rekuperatoreinlass 251 des Rekuperators 250 und
• ein Niederdrucktemperatursensor 501 misst vorteilhaft eine Verdichtereintrittstemperatur oder dient vorteilhaft zur Erfassung der Kältemittelniederdrucktemperatur T_ND oder vorteilhaft einer Verdichtereintrittstemperatur T_KE am Verdichtereintritt 211, oder zwischen dem niederdruckseitigem Rekuperatorauslass 252 des Kältemittels aus dem Rekuperator 250 und dem Verdichtereintritt 211.

The following sensors are particularly advantageous for carrying out the method according to the invention:

• A low pressure sensor 502 for detecting the low pressure ND of the refrigerant at the compressor inlet 211, or between the expansion valve 230 and the compressor inlet 211,
• an evaporator outlet temperature sensor 508 for detecting the evaporator outlet temperature T _{Va of} the refrigerant at the evaporator outlet 242 or between the evaporator outlet 242 and the low-pressure side 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 detect the low-pressure refrigerant 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.

The process variable, which has a significant influence on the overall efficiency of the vapor compression circuit 200 as the quotient between the heating power Q _{H transferred by the vapor compression circuit 200 to the electrical power P e} consumed by the compressor 210, is the overheating of the refrigerant at the compressor inlet 211 Compressor operating conditions, however, restrictions with regard to the permitted overheating range of the refrigerant at the compressor inlet are advantageously observed. Overheating that is too low endangers the lubricating properties of the machine oil in particular, while overheating that is too high particularly results in a hot gas temperature that is too high.

The overheating describes the temperature difference between the recorded compressor _{inlet temperature T KE of} the refrigerant and the evaporation temperature of the refrigerant in the case of saturated steam.

According to the invention, the compressor inlet overheating is preferably regulated in such a way that no condensate precipitates due to the water vapor content in the ambient air falling below the dew point in components of the refrigeration circuit, particularly in the section between the refrigerant outlet of the recuperator 252 and the compressor inlet 211. The refrigeration circuit section between evaporator outlet 242 and recuperator inlet 251 is usually colder, because it is typically only a short pipe section, better insulation is possible compared to the section between the refrigerant outlet of recuperator 252 and compressor inlet 211. For example, sits at the point of the compressor inlet 211 on the compressor is the refrigerant separator that is to be protected. This is difficult to enclose, so the temperature should be kept so high that nothing condenses. The problem of condensation does not usually occur on the high pressure side. The passage between the high-pressure side recuperator outlet 252 and the entry into the expansion valve 231 also cools regularly, depending on the operating point, with 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 typically short and can be isolated very well, this section is usually not problematic either. It should be noted, however, that the method according to the invention can in principle prevent a condensate drop over the entire circuit of the heat pump.

If - for the purpose of a numerical example - an evaporation temperature level of approx. -10 ° C is assumed and the temperature at the brine inlet 330 is approx. -10 ° C, at the brine outlet 310 approx. -13 ° C and at the compressor inlet 5 ° C, the overheating is concerned 15K.

In many systems, room temperature sensors and room humidity sensors are advantageous, since they enable precise determination of the condensation conditions of the air, for example at 21 ° C and 60% rel. Moisture the condensation temperature in the range of 13 ° C. Under these conditions, no condensation takes place as long as the pipe temperature is above 13 ° C plus, if necessary, a buffer, e.g. 1K.

Using the numerical example, which of course does not limit, is now achieved overheating of 15 K at a compressor inlet temperature of 5 ° C. This temperature is below 13 ° C, which is determined for the current ambient conditions as the condensation temperature of the water vapor in the ambient air. Accordingly, condensation takes place. If the compressor inlet temperature is to be at least 14 ° C, ie condensation temperature plus buffer, the overheating must be 9K greater, ie an overheating of 24K must be maintained.

Limit values, in particular for overheating, define the permissible overheating range of the components at the compressor inlet 211 as a function of the operating point. Furthermore, there are also dependencies between the compressor _{inlet overheating} dT ÜE and the overall efficiency of the vapor compression _{circuit 200 or between the compressor inlet overheating} dT ÜE and a stability S of a control value R, which is advantageous when regulating the compressor inlet overheating.

To take into account all these requirements, depending on the operating point of the vapor 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 are used to calculate the compressor _{inlet superheat} dT ÜE. Alternatively or additionally, a calculation of the target value Z as a default value for the compressor _{inlet superheating} dT ÜE can be carried out from the refrigeration circuit measurement variables that are dependent on the operating point, such as heat source medium temperature, heating medium temperature, compressor power P _{e and parameterizable coefficients that are adapted to the behavior of the respective refrigeration circuit components.} In the simplest case, the target value for the compressor _{inlet superheat dT ÜE is constant} regardless of all operating conditions, eg 10 Kelvin. In the case of a more complex adaptation it is varied as a function of an operating point variable, for example the compressor power P _e , or in the case of an even more complex adaptation it varies as a function of several operating point variables.

• Regelabweichung der Verdichtereintrittsüberhitzung dT_ÜE = Messwert Verdichtereintrittsüberhitzung - Zielwert Verdichtereintrittsüberhitzung Z_TÜE
• Regelabweichung der Verdampferaustrittsüberhitzung dT_ÜA = Messwert Verdampferaustrittsüberhitzung - Zielwert Verdampferaustrittsüberhitzung Z_TÜA

A control deviation of the compressor _{inlet superheat} dT ÜE and a control deviation of the evaporator _{outlet superheat dT ÜA} are weighted combined, from which a total control deviation is calculated in the controller 500, which is fed in to control the vapor compression circuit 200. The control deviations from the compressor _{inlet superheat} dT ÜE and evaporator _{outlet superheat dT ÜA} are advantageously formed more precisely by forming 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

Then, from the weighted influence of the control deviation of the compressor _{inlet overheating} dT ÜE and the weighted influence of the control deviation of the evaporator _{outlet overheating dT ÜA} in the controller 500, the total control deviation is calculated, which is fed in to regulate the vapor compression circuit 200.

