EP3922931B1 - Compression cooling system and method for operating the same - Google Patents

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, an internal heat exchanger, a throttle device and a control unit.

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, where 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 under low pressure from the evaporator to the compressor.

The refrigerant is then guided 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 depending on a control value. The refrigerant at low pressure evaporates in the evaporator by absorbing source heat.

In document EP 1 014 013 A1 discloses a vapor compression type refrigeration cycle, wherein in a vapor compression type refrigeration cycle using carbon dioxide as a refrigerant, a superheat control valve is connected between an evaporator and an internal heat exchanger, the superheat control valve serving to adjust the flow rate of a liquid phase portion of the refrigerant supplied to the evaporator in accordance with a control signal to maintain a degree of superheat of a gas phase portion of the refrigerant supplied to a compressor. Thermistors detect temperatures of the gas phase portion and an evaporated portion of the refrigerant, flowing out of the evaporator to generate first and second temperature signals, respectively. A temperature/pressure sensor detects a condition of a radiated portion of the refrigerant flowing out of a radiator to generate a refrigerant condition value. A selector selects as a selected signal one of the first and second temperature signals in accordance with the refrigerant condition value to supply the selected signal as the control signal to the superheat control valve.

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

Out of EN 11 2016 005264 T5 It is known to provide a refrigeration circuit of an air conditioning system for vehicles, which has an internal heat exchanger, which makes it possible to suppress a deterioration in the coefficient of performance without additional components. The refrigeration circuit comprises a refrigerant circuit including a compressor, a condenser, an electronically controlled expansion valve and an evaporator, an internal heat exchanger, a pressure sensor that detects the pressure of a refrigerant between the condenser and the internal heat exchanger, a first temperature sensor that detects the temperature of the evaporator or a second temperature sensor that has a measuring point between the evaporator and the internal heat exchanger, and a valve opening controller, wherein the valve opening controller uses as the superheat degree the value of a calculated superheat degree that is calculated based on parameters such as the detection value of the pressure sensor and the detection value of the first temperature sensor or the second temperature sensor.

Out of DE 101 59 892 A1 It is known to use a recuperator in a refrigeration machine, especially in a heat pump, to increase the heating output in a structurally simple manner at low outside temperatures. For this purpose, the recuperator is dimensioned in such a way that at low evaporation temperatures it can at least 15% of the heat pump's heating capacity is transferred 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.

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

The control of the compression refrigeration system must meet various requirements, for example, the coefficient of performance must be as high as possible in order to allow the most energy-efficient operation possible. However, it is also crucial that the operating limits of the components are adhered to.

In this context, the compressor is particularly important, as it compresses the gaseous refrigerant from low pressure (LP) to high pressure (HD). It is known that compressors that allow operation at different speeds can be used to optimize performance. In addition to other operating limits, the manufacturers of such variable-speed compressors stipulate, for example, that both upper and lower limits for the low pressure (LP) and high pressure (HD) must be observed depending on the speed. If the limits are exceeded on either the low pressure side or the high pressure side, the compressor must be switched off in the worst case scenario.

Many compressors are unsuitable for sucking in saturated vapor at the compressor inlet, which is why many compressor manufacturers specify that a minimum superheat of the refrigerant at the compressor inlet must be ensured, for example a minimum superheat of 2 Kelvin (K), whereby a minimum superheat must also be observed for refrigerants with a temperature glide.

If the minimum superheat requirement is not met, this can result in increased compressor wear due to insufficient lubrication of the compression components or damage caused by cavitation due to post-evaporation in the compressor.

In refrigeration circuits with an internal heat exchanger, for example in the form of a recuperator, it is known to use an evaporator outlet superheat as an auxiliary control variable for controlling the compressor inlet superheat. When comparing the two superheat values, the evaporator outlet superheat reacts much faster and thus makes it possible to keep vibrations in the refrigeration circuit as low as possible.

The dew point temperature of the refrigerant is pressure-dependent and can be tabulated or calculated in the form of characteristic curves, called wet vapor characteristic curves. To regulate the evaporator outlet superheat, a low pressure in the evaporator can be regulated, for example by setting the opening degree of a throttle device.

However, it has been found that, for example, tolerances in the wet steam characteristics, i.e. the course of the pressure-dependent dew point temperature, which is used to control the evaporator outlet superheat based on a control of the opening degree of a pressure reduction unit such as an expansion valve, can exist, which make reliable control of the compressor inlet superheat difficult.

Against this background, it is an object of the present invention to propose a control method for the compression refrigeration system mentioned at the outset, with which the control of the compressor inlet superheat is improved.

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

A control value is influenced in a commissioning phase of the vapor compression system depending on a control deviation of an evaporator outlet superheat, which is used to control the expansion valve. The control value is further determined after the commissioning phase, during a run-in operating state of the vapor compression system, depending on a compressor inlet superheat and the expansion valve is controlled after the commissioning phase depending on the determined evaporator outlet superheat and the compressor inlet superheat.

A controlled variable for the evaporator outlet superheat is calculated. A control deviation for the evaporator outlet superheat is calculated using a target value for the evaporator outlet superheat. A controlled variable for the compressor inlet superheat is calculated. A control deviation for the compressor inlet superheat is calculated using a target value for the compressor inlet superheat. The control value R is calculated from a weighted influence of the control deviation for the evaporator outlet superheat and a weighted influence of the control deviation for the compressor inlet superheat. The expansion valve is controlled using the control value.

Accordingly, in refrigeration circuits with an internal heat exchanger between the refrigerant path for subcooled refrigerant after exiting the heat sink-side heat exchanger and the refrigerant path after exiting the heat source-side heat exchanger and the compressor inlet, the compressor inlet superheat as well as the evaporator outlet superheat can be included proportionally to control the superheating of the refrigerant.

This method enables a faster control reaction to operating point changes in the refrigeration circuit because the reaction of the evaporator outlet superheat to disturbances such as operating mode/operating point changes is up to 10 times faster than the reaction of the compressor inlet superheat, depending on the refrigeration circuit operating point.

