METHOD FOR OPERATING A HEAT PUMP WITH A STEAM COMPRESSION SYSTEM

Description

[0001] The invention relates to a method for operating a heat pump with a vapor compression system in which a gaseous refrigerant is compressed from low pressure to high pressure by a compressor controlled by a controller. The refrigerant is forced through a condenser, where it releases heat to a heating medium located in a heat sink system. Internal heat is transferred in a recuperator from the refrigerant flowing under high pressure from the condenser to the expansion valve to 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 controller, where the refrigerant is expanded from high pressure to low pressure depending on a control value. The refrigerant at low pressure evaporates in the evaporator, absorbing heat from the source.

[0002] From DE 101 59 892 A1, it is known to use a recuperator in a refrigeration machine, particularly in a heat pump, to increase the heating capacity in a structurally simple manner at low outside temperatures. For this purpose, the recuperator is dimensioned such that at low evaporation temperatures it transfers at least approximately 15% of the heat pump's heating capacity from the liquid refrigerant to the gaseous refrigerant. An injection valve injects liquid refrigerant into the compressor so that the final compression temperature remains below 120 °C.

[0003] A heat pump system with a refrigerant circuit is known from DE 10 2005 061 480 B3. It is equipped with a compressor, a first heat exchanger, an throttling device, an evaporator, and a 4-2-way valve unit for switching between a first (heating) and a second (cooling) operating mode. The flow direction of the refrigerant in the refrigerant circuit can be reversed such that the first heat exchanger serves to condense the refrigerant in the first operating mode and to evaporate it in the second operating mode, and the second heat exchanger serves to evaporate the refrigerant in the first operating mode and to condense it in the second operating mode, wherein the first heat exchanger in the refrigerant circuit is configured to operate as a counterflow heat exchanger in both heating and cooling modes.

[0004] US 2008 / 223056 A1 discloses a method according to the preamble of claim 1 and relates to a refrigeration circuit for circulating a refrigerant in a predetermined flow direction, comprising, in the flow direction, a heat-emitting heat exchanger, an evaporator throttle valve, an evaporator, a compressor, an internal heat exchanger arranged with its "cold side" between the evaporator and the compressor, an inlet temperature sensor arranged between the evaporator and the internal heat exchanger, and an outlet temperature sensor located between the internal heat exchanger and the compressor, and a controller for controlling the evaporator throttle valve based on the measurement of the inlet and outlet temperature sensors, wherein the controller is adapted to control the evaporator throttle valve based on an inlet temperature setpoint at the inlet temperature sensor; and shifting the inlet temperature setpoint based on the outlet temperature sensor measurement.

[0005] JP H01 102255 A provides a fuel gas metering device for a gas engine, enabling simple and safe engine starting. It includes a pressure regulator with a vacuum locking mechanism and a centrifugal governor. The vacuum locking mechanism is operated by the engine's intake vacuum. Fuel gas is also supplied to a mixer; if the intake vacuum falls below a specified value, the fuel gas supply is interrupted. The pressure regulator also incorporates an electrical release device. This release device switches the engine's starter switch to the start position, thus reversing the interruption of the fuel gas supply by the vacuum locking mechanism.

[0006] EP 1 014 013 A1 discloses that in a vapor compression 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 serves to adjust the flow rate of the liquid phase portion of the refrigerant supplied to the evaporator according to a control signal in order to maintain a superheat level of the gas phase portion of the refrigerant supplied to a compressor. Thermistors detect the temperatures of the gas phase portion and of the evaporated portion flowing from the evaporator to generate a first and a second temperature signal, respectively. A temperature / pressure sensor detects the state of a radiated portion of the refrigerant flowing from a cooler to generate a refrigerant state value.A selector chooses one of the first and second temperature signals in accordance with the refrigerant state value as the selected signal in order to supply the selected signal as the control signal to the superheat control valve.