In the vapor compression circuit 200, after the expansion through the expansion valve 230, the refrigerant passes two sequentially arranged heat exchangers, the evaporator 240 and the recuperator 250, in which the refrigerant is supplied with _{thermal energy Q Q} and Q _i.

In the evaporator 250, the refrigerant is supplied _{with source heat energy Q Q} from the heat source system 300. The temperature level of the supplied source heat Q _Q is at a temperature level of the heat source, in particular such as that of the ground or the outside air.

In the recuperator 250 following in the refrigerant high-pressure flow direction S _{HD, thermal energy Q i is} withdrawn from the refrigerant after it has left the condenser 220. The temperature level of the refrigerant at the outlet of the condenser is approximately at the level of the return temperature of the heating medium.

This interconnection of the evaporator 240 with the recuperator 250 in series has a decisive influence on the transfer function of the controlled system for the regulation of compressor _{inlet overheating} dT ÜE.

The control value R is advantageously the weighted link between the control deviation of the compressor _{inlet superheat} dT ÜE and the control deviation of the evaporator outlet superheat.

_{Actuators or operating state variables} with an influence on the control value R, in particular the compressor inlet overheating dT ÜE, are the compressor speed and / or the degree of opening of the in the relevant vapor compression circuit 200 Expansion valve 230, which also advantageously determines the low pressure LP and the evaporation temperature level.

Actuators have a particularly advantageous influence on the control value R, in particular on the weighted linkage of the control deviation of the compressor inlet overheating with the control deviation of the evaporator outlet overheating. In the relevant vapor compression circuit 200, in particular the compressor 210 by varying the compressor speed and the expansion valve 230 by influencing the degree of opening are such actuators. These two actuators influence the low pressure LP and the evaporation temperature level.

Not all influences are desired here. For example, a change in the compressor speed to regulate the desired heating power without further compensatory changes in the degree of opening of the expansion valve advantageously changes the control value R into undesired ranges, so that a model-based, supported change in the degree of opening of the expansion valve to regulate R is advantageous, if necessary, even necessary .

The compressor speed is advantageously set in the vapor compression circuit 200 such that the heating power QH transferred from the vapor compression circuit 200 to the heating medium corresponds to the requested target value Z. To comply with this requirement, influencing the compressor _{speed to control the compressor inlet superheating} dT ÜE is advantageously subordinate or not appropriate.

The degree of opening of the expansion valve 230 is advantageously used as a control _{value for regulating the superheating of} the compressor inlet dT ÜE. The influence of the degree of opening of the expansion valve 230 on the compressor _{inlet superheat} dT ÜE takes place as follows:
The expansion valve 230 acts as a nozzle with a nozzle cross section that can be adjusted by an electric motor, in which a needle-shaped nozzle needle is usually threaded into a nozzle seat by means of a stepper motor.

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

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

If the degree of opening of the expansion valve 230 is reduced, less refrigerant passes through the expansion valve 230 at a constant high pressure HD and initially still a constant low pressure LP. However, since the compressor 210 initially continues to deliver the same refrigerant mass flow, there is _{less in the high pressure flow direction S HD} through the expansion valve 230 Refrigerant supplied than is sucked out by the compressor 210.

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

If the degree of opening of the expansion valve 230 is increased, more refrigerant passes through the expansion valve 230 at a constant high pressure HP and initially still a constant low pressure LP.As the compressor 210 initially continues to convey the same refrigerant mass flow, the low pressure side LP of the refrigeration circuit becomes more refrigerant through the expansion valve 230 fed as is sucked off by the compressor 210. Since the refrigerant vapor is a compressible medium, the low pressure ND on the low pressure side of the vapor compression circuit 200 rises. When the low pressure ND rises, the mass flow rate of the compressor 210 increases approximately proportionally, since its rate can be approximately described as volume / time, and it A correspondingly increased low pressure ND is established, at which the refrigerant mass flow supplied by the expansion valve 230 is equal to the refrigerant mass flow discharged by the compressor 210.

The low pressure ND in turn significantly influences the heat transfer between the heat source medium and the refrigerant in the evaporator 240. The heat flow Q _Q from the heat source system 300 is transferred between the heat source medium and the refrigerant at different temperatures, the heat flow Q _Q depending on the temperature difference between the heat source medium and the refrigerant and the heat transfer resistance of a heat transfer layer of the evaporator 240.

The heat transfer resistance between the heat source media path of the evaporator and the refrigerant path of the evaporator is to be assumed to be approximately constant in a respective vapor compression circuit 200. The size of the heat transfer capacity in the evaporator 240 is therefore critically dependent on the integral of the temperature differences of all surface elements of the heat transfer layer.

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

If the state of aggregation of the refrigerant when flowing through the evaporator 240 is saturated vapor, a refrigerant temperature is established which, as a material property of the refrigerant, is a function of the low pressure ND of the refrigerant as a result of the saturation vapor characteristic. Thus, by a control of the low pressure LP or also an evaporation pressure indirectly control a control of the evaporation temperature of the refrigerant as it flows through the recuperator 250.

The thermal energy Q _Q , which is transferred from the heat source system to the refrigerant flowing through the evaporator 240, influences the state of the refrigerant.

• Nassdampfanteil bei Eintritt in den Verdampfer 240,
• Kältemittelmassenstrom,
• Übertragener Wärmeleistung Q_Q, und von einer
• Enthalpiedifferenz im Nassdampfgebiet beim jeweiligen Niederdruck ND, welche das Kältemittel als eine zugeordnete Funktion des Drucks aufweist.