Therefore, the control deviation of the evaporator outlet superheat is usually included in the control to a greater extent than the control deviation of the compressor inlet superheat in order to optimize the overall response time of the control.

In order to achieve the highest possible refrigeration circuit efficiency, one goal is to regulate the superheat at the evaporator outlet as low as possible and to shift the superheat of the refrigerant as completely as possible into the internal heat exchanger.

Thus, the setpoint for evaporator outlet superheat will always be set approximately close to 0 Kelvin relative to the dew line of the refrigerant.

However, when the evaporator outlet superheat is included proportionally, it is necessary to adapt the respective setpoint for evaporator outlet superheat to be controlled as closely as possible to the respective setpoint for compressor inlet superheat to be controlled on a model-based basis.

• Toleranzen bei der Erfassung der Verdampferaustrittstemperatur,
• Toleranzen bei der Erfassung des Verdampfungsdruckes und der daraus erfolgenden Berechnung der Verdampfungstemperatur,
• Toleranzen bei der modellbasierten Berechnung des Druckabfalls zwischen Messstelle am Verdichtereingang und Referenzort am Verdampferausgang,
• Mischungsverhältnisabweichungen der Zusammensetzung des Kältemittels bei Kältemitteln mit Temperaturgleit.

However, deviations in the measured value of the actual value for evaporator outlet superheat arise from

• Tolerances in the detection of the evaporator outlet temperature,
• Tolerances in the recording of the evaporation pressure and the resulting calculation of the evaporation temperature,
• Tolerances in the model-based calculation of the pressure drop between the measuring point at the compressor inlet and the reference point at the evaporator outlet,
• Mixing ratio deviations of the composition of the refrigerant for refrigerants with temperature glide.

According to the invention, a method is now proposed for adaptive compensation of the tolerances that are individual for each heat pump but are usually systematic. This means that the tolerances are caused by, for example, component tolerances that are individual for each heat pump but generally do not change on a daily basis.

According to the invention, a solution is proposed to design the setpoint values for evaporator outlet superheat and setpoint for evaporator outlet superheat such that in the steady state of the superheat control both the control deviation for the evaporator outlet superheat and the control deviation for the compressor inlet superheat are equally zero.

Accordingly, in one aspect of the invention it is proposed that the method for controlling the compression refrigeration system includes the following steps: determining a target value of the evaporator outlet superheat and a target value of the compressor inlet superheat, calculating a correction value based on a control deviation of the compressor inlet superheat from the target value of the compressor inlet superheat, correcting the target value of the evaporator outlet superheat with the calculated correction value, calculating a control value after a commissioning phase of the compression refrigeration system depending on the target value of the evaporator outlet superheat and the target value of the compressor inlet superheat, and controlling the expansion valve on the basis of the control value.

By means of the correction value with which the target value of the evaporator outlet superheat is corrected, the method according to the invention can react to tolerances of the wet steam characteristic curve and control the compression refrigeration system more precisely.

Tolerances in the wet steam characteristic curve regularly arise, for example, due to a separation of refrigerant components when filling the compression refrigeration system in a case where a refrigerant composed of several refrigerants is used. For example, a refrigerant known as R454C, which is usually a mixture of 21.5% R32 and 78.5% R1235yf, can be used.

If compression refrigeration systems are filled with the refrigerant, for example from a provided refrigerant container, the gravity alone causes a separation during the filling process, which leads to a deviation in the composition of the refrigerant mixture in the compression refrigeration system from the composition in the refrigerant storage container. In other words, even if all compression refrigeration systems are filled from the same storage container in a manufacturing process, different mixing ratios of the refrigerant components arise between the compression refrigeration systems. These deviations in the proportions of the refrigerant components can have significant effects on the wet steam characteristic curve when superheating is controlled, which can be corrected using the method according to the invention.

Preferably, the correction value is calculated proportional to a time integral of the control deviation of the compressor inlet superheat to the target value of the compressor inlet superheat.

Preferably, the correction value is corrected in discrete time steps by a proportional part of the control deviation of the compressor inlet superheat.

Preferably, the correction value is limited to a permissible range of values. Preferably, the target value of the evaporator outlet superheat is corrected by adding the calculated correction value.

Preferably, the refrigerant has a temperature glide, wherein the refrigerant in particular comprises or consists of R454C, and wherein the compression refrigeration system in particular contains an internal heat exchanger for transferring thermal energy of the refrigerant before entering the throttle element to the refrigerant before entering the compressor. This is particularly relevant because powerful internal heat exchangers are used in refrigeration circuits with R454C.

The object is further achieved according to the invention by a compression refrigeration system and a heat pump with a compression refrigeration system according to the invention.

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.

Fig. 1

Wärmepumpe 100 mit einem Dampfkompressionskreislauf 200

Fig. 2

log p / h - Diagramm des Dampfkompressionsprozesses mit Rekuperator 250

Fig. 3

schematisch und exemplarisch Nassdampfkennlinien verschiedener Kältemittelmischungen und

Fig. 4

schematisch und exemplarisch eine Adaption eines Verdampferausgangsüberhitzungssollwertes

Further advantages and embodiments are explained below with reference to the figures. Here:

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

Fig.3

schematic and exemplary wet steam characteristics of various refrigerant mixtures and

Fig.4

schematic and exemplary adaptation of an evaporator outlet superheat setpoint

• 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 optionalen 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 exemplarily a heat pump 100. The heat pump 100 essentially consists of a vapor compression system 200 forming a compression refrigeration system, which contains the following components:

