DE102020115274A1 - Method for operating a compression refrigeration system - Google Patents

Abstract

The present invention relates to a method for regulating a compression refrigeration system (200), in particular a heat pump (100), as well as an associated compression refrigeration system, containing the process steps: determining a target value (ZÜA) of the evaporator outlet superheating (TÜA) and a target value (ZÜE) for the compressor inlet superheating ( TÜE), calculating a correction value based on a control deviation of the compressor inlet superheat (TÜE) to the target value (ZÜE) of the compressor inlet superheat (TÜE), correcting the target value (ZÜA) of the evaporator outlet superheat (TÜA) with the calculated correction value (R), calculating a control value (R) a commissioning phase of the compression refrigeration system (200) depending on the target value (ZÜA) of the evaporator outlet superheat (TÜA) and the target value (ZÜE) of the compressor inlet superheat (TÜE), and regulation of the expansion valve (230) based on the control value (R).

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 element 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, 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.

The regulation of the compression refrigeration system has to meet various requirements, for example it is required that a coefficient of performance is as high as possible in order to allow the most energy-efficient operation possible. But it is also crucial that the operating limits of the components are observed.

In this context, particular emphasis should be placed on the compressor, which, as stated, compresses the gaseous refrigerant from low pressure (LP) to high pressure (HP). It is known to provide compressors to optimize performance, which allow operation at different speeds. 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 the high pressure (HP) must be adhered to, depending on the speed. If the limits are exceeded either on the low pressure side or on the high pressure side, the operation of the compressor must be switched off in the worst case.

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

Failure to comply with the minimum overheating requirement can result in increased compressor wear due to insufficient lubrication of the compression components or damage due to cavitation due to re-evaporation in the compressor.

In the case of refrigeration circuits with an internal heat exchanger, for example in the form of a recuperator, it is known to use an evaporator outlet overheating as an auxiliary control variable for regulating the compressor inlet overheating. In a comparison of the two superheating values, the evaporator outlet overheating reacts much faster and thus enables the oscillations of the refrigeration circuit to be kept as low as possible.

The dew point temperature of the refrigerant is pressure-dependent and can be tabulated or calculated in the form of characteristics, called wet steam characteristics. To regulate the overheating of the evaporator outlet, a low pressure in the evaporator can therefore be regulated, for example, by adjusting the degree of opening of a throttle element.

However, it has been found here that, for example, there may be tolerances in the wet steam characteristics, i.e. the course of the pressure-dependent dew point temperature, which is used to regulate the overheating of the evaporator outlet based on a control of the degree of opening of a pressure reduction unit such as an expansion valve, which make reliable regulation of the overheating of the compressor inlet more 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 beginning, with which the control of the compressor inlet overheating is improved.

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

In a commissioning phase of the vapor compression system, a control value is influenced as a function of a control deviation of an evaporator outlet overheating, with which the expansion valve is controlled. The control value is still determined after the commissioning phase, during a retracted operating state of the vapor compression system, depending on compressor inlet overheating and the expansion valve is controlled after the commissioning phase depending on the determined evaporator outlet overheating and the compressor inlet overheating.

A controlled variable for the superheating of the evaporator outlet is calculated. With a target value for evaporator outlet superheating, a control deviation for evaporator outlet superheating is calculated. A controlled variable for compressor inlet superheating is calculated. With a target value for compressor inlet overheating, a control deviation for compressor inlet overheating is calculated. The control value R is calculated from a weighted influence of the control deviation evaporator outlet overheating and a weighted influence of the control deviation compressor inlet overheating. The expansion valve is regulated with the control value.

Accordingly, in refrigeration circuits with an internal heat exchanger between the refrigerant path for supercooled 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 superheating and the evaporator outlet can be proportionally included to regulate the overheating of the refrigerant.

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

Therefore, the control deviation of the evaporator outlet overheating is proportionally more included in the control than the control deviation of the compressor inlet overheating in order to optimize the overall reaction time of the control.

To achieve the highest possible refrigeration cycle efficiency, one goal is to regulate the lowest possible overheating at the evaporator outlet and to shift the overheating of the refrigerant as completely as possible into the internal heat exchanger.

This means that the setpoint for evaporator output overheating is always set in the vicinity of 0 Kelvin in relation to the dew line of the refrigerant.

However, if the evaporator outlet overheating is included, it is necessary to adapt the respective setpoint for evaporator outlet overheating to be regulated, based on the model, as far as possible to the respective setpoint for compressor inlet overheating to be regulated.

However, deviations in the measured variable of the actual value for evaporator output overheating result from

• • Tolerances when recording the evaporator outlet temperature,
• • Tolerances when recording 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 location at the evaporator outlet,
• • Mixing ratio deviations in the composition of the refrigerant for refrigerants with a temperature glide.

According to the invention, a method is now proposed for adaptive compensation of the tolerances, which are individual, but mostly systematically, for each heat pump. What is meant is that the tolerances are caused by component tolerances, for example, which 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 setpoints for evaporator outlet overheating and setpoint for evaporator outlet overheating in such a way that when the overheating control is in a steady state, both the control deviation for the evaporator output overheating and the control deviation for the compressor inlet overheating are equally zero.