[0007] US 2013 / 074535 A1 discloses a vapor compression system comprising a compressor, a condenser, an expansion device, e.g., in the form of an expansion valve, and an evaporator arranged along a refrigerant path. A method for operating the vapor compression system comprises the steps of: obtaining a superheat value representative of the superheat of the refrigerant entering the compressor; obtaining a subcooling value representative of the subcooling of the refrigerant entering the expansion device; and operating the expansion device based on the obtained superheat value and the obtained subcooling value.The subcooling value is taken into account during the operation of the expansion unit, as fluctuations in the subcooling value at a given expansion unit opening degree have a significant impact on the evaporator's cooling capacity and thus lead to more stable system operation. The system may also include an internal heat exchanger.

[0008] WO 03 / 106900 A1 relates to a method for controlling a Carnot cycle in a system in which at least the following components are arranged in a working fluid circuit: a compressor, a condenser, an expansion valve, an evaporator, and an internal heat exchanger. This heat exchanger is traversed by the condensed working fluid flowing from the condenser and the working fluid flowing from the evaporator. In the internal heat exchanger, the gaseous working fluid is superheated and the condensed working fluid is subcooled, with the superheating of the working fluid being controlled by the expansion valve. To improve performance, for example, 10 to 30% of the total evaporation of the working fluid is carried out in the internal heat exchanger, which is designed as a multi-pass auxiliary evaporator.

[0009] Many compressors are unsuitable for drawing in saturated vapor at the compressor inlet; therefore, 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, whereby a minimum superheat must also be maintained for refrigerants with a temperature glide.

[0010] Failure to comply with minimum superheating requirements can result in increased compressor wear due to insufficient lubrication of the compression components or damage from cavitation caused by post-evaporation in the compressor.

[0011] An increase in the superheating of the refrigerant at the compressor inlet corresponds approximately to the same extent to an increase in the refrigerant outlet temperature from the compressor.

[0012] The refrigerant outlet temperature affects the thermal stress on the compressor components and the oil in the refrigeration circuit, so the compressor manufacturer specifies an upper limit for the permissible refrigerant outlet temperature from the compressor.

[0013] The object of the invention is to propose a control method which operates a steam compression circuit in such a way that an optimal overall efficiency is achieved.

[0014] A further object of the invention is to achieve good control behavior even with refrigerants consisting of at least two substances, in particular when subjected to disturbances such as changes in operating conditions due to temperature changes in the heat source circuit, heat sink system or changes in compressor power.

[0015] The problem is solved by the process features of claim 1. Preferred embodiments are described in the dependent claims.

[0016] A controlled variable for evaporator outlet superheat is calculated. Using a target value for evaporator outlet superheat, a control deviation for evaporator outlet superheat is calculated. A controlled variable for compressor inlet superheat is also calculated. Using a target value for compressor inlet superheat, a control deviation for compressor inlet superheat is calculated. The control value R is calculated from a weighted influence of the control deviation for evaporator outlet superheat and a weighted influence of the control deviation for compressor inlet superheat. The expansion valve is controlled using this control value.

[0017] Another advantageous method is to perform the subtraction "inversely," whereby the actual value is subtracted from the target value to calculate the control deviations. The control deviation is advantageously calculated by determining the difference between an "actual value" and a "target value."

[0018] According to a method step of the invention, the evaporator outlet superheat is weighted against the compressor inlet superheat using a parameter to determine the control value. It is particularly advantageous to weight the control deviations of an evaporator outlet superheat or the compressor inlet superheat from a setpoint.

[0019] In another advantageous process step, the evaporator outlet superheat is weighted against the compressor inlet superheat to determine the controlled variable using a parameter. Particularly advantageously, the control deviations of an evaporator outlet superheat or a compressor inlet superheat are weighted with a setpoint.

[0020] In a further advantageous embodiment of the method, the evaporator outlet superheat is determined from an evaporator outlet temperature measurement value, which is measured with an evaporator outlet temperature sensor, and from a low pressure, which is measured with a low pressure sensor.

[0021] In a further preferred process step, the compressor inlet superheat is determined from a compressor inlet temperature, measured with a compressor inlet temperature sensor, and a low pressure, measured with a low pressure sensor.

[0022] In an advantageous process step, a first control deviation of the compressor inlet superheat is calculated in the controller together with a second control deviation of the evaporator outlet superheat to form a total control deviation, and the total control deviation is used to adjust the expansion valve.