• A compressor 210 for compressing the superheated refrigerant,
• a condenser 220, with a refrigerant-side condenser inlet 221 and a condenser outlet 222 for transferring thermal energy Q _H from the vapor compression system 200 to a heating medium of a heating system 400, with a heating medium inlet 401, a heating medium outlet 402 and a heating medium pump 410, to a building heating system or a system for hot water heating,
• advantageously a refrigerant collector 260, which is used as a refrigerant reservoir to compensate for different refrigerant quantity requirements depending on the operating conditions,
• a throttle element 230 designed as an expansion valve for expanding the refrigerant,
• an evaporator 240, with an evaporator inlet 241, for transferring source energy 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 can in particular be a brine system which absorbs heat energy Q _Q from the ground or an air system which absorbs heat energy Q _Q from the ambient air and releases it to the vapor compression system 200 or any other heat source,
• a recuperator as an example of an optional internal heat exchanger 250, which is intended to transfer internal heat energy Q _i between the refrigerant flowing from the condenser 220 to the expansion valve 230 to the refrigerant flowing from the evaporator 240 to the compressor 210 and
• a refrigerant, in particular a refrigerant mixture of at least two substances or two refrigerants which flows in a flow direction S _HD and S _ND through the vapor compression circuit 200, wherein in the vapor compression circuit 200 refrigerant vapor is brought to a high pressure HD by the compressor 210 and is led to a condenser 220, wherein a high pressure path with the high pressure flow direction S _HD is formed from the compressor 210 to the expansion valve 230. After the expansion valve 230 to the compressor 210, a low pressure path with a low pressure flow direction S _ND of the refrigerant is formed, in which the evaporator 240 is located.

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

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

The compressor 210 serves to compress the superheated refrigerant from an inlet connection 211 to a compressor outlet pressure P _Va at a compressor outlet temperature corresponding to the hot gas temperature at the compressor outlet 212. The compressor 210 usually contains a drive unit with an electric motor, a compression unit and advantageously the electric motor can be operated at variable speed. The compression unit can be designed as a rotary piston unit, scroll unit or otherwise. At the compressor outlet 212 the compressed superheated refrigerant is at the compressor outlet pressure P _Va 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 compressed inlet temperature T _VE , which describes the state of the refrigerant temperature at the inlet connection 211 when entering a compression chamber.

In the condenser 220, the transfer of thermal energy Q _H from the refrigerant of the vapor compression system 200 to a heating medium of the heat sink system 400 takes place. First, the refrigerant is deheated in the condenser 220, with superheated refrigerant vapor transferring part of its thermal energy to the heating medium of the heat sink system 400 by reducing its temperature.

After the refrigerant vapor has been deheated, a further heat transfer Q _H advantageously takes place in the condenser 220 through condensation of the refrigerant during the phase transition from the gas phase of the refrigerant to the liquid phase of the refrigerant. In this case, 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 during operation of the compressor 210 corresponds approximately to a condensation pressure of the refrigerant at a heating medium temperature Tws in the heat sink system.

The heating medium, in particular water, is pumped by means of a heating medium pump 410 through the heat sink system 400 in a direction SW through the condenser 220, whereby the thermal energy Q _H is transferred from the coolant to the heating medium.

The subsequent collector 260 stores coolant exiting from the condenser 220, which should not be fed into the circulating coolant depending on the operating point of the vapor compression circuit 200. If more coolant is fed in from the condenser 220 than is passed on through the expansion valve 230, the collector 260 fills up, otherwise it becomes emptier or emptied.

In the subsequent recuperator 250, which can also be referred to as an internal heat exchanger, internal heat energy Q _i is transferred from the refrigerant under the high pressure HD, which flows from the condenser 220 to the expansion valve 230 in a high pressure flow direction S _HD , to the refrigerant flowing under the low pressure ND, which flows from the evaporator to the compressor in a low pressure flow direction S _ND flows. The refrigerant flowing from the condenser to the expansion valve 230 is advantageously subcooled.

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 HD to the low pressure ND by the refrigerant advantageously passing through a nozzle arrangement or throttle with an advantageously variable opening cross-section, whereby the low pressure advantageously corresponds approximately to a suction pressure of the compressor 210. Instead of an expansion valve 230, any other desired pressure reduction device can also be used. Pressure reduction pipes, turbines or other expansion devices are advantageous.

The degree of opening of the expansion valve 230 is set by an electric motor, which is usually designed as a stepper motor, which is controlled by the control unit or regulator 500. The low pressure ND at the expansion valve outlet 232 of the refrigerant from the expansion valve 230 is controlled so that the resulting low pressure ND of the refrigerant during operation of the compressor 210 corresponds approximately to the evaporation pressure of the refrigerant at 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, a transfer of evaporation heat energy Qv takes place 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 releases the source energy Q _Q to the vapor compression system 200.

The coolant flowing into the evaporator 240 reduces its wet vapor content as it flows through the evaporator 240 by absorbing heat Q _Q and leaves the evaporator 240 advantageously with a low wet vapor content or advantageously also as a superheated gaseous coolant. The heat source medium is conveyed through the heat source medium 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, whereby the heat energy Q _Q is extracted 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 the refrigerant flowing from the evaporator 240 to the compressor 210, wherein the refrigerant flowing from the evaporator 240 to the compressor 210 in particular continues to superheat.

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

The recuperator 250 is used in the vapor compression circuit 200 to increase the overall efficiency as a quotient of the delivered heating power Q _H and the absorbed electrical power P _e to drive the compressor motor.

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

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

The heat energy Q _i extracted in the high-pressure path is then fed back to the refrigerant after the evaporator outlet 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 supply of energy advantageously causes a reduction in the wet steam portion to a state without a wet steam portion and then superheating occurs through further 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 feedforward control is implemented, in particular to safeguard and optimize the operating conditions of the vapor compression system 200, in particular in the event of operating state changes.

• 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 Kältemittel 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, the process values recorded by sensors are used to ensure the permissible working ranges of the components, such as the compressor 210 in particular, and on the other hand, model-based Pre-controls, in particular of a speed of the compressor 210 and/or a valve opening degree of the expansion valve, so that the controller only has to make small corrections to compensate for a control deviation that is nevertheless smaller due to the pre-control:

• A high pressure sensor 503 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 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 advantageous 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 inlet 231. The internal temperature is advantageously also referred to as "recuperator outlet temperature high pressure path" and
• advantageously a recuperator internal temperature sensor 505. The recuperator internal temperature sensor 505 advantageously detects the condenser outlet temperature T _FA of the refrigerant in the flow direction at the condenser outlet or the high-pressure side recuperator inlet and therefore the condenser outlet temperature T _FA is advantageously measured by the recuperator internal temperature sensor 505.