Accordingly, it is proposed in one aspect of the invention that the method for regulating the compression refrigeration system contains the following steps: determining a target value for the evaporator outlet superheating and a target value for the compressor inlet superheating, calculating a correction value based on a control deviation of the compressor inlet superheating from the target value for the compressor inlet superheating, correcting the target value the evaporator outlet superheat with the calculated correction value, calculation of 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 regulation of the expansion valve based on the control value.

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

Tolerances of the wet steam characteristic arise, for example, regularly as a result of the segregation of refrigerant components when the compression refrigeration system is filled in a case in which 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 force of gravity during the filling process already results in segregation, which leads to a deviation of 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 tank 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 already have significant effects on the wet steam characteristic curve in the case of an overheating control, which are corrected with the method according to the invention.

The correction value is preferably calculated proportionally to a time integral of the control deviation of the compressor inlet overheating from the target value of the compressor inlet overheating.

The correction value is preferably corrected in discrete time steps by a proportional portion of the control deviation of the compressor inlet overheating.

The correction value is preferably limited to a permissible range of values.

The target value for the superheating of the evaporator outlet is preferably corrected by adding the calculated correction value.

Another advantageous method is to perform the addition for correcting the target value of the evaporator outlet overheating "inversely", with the actual value being subtracted to form the control deviations from the target value. The control deviation is advantageously calculated by calculating or creating a difference between an “actual value” and a “setpoint”.

According to an advantageous method step, the superheating of the evaporator outlet is weighted with a parameter in relation to the superheating of the compressor in order to determine the control value. The control deviations of an evaporator outlet overheating or the compressor inlet overheating from a setpoint value are particularly advantageously weighted.

In another advantageous method step, the evaporator outlet overheating is weighted against a weighting of the compressor inlet overheating in order to determine the controlled variable by means of a parameter. The control deviations of an evaporator outlet overheating or a compressor inlet overheating are particularly advantageously weighted with a setpoint value.

In a further advantageous embodiment of the method, the evaporator outlet overheating is determined from a measured evaporator outlet temperature value, which is measured with an evaporator outlet temperature sensor, and from a low pressure, which is measured with a low pressure sensor.

In a further preferred method step, the compressor inlet overheating is determined from a compressor inlet temperature, measured with a compressor inlet temperature sensor, and a low pressure, which is measured with a low pressure sensor.

In an advantageous process step, a first control deviation of the compressor inlet superheat is calculated in the controller with a second control deviation of the evaporator outlet superheat to a total control deviation and the total control deviation is used to adjust the expansion valve.

The method is advantageously carried out in such a way that the parameter is gradually changed, preferably after the commissioning phase, in particular a weighting of the compressor inlet overheating begins with an influence of less than 20% compared to the evaporator outlet overheating and the weighting then changes, in particular increases, taken into account when determining the control value is, in particular up to a predetermined target weighting of the parameter. During the commissioning phase, the weighting coefficient of the compressor inlet overheating is particularly advantageously set to a value between 0% and 20%, which is small compared to the operating phase, and after the commissioning phase it rises in a ramp over time to a target value for the operating phase.

In a time-limited commissioning phase after the compressor has been switched on, the parameter value is advantageously controlled in such a way that the weighting of the compressor inlet overheating for determining the controlled variable has an influence of 0% or close to 0%, in particular below 20%. This means that the evaporator outlet overheating is included exclusively or primarily in the controlled variable. In a subsequent transition phase, the parameter P is controlled in such a way that the weighting of the controlled variable of the compressor inlet overheating to determine the controlled variable is increased up to a target value and in the subsequent operating phase the weighting of the compressor inlet overheating reaches the target value.

Furthermore, it is advantageous to vary the target value for the weighting of the controlled variable as a function of compressor inlet overheating in operating phases in the steady state of the vapor compression system.

The refrigerant preferably has a temperature glide, the refrigerant in particular having or consisting of R454C, and where the compression refrigeration system contains in particular an internal heat exchanger for transferring thermal energy of the refrigerant to the refrigerant before it enters the throttle element before it enters the compressor. This is particularly relevant because high-performance internal heat exchangers are used in refrigeration circuits with R454C.

The object is also 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 the installation.

• 1 Wärmepumpe 100 mit einem Dampfkompressionskreislauf200
• 2 log p / h - Diagramm des Dampfkompressionsprozesses mit Rekuperator 250
• 3 schematisch und exemplarisch Nassdampfkennlinien verschiedener Kältemittelmischungen und
• 4 schematisch und exemplarisch eine Adaption eines Verdampferausgangsüberhitzungssollwertes

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

• 1 Heat pump 100 with a vapor compression cycle200
• 2 log p / h - diagram of the vapor compression process with recuperator 250
• 3 schematic and exemplary wet steam characteristics of various refrigerant mixtures and
• 4th Schematic and exemplary an adaptation of an evaporator output superheating 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.