The proportion of wet steam in saturated refrigerant vapor decreases at constant low pressure when heat is transferred to the refrigerant. In the event of incomplete evaporation, the proportion of wet steam and thus also the internal energy state of the refrigerant when it exits the heat exchanger is a function of:

• Wet steam content when entering the evaporator 240,
• Refrigerant mass flow rate,
• Transferred heat output Q _Q , and from one
• Enthalpy difference in the wet steam area at the respective low pressure LP, which the refrigerant has as an assigned function of the pressure.

For complete evaporation, an additional energy supply takes place in the recuperator 250 in order to overheat the refrigerant beyond the state of saturated vapor.

With the method, given the operating conditions of the vapor compression circuit 200, a corresponding refrigerant state when exiting the evaporator 240 is set as a function of the manipulated variable “degree of opening of the expansion valve 230”.

In the steady-state state, a controlled system behavior with moderate steepness results with regard to a control path steepness of the "isolated" controlled system "evaporator 240". The control system behavior is characterized in particular by the control system output value of the evaporator outlet overheating as a function of the control system input value of the expansion valve opening degree. A refrigerant is advantageously used, in particular a refrigerant mixture as refrigerant, which has a “temperature glide”, in particular R454C is advantageously used. In the case of a refrigerant mixture with a temperature glide, a relative change in the degree of opening of the expansion valve actuator of 1% rel. at the outlet of the refrigerant from the evaporator usually set with a superheating change of advantageously approximately less than 1 K.

• Eine erste Zeitkonstante bewirkt vorteilhaft eine Verzögerung der mechanischen Öffnungsgradänderung des Expansionsventils 230 durch die Begrenzung der Verfahrgeschwindigkeit durch den Regler 500, der Regelwert R wird in dieser ersten Zeitkonstante Z in der Verfahrgeschwindigkeit durch einen Bremswert reduziert. Der Bremswert kann beispielsweise die reglertechnische Zykluszeit, in welcher ein Verfahrschritt des Expansionsventils 230 gesteuert wird, umfassen.
• Eine zweite Zeitkonstante wirkt durch den Regler 500 vorgegeben vorteilhaft auf eine verzögerte Einstellung eines korrespondierenden Niederdruckes bei Öffnungsgradänderungen des Expansionsventils 230 aufgrund der Kompressibilität des Kältemitteldampfes bei Niederdruck ND im Niederdruckpfad.
• Eine dritte Zeitkonstante ist vorteilhaft eine thermische Zeitkonstante der Wärmeübertragungsschicht des Verdampfers 240, wobei eine Änderung des Verdampfungsdruckes und damit der Verdampfungstemperatur eine verzögerte Temperaturänderung der Wärmeübertragungsschicht des Verdampfers, welcher oft mehrere Kilogramm Metall hat und des Wärmequellenmediums.
• Eine vierte Zeitkonstante ergibt sich vorteilhaft aus verzögerten Aggregatzustandsänderungen des Kältemittels bei Verdampfungstemperaturänderungen.
• Eine fünfte Zeitkonstante ergibt sich vorteilhaft aus dem Transport des Kältemittels durch den Verdampfer 240 mit einer endlichen Strömungsgeschwindigkeit.

This state is advantageously also set by influencing at least one or more of the various subsequent time constants in terms of control technology; which ultimately influence the process variable refrigerant overheating at the evaporator outlet 242:

• A first time constant advantageously causes a delay in the mechanical change in the degree of opening of the expansion valve 230 by limiting the travel speed by the controller 500; the control value R is reduced in the travel speed in this first time constant Z by a braking value. The braking value can include, for example, the controller cycle time in which a displacement step of the expansion valve 230 is controlled.
• A second time constant, predetermined by the controller 500, advantageously acts on a delayed setting of a corresponding low pressure when the opening degree of the expansion valve 230 changes due to the compressibility of the refrigerant vapor at low pressure 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, a change in the evaporation pressure and thus the evaporation temperature causing a delayed temperature change in the heat transfer layer of the evaporator, which often has several kilograms of metal, and the heat source medium.
• A fourth time constant results advantageously from delayed changes in the physical state of the refrigerant in the event of changes in the evaporation temperature.
• A fifth time constant results advantageously from the transport of the refrigerant through the evaporator 240 at a finite flow speed.

After changing the manipulated variable "degree of opening of expansion valve 230", there is a delay in the corresponding change in refrigerant condition when exiting the evaporator outlet 242 and an overall time constant Z _tot is advantageously in the range of 30 seconds to about 5 minutes, depending on the operating point.

After flowing through the evaporator 240, the refrigerant enters the low-pressure path of the recuperator 250 at low pressure LP.

If the state of aggregation of the refrigerant when flowing into the recuperator 250 is in a normal operating case, i.e. advantageously either saturated steam with a low vapor content between 0 to 20% or, in particular, also advantageously already superheated refrigerant.

In the case of advantageously saturated steam, a refrigerant temperature is established which, due to the saturation vapor characteristic curve of the refrigerant, is a function of the refrigerant pressure. When superheated refrigerant enters, the refrigerant temperature will at most assume a size which corresponds to the entry temperature of the heat source medium. In this case, the size preferably corresponds to the inlet temperature of the refrigerant into the high-pressure path of the recuperator 250, that is to say the temperature of the refrigerant after it exits the condenser 220.

In order to be able to transfer a sufficient amount of thermal energy from the refrigerant of the high-pressure-side refrigerant path to the refrigerant of the low-pressure-side refrigerant path in recuperator 250, it must be ensured that the temperature of the refrigerant of the high-pressure-side refrigerant path at high pressure HD in as many surface elements of the transfer layer of recuperator 250 as possible is greater than is the temperature of the refrigerant of the low-pressure side refrigerant path at low pressure LP on the respective surface element.

The corresponding temperatures of the heating system 400 of the vapor compression system 200 are higher in a heating case than the corresponding temperatures of the heat source such as the ground or the outside air.