• 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 inlet of the refrigerant into the recuperator inlet 251 of the recuperator 250 and
• a low-pressure temperature sensor 501 advantageously measures a compressor inlet temperature or is advantageously used to detect the refrigerant low-pressure temperature T _ND or advantageously a compressor inlet temperature T _KE at the compressor inlet 211, or between the low-pressure side recuperator outlet 252 of the refrigerant from the recuperator 250 and the compressor inlet 211.

The process variable which has a significant influence on the overall efficiency of the vapor compression circuit 200 as a quotient between the heating power Q _H transmitted by the vapor compression circuit 200 and an electrical power P _e absorbed by the compressor 210 is the superheat of the refrigerant at the compressor inlet 211. In order to comply with permissible compressor operating conditions, however, it is advantageous to observe restrictions with regard to the permitted superheat range of the refrigerant at the compressor inlet. Overheating that is too low endangers the lubricating properties of the machine oil in particular, while overheating that is too high causes the hot gas temperature to be too high.

The superheat describes the temperature difference between the measured compressor inlet temperature T _KE of the refrigerant and the evaporation temperature of the refrigerant at saturated vapor.

The compressor inlet superheat is controlled in such a way that no condensate is formed on components of the refrigeration circuit due to the water vapor content in the ambient air falling below the dew point, particularly in the section between the refrigerant outlet of the recuperator 252 and the compressor inlet 211. Although the refrigeration circuit section between the evaporator outlet 242 and the recuperator inlet 251 is usually colder because this is typically only a short pipe section, better insulation is possible compared to the section between the refrigerant outlet of the recuperator 252 and the compressor inlet 211. For example, the refrigerant separator that needs to be protected is located at the location of the compressor inlet 211 on the compressor. This is difficult to enclose, so the temperature here needs to be kept high enough 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 inlet into the expansion valve 231 also 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 not problematic in general. However, it should be noted that the This process can fundamentally prevent condensate loss throughout the entire heat pump circuit.

If - for the purposes of a numerical example - an evaporation temperature level of approximately -10°C is assumed and the temperature at the brine inlet 330 is approximately -10°C, at the brine outlet 310 approximately -13°C and at the compressor inlet 5°C, the superheat is 15K. In many systems, room temperature sensors and room humidity sensors are advantageous, as they enable the condensation conditions of the air to be determined precisely. For example, at 21°C and 60% relative humidity, the condensation temperature is in the range of 13°C. Under these conditions, as long as the pipe temperature is above 13°C plus a buffer if necessary, e.g. 1K, no condensation takes place.

The numerical example, which is of course not restrictive, is used to achieve a superheat of 15K at a compressor inlet temperature of 5°C. This temperature is below the 13°C that is determined for the current ambient conditions as the condensation temperature of the water vapor in the ambient air. Condensation therefore takes place. If the compressor inlet temperature is to be at least 14°C, i.e. condensation temperature plus buffer, the superheat must be increased by 9K, i.e. a superheat of 24K must be maintained.

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

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

• $$\begin{array}{l}\mathrm{Regelabweichung}\mathrm{der}\mathrm{Verdichtereintrittsüberhitzung}{\mathrm{dT}}_{\mathrm{ÜE}}=\mathrm{Messwert}\mathrm{Verdich}-\\ \mathrm{tereintrittsüberhitzung}-\mathrm{Zielwert}\mathrm{Verdichtereintrittsüberhitzung}{\mathrm{Z}}_{\mathrm{TÜE}}\end{array}$$
  
• $$\begin{array}{l}\mathrm{Regelabweichung}\mathrm{der}\mathrm{Verdampferaustrittsüberhitzung}{\mathrm{dT}}_{\mathrm{ÜA}}=\mathrm{Messwert}\mathrm{Verdamp}-\\ \mathrm{feraustrittsüberhitzung}-\mathrm{Zielwert}\mathrm{Verdampferaustrittsüberhitzung}{\mathrm{Z}}_{\mathrm{TÜA}}\end{array}$$
  

A control deviation of the compressor inlet superheat dT _ÜE and a control deviation of the evaporator outlet superheat dT _ÜA are combined in a weighted manner, from which a total control deviation is calculated in the controller 500, which is fed into the control of the vapor compression circuit 200. Advantageously, the control deviations of the compressor inlet superheat dT _ÜE and evaporator outlet superheat dT _ÜA are first formed more precisely by forming the differences between the respective measured values and target values.

• $$\begin{array}{l}\mathrm{Control\; deviation}\mathrm{the}\mathrm{Compressor\; inlet\; superheat}{\mathrm{dT}}_{\mathrm{ÜE}}=\mathrm{Measured\; value}\mathrm{Compact}-\\ \mathrm{ter\; inlet\; superheat}-\mathrm{Target\; value}\mathrm{Compressor\; inlet\; superheat}{\mathrm{Z}}_{\mathrm{TUE}}\end{array}$$
  
• $$\begin{array}{l}\mathrm{Control\; deviation}\mathrm{the}\mathrm{Evaporator\; outlet\; superheat}{\mathrm{dT}}_{\mathrm{ÜA}}=\mathrm{Measured\; value}\mathrm{Evaporation}-\\ \mathrm{outlet\; superheat}-\mathrm{Target\; value}\mathrm{Evaporator\; outlet\; superheat}{\mathrm{Z}}_{\mathrm{TÜA}}\end{array}$$
  

Then, the total control deviation is advantageously calculated from the weighted influence of the control deviation of the compressor inlet superheat dTüε and the weighted influence of the control deviation of the evaporator outlet superheat dT _ÜA in the controller 500, which is fed into the control of the vapor compression circuit 200.