1 shows schematically and as an example a heat pump 100 . The heat pump 100 consists essentially of a vapor compression system forming a compression refrigeration system 200 , which contains the following components:

• • A compressor 210 to compress the superheated refrigerant,
• • a condenser 220 , with a refrigerant-side condenser inlet 221 and a condenser outlet 222 for the transfer of 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 quantities depending on the operating conditions,
• • a throttle device designed as an expansion valve 230 to expand the refrigerant,
• • an evaporator 240 , with an evaporator inlet 241 , for the transmission of source energy Q _Q from a heat source system 300 , with a heat source inlet 320 and a heat source outlet 310 , being the heat source system 300 in particular can be a brine system, which heat energy Q _Q from the ground or an air system that absorbs heat energy Q _Q from the ambient air and to the vapor compression system 200 emits or any other heat source,
• • a recuperator as an example of an optional internal heat exchanger 250 which is intended to be internal heat energy Q _i between that of the condenser 220 to the expansion valve 230 flowing refrigerant to that from the evaporator 240 to the compressor 210 transferring flowing refrigerant and
• • a refrigerant, in particular a refrigerant mixture of at least two substances or two refrigerants which in one flow direction S _HD and S _ND through the vapor compression circuit 200 flows, being in the vapor compression cycle 200 Refrigerant vapor through the compressor 210 is brought to a high pressure HD and to a condenser 220 is guided, 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 to the compressor 210 is a low pressure path with a low pressure flow direction S _ND of the refrigerant formed in which the evaporator 240 located.

The actuators listed below are advantageously at least partially connected to the controller via a data connection 510 that can be done 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, an in the 1 evaporator inlet temperature sensor, not shown, the temperature at the evaporator inlet 241 determine.

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

The compressor 210 serves to compress the superheated refrigerant from an inlet connection 211 to a compressor discharge pressure P _Va at a compressor outlet temperature corresponding to 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 rolling piston unit, scrolling unit or otherwise. At the compressor outlet 212 is the compressed, superheated refrigerant at the compressor discharge 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 LP, at a compressed inlet temperature T _VE what the state of the refrigerant temperature at the inlet connection 211 describes when entering a compression chamber.

In the condenser 220 the transfer of thermal energy takes place Q _H from the refrigerant of the vapor compression system 200 to a heating medium of the heat sink system 400 . First takes place in the liquefied 220 the desuperheating of the refrigerant takes place, whereby superheated refrigerant vapor, by reducing the temperature, transfers part of its thermal energy to the heating medium of the heat sink system 400 transmits.

After the refrigerant vapor has been de-heated, it is advantageous to do so in the condenser 220 another heat transfer Q _H by condensation of the refrigerant during the phase transition from the gas phase of the refrigerant to the liquid phase of the refrigerant. This will add more heat Q _H from the refrigerant from the vapor compression system 200 to the heating medium of the heat sink system 400 transfer.

Which is in the liquefier 220 The setting high pressure HD of the refrigerant corresponds to the operation of the compressor 210 approximately with a condensation pressure of the refrigerant at a heating medium temperature Tws in the heat sink system.

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

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

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

First, the refrigerant flows through an expansion valve inlet 231 into the expansion valve. In the expansion valve 230 the refrigerant pressure is throttled from high pressure HP to low pressure LP by the refrigerant advantageously passing through a nozzle arrangement or throttle with an advantageously variable opening cross section, the low pressure advantageously being approximately the suction pressure of the compressor 210 is equivalent to. 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 stepper motor, which is controlled by the control unit or regulation 500 is controlled. The low pressure is LP at the expansion valve outlet 232 of the refrigerant from the expansion valve 230 controlled in such a way that the low pressure ND of the refrigerant that is established when the compressor is in operation 210 corresponds _{roughly 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 Q _V from the heat source fluid of the heat source system 300 , which is a brine system, a geothermal system for the use of thermal energy Q _Q from the ground, an air system for the use of energy Q _Q from the ambient air or another heat source that is the source energy Q _Q to the vapor compression system 200 gives away.

That in the vaporizer 240 Incoming refrigerant is reduced when flowing through the evaporator 240 through heat absorption Q _Q its wet steam content and leaves the evaporator 240 advantageously with a low proportion of wet steam or advantageously also as a superheated gaseous refrigerant. The heat source medium is supplied 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 through the heat source media path of the evaporator 240 promoted, with the heat source medium the heat energy as it flows through the evaporator Q _Q is withdrawn.

In the recuperator 250 becomes thermal energy Q _i between that of the condenser 220 to the expansion valve 230 flowing refrigerant to that from the evaporator 240 to the compressor 210 Transferring flowing refrigerant, taking that from the evaporator 240 to the compressor 210 flowing refrigerant in particular continues to overheat.

This superheated refrigerant which has an overheating temperature T _Ke from the recuperator 250 becomes the refrigerant inlet connection 211 of the compressor 210 directed.

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

For this purpose, the refrigerant in the condenser 220 Thermal energy Q _H at a heat sink side temperature level to the heating medium, in the high pressure path of the recuperator 250 further thermal energy through subcooling Q _i withdrawn.

The internal energy state of the refrigerant when it enters the evaporator 240 is through this deprivation of heat Q _i reduced so that the refrigerant has more heat energy at the same evaporation temperature level Q _Q from the heat source 300 can accommodate.