[0023] According to the invention, the method is implemented such that the parameter is changed gradually after the commissioning phase, with weighting initially being applied using a weighting coefficient for the compressor inlet superheat, which has an influence of less than 20% compared to the evaporator outlet superheat. The weighting is then taken into account in a changing, and in particular increasing, manner when determining the control value, especially up to a predetermined target weighting of the parameter. According to the invention, during the commissioning phase, the weighting coefficient for the compressor inlet superheat is set to a value between 0% and 20%, which is low compared to the operating phase, and after the commissioning phase, it increases ramp-wise over time to a target value for the operating phase.

[0024] During a time-limited commissioning phase after the compressor is switched on, the parameter value is advantageously controlled such that the weighting of the compressor inlet superheat for determining the controlled variable has an influence of 0% or close to 0%, particularly below 20%. This ensures that only or predominantly the evaporator outlet superheat is included in the controlled variable. In a subsequent transition phase, the parameter P is controlled such that the weighting of the compressor inlet superheat for determining the controlled variable is increased to a target value, and in the subsequent operating phase, the weighting of the compressor inlet superheat reaches this target value.

[0025] Furthermore, it is advantageous to vary the target value for the weighting of the controlled variable depending on compressor inlet superheat during operating phases in the steady state of the vapor compression system.

[0026] The figures show an example of implementation: 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

[0027] The heat pump 100 essentially consists of a vapor compression system 200, 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 heat energy QH 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, for building heating or a hot water heating system, advantageously a refrigerant receiver 260, which is used as a refrigerant reservoir to compensate for varying refrigerant quantity requirements depending on operating conditions, an expansion valve 230 for expanding the refrigerant, an evaporator 240, with an evaporator inlet 241, for transferring source energy QQ 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 QQ from the ground or an air system which absorbs heat energy QQ from the ambient air and transfers it to the vapor compression system 200 or any other heat source, a recuperator 250 which is designed to transfer internal heat energy Q i between the refrigerant flowing from the condenser 400 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 high-pressure flow direction SHD is formed from the compressor 210 to the expansion valve 230. From the expansion valve 230 to the compressor 210, a low-pressure path with low-pressure flow direction SND of the refrigerant is formed, in which the evaporator 240 is located.

[0028] The actuators listed below are advantageously at least partially connected to the controller via a data connection 510, which can be via cable, radio or other technologies: compressor 210, heating medium pump 410, brine pump 330, expansion valve 230, compressor inlet temperature sensor 501, low pressure sensor 502, high pressure sensor 503, hot gas temperature sensor 504, recuperator inlet temperature sensor 505, recuperator outlet temperature sensor 506, evaporator inlet temperature sensor 507 and / or evaporator outlet temperature sensor 508.

[0029] The compressor 210 serves to compress the superheated refrigerant from an inlet port 211 to a compressor outlet pressure Pva at a compressor outlet temperature Tva at the compressor outlet 212. The compressor 210 typically includes a drive unit with an electric motor, a compression unit, and the electric motor can advantageously be operated at variable speed. The compression unit can be designed as a rotary piston unit, scroll unit, or otherwise. At the compressor outlet 212, the compressed superheated refrigerant is at a higher pressure, in particular a high pressure (HP), than at the inlet port 211 with a compressor inlet pressure Pve, in particular a low pressure (LP), at a compressed inlet temperature Tve, which describes the state at the inlet port 211 in a compression chamber.

[0030] In the condenser 220, heat energy QH is transferred from the refrigerant of the vapor compression system 200 to a heating medium of the heat sink system 400. First, the refrigerant is deheated in the condenser 220, whereby superheated refrigerant vapor transfers some of its heat energy to the heating medium of the heat sink system 400 through a temperature reduction.

[0031] After the refrigerant vapor has been deheated, a further heat transfer QH advantageously takes place in the condenser 220 through condensation of the refrigerant during the phase transition from the gaseous phase of the refrigerant to the liquid phase. In this process, further heat QH is transferred from the refrigerant in the vapor compression system 200 to the heating medium of the heat sink system 400.