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

In the evaporator 250, source heat energy Q _Q from the heat source system 300 is supplied to the refrigerant. The temperature level of the supplied source heat Q _Q is at a temperature level of the heat source, in particular the ground or the outside air.

In the recuperator 250 following in the high-pressure refrigerant flow direction S _HD , thermal energy Q _i is extracted from the refrigerant after it leaves the condenser 220.

The temperature level of the refrigerant at the outlet of the condenser is approximately equal to the return temperature of the heating medium.

This connection of the evaporator 240 with the recuperator 250 in series has a decisive influence on the transfer function of the control system for controlling the compressor inlet superheat dTüε.

The control value R is advantageously the weighted combination of the control deviation of the compressor inlet superheat dTüε with the control deviation of the evaporator outlet superheat.

Actuator operating state variables with an influence on the control value R, in particular the compressor inlet superheat dTüε, are the compressor speed and/or the degree of opening of the expansion valve 230 in the relevant vapor compression circuit 200, which also advantageously determines the low pressure ND and the evaporation temperature level.

Actuators have a particularly advantageous influence on the control value R, in particular on the weighted link between the control deviation of the compressor inlet superheat and the control deviation of the evaporator outlet superheat. In the vapor compression circuit 200 in question, 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 ND and the evaporation temperature level.

Not all influences are desired here. For example, a change in the compressor speed to regulate the desired heating output without further compensatory changes in the degree of opening of the expansion valve changes the control value R into undesirable ranges, so that a model-based supported change in the degree of opening of the expansion valve to regulate R, which accompanies the change in compressor speed, is advantageous, and may even be necessary.

Advantageously, the compressor speed in the vapor compression circuit 200 is set such that the heating power QH transferred from the vapor compression circuit 200 to the heating medium corresponds to the requested target value Z. In order to comply with this requirement, influencing the compressor speed to control the compressor inlet superheat 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 controlling the compressor inlet superheat dTüε. The influence of the degree of opening of the expansion valve 230 on the compressor inlet superheat dT _ÜE is as follows:
The expansion valve 230 acts as a nozzle with an electric motor-adjustable nozzle cross-section, 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 flow rate through the expansion valve is approximately proportional to the square root of the pressure difference between the expansion valve inlet 231 and outlet 232 multiplied by a current relative value of the nozzle cross-section or degree of opening and, advantageously, a constant dependent on the refrigerant and a geometry of the expansion valve 230.

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 when entering the expansion valve 230 can also be assumed to be constant, the degree of opening of the expansion valve 230 significantly influences only the low pressure ND, i.e. the outlet pressure from the expansion valve 230.

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

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

If the degree of opening of the expansion valve 230 is increased, more refrigerant passes through the expansion valve 230 at constant high pressure HD and initially constant low pressure ND. Since the compressor 210 continues to initially deliver the same refrigerant mass flow, more refrigerant is supplied to the low pressure side ND of the refrigeration circuit through the expansion valve 230 than is 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 increases. As the low pressure ND increases, the mass flow rate of the compressor 210 increases approximately proportionally, since its delivery rate can be approximately described as volume / time, and a correspondingly increased low pressure ND is set in, at which the refrigerant mass flow supplied through 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, whereby the heat flow Q _Q is dependent on the temperature difference between the heat source medium and the refrigerant and the heat transfer resistance of a heat transfer layer of the evaporator 240.

The heat transfer resistance between the heat source medium path of the evaporator and the refrigerant path of the evaporator is assumed to be approximately constant in a respective vapor compression circuit 200. Therefore, the magnitude of the heat transfer performance in the evaporator 240 is largely 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 heat energy Q _Q from the heat source system 300 to the coolant, it must be ensured that the temperature of the heat source medium in as many surface elements of the transfer layer of the heat exchanger, here the evaporator 240, as possible is greater than the temperature of the coolant at the respective surface element.

If the state of aggregation of the refrigerant is saturated vapor as it flows through the evaporator 240, a refrigerant temperature is established which is a function of the low pressure ND of the refrigerant due to the saturation vapor characteristic curve as a material property of the refrigerant. Thus, by controlling the low pressure ND or also an evaporation pressure can indirectly control 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 aggregation 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 Stoffkonstante als Funktion des Drucks aufweist.

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

• Wet steam content entering the evaporator 240,
• Refrigerant mass flow,
• Transferred heat power Q _Q , and of a
• Enthalpy difference in the wet steam region at the respective low pressure ND, which the refrigerant has as a material constant as a function of pressure.

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

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

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

Advantageously, a refrigerant, in particular a refrigerant mixture, is used which has a "temperature glide", in particular R454C is advantageously used. Advantageously, in the case of a refrigerant mixture with a temperature glide, With a relative change in the degree of opening of the expansion valve actuator of 1% rel., a superheat change of less than 1 K is usually set at the outlet of the refrigerant from the evaporator.

• 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.

The setting of this state is also advantageously carried out by a control-technical influence on at least one or more of the following various time constants; which ultimately influence the process variable refrigerant superheat at the evaporator outlet 242:

• A first time constant advantageously causes a delay in the mechanical opening degree change of the expansion valve 230 by limiting the travel speed by the controller 500, the control value R is reduced in the travel speed by a braking value in this first time constant Z. The braking value can, for example, include the controller cycle time in which a travel step of the expansion valve 230 is controlled.
• A second time constant specified by the controller 500 has an advantageous effect 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 ND in the low-pressure path.
• A third time constant is advantageously a thermal time constant of the heat transfer layer of the evaporator 240, wherein a change in the evaporation pressure and thus the evaporation temperature causes a delayed temperature change of the heat transfer layer of the evaporator, which often has several kilograms of metal, and of the heat source medium.
• A fourth time constant advantageously results from delayed changes in the state of aggregation of the refrigerant when the evaporation temperature changes.
• A fifth time constant advantageously results from the transport of the refrigerant through the evaporator 240 at a finite flow velocity.