Then the refrigerant, after the evaporator outlet 242 from the vaporizer 240 , in the low pressure path at low pressure LP and at a low pressure temperature corresponding to an evaporator outlet temperature T _Va in the recuperator 250 the thermal energy extracted in the high pressure path Q _i fed back. The supply of energy has the advantageous effect of reducing the proportion of wet steam to a state without a proportion of wet steam and then overheating occurs through further supply of energy.

Furthermore, to record the operating status of the vapor compression system 200 advantageously arranged the following sensors, with which in particular to safeguard and optimize the operating conditions of the vapor compression system 200 A model-based pre-control is implemented, especially in the event of changes in the operating status.

• • 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_Ie 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, with the aid of the process values recorded by sensors, safeguards are advantageously made with regard to permissible working areas of the components, such as the compressor in particular 210 On the other hand, based on the sensor data, model-based precontrols, in particular a speed of the compressor, take place 210 and / or a valve opening degree of the expansion valve, so that the controller only needs to make minor corrections to compensate for a control deviation that is nevertheless smaller due to the precontrol:

• • A high pressure sensor 503 advantageous for recording 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 indoor temperature sensor 506 advantageous for detecting the internal temperature T _{Ie of} the refrigerant between the internal recuperator outlet on the high-pressure side 252 of the refrigerant from the recuperator 250 and the expansion valve entry 231 . The internal temperature is advantageously also known as the "recuperator outlet temperature high-pressure path" and
• • Advantageously, a recuperator internal temperature sensor 505 . The internal recuperator 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 from the recuperator internal temperature sensor is advantageous 505 measured.

• • 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 recording the low pressure LP of the refrigerant at the compressor inlet 211 , or between the expansion valve 230 and the compressor inlet 211 ,
• • an evaporator outlet temperature sensor 508 for recording 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 advantageously serves 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 recuperator outlet on the low pressure side 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 cycle 200 as the quotient between that of the vapor compression circuit 200 transferred heating power Q _H to one from the compressor 210 The electrical power P _e consumed is the overheating of the refrigerant at the compressor inlet 211 . In order to comply with permissible compressor operating conditions, however, it is advantageous to comply with restrictions with regard to the permissible overheating 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 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 superheating of the compressor inlet is preferably regulated in such a way that no condensate 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 compressor inlet 211 fails. The refrigeration circuit section between the evaporator outlet 242 and recuperator inlet 251 is usually colder because this is typically only a short pipe section, there is better insulation compared to the section between the refrigerant outlet of the recuperator 252 and compressor inlet 211 possible. For example, sits at the point of the compressor inlet 211 the refrigerant separator on the compressor 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. Also the passage between the recuperator outlet on the high pressure side 252 and entry into the expansion valve 231 cools regularly depending on the operating point with ideal heat transfer conditions in the recuperator 250 on the temperature level of the refrigerant at the evaporator outlet 242 away. 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 at about -10 ° C, at the brine outlet 310 is around -13 ° C and 5 ° C at the compressor inlet, this means overheating 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.

The numerical example, which is of course non-restrictive, is now used to achieve overheating of 15K 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, i.e. the condensation temperature plus buffer, the overheating must be 9K greater, i.e. an overheating of 24K must be maintained.

Limit values, especially for overheating, determine the permissible overheating range of the components at the compressor inlet depending on the operating point 211 fixed. Furthermore, there are also dependencies between the compressor inlet superheating _{dT U ̈ E} and the overall efficiency of the Vapor compression circuit 200 or between the compressor inlet overheating _{dT U ̈ E} and a stability S of a control value R, which is advantageous when regulating the compressor inlet overheating.

In order to take into account all these requirements, it is advantageous to depend 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 are used to calculate the compressor inlet superheat _{dT U ̈ E.} Alternatively or additionally, a calculation of the target value Z as a default value for the compressor inlet superheating _{dT U ̈ E} can be carried out from the refrigerant circuit parameters 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 U ̈ E 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_Ü_E = 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 U ̈ E} and a control deviation of the evaporator _{outlet superheat dT ÜA} are combined with each other weighted, resulting in the controller 500 a total control deviation is calculated, which is used to control the vapor compression circuit 200 is fed in. The system deviations from the compressor inlet superheat _{dT U ̈ E} and the 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 U ̈ E} = 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

The weighted influence of the control deviation of the compressor inlet superheat _{dT U ̈ E} and the weighted influence of the control deviation of the evaporator _{outlet superheat dT ÜA} in the controller are then advantageous 500 the total control deviation is calculated, which is used to regulate the vapor compression circuit 200 is fed in.

With the vapor compression circuit 200 the refrigerant passes through the expansion valve after the expansion 230 two sequentially arranged heat exchangers, the evaporator 240 and the recuperator 250 in which the refrigerant heat energy Q _Q and Q _i is fed.