The thermal energy Q _i , which is transferred from the refrigerant at high pressure HD of the high-pressure side refrigerant path to the refrigerant at low pressure in the low-pressure side refrigerant path of the recuperator 250, influences the physical state of the refrigerant on the low-pressure side. The wet steam proportion of the refrigerant flowing through the recuperator 250 on the low pressure side at low pressure LP decreases when heat is transferred to the refrigerant and, after complete evaporation, the refrigerant is advantageously overheated.

• Nassdampfanteil bei Eintritt in den Rekuperator 250,
• Kältemittelmassenstrom,
• übertragene Wärmeleistung Q_i, womit vorteilhaft abhängig von der Temperaturdifferenz zwischen der Temperatur des Kältemittels bei Hochdruck HD im hochdruckseitigen Kältemittelpfad und der Temperatur des Kältemittels des niederdruckseitigen Kältemittelpfades bei Niederdruck ND geregelt wird, und/oder
• eine Enthalpiedifferenz im Nassdampfgebiet beim jeweiligen Niederdruck ND.

The internal energy state of the refrigerant when it emerges from the low-pressure side path of the recuperator is advantageously influenced as a function of one or more of the following factors. It should be noted that the change in energy state is based exclusively on physical dependencies, with the controller influencing the control of the actuators, which of course also influences the physical variables such as the refrigerant mass flow:

• Wet steam content when entering the recuperator 250,
• Refrigerant mass flow rate,
• transferred heat output Q _i , which is advantageously regulated depending on the temperature difference between the temperature of the refrigerant at high pressure HD in the high-pressure side refrigerant path and the temperature of the refrigerant of the low-pressure side refrigerant path at low pressure LP, and / or
• an enthalpy difference in the wet steam area at the respective low pressure LP.

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

In the steady-state condition, the "isolated" controlled system at low pressure ND of the refrigerant in the low-pressure side path of the recuperator 250 results in a control system behavior with high steepness, with an approximately constant internal energy state of the refrigerant upon entry 251 into the low-pressure side LP path of the recuperator 250. With a particularly relative change in the degree of opening of the expansion valve of 1%, there is a change in superheating at the outlet of the refrigerant from the recuperator 250 of approximately 10 K or even more than 10 K.

Compared to the recuperator 250, there is advantageously a significantly higher heat transfer in the evaporator 240 between the source medium and the refrigerant in the evaporator 240.

A significantly higher heat transfer takes place in the evaporator 240 than in the recuperator 250, which is also necessary since a much greater amount of energy is to be withdrawn from the environment by means of the evaporator 240 than is only transferred in the recuperator 250 within the refrigeration circuit. However, the driving temperature difference in the recuperator can be between 20 and 60 K, while it is only between 3 and 10 K in the evaporator. In order to be able to transfer the desired energies in spite of different driving temperature differences, the exchanger surface of the evaporator is, for example, approximately 5 to 20 times larger than that of the recuperator 250.

• Mit einer elften Zeitkonstante Z₁₁ wird vorteilhaft eine Verzögerung der mechanischen Öffnungsgradänderung des Expansionsventils 230 durch die Begrenzung einer Verfahrgeschwindigkeit vorgegeben.
• Eine zwölfte Zeitkonstante Z₁₂ wirkt vorteilhaft auf die verzögerte Einstellung eines korrespondierenden Niederdruckes ND bei Öffnungsgradänderungen des Expansionsventils 230 aufgrund der Kompressibilität des Kältemitteldampfes im Niederdruckpfad ND.
• Eine 13. Zeitkonstante Z13 ist eine thermische Zeitkonstante der Wärmeübertragungsschicht des Verdampfers. Somit bewirkt eine Änderung des Verdampfungsdruckes und damit der Verdampfungstemperatur eine verzögerte Temperaturänderung der Wärmeübertrageschicht, welche oft mehrere Kilogramm Metall beinhaltet, und des Kältemittels im Niederdruckpfad des Verdampfers 240.
•  
• Eine 14. Zeitkonstante Z₁₄ wird vorteilhaft aus verzögerten Aggregatzustandsänderungen des Kältemittels bei Verdampfungstemperaturänderungen ermittelt oder vorgegeben.
• Eine 15. Zeitkonstante Z₁₅ ergibt sich vorteilhaft aus dem Transport des Kältemittels durch den Verdampfer 240 mit einer endlichen Strömungsgeschwindigkeit und wird berücksichtigt.

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 change in the degree of opening of the expansion valve 230 is advantageously specified by limiting a travel speed.
• A twelfth time constant Z ₁₂ has an advantageous effect on the delayed setting of a corresponding low pressure ND when the degree of opening of the expansion valve 230 changes due to the compressibility of the refrigerant vapor in the low pressure path ND.
• A 13th time constant Z13 is a thermal time constant of the heat transfer layer of the evaporator. A change in the evaporation pressure and thus the evaporation temperature thus causes a delayed change in temperature of the heat transfer layer, which often contains several kilograms of metal, and of the refrigerant in the low-pressure path of the evaporator 240.
• A 14th time constant Z ₁₄ is advantageously determined or specified from delayed changes in the physical state of the refrigerant in the event of changes in the evaporation temperature.
• A 15th time constant Z ₁₅ is advantageously obtained from the transport of the refrigerant through the evaporator 240 with a finite flow rate and is taken into account.

The low-pressure side refrigerant path of the recuperator 250 is fed from the evaporator outlet 242 of the evaporator 240. The internal energy state of the refrigerant is already delayed here by at least two time constants Z, Z ₁₁ , Z ₁₂ , Z ₁₃ , Z ₁₄ , Z ₁₅ , Z _tot after the manipulated variable "expansion valve opening degree" has been changed.

After changing the manipulated variable “opening degree of expansion valve 230”, there is a further delay in the corresponding change in refrigerant state due to the time behavior of recuperator 250 when it exits the low-pressure side refrigerant path of recuperator 250.