[0032] The high pressure (HD) of the refrigerant that develops in the condenser 220 corresponds approximately to a condensation pressure of the refrigerant at a heating medium temperature T WS in the heat sink system during the operation of the compressor 210.

[0033] The heating medium, in particular water, is conveyed by means of a heating medium pump 410 through the heat sink system 400 in a SW direction through the condenser 220, whereby the heat energy QH is transferred from the refrigerant to the heating medium.

[0034] In the downstream collector 260, refrigerant exiting the condenser 220 is stored. Depending on the operating point of the vapor compression circuit 200, this refrigerant is not to be fed into the circulating refrigerant. If more refrigerant is fed from the condenser 220 than is passed through the expansion valve 230, the collector 260 fills up; otherwise, it empties or empties.

[0035] In the subsequent recuperator 250, which can also be referred to as an internal heat exchanger, internal thermal energy Qi is transferred from the high-pressure refrigerant (HD) flowing from the condenser 220 to the expansion valve 230 in a high-pressure flow direction (SHD) to the low-pressure refrigerant (ND) flowing from the evaporator to the compressor in a low-pressure flow direction (SND). In this process, the refrigerant flowing from the condenser to the expansion valve 230 is advantageously subcooled.

[0036] The refrigerant initially flows into the expansion valve through an expansion valve inlet 231. In the expansion valve 230, the refrigerant pressure is reduced from high pressure (HP) to low pressure (LP) by passing the refrigerant advantageously through a nozzle arrangement or throttle with an advantageously variable opening cross-section, the low pressure advantageously corresponding approximately to the suction pressure of the compressor 210. Instead of an expansion valve 230, any other pressure-reducing device can also be used. Pressure-reducing tubes, turbines, or other expansion devices are advantageous.

[0037] The opening degree of the expansion valve 230 is set by an electric motor, usually a stepper motor, which is controlled by the controller 500. The low pressure (LP) at the expansion valve outlet 232 of the refrigerant from the expansion valve 230 is controlled such that the resulting low pressure (LP) of the refrigerant during operation of the compressor 210 corresponds approximately to the evaporation pressure of the refrigerant at the heat source medium temperature TWQ. Advantageously, the evaporation temperature of the refrigerant will be a few Kelvin below the heat source medium temperature TWQ, so that the temperature difference drives heat transfer.

[0038] In the evaporator, heat of vaporization energy QV is transferred from the heat source fluid of the heat source system 300, which may be a brine system, a geothermal system for utilizing heat energy QQ from the ground, an air system for utilizing energy QQ from the ambient air, or another heat source that delivers the source energy QQ to the vapor compression system 200.

[0039] The refrigerant flowing into the evaporator 240 reduces its wet vapor content by absorbing heat QQ as it passes through the evaporator 240 and advantageously exits the evaporator 240 with a low wet vapor content or, even more advantageously, as a superheated gaseous refrigerant. The heat source medium is circulated through the heat source medium path of the evaporator 240 by means of a brine pump 330 in brine-to-water heat pumps or an outdoor air fan in air-to-water heat pumps, whereby thermal energy QQ is extracted from the heat source medium as it flows through the evaporator.

[0040] In the recuperator 250, heat energy Q i is transferred between the refrigerant flowing from the condenser 220 to the expansion valve 230 and the refrigerant flowing from the evaporator 240 to the compressor 210, whereby the refrigerant flowing from the evaporator 240 to the compressor 210 is further superheated.

[0041] This superheated refrigerant, which exits the recuperator 250 at a superheat temperature T Ke, is directed to the refrigerant inlet connection 211 of the compressor 210.

[0042] The recuperator 250 is used in the steam compression circuit 200 to increase the overall efficiency as the quotient of the delivered heating power QH and the input electrical power P e for driving the compressor motor.

[0043] For this purpose, further heat energy Q i is extracted from the refrigerant, which releases heat energy QH 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.

[0044] The internal energy state of the refrigerant upon entering the evaporator 240 is reduced by this heat extraction Q i, so that the refrigerant can absorb more heat energy QQ from the heat source 300 at the same evaporation temperature level.