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

In the high pressure refrigerant flow direction S _HD subsequent recuperator 250 heat energy is added to the refrigerant Q _i after leaving the condenser 220 withdrawn. 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 control of compressor inlet superheating _{dT U ̈ E.}

The control value R is advantageously the weighted link between the control deviation of the compressor inlet superheat _{dT U ̈ E} and 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 overheating dT ÜE, are in the relevant vapor compression circuit 200 the compressor speed and / or the degree of opening of the 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 are especially the compressor 210 by varying the compressor speed and the expansion valve 230 by influencing the Degree of opening 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 output without further compensatory changes in the degree of opening of the expansion valve 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.

It is advantageous in the steam compression cycle 200 the compressor speed is set so that that of the vapor compression circuit 200 heat output transferred to the heating medium QH corresponds to the requested target value Z. To comply with this requirement, influencing the compressor _{speed to regulate the superheating dT U ̈ E} at the compressor inlet is advantageously subordinate or not appropriate.

• Das Expansionsventil 230 agiert als Düse mit elektromotorisch verstellbarem Düsenquerschnitt, bei welchem üblicherweise mittels eines Schrittmotor eine nadelförmige Düsennadel per Gewinde in einen Düsensitz gefahren wird.

The degree of opening of the expansion valve is advantageous 230 used as the control value for the regulation of the compressor inlet superheating _{dT U ̈ E.} The influence of the degree of opening of the expansion valve 230 on the compressor inlet overheating _{dT U ̈ E} takes place as follows:

• The expansion valve 230 acts as a nozzle with an electrically adjustable nozzle cross-section, in which a needle-shaped nozzle needle is usually threaded into a nozzle seat by means of a stepper motor.

The refrigerant throughput through the expansion valve is when operating with liquid refrigerant at the expansion valve inlet 231 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 when entering the expansion valve 230 can be assumed to be constant, affects the degree of opening of the expansion valve 230 relevant only the low pressure LP, i.e. the outlet pressure from the expansion valve 230 .

Becomes the opening degree of the expansion valve 230 reduced, less refrigerant passes through the expansion valve at constant high pressure HP and initially still constant low pressure LP 230 . Because the compressor 210 but continues to initially deliver the same refrigerant mass flow, it is in the high-pressure flow direction S _HD through the expansion valve 230 less refrigerant supplied than from the compressor 210 is sucked off.

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

Becomes the opening degree of the expansion valve 230 increased, more refrigerant passes through the expansion valve at constant high pressure HP and initially still constant low pressure LP 230 . Because the compressor 210 but continues to deliver the same refrigerant mass flow initially, the low-pressure side LP of the refrigeration circuit is passed through the expansion valve 230 more refrigerant supplied than from the compressor 210 is sucked off. Since refrigerant vapor is a compressible medium, the low pressure LP increases on the low pressure side of the vapor compression circuit 200 . With increasing low pressure LP, the mass flow rate of the compressor increases 210 roughly proportional, since its delivery rate can be roughly described as volume / time, and a correspondingly increased low pressure LP is set at which the pressure caused by the expansion valve 230 supplied refrigerant mass flow equal to that from the compressor 210 discharged refrigerant mass flow is.

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

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

To get a sufficient amount of thermal energy Q _Q from the heat source system 300 To be able to transfer 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 on the respective surface element.

Is the physical state of the refrigerant as it flows through the evaporator 240 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 controlling the low pressure LP or also an evaporation pressure, it is possible to indirectly control the evaporation temperature of the refrigerant as it flows through the recuperator 250 steer.

The thermal energy Q _Q , which from the heat source system to the the evaporator 240 The refrigerant flowing through is transferred, affects the physical 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 Stoffkonstante als 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,
• • Transferred heat output Q _Q , and from one
• • Enthalpy difference in the wet steam area at the respective low pressure LP, which the refrigerant exhibits as a material constant as a function of pressure.

Additional energy is supplied to the recuperator for complete evaporation 250 to superheat the refrigerant beyond the saturated vapor state.

With the method, the vapor compression cycle is created under given operating conditions 200 Depending on the manipulated variable "degree of opening expansion valve 230", a corresponding refrigerant condition when exiting the evaporator 240 set.

In the steady state, the "isolated" controlled system "evaporator." 240 “A controlled system behavior with moderate steepness. 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 the refrigerant, which has a “temperature glide”, in particular R454C is advantageously used. In the case of a refrigerant mixture with a temperature glide, it is advantageous if the relative degree of opening of the actuator expansion valve changes by 1% rel. A change in superheating of approximately 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.

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 is ultimately the process variable refrigerant overheating at the evaporator outlet 242 influence:

• 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, for example, be the controller cycle time in which a travel step of the expansion valve 230 is controlled, include.
• • A second time constant acts through the controller 500 predetermined advantageously to a delayed setting of a corresponding low pressure in the case of changes in the degree of opening of the expansion valve 230 due to the compressibility of the refrigerant vapor at low pressure LP in the low pressure path.
• A third time constant is advantageously a thermal time constant of the heat transfer layer of the evaporator 240 , whereby a change in the evaporation pressure and thus the evaporation temperature is a delayed change in temperature 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 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 with a finite flow velocity.

Thus, after changing the manipulated variable “degree of opening of expansion valve 230”, there is advantageously a delay in the corresponding change in refrigerant state when it exits the evaporator outlet 242 one and one total time constant Ztot, depending on the operating point, is advantageously in the range from 30 seconds to about 5 minutes.