Thus, after changing the manipulated variable "degree of opening of the expansion valve 230", a delay in the corresponding change in the state of the refrigerant when exiting the evaporator outlet 242 occurs, and a total time constant Z _total 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 ND.

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

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

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

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

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, has an effect on the state of aggregation of the refrigerant on the low-pressure side. The wet vapor portion of the refrigerant flowing through the recuperator 250 on the low-pressure side at low pressure ND 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 leaving the low-pressure side path of the recuperator, is advantageously influenced depending on one or more of the following factors. It should be noted that the energy state change is based exclusively on physical dependencies, whereby the controller influences the control of the actuators, which then of course also influences the physical quantities such as the refrigerant mass flow:

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

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

In the steady state, with regard to the control system slope of the "isolated" control system at the low pressure ND of the refrigerant in the low-pressure side path of the recuperator 250, a control system behavior with a high slope results, with an approximately constant internal energy state of the refrigerant at the inlet 251 into the low-pressure side ND path of the recuperator 250. With a relative opening degree change of the expansion valve of 1% in particular, a superheat change at the outlet of the refrigerant from the evaporator 230 of advantageously about 10 K or even more than 10 K is set.

Compared to the recuperator 250, a much higher heat transfer in the evaporator 240 between the source medium and the refrigerant in the evaporator 240 advantageously takes place.

For example, a significantly higher heat transfer is set in the evaporator 240 than in the recuperator 250, since a significantly greater amount of energy is to be extracted from the environment by means of the evaporator 240 than is transferred within the refrigeration circuit in the recuperator 250 alone. The driving temperature difference in the recuperator is, for example, between 20 K and 60 K, while in the evaporator it is only between 3 K and 10 K. In order to be able to transfer the desired energies despite different driving temperature differences, the exchange surface of the evaporator is designed to be 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.

The setting of this state is advantageously carried out using at least one of the following time constants Z:

• With an eleventh time constant Z ₁₁ , a delay of the mechanical opening degree change of the expansion valve 230 is advantageously specified by limiting a travel speed.
• A twelfth time constant Z ₁₂ has a beneficial effect on the delayed setting of a corresponding low pressure ND when the opening degree of the expansion valve 230 changes due to the compressibility of the refrigerant vapor in the low pressure path ND.
• A 13th time constant Z ₁₃ is a thermal time constant of the heat transfer layer of the evaporator. Thus, a change in the evaporation pressure and thus the evaporation temperature causes a delayed temperature change of the heat transfer layer, which often contains several kilograms of metal, and of the refrigerant in the low-pressure path of the evaporator 240.
• A 14th time constant Z ₁₄ is advantageously determined or specified from delayed changes in the state of aggregation of the refrigerant during changes in evaporation temperature.
• A 15th time constant Z ₁₅ advantageously results from the transport of the refrigerant through the evaporator 240 with a finite flow velocity 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. Here too, the internal energy state of the refrigerant is already delayed by at least two time constants Z, Z ₁₁ , Z ₁₂ , Z ₁₃ , Z ₁₄ , Z ₁₅ , Z _total after changing the manipulated variable "expansion valve opening degree".

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

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

A weighted combination of compressor inlet superheat dTüε and evaporator outlet superheat dT _ÜA is advantageously carried out by calculating the total control deviation, which is fed into the controller 500 for controlling the vapor compression circuit 200, 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 compressor inlet superheat dTüε 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, the process variables compressor inlet superheat dT _ÜE are advantageously measured as the main control variable and the evaporator outlet superheat dT _ÜA are advantageously measured as the 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

• directly determined by measurement, with a temperature sensor positioned to detect a temperature corresponding to the refrigerant temperature in the wet steam region or
• determined indirectly by measurement, with a pressure sensor which detects a refrigerant pressure of the refrigerant evaporating in the wet steam region and the evaporation temperature is then calculated from the refrigerant-specific dependence between pressure and temperature in the wet steam region.

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

The starting values for the calculation in step 1 are then the compressor inlet superheat dTüε and the evaporator outlet superheat dT _ÜA .

Step 2: The process variables compressor inlet superheat dTüε and evaporator outlet superheat dT _ÜA are advantageously offset in a second step to form assigned control deviations with respective assigned setpoints:
The setpoint for the compressor inlet superheat dTüε is advantageously varied in the range between approx. 5 K and 20 K to ensure the permissible compressor operating range and the highest possible efficiency of the refrigeration circuit.

The setpoint for the evaporator outlet superheat dT _ÜA at the evaporator outlet 242 is then varied depending on the refrigeration circuit operating mode and the refrigeration circuit operating point so that the evaporator superheat in the steady state control case approximately corresponds to the resulting process value of the evaporator outlet superheat dT _ÜA . This setpoint for the evaporator outlet superheat dT _ÜA can be precalculated on a model-based basis 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 dTüε at the evaporator outlet 242 and/or component properties and can be adaptively corrected.

The control deviation of the compressor inlet superheat dT _ÜE is then calculated by subtracting the setpoint value of the compressor inlet superheat dTüε from the process value of the compressor inlet superheat dT _ÜE .

The control deviation of the evaporator outlet superheat dT _ÜA is then calculated by subtracting the setpoint value of the evaporator outlet superheat dT _ÜA from the process value of the evaporator outlet superheat dT _ÜA .

Step 3: In a third process step, the control deviation of the compressor inlet superheat dT _ÜE and the control deviation of the evaporator outlet superheat dT _ÜA are advantageously combined to form a total control deviation superheat.

The combination is carried out 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 can, in extreme cases, result in the exclusive inclusion of only one individual control deviation, but usually the weighted inclusion of both individual control deviations.