[0045] Subsequently, after the evaporator outlet 242 from the evaporator 240, the heat energy Qi extracted in the high-pressure path is supplied back to the refrigerant in the low-pressure path at low pressure ND and a low-pressure temperature T Va in the recuperator 250. The supply of energy advantageously reduces the wet vapor content to a state without wet vapor, and then superheating occurs through further energy supply.

[0046] Furthermore, the following sensors are advantageously arranged to monitor the operating state of the steam compression system 200, enabling model-based feedforward control, particularly for safeguarding the operating conditions of the steam compression system 200. On the one hand, the process values ​​acquired by the sensors are advantageously used to safeguard the permissible operating ranges of the components, such as the compressor 210. On the other hand, model-based feedforward control is performed based on the sensor data, specifically for the speed of the compressor 210 and / or the valve opening degree of the expansion valve. This allows the controller to make only minor corrections to compensate for any remaining, albeit smaller, control deviations. A high-pressure sensor 503 is advantageously used to detect the high pressure HD of the refrigerant at the compressor outlet 212, or between the compressor outlet 212 and the expansion valve inlet 231; a hot gas temperature sensor 504 is advantageously used to detect 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 is advantageously used to detect the internal temperature Tle 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 the "recuperator outlet temperature high-pressure path" and is 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.

[0047] The following sensors are particularly required 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 advantageously serves 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.

[0048] The process parameter that has a significant influence on the overall efficiency of the vapor compression circuit 200, as the quotient between the heating power QH transferred by the vapor compression circuit 200 and the electrical power Pe consumed by the compressor 210, is the superheat of the refrigerant at the compressor inlet 211. To maintain permissible compressor operating conditions, however, it is advantageous to adhere to limits regarding the allowed superheat range of the refrigerant at the compressor inlet. Excessively low superheats particularly compromise the lubricating properties of the machine oil, while excessive superheats result in an excessively high hot gas temperature.

[0049] Superheating describes the temperature difference between the measured compressor inlet temperature T KE of the refrigerant and the evaporation temperature of the refrigerant at saturated vapor.

[0050] These limit values, particularly for superheating, define the permissible superheating range of the components at the compressor inlet 211, depending on the operating point. Furthermore, dependencies also exist between the compressor inlet superheat TÜE and the overall efficiency of the vapor compression circuit 200, or between the compressor inlet superheat TÜE and the stability S of a control value R, which is advantageous for controlling the compressor inlet superheat.

[0051] To meet all these requirements, the heat source medium temperature, the heating medium temperature, the compressor power Pe, and target values ​​Zi or Zi for calculating the compressor inlet superheat Ti are advantageously used, depending on the operating point of the steam compression circuit 200. In the simplest case, the target value for the compressor inlet superheat Ti is constant regardless of all operating conditions, e.g., 10 Kelvin. With a more complex adaptation, it is varied as a function of an operating point parameter, e.g., the compressor power Pe, or with an even more complex adaptation, it varies as a function of several operating point parameters.

[0052] A compressor inlet superheat TÜE is weighted and determined in conjunction with the evaporator outlet superheat TÜA, from which a total control deviation is calculated in the controller 500. This total deviation is then used to control the steam compression circuit 200. Advantageously, the control deviations of the compressor inlet superheat TÜE and evaporator outlet superheat TÜA are first calculated more precisely by determining the differences between the respective measured values ​​and target values. Compressor inlet superheat TÜE = Measured value compressor inlet superheat - Target value compressor inlet superheat ZTÜE Evaporator outlet superheat TÜA = Measured value evaporator outlet superheat - Target value evaporator outlet superheat ZTÜA

[0053] Then, the total control deviation is advantageously calculated from the weighted influence of the control deviation of the compressor inlet superheat T ÜE and the weighted influence of the control deviation of the evaporator outlet superheat T ÜA in the controller 500, which is fed into the control of the steam compression circuit 200.

[0054] In the vapor compression circuit 200, after expansion through the expansion valve 230, the refrigerant passes through two sequentially arranged heat exchangers, the evaporator 240 and the recuperator 250, in which heat energy QQ and Qi is supplied to the refrigerant.