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

The physical state of the refrigerant as it flows into the recuperator 250 is in a normal operating case, that is, advantageously either saturated steam with a low steam content between 0 to 20% or, in particular, advantageously also 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 in the high-pressure path of the recuperator 250 , i.e. the temperature of the refrigerant after it leaves the condenser 220 .

To get 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 To be able to transmit, it must be ensured that the temperature of the refrigerant of the high-pressure side refrigerant path is at high pressure HD in as many surface elements as possible of the transfer layer of the recuperator 250 is greater than 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 than the corresponding temperatures of the heat source such as the ground or the outside air when heating.

The thermal energy Q _i , which is 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 is transmitted, affects the physical state of the refrigerant on the low-pressure side. The wet steam content of the recuperator 250 On the low-pressure side, the refrigerant flowing through 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,
• • 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, the vapor compression circuit 200 as well as, depending on the manipulated variable "degree of opening expansion valve 230", a corresponding refrigerant state at the outlet 252 from the recuperator 250 sets at low pressure LP.

In the steady state, the steepness of the controlled system results in the "isolated" controlled system at low pressure LP of the refrigerant in the low-pressure side path of the recuperator 250 a control system behavior with a high slope, with an approximately constant internal energy state of the refrigerant at the inlet 251 in the low pressure side LP path of the recuperator 250 . 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 evaporator 230 set of advantageously about 10 K or above 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 .

This is what happens in the vaporizer 240 a much higher heat transfer than in the recuperator 250 set as the environment by means of evaporator 240 a much greater amount of energy is to be withdrawn than it is only in the recuperator 250 to be transferred within the refrigeration circuit. The driving temperature difference in the recuperator is between 20 K and 60 K, for example, while it is only between 3 K 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, for example, is designed to be approx. 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 is advantageous 230 specified by the limitation of 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 opening degree of the expansion valve changes 230 due to the compressibility of the refrigerant vapor in the low pressure path LP.
• • One 13th . Time constant Z ₁₃ 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 .
• • One 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.
• • One 15th . Time constant Z ₁₅ results advantageously from the transport of the refrigerant through the evaporator 240 with a finite flow velocity and is taken into account.

The low-pressure side refrigerant path of the recuperator 250 is from the evaporator outlet 242 of the evaporator 240 fed. Here, too, the internal energy state of the refrigerant is _{delayed 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 "degree of opening expansion valve 230", there is a further delay in the corresponding change in the refrigerant state due to the time behavior of the recuperator 250 when exiting the low-pressure side refrigerant path of the recuperator 250 one.

The time behavior of the recuperator 250 can advantageously be taken into account as the total recuperator time constant Ztot depending on the respective working point of the steam compression circuit in the range between about 1 minute and 30 minutes.

There is advantageously a weighted combination of compressor inlet superheating _{dT U ̈ E} and that of evaporator _{outlet superheating dT ÜA} , in that the overall control deviation is calculated in the controller using a weighted combination of the control deviation of the compressor overheating and the control deviation of the evaporator outlet _{superheating dT ÜA} 500 to regulate the steam compression circuit 200 is fed in.

• Schritt 1: Zunächst werden die Prozessgrößen Verdichtereintrittsüberhitzung dT_Ü_E vorteilhaft als Hauptregelgröße und die Verdampferaustrittsüberhitzung dT_ÜA vorteilhaft als Hilfsgröße in einem ersten Verfahrensschritt messtechnisch erfasst.

The compressor inlet overheating _{dT U ̈ E} 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 overheating _{dT U ̈ E are} advantageously recorded as the main control _{variable and the evaporator outlet overheating dT ÜA} advantageously 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 in such a way that it detects a temperature corresponding to the refrigerant temperature in the wet steam area or
• • Determined indirectly by measurement, with a pressure sensor which detects the 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, at the respective overheating measuring point, in particular at the evaporator outlet 242 and / or at the compressor inlet 211 assigned temperatures of the refrigerant temperature by means of temperature sensors 501 , 508 recorded. 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 U ̈ E} and the evaporator _{outlet superheat dT ÜA} .

• Der Sollwert für die Verdichtereintrittsüberhitzung dT_Ü_E wird vorteilhaft zur Sicherstellung des zulässigen Verdichtersbetriebsbereiches und eines möglichst hohen Wirkungsgrades des Kältekreises im Bereich zwischen ca. 5 K bis 20 K variiert.

Step 2: The process variables compressor inlet superheating _{dT U ̈ E} and evaporator _{outlet superheating dT ÜA} are advantageously offset in a second step to create assigned control deviations with the respective assigned setpoints:

• The nominal value for the compressor inlet superheating _{dT U ̈ E} 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 _superheat dT ÜA at the evaporator outlet 242 is then varied as a function of the refrigeration circuit operating mode and the refrigeration circuit operating point so that the evaporator overheating in the steady normal case roughly corresponds to the process value of _{the evaporator outlet overheating dT ÜA.} This setpoint for the evaporator outlet _superheating dT ÜA can be model-based 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 superheating _{dT U ̈ E} at the evaporator outlet 242 and / or are precalculated from component properties and corrected adaptively.