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

Takes place, it advantageously a weighted combination of the compressor inlet superheating dT _UE and the evaporator outlet superheat dT _ÜA by dividing the total deviation is calculated in particular by means of a weighted combination of the control deviation of the compressor overheating and the deviation of the evaporator outlet superheat dT _ÜA which is fed in the controller 500 for controlling the vapor compression cycle 200th

The compressor _{inlet overheating} dT ÜE is advantageously used as the main control variable and the corresponding signal flows and signal processing takes place in particular in the following process steps:

Step 1

First of all, the process variables compressor _{inlet overheating} dT ÜE are advantageously measured as the main control _{variable and the evaporator outlet overheating dT ÜA is} advantageously recorded as an auxiliary variable in a first process step.

• direkt messtechnisch ermittelt, mit einem Temperatursensor, welcher so positioniert ist, dass er eine der Kältemitteltemperatur im Nassdampfgebiet entsprechende Temperatur erfasst oder
• indirekt messtechnisch ermittelt, mit einem Drucksensor, welcher einen Kältemitteldruck des im Nassdampfgebiet verdampfenden Kältemittels erfasst und aus der kältemittelspezifischen Abhängigkeit zwischen Druck und Temperatur im Nassdampfgebiet dann die Verdampfungstemperatur berechnet wird.

For this purpose, an evaporation temperature of the refrigerant at the respective detection point is either

• determined directly by measurement, with a temperature sensor which is positioned so that it detects a temperature corresponding to the refrigerant temperature in the wet steam area or
• indirectly determined by measurement, with a pressure sensor which detects a refrigerant pressure of the refrigerant evaporating in the wet steam area and then calculates the evaporation temperature from the refrigerant-specific dependency between pressure and temperature in the wet steam area.

Furthermore, the temperatures of the refrigerant temperature assigned to the overheating measuring point, in particular at the evaporator outlet 242 and / or at the compressor inlet 211, are recorded by means of temperature sensors 501, 508. The temperature difference between the refrigerant at the respective measuring point and the evaporation temperature is then calculated and this temperature difference value then corresponds to the respective overheating of the refrigerant at the measuring point.

The output variables of the calculation in step 1 are then the compressor _{inlet superheat} dT ÜE and the evaporator _{outlet superheat dT ÜA} .

step 2

The process variables compressor _{inlet superheating} dT ÜE and evaporator _{outlet superheating dT ÜA} are advantageously offset in a second step to form assigned control deviations with the respectively assigned setpoints:

The setpoint value for the compressor _{inlet superheat} dT ÜE is advantageously varied in the range between approx. 5 K to 20 K in order to ensure the permissible compressor operating range and the highest possible efficiency of the refrigeration circuit.

The setpoint for the evaporator outlet _overheating dT ÜA at the evaporator outlet 242 is then varied depending on the refrigeration circuit operating mode and the refrigeration circuit operating point so that the evaporator overheating in the steady normal case corresponds _{approximately to the process value of the evaporator outlet overheating dT ÜA.} This setpoint for the evaporator outlet _superheat dT ÜA can be precalculated and adaptively corrected based on a model depending on an operating mode or an operating point depending on the evaporation temperature, the condensation temperature, the compressor _{output, a setpoint for the compressor inlet superheat} dT ÜE at the compressor inlet 211 and / or on component properties.

The control deviation of the compressor _{inlet superheat dT ÜE is then calculated by subtracting} the setpoint of the compressor _{inlet superheat} dT ÜE from the process value of the compressor inlet superheat dT ÜE.

It will then calculate the difference of the evaporator outlet overheat dT _ÜA by the target value of the evaporator outlet overheat dT _ÜA is subtracted from the process value of the evaporator outlet overheat dT _Prob.

step 3

In a third method step, the control deviation of the compressor _{inlet superheating} dT ÜE and the control deviation of the evaporator _{outlet superheating dT ÜA are} advantageously combined to form an overall control deviation - superheating.

The combination takes place in particular by means of a weighted addition of the individual control deviations.

The weighting influence is a measure of the proportional combination of the individual control deviations and, in extreme cases, can only be included exclusively an individual control deviation, but usually the weighted inclusion of both individual control deviations.

The weighting influence is advantageously estimated as a value between 0 to 1, i.e. 0 to 100%, and this value is included in the degree of inclusion of the control deviation of the compressor _{inlet superheat} dT ÜE in the total control deviation, which results in the following dependency for the calculation of the total control deviation results in: $$\mathrm{total}-\mathrm{Control\; deviation}\ddot{\mathrm{U}}\mathrm{overheating}=\left(\mathrm{Weighting\; influence}*\mathrm{Control\; deviation\; at\; compressor\; inlet}\ddot{\mathrm{u}}\mathrm{overheating}\right)+\left(\left(\mathrm{1}-\mathrm{Weighting\; influence}\right)*\mathrm{Deviation\; from\; the\; evaporator\; outlet}\ddot{\mathrm{u}}\mathrm{overheating}\right)$$

• Beim Betriebsartübergang zwischen Betriebsart = Betrieb mit ausgeschaltetem Verdichter 210 und Betriebsart = Betrieb mit eingeschaltetem Verdichter 210 im Heizbetrieb wird aufgrund der dynamischen Prozesswerte-änderungen beim Anfahren des Dampfkompressionssystems 200 vorteilhaft ausschließlich zunächst die Regelabweichung Verdampferaustrittsüberhitzung dT_ÜA in die Gesamt - Regelabweichung einbezogen, insbesondere ist der Wert eines Gewichtungseinflusses dann zunächst = 0 oder ein Wert vorteilhaft unter 20 %.
• Nach einer Stabilisierungsphase des Dampfkompressionssystems 200 ist es vorteilhaft, nicht spontan auf den für den Regelbetrieb ausgelegten Wert des Gewichtungseinflusses umzuschalten, sondern den Übergang rampenförmig zu gestalten. In diesem Fall ist es vorteilhaft, dass der Wert vom Gewichtungseinfluss vom Startwert = 0, oder einem Wert insbesondere unter 20%, vorteilhaft rampenförmig auf den vorgesehenen Zielwert angehoben werden. Hiermit wird insbesondere eine Werteunstetigkeit bei einem spontanen Umschalten vermieden und somit Regelschwingungen vermieden.
• Der Zielwert des Gewichtungseinflusses wird vorteilhaft an die jeweilige Betriebsart und den Arbeitspunkt angepasst. Betriebspunkte, welche sich durch erhöhte Schwingneigung auszeichnen bedürfen vorteilhaft einer geringeren Gewichtung der Regelabweichung der Verdichtereintrittsüberhitzung dT_ÜE, insbesondere wird hiermit ein regeltechnisch kritisches Signalverhalten der Verdichtereintrittsüberhitzung dT_ÜE aufgrund der gegenüber der Verdampferaustrittsüberhitzung dT_ÜA größeren Signalverzögerung und größeren Streckensteilheit eine Schwingneigung vermieden.