[0055] In the evaporator 250, source heat energy QQ from the heat source system 300 is supplied to the refrigerant. The temperature level of the supplied source heat QQ is at the same temperature level as the heat source, in particular the ground or the outside air.

[0056] In the recuperator 250, which follows the refrigerant in the high-pressure flow direction SHD, thermal energy Qi is supplied to the refrigerant after it leaves the condenser 220. The temperature level of the refrigerant at the condenser outlet is approximately equal to the return temperature of the heating medium.

[0057] This connection of the evaporator 240 with the recuperator 250 in series has a decisive influence on the transfer function of the control loop for the control of compressor inlet superheat T ÜE .

[0058] The control value R is advantageously the weighted combination of the control deviation of the compressor inlet superheat T ÜE with the control deviation of the evaporator outlet superheat.

[0059] Actuators with an influence on the control value R, in particular the compressor inlet superheat T ÜE, are influenced in the relevant steam compression circuit 200 by the compressor speed and / or the opening degree of the expansion valve 230, which also advantageously determines the low pressure ND and the evaporation temperature level.

[0060] Actuators have a particularly advantageous influence on the control value R, especially on the weighted relationship between the control deviation of the compressor inlet superheat and the control deviation of the evaporator outlet superheat. In the relevant steam compression circuit 200, the compressor 210, through the variation of the compressor speed, and the expansion valve 230, through the influence on the degree of opening, are such actuators. These two actuators influence the low pressure LP and the evaporation temperature level.

[0061] Advantageously, in steam compression circuit 200, the compressor speed is set so that the heating power QH transferred from steam compression circuit 200 to the heating medium corresponds to the required target value Z. To comply with this requirement, influencing the compressor speed to control the compressor inlet superheat T ÜE is advantageously subordinate or unnecessary.

[0062] Advantageously, the opening degree of the expansion valve 230 is used as a control parameter for regulating the compressor inlet superheat TÜE. The influence of the opening degree of the expansion valve 230 on the compressor inlet superheat TÜE is as follows: The expansion valve 230 acts as a nozzle with an electrically adjustable nozzle cross-section, in which a needle-shaped nozzle needle is typically moved into a nozzle seat via a thread by means of a stepper motor.

[0063] The refrigerant flow rate through the expansion valve, when operating with liquid refrigerant at the expansion valve inlet 231, 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 opening degree and advantageously by a constant that depends on the refrigerant and the geometry of the expansion valve 230.

[0064] Since, assuming a constant heating medium temperature T WS at an operating point, the corresponding low pressure ND of the refrigerant at the inlet to the expansion valve 230 can also be assumed to be constant, the degree of opening of the expansion valve 230 only significantly influences the low pressure ND, i.e., the outlet pressure from the expansion valve 230.

[0065] If the opening degree of the expansion valve 230 is reduced, less refrigerant passes through the expansion valve 230 at a constant high pressure (HP) and initially constant low pressure (LP). Since the compressor 210 continues to deliver the same refrigerant mass flow rate, less refrigerant is supplied through the expansion valve 230 in the high-pressure flow direction (HP) than is extracted by the compressor 210. Because refrigerant vapor is a compressible medium, the low pressure (LP) on the low-pressure side of the vapor compression circuit 200 then decreases.As the low pressure ND decreases, the mass flow rate of refrigerant through the compressor 210 decreases approximately proportionally, since its delivery rate can be described approximately 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 rate supplied through the expansion valve 230 is equal to the refrigerant mass flow rate discharged by the compressor 210.

[0066] If the opening of the expansion valve 230 is increased, more refrigerant passes through the expansion valve 230 at a constant high pressure (HP) and initially constant low pressure (LP). Since the compressor 210 continues to deliver the same refrigerant mass flow rate, more refrigerant is supplied to the low-pressure side (LP) of the refrigeration circuit through the expansion valve 230 than is extracted by the compressor 210. Because refrigerant vapor is a compressible medium, the low pressure (LP) on the low-pressure side of the vapor compression circuit 200 increases. As the low pressure (LP) increases, the mass flow rate of the compressor 210 increases approximately proportionally, since its flow rate can be approximated as volume / time. This results in a correspondingly higher low pressure (LP) at which the refrigerant mass flow rate supplied through the expansion valve 230 equals the refrigerant mass flow rate discharged by the compressor 210.