It is advantageous to estimate the weighting influence as a value between 0 and 1, i.e. 0 to 100%, and this value is included in the degree to which the control deviation of the compressor inlet superheat dT _ÜE is included in the total control deviation, which results in the following dependency for the calculation of the total control deviation: $$\begin{array}{l}\mathrm{In\; total}-\mathrm{Control\; deviation}\mathrm{Overheating}=\\ \left(\mathrm{Weighting\; influence}*\mathrm{Control\; deviation}\mathrm{Compressor\; inlet\; superheat}\right)+\\ \left(\left(1-\mathrm{Weighting\; influence}\right)*\mathrm{Control\; deviation}\mathrm{Evaporator\; outlet\; superheat}\right)\end{array}$$

• Beim Betriebsartübergang zwischen Betriebsart = Betrieb mit ausgeschaltetem Verdichter 210 und Betriebsart = Betrieb mit eingeschaltetem Verdichter 210 im Heizbetrieb wird aufgrund der dynamischen Prozesswerteänderungen 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 dTüε aufgrund der gegenüber der Verdampferaustrittsüberhitzung dT_ÜA grö-ßeren 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:

• During the 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, it is advantageous to initially only include the control deviation evaporator outlet superheat dT _ÜA in the total control deviation, in particular the value of a weighting influence is then initially = 0 or a value advantageously below 20%.
• After a stabilization phase of the vapor compression system 200, it is advantageous not to switch spontaneously to the value of the weighting influence designed for control operation, but to make the transition in a ramp-like manner. In this case, it is advantageous for the value of the weighting influence to be increased from the starting value = 0, or a value in particular below 20%, to the intended target value in a ramp-like manner. This in particular avoids value discontinuity in the event of a spontaneous switchover and thus avoids control oscillations.
• The target value of the weighting influence is advantageously adapted to the respective operating mode and the operating point. Operating points which are characterized by an increased tendency to oscillate advantageously require a lower weighting of the control deviation of the compressor inlet superheat dT _ÜE , in particular this avoids a tendency to oscillate due to the control-technically critical signal behavior of the compressor inlet superheat dTüε due to the greater signal delay and greater path steepness compared to the evaporator outlet superheat dT _ÜA .

Step 4: In a fourth method step, the calculated total control deviation of the superheat is then processed in the controller 500, which controls the corresponding actuators of the refrigeration circuit, in particular the expansion valve 230 with the adjustable opening degree and/or the compressor 210 with adjustable compressor speed, such that in the regulated case a control deviation of the superheat is set equal to approximately 0 Kelvin.

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.

Accordingly, in refrigeration circuits with an internal heat exchanger, in this example the recuperator 250, between the refrigerant path for subcooled refrigerant after exiting the heat sink-side heat exchanger, here the condenser 220, and the refrigerant path after exiting the heat source-side heat exchanger, here the evaporator 210, and the compressor inlet 211, the compressor inlet superheat T _ÜE as well as the evaporator outlet superheat T _ÜA can be included proportionally to regulate the superheating of the refrigerant.

In this idealized case, the total control deviation combined from the individual control deviations would also be zero, but due to component or control tolerances, this case will hardly occur.

Usually, a steady-state control operation will occur in which the individual control deviations are not equal to zero, but the linked total control deviation is zero. In this case, the actual superheat at the compressor inlet is not the target superheat at the compressor inlet, so the refrigeration circuit is not operating at the optimal operating point.

If, at the optimum operating point for the compressor inlet superheat (actual value of compressor inlet superheat equals setpoint value of compressor inlet superheat), the setpoint value for the evaporator outlet superheat were adapted so that it was equal to the current actual value of the evaporator outlet superheat, the control would settle without any control deviation with regard to the respective superheat values and would be operated at the optimum operating point.

According to the invention, this is implemented by correcting the setpoint value for the evaporator outlet superheat. The control deviation of the compressor inlet superheat, i.e. the difference between the setpoint superheat of the compressor inlet and the actual superheat of the compressor inlet, is used as the input variable for the correction. Using an adaptation time constant as a parameter, a compensation variable for the temperature difference adaptation is calculated by using a time function to make the value of the adaptation time constant take on the inverse value of the control deviation of the compressor inlet superheat. In other words, the adaptation time constant, together with the time function, determines the duration within which the temperature difference adaptation follows the control deviation. Other forms of filtering rapid changes in the control deviation, such as low-pass filters, are also conceivable.

The need for adaptation or correction arises, for example, from tolerances in the wet vapor characteristic curve of the refrigerant.

Fig.3 shows schematically and as an example three wet steam characteristic curves 1010, 1020 and 1030 for different mixing ratios of refrigerant components, in this example of R32 and R1234yf. The wet steam characteristic curve 1020 corresponds in this example to the wet steam characteristic curve of R454C. For the wet steam characteristic curve 1010, the proportion of R32 is reduced compared to R454C, while it is increased for the wet steam characteristic curve 1030. The three wet steam characteristic curves 1010, 1020 and 1030 are therefore typical mixture values obtained from filling a compression refrigeration system from a container with R454C.

It can be seen that the dew point temperature as a function of the pressure for lower proportions of R32, see wet steam characteristic curve 1010, is higher than the dew point temperature for higher proportions of R32, see wet steam characteristic curve 1020, 1030. According to the invention, these tolerances of the wet steam characteristic curves can be compensated by correcting and including the correction value.

The temperature difference adaptation can be calculated in two consecutive steps. In a first step, an unlimited new value of the temperature difference adaptation evaporator output superheat setpoint (unlimited) is calculated using the value of the temperature difference correction evaporator output superheat setpoint calculated in the last loop run. In the second step, this value is limited to the parameterizable range limit and then processed further as a newly calculated process value. The temperature difference adaptation evaporator output superheat setpoint is preferably adapted in such a way that it would assume the value of the control deviation of the superheat at the compressor inlet in a specified time, called the adaptation time constant evaporator output superheat setpoint.

Subsequently, the range of the temperature difference correction evaporator outlet superheat setpoint is limited to the range set with an adaptation range parameter by limiting the previously calculated value to the range +/- adaptation range parameter (in Kelvin).