The value of the weighting influence can advantageously be varied depending on the operating mode and / or the operating point of the heat pump 100:

• When operating mode transition between operating mode = operation with compressor 210 switched off and operating mode = operation with compressor 210 switched on in heating mode, due to the dynamic process value changes when starting up the vapor compression system 200, advantageously only the control deviation evaporator outlet _{overheating dT ÜA is} initially included in the overall control deviation, in particular the The value of a weighting influence then initially = 0 or a value advantageously below 20%.
• After a stabilization phase of the vapor compression system 200, it is advantageous not to switch over spontaneously to the value of the weighting influence designed for regular operation, but rather to design the transition in a ramp-shaped manner. In this case it is advantageous that the value of the weighting influence from the starting value = 0, or a value in particular below 20%, is advantageously increased in a ramp-shaped manner to the intended target value. This in particular avoids a value discontinuity in the event of a spontaneous switchover and thus avoids control fluctuations.
• The target value of the weighting influence is advantageously adapted to the respective operating mode and the operating point. Operating points, which characterized by an increased tendency to oscillate advantageously require a lower weighting of the control deviation of the compressor _{inlet overheating} dT ÜE, in particular a control-technically critical signal _{behavior of the compressor inlet overheating dT ÜE is avoided} due to the greater signal delay and greater steepness of the route compared to the evaporator _{outlet overheating dT ÜA.}

Step 4:

In a fourth method step, the calculated overall control deviation of the overheating is then processed in the controller 500, which controls the corresponding actuators of the refrigeration circuit, in particular the expansion valve 230 with the adjustable degree of opening and / or the compressor 210 with adjustable compressor speed, so that the If a control deviation of the overheating is equal to about 0 Kelvin, if possible.

A P, I, PI, PID controller can be used, the control components being advantageously dynamically adapted to the respective operating mode and the operating point.

To protect components (mechanical components, refrigeration machine oil), the compressor must be protected from excessively high refrigerant gas temperatures when exiting the compression chamber; the compressor manufacturer can also ensure compliance with an absolute (applicable to all operating states) or relative (applicable depending on operating states) prescribe maximum hot gas temperature (e.g. 120 ° C).

In order to prevent the hot gas temperature from exceeding the limit value in an impermissible manner, a hot gas temperature limit value relevant to the control technology (which is below the hot gas temperature limit specified by the compressor manufacturer) is specified (e.g. 110 ° C), which is used as a limit for the controller activities that limit the hot gas temperature.

On the basis of the hot gas temperature limit value relevant to the control, a (simplified) thermodynamic calculation is made for the current Operating point of the refrigeration circuit calculates the overheating of the refrigerant at the compressor inlet, known as compressor inlet overheating, when operating with the hot gas temperature limit value relevant to the control technology.

Influencing variables in this model-based calculation can be low pressure, high pressure, (compressor speed) or dew point temperature ND, boiling point temperature HD, (compressor speed).

From a purely thermodynamic point of view, the compressor speed does not play a role in the calculation, but the dependency of the compression losses of the compressor as a function of the compressor speed can have an influence on the thermodynamic behavior.

A parameterizable approximation of the compressor behavior with regard to the hot gas temperature is preferably used in order to estimate a temperature difference between the hot gas temperature and the superheat at the compressor inlet at the respective operating point of the compressor based on the process data for the evaporation temperature and the condensation temperature. $$\mathbf{Temperature\; difference\; hot\; gas\; temperature\; minus}\ddot{\mathbf{U}}\mathbf{Overheating\; compressor\; inlet\; V}-\mathbf{ND\; sch}\ddot{\mathbf{a}}\mathbf{tzt}=\left(\left(\underline{\mathrm{Parameter\; Q}-\mathrm{Proportion\; of\; condensation\; temperature\; for\; hot\; gas\; temperature\; calculation}}/1000\right)*\mathbf{Boiling\; point\; HD}*\mathbf{Boiling\; point\; HD}\right)+\left(\left(\underline{\mathrm{Parameter\; factor\; condensation\; temperature\; hot\; gas\; temperature\; calculation}}/1000\right)*\mathbf{Boiling\; point\; HD}\right)+\left(\left(\underline{\mathrm{Parameter\; Q}-\mathrm{Proportion\; of\; evaporation\; temperature,\; hot\; gas\; temperature\; calculation}}/1000\right)*\mathbf{Dew\; point\; temperature\; ND}*\mathbf{Dew\; point\; temperature\; ND}\right)+\left(\left(\mathrm{Parameter\; factor\; evaporation\; temperature\; hot\; gas\; temperature\; calculation}/1000\right)*\mathbf{Dew\; point\; temperature\; ND}\right)+\left(\left(\underline{\mathrm{Parameter\; factor\; condensation\; temperature\; times\; evaporation\; temperature\; hot\; gas\; temperature\; calculation}}/1000\right)*\mathbf{Boiling\; point\; HD}*\mathbf{Dew\; point\; temperature\; ND}\right)+\underline{\mathrm{Parameter\; offset\; hot\; gas\; temperature\; calculation}\mathrm{.}}$$

On the basis of this calculation, it is then possible in turn to calculate a corresponding superheating of the compressor inlet for a target hot gas temperature.