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

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 process step, the control deviation of the compressor inlet superheating _{dT U ̈ E} and the control deviation of the evaporator _{outlet superheating dT ÜA are} advantageously combined to form an overall control deviation for 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 system deviations and, in extreme cases, can result in the exclusive inclusion of only one individual system deviation, but usually the weighted inclusion of both individual system 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 U ̈ E} in the total control deviation, which is used for the calculation of the total control deviation the following dependency results: $$\begin{array}{l}\text{total}-\text{Control deviation}\ddot{\text{U}}\text{overheating}=\\ \left(\text{Weighting influence}*\text{Control deviation at compressor inlet}\ddot{\text{u}}\text{overheating}\right)+\\ \left(\left(1-\text{Weighting influence}\right)*\text{Control deviation at compressor inlet}\ddot{\text{u}}\text{overheating}\right)\end{array}$$

• • 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 dT_Ü_E aufgrund der gegenüber der Verdampferaustrittsüberhitzung dT_ÜA größeren Signalverzögerung und größeren Streckensteilheit eine Schwingneigung vermieden.

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

• • When the operating mode is changed between operating mode = operation with the compressor switched off 210 and operating mode = operation with the compressor switched on 210 in heating mode, due to the dynamic process value changes when starting up the vapor compression system 200 Advantageously, initially only the control deviation evaporator outlet _{overheating dT ÜA is included} 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 over spontaneously to the value of the weighting influence designed for regular operation, but 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 are 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, this _{avoids a signal behavior of the compressor inlet overheating dT U ̈ E} that is critical in terms of control technology due to the higher degree of signal delay compared to the evaporator _{outlet overheating dT ÜA and a} greater signal delay dT ÜA.

Step 4: In a fourth process step, the calculated total control deviation of the overheating is then entered in the controller 500 processed, which 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, controls in such a way that in the regulated case a control deviation of the overheating is set as close as possible to about 0 Kelvin.

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.

Accordingly, in the case of cooling circuits with an internal heat exchanger, in this example the recuperator 250 , between the refrigerant filament for subcooled refrigerant after exiting the heat sink on the 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 , to regulate the overheating of the refrigerant proportionally the compressor inlet _{superheating T U ̈ E} and proportionally the evaporator _{outlet superheating T ÜA are} included.

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

Usually we find a steady control operation, in which, however, the individual control deviations are not equal to zero, but the linked total control deviation is equal to zero. In this case, the actual superheating of the compressor inlet is not the target superheating of the compressor inlet, so the refrigeration circuit is not operated at the optimal operating point.

If at the optimal working point for the compressor inlet superheat (actual value of compressor inlet superheat equal to setpoint for compressor inlet superheat) the setpoint for evaporator outlet superheat would be adapted in such a way that it corresponds to the current actual value of evaporator outlet superheat, the control would be operated at the optimal operating point without any deviation with respect to the respective superheat values.

According to the invention, this is implemented by the adaptation of the setpoint value for the evaporator outlet superheating that has been carried out. The control deviation for superheating compressor inlet, i.e. the difference between the target superheating at the compressor inlet and the actual superheating at the compressor inlet, is used as the input variable for the adaptation. With the help of an adaptation time constant as a parameter, a compensation variable temperature difference adaptation is calculated in that the value of the adaptation time constant assumes the inverse value of the control deviation overheating compressor inlet with the help of a time function. 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, for example low-pass filters, are also conceivable.

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

3 shows schematically and by way of example three wet steam characteristics 1010 , 1020 and 1030 for different mixing ratios of refrigerant components, in this example R32 and R1234yf. The wet steam characteristic 1020 in this example corresponds to the wet steam characteristic of R454C. For the wet steam characteristic 1010 the proportion of R32 is reduced compared to R454C, while it is for the wet steam characteristic 1030 is increased. With the three wet steam characteristics 1010 , 1020 and 1030 These are therefore typical mixture values that are obtained from a container filled with R454C when a compression refrigeration system is filled.

It can be seen that the dew point temperature depends on the pressure for lower proportions of R32, see wet steam characteristic 1010 , is higher than the dew point temperature for higher proportions of R32, see wet steam characteristic 1020 , 1030 . According to the invention, these tolerances of the wet steam characteristics can be compensated for by adapting and including the correction value.

The temperature difference adaptation can be calculated in two successive steps. In a first step, taking into account the value of the temperature difference adaption evaporator output overheating target value calculated in the last loop pass, an unlimited new value of temperature difference adaption evaporator output overheating target value (unlimited) is calculated, 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 of the 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 within a specified time, called the adaptation time constant for the evaporator output superheat setpoint.

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

4th shows schematically and by way of example a course 2010 the control deviation of the compressor inlet superheating and a curve 2020 the adaptation of the evaporator output superheat setpoint over the time in minutes on the horizontal axis as a result of the control deviation.

In the first four minutes, for example, there is no control deviation of the compressor inlet overheating, so that there is also no correction or adaptation. In a time range 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 time 2040 the correction reaches an exemplary maximum value shown at - 1 K, to which it remains limited even though the system deviation exists.