Fig.4 shows schematically and exemplarily a curve 2010 of the control deviation of the compressor inlet superheat and a curve 2020 of the correction of the evaporator outlet superheat setpoint resulting from the control deviation over time in minutes on the horizontal axis.

In the first four minutes, for example, there is no control deviation of the compressor inlet superheat, so that no correction or adaptation occurs. In a time range of 2030 there is a control deviation of 0.5 K. As long as the control deviation remains at + 0.5 K, the adaptation or correction increases in the opposite direction. At a At time 2040, the correction reaches an exemplary maximum value shown at -1 K, to which it remains limited, although the control deviation exists.

In a time range 2050, there is now a control deviation that has deviated in the opposite direction. The value of the adaptation increases gradually to counteract the control deviation. The slope of the adaptation 2020 in the time range 2050 is greater than the slope in the time range 2030, since the slope is preferably selected to be proportional to the value of the control deviation. After a decrease in a time range 2060, the correction remains constant in a time range 2070, since the control deviation in this time range is zero.

In the case of multi-component refrigerants with "temperature glide", for example R454C, overheating in the negative range relative to the dew point temperature is also possible at operating points where not all refrigerant components have completely evaporated.

Claims (8)

1. Method for controlling a compression refrigeration system (200), in particular a heat pump (100), in which

   - a gaseous refrigerant is compressed from a low pressure (ND) to a high pressure (HD) by means of a compressor (210) controlled by a controller (500),

   - the refrigerant is driven through a condenser (220), in which it transfers a heating energy (Q_h) to a heating medium located in a heat sink system (400),

   - an internal heat (Q_i) of the refrigerant is exchanged in a recuperator (250) while the refrigerant flows under high pressure (HD) from the condenser (220) to the expansion valve (230) with the refrigerant flowing under low pressure (ND) from the evaporator (240) to the compressor (210),

   - the refrigerant is guided in a high pressure flow direction (S_HD) to an expansion valve (230) controlled by the controller (500), in which the refrigerant is expanded from the high pressure (HD) to the low pressure (ND) in a controlled manner depending on a control value (R), wherein

   - the refrigerant at the low pressure ND evaporates in the evaporator (240) upon absorption of source heat Q_Q,

   including the method steps:

   - determining a target value (Z_ÜA) of the evaporator discharge superheating (T_ÜA) and a target value (Z_ÜE) of the compressor intake superheating (T_ÜE),

   - calculating a correction value based on a control deviation of the compressor intake superheating (T_ÜE) from the target value (Z_ÜE) of the compressor intake superheating (T_ÜE),

   - correcting the target value (Z_ÜA) of the evaporator discharge superheating (T_ÜA) with the calculated correction value,

   - calculating a control value (R) after a commissioning phase of the compression refrigeration system (200) depending on the target value (Z_ÜA) of the evaporator discharge superheating (T_ÜA) and the target value (Z_ÜE) of the compressor intake superheating (T_ÜE), and

   - controlling the expansion valve (230) based on the control value (R).
2. Method according to claim 1, wherein the correction value is calculated in proportion to a time integral of the control deviation of the compressor intake superheating (T_ÜE) to the target value (Z_ÜE) of the compressor intake superheating (T_ÜE).
3. Method according to claim 2, wherein the correction value is corrected in discrete time steps by a proportional component of the control deviation of the compressor intake superheating (T_ÜE).
4. Method according to any one of the preceding claims, wherein the correction value is limited to a permissible value range.
5. Method according to one of the preceding claims, wherein the target value (Z_ÜA) of the evaporator discharge superheating (T_ÜA) is corrected by adding the calculated correction value.
6. Compression refrigeration system (200), in particular heat pump (100), with

   - a refrigerant circuit with refrigerant,

   - a compressor (210),

   - a condenser (220),

   - an evaporator (240),

   - an expansion valve (230),

   - an internal heat exchanger (250), referred to as a recuperator, and

   - a controller (500),

   wherein during operation of the compression refrigeration system (200)

   - gaseous refrigerant is compressed from a low pressure (ND) to a high pressure (HD) by means of a compressor (210) controlled by the controller (500),

   - the refrigerant is driven through the condenser (220), wherein the condenser (220) is configured such that the refrigerant in the condenser (220) transfers heating energy (Q_h) to a heating medium located in a heat sink system (400),

   - an internal heat (Q_i) of the refrigerant is exchanged in a recuperator (250) while the refrigerant flows under high pressure (HD) from the condenser (220) to the expansion valve (230) with the refrigerant flowing under low pressure (ND) from the evaporator (240) to the compressor (210),

   - the refrigerant is guided in a high pressure flow direction (S_HD) to an expansion valve (230) controlled by the controller (500), in which the refrigerant is expanded from the high pressure (HD) to the low pressure (ND) in a controlled manner depending on a control value (R), wherein

   - the refrigerant at the low pressure ND evaporates in the evaporator (240) upon absorption of source heat Q_Q,

   wherein the controller (500) is configured for:

   - determining a target value (Z_ÜA) of the evaporator discharge superheating (T_ÜA) and a target value (Z_ÜE) of the compressor intake superheating (T_ÜE),

   - calculating a correction value based on a control deviation of the compressor intake superheating (T_ÜE) from the target value (Z_ÜE) of the compressor intake superheating (T_ÜE),

   - correcting the target value (Z_ÜA) of the evaporator discharge superheating (T_ÜA) with the calculated correction value,

   - calculating a control value (R) after a commissioning phase of the compression refrigeration system (200) depending on the target value (Z_ÜA) of the evaporator discharge superheating (T_ÜA) and the target value (Z_ÜE) of the compressor intake superheating (T_ÜE), and

   - controlling the expansion valve (230) based on the control value (R).
7. Compression refrigeration system (200) according to claim 6, wherein the refrigerant comprises a mixture of several refrigerant components.
8. Compression refrigeration system (200) according to claim 7, wherein a mixture containing R32 and R1234yf, in particular R454C, is used as the refrigerant.

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