With the help of the parameter Limitation of hot gas temperature suction gas superheating V-ND , a limitation of the hot gas temperature can be set, which is then controlled by a corresponding limitation of the compressor inlet superheating.

The corresponding compressor inlet superheat calculated on the basis of an estimated hot gas temperature is managed as a process variable, setpoint superheat compressor inlet V-ND, estimated as a maximum .

A calculation of the maximum permissible setpoint superheating compressor inlet V-ND estimated on the basis of the hot gas temperature limit set with the parameter Limitation hot gas temperature Suction gas superheating V-ND is then estimated with the help of the temperature difference hot gas temperature minus superheating compressor inlet V-ND estimated as an auxiliary variable as follows: $$\mathbf{Setpoint}\ddot{\mathbf{U}}\mathbf{Overheating\; compressor\; inlet\; V}-\mathbf{ND\; max}\ddot{\mathbf{a}}\mathbf{tzt}=\mathrm{maximum}\left[\left(\underline{\mathrm{Parameter\; limit\; hot\; gas\; temperature\; suction\; gas}\ddot{\mathrm{u}}\mathrm{overheating\; V}-\mathrm{ND}}-\mathbf{Temperature\; difference\; hot\; gas\; temperature\; minus}\ddot{\mathbf{U}}\mathbf{Overheating\; compressor\; inlet\; V}-\mathbf{ND\; sch}\ddot{\mathbf{a}}\mathbf{tzt}\right);0\right]$$

As a further refinement, the current hot gas temperature can also be included in the calculation, since the model-based calculation can be incorrect with regard to the time behavior, with regard to tolerances in the acquisition and processing of the included process values, component tolerances (compressor, refrigerant), ambient conditions, e.g. machine room temperature, so that a correction of this calculation based on a detection and inclusion of the actual hot gas temperature is helpful.

• Ist die Heißgastemperatur größer als Parameter Begrenzung Heißgastemperatur Sauggasüberhitzung V-ND, also die Differenz negativ, wird der vorgerechnete Sollwert Überhitzung Verdichtereintritt V-ND maximal geschätzt um die Differenz multipliziert mit Parameter P-Faktor Begrenzung Heißgastemperatur Sauggasüberhitzung V-ND reduziert und als Sollwert Überhitzung Verdichtereintritt V-ND maximal weiterverarbeitet
• Ist die Heißgastemperatur kleiner als Parameter Begrenzung Heißgastemperatur Sauggasüberhitzung V-ND, also die Differenz positiv, wird der vorgerechnete Sollwert Überhitzung Verdichtereintritt V-ND maximal geschätzt um die Differenz multipliziert mit Parameter P-Faktor Begr. Heißgastemperatur Sauggasüberhitzung V-ND erhöht und als Sollwert Überhitzung Verdichtereintritt V-ND maximal weiterverarbeitet

The correction is made by calculating the difference between the parameter Limitation of the hot gas temperature, suction gas superheating V- LP and the hot gas temperature,

• If the hot gas temperature is higher than the parameter Limitation of hot gas temperature suction gas superheating V-ND, i.e. the difference is negative, the precalculated setpoint for superheating compressor inlet V-ND is estimated as a maximum by the difference multiplied by parameter P-factor Limitation of hot gas temperature for suction gas superheating V-ND and reduced as the setpoint for superheating compressor inlet V-ND further processed to the maximum
• If the hot gas temperature is lower than the parameter limit hot gas temperature suction gas overheating V-ND, i.e. the difference is positive, the precalculated setpoint superheating compressor inlet V-ND is estimated as a maximum by the difference multiplied by parameter P-factor Lim. Hot gas temperature suction gas overheating V-ND increased and further processed as a setpoint for superheating compressor inlet V-ND to the maximum

The following applies: $$\mathbf{Setpoint}\ddot{\mathbf{U}}\mathbf{Overheating\; compressor\; inlet\; V}-\mathbf{ND\; maximum}=\mathrm{maximum}\left[\left(\underline{\mathrm{Parameter\; limit\; hot\; gas\; temperature\; suction\; gas}\ddot{\mathrm{u}}\mathrm{overheating\; V}-\mathrm{ND}}-\mathbf{Hot\; gas\; temperature}\right)*\underline{\mathrm{Parameter\; P}-\mathrm{Factor\; Lim}\mathrm{.}\mathrm{Hot\; gas\; temperature\; suction\; gas}\ddot{\mathrm{u}}\mathrm{overheating\; V}-\mathrm{ND}};0\right]$$

Another alternative embodiment is the maximum integrated in the formulas - formation of the overheating setpoint calculated for limiting the hot gas temperature and a lower limit provided for limiting the value range of this overheating setpoint, here the value 0 Kelvin is implemented as an example.

Such a limitation can be advantageous because the result of the model-based calculation of the superheating setpoint calculated for the limitation of the hot gas temperature can also result in negative superheating setpoints which, if regulated to these values, would cause undesired wet steam suction of the compressor. This is avoided by limiting the overheating to a minimum, which can also be interpreted in the (slightly) negative value range if necessary; Priority is subordinate) that the hot gas temperature is not limited exactly to the desired limit value, but may exceed it to a small extent.