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

In the case of multi-fuel refrigerants with a “temperature glide”, for example R454C, overheating related to the dew point temperature is also possible in the negative range, at operating points at which not all refrigerant components have completely evaporated.

QUOTES INCLUDED IN THE DESCRIPTION

This list of the documents listed by the applicant was generated automatically and is included solely for the better information of the reader. The list is not part of the German patent or utility model application. The DPMA assumes no liability for any errors or omissions.

Patent literature cited

• DE 10159892 A1 [0005]
• DE 102005061480 B3 [0006]

Claims (8)

Method for regulating a compression refrigeration system (200), in particular a heat pump (100), in which - a gaseous refrigerant is compressed from a low pressure (LP) to a high pressure (HP) by a compressor (210) controlled by means of a regulator (500), - the refrigerant is driven through a condenser (220), in which it _{gives off a thermal heat (Q h} ) to a heating medium located in a heat sink system (400), - an internal heat (Q _i ) of the refrigerant in a recuperator (250) during refrigerant flowing under high pressure (HP) from the condenser (220) to the expansion valve (230) is exchanged with the refrigerant flowing under low pressure (LP) from the evaporator (240) to the compressor (210), - the refrigerant in a high pressure flow direction (S _HP ) to an expansion valve (230) controlled by the regulator (500), in which the refrigerant is relaxed from high pressure (HP) to low pressure (LP) in a controlled manner as a function of a control value (R), wherein - The refrigerant at the low pressure LP evaporates in the evaporator (240) when source heat Q _{Q is} absorbed, including the process steps: - Determination of a target value (Z _ÜA ) of the evaporator _{outlet superheating (T ÜA} ) and a target value (Z ÜE) of the compressor inlet superheating ( T _ÜE ), - Calculation of a correction value based on a control deviation of the compressor _{inlet superheat} (T ÜE) to the target value (Z _ÜE ) of the compressor _{inlet superheat} (T ÜE), - Correcting the target value (Z _ÜA ) of the evaporator _{outlet superheat (T ÜA} ) with the calculated correction value - Calculation of a control value (R) after a commissioning phase of the compression refrigeration system (200) depending on the target value (Z _ÜA ) of the evaporator _{outlet superheat (T ÜA} ) and the target value (Z ÜE) of the compressor _{inlet superheat} (T ÜE), and - Control of the expansion valve ( 230) based on the control value (R). Procedure according to Claim 1 , wherein the correction value is calculated proportionally to a time integral of the control deviation of the compressor _{inlet superheat} (T ÜE) to the target value (Z _ÜE ) of the compressor _{inlet superheat} (T ÜE). Procedure according to Claim 2 _{, whereby the correction value is} corrected in discrete time steps by a proportional part of the control deviation of the compressor inlet overheating (T ÜE). Method according to one of the preceding claims, wherein the correction value is limited to a permissible range of values. Method according to one of the preceding claims, wherein the target value (Z _ÜA ) of the evaporator _{outlet superheating} (T ÜA) is corrected by adding the calculated correction value. Compression refrigeration system (200), in particular heat pump (100), with - a refrigeration circuit with refrigerant, - a compressor (210), - a condenser (220), - an evaporator (240), - an expansion valve (230), - an internal heat exchanger (250), called recuperator, and - a controller (500), whereby during operation of the compression refrigeration system (200) - gaseous refrigerant from the compressor (210) controlled by means of the controller (500) from a low pressure (LP) to a high pressure (HP ) is compressed, - the refrigerant is driven through the condenser (220), in which it _{gives off a thermal heat (Q h} ) to a heating medium located in a heat sink system (400), - an internal heat (Q _i ) of the refrigerant in a recuperator (250) when the refrigerant flowing under high pressure (HP) from the condenser (220) to the expansion valve (230) is exchanged with the refrigerant flowing under low pressure (LP) from the evaporator (240) to the compressor (210), backflow direction (S _HD ) is led to an expansion valve (230) controlled by the controller (500), in which the refrigerant is relaxed from high pressure (HD) to low pressure (ND) in a regulated manner depending on a control value (R), whereby - the Refrigerant located at the low pressure LP evaporates in the evaporator (240) when source heat Q _{Q is} absorbed, the controller (500) being designed to: Determination of a target value (Z _ÜA ) of the evaporator _{outlet superheat (T ÜA} ) and a target value (Z ÜE) of the compressor _{inlet superheat} (T ÜE), - Calculation of a correction value based on a control deviation of the compressor _{inlet superheat} (T ÜE) to the target value (Z ÜE) of the compressor _{inlet superheat} (T ÜE), - Correcting the target value (Z _ÜA ) of the evaporator _{outlet 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 outlet superheating ( T _ÜA ) and the target value (Z _ÜE ) of the compressor _{inlet superheat} (T ÜE), and - control of the expansion valve (230) on the basis of the control value (R). Compression refrigeration system (200) Claim 6 , wherein the refrigerant contains a mixture of several refrigerant components. Compression refrigeration system (200) Claim 7 , whereby a mixture containing R32 and R1234yf, in particular R454C, is used as the refrigerant.

Images