[0067] 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 QQ from the heat source system 300 is transferred between the heat source medium and the refrigerant at different temperatures, whereby the heat flow QQ depends on the temperature difference between the heat source medium and the refrigerant and the thermal resistance of a heat transfer layer of the evaporator 240.

[0068] The thermal resistance between the heat source media path of the evaporator and the refrigerant path of the evaporator can be assumed to be approximately constant in a given vapor compression circuit 200. Therefore, the magnitude of the heat transfer power in the evaporator 240 depends significantly on the integral of the temperature differences of all surface elements of the heat transfer layer.

[0069] In order to transfer a sufficient amount of heat energy QQ from the heat source system 300 to the refrigerant, it must be ensured that the temperature of the heat source medium in as many surface elements as possible of the transfer layer of the heat exchanger, here the evaporator 240, is greater than the temperature of the refrigerant at the respective surface element.

[0070] If the refrigerant is in a saturated vapor state when flowing through the evaporator 240, a refrigerant temperature is established which, according to the saturation vapor characteristic curve, is a function of the refrigerant's low pressure (LP). Therefore, by controlling the low pressure (LP) or the evaporation pressure, the evaporation temperature of the refrigerant as it flows through the recuperator 250 can be indirectly controlled.

[0071] The thermal energy QQ, which is transferred from the heat source system to the refrigerant flowing through the evaporator 240, causes a change in the state of matter of the refrigerant.

[0072] The wet vapor fraction in saturated refrigerant vapor decreases at constant low pressure during heat transfer to the refrigerant. With incomplete evaporation, the wet vapor fraction, and thus also the internal energy state of the refrigerant upon exiting the heat exchanger, is a function of: wet vapor fraction at the inlet to the evaporator 240, refrigerant mass flow, heat transfer QQ, and an enthalpy difference in the wet vapor region at the respective low pressure ND, wherein the refrigerant has a material constant as a function of pressure.

[0073] For complete evaporation, additional energy is supplied in the recuperator 250 to superheat the refrigerant beyond the saturated vapor state.

[0074] With this method, given the operating conditions of the vapor compression circuit 200, a corresponding refrigerant state is set at the exit from the evaporator 240 depending on the control variable "opening degree of expansion valve 230".

[0075] In steady state, the control loop behavior of the "isolated" control loop "evaporator 240" is moderately steep.

[0076] Advantageously, a refrigerant, in particular a refrigerant mixture, is used which exhibits a "temperature glide," especially R454c. Advantageously, with a refrigerant mixture exhibiting a temperature glide, a relative opening degree change of the expansion valve actuator of 1% relative is set with a superheat change at the refrigerant outlet from the evaporator, typically with a superheat change of less than approximately 1 K.

[0077] This state can advantageously also be achieved by influencing at least one or more of the following various time constants through control engineering; these ultimately influence the process variable refrigerant superheat at the evaporator outlet 242: A first time constant advantageously delays the mechanical change in the opening degree of the expansion valve 230 by limiting the travel speed via the controller 500. The control value R is reduced in this first time constant Z by a braking value. A second time constant, also specified by the controller 500, advantageously delays the adjustment 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, whereby a change in the evaporation pressure, and thus the evaporation temperature, results in a delayed temperature change of the heat transfer layer of the evaporator, which often contains several kilograms of metal, and of the heat source medium.A fourth time constant is advantageously derived from delayed changes of state of the refrigerant during changes in evaporation temperature. A fifth time constant is advantageously derived from the transport of the refrigerant through the evaporator 240 at a finite flow velocity.

[0078] Therefore, after changing the control variable "opening degree of the expansion valve 230", a delay in the corresponding refrigerant state change occurs at the exit from the evaporator outlet 242, and a total time constant Zges is advantageously in the range of 30 seconds to about 5 minutes, depending on the operating point.

[0079] After passing through the evaporator 240, the refrigerant enters the low-pressure path of the recuperator 250 at low pressure ND