In a final step, the compressor inlet superheating setpoint designed for optimum efficiency and the superheating setpoint calculated for limiting the hot gas temperature are combined in such a way that, in the event of a required hot gas temperature limitation, an overheating setpoint that is optimized in terms of efficiency can be reduced: $$\mathbf{Setpoint}\ddot{\mathbf{U}}\mathbf{Overheating\; compressor\; inlet\; V}-\mathbf{ND}=\mathrm{minimum}\left[\mathbf{Setpoint}\ddot{\mathbf{U}}\mathbf{Overheating\; compressor\; inlet\; V}-\mathbf{LP\; efficiency\; setpoint}\ddot{\mathbf{U}}\mathbf{Overheating\; compressor\; inlet\; V}-\mathbf{ND\; maximum}\right]$$

Calculated with exemplary values without the influence of hot gas temperature:
Efficiency-optimized setpoint specification:
Setpoint superheating compressor inlet V-LP efficiency = 30 K

Calculated from Boiling HD, dew point temperature difference ND hot gas temperature minus overheating compressor inlet V-ND estimated = 100 K.

With parameter limitation of hot gas temperature suction gas superheating V-LP = 110 ° C calculated $$\mathbf{Setpoint}\ddot{\mathbf{U}}\mathbf{Overheating\; compressor\; inlet\; V}-\mathbf{ND\; max}\ddot{\mathbf{a}}\mathbf{tzt}=\mathrm{maximum}\left[\left(\underline{\mathrm{Parameter\; limit\; hot\; gas\; temperature\; suction\; gas}\ddot{\mathrm{u}}\mathrm{overheating\; V}-\mathrm{ND}}-\mathbf{Temperature\; difference\; hot\; gas\; temperature\; minus}\ddot{\mathbf{U}}\mathbf{Overheating\; compressor\; inlet\; V}-\mathbf{ND\; sch}\ddot{\mathbf{a}}\mathbf{tzt}\right);0\right]$$

Compressor inlet superheat setpoint V-LP estimated maximum = 10 K

$$\mathbf{Sollwert}\ddot{\mathbf{U}}\mathbf{berhitzung\; Verdichtereintritt\; V}-\mathbf{ND\; maximal}\left(\mathbf{gesch}\ddot{\mathbf{a}}\mathbf{tzt}\right)]$$

Calculated with the above calculation results $$\mathbf{Setpoint}\ddot{\mathbf{U}}\mathbf{Overheating\; compressor\; inlet\; V}-\mathbf{ND}=\mathrm{minimum}\left[\mathbf{Setpoint}\ddot{\mathbf{U}}\mathbf{Overheating\; compressor\; inlet\; V}-\mathbf{LP\; efficiency}\right]$$

$$\mathbf{Setpoint}\ddot{\mathbf{U}}\mathbf{Overheating\; compressor\; inlet\; V}-\mathbf{ND\; maximum}\left(\mathbf{sh}\ddot{\mathbf{a}}\mathbf{tzt}\right)]$$

Compressor inlet superheat setpoint V-LP = 10 K.

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

Claims (9)

Method for regulating a compression refrigeration system (200) with - a refrigeration cycle, with - a refrigerant, - an evaporator (240), - a compressor (210), - a condenser (220), - a throttle device (230), - An internal heat exchanger (250) for transferring thermal energy of the refrigerant before it enters the throttle element (230) to the refrigerant before it enters the compressor (210), and - a control unit (500) a) for detecting overheating (T _ÜE ) of the refrigerant upon entry into the compressor, the superheating (T _ÜE ) being defined as a difference between a dew point temperature and a temperature of the refrigerant, and b) for regulating the throttle element (230) based on the overheating (T _ÜE ), the method comprising the following steps: - Determination of a first target overheating of the refrigerant on entry into the compressor, the first target overheating as a function of an operating point of the compression refrigeration system maximizing the efficiency of the compression refrigeration system, - Determining a second setpoint superheating of the refrigerant upon entry into the compressor on the basis of a maximum permissible hot gas temperature at the exit of the compressor, and - Regulating the throttle element (230) based on the lower value from the first target overheating and the second target overheating. Method according to claim 1, wherein the determination of the second setpoint superheating is based on a refrigeration model that establishes a relationship between the superheating at the inlet of the compressor and the hot gas temperature at the outlet of the compressor. Method according to claim 2, wherein the refrigeration model contains at least the following input variables: low pressure, high pressure, speed of the compressor, or the following input variables: dew point temperature in low pressure, boiling point temperature in high pressure, speed of the compressor. Method according to claim 2 or 3, wherein the refrigeration model comprises a linear function of quadratic order of the boiling point temperature in high pressure and the dew point temperature in low pressure. Method according to one of the preceding claims, wherein the maximum permissible hot gas temperature at the outlet of the compressor is defined as a temperature which is below a hot gas temperature limit defined for the compressor, in particular defined by the manufacturer. Method according to one of the preceding claims, wherein the regulation of the throttle element (230) also takes place based on a currently measured hot gas temperature. Compression refrigeration system (200) with - a refrigeration cycle with - a refrigerant, - an evaporator (240), - a compressor (210), - a condenser (220), - a throttle device (230), - An internal heat exchanger (250) for transferring thermal energy of the refrigerant before it enters the throttle element (230) to the refrigerant before it enters the compressor (210), and - a control unit (500) a) for detecting overheating (T _ÜE ) of the refrigerant upon entry into the compressor, the superheating (T _ÜE ) being defined as a difference between a dew point temperature and a temperature of the refrigerant, and b) for regulating the throttle element (230) based on the overheating (T _ÜE ), wherein the control unit (500) is designed to: - Determination of a first target overheating of the refrigerant on entry into the compressor, the first target overheating as a function of an operating point of the compression refrigeration system maximizing the efficiency of the compression refrigeration system, - Determining a second setpoint superheating of the refrigerant upon entry into the compressor on the basis of a maximum permissible hot gas temperature at the exit of the compressor, and - Regulating the throttle element (230) based on the lower value from the first target overheating and the second target overheating. Compression refrigeration system according to claim 7, wherein the refrigerant has a temperature glide, the refrigerant in particular having or consisting of R454C. Heat pump (100), in particular heat pump (100) installed inside a building, with a compression refrigeration system (200) according to claim 7 or 8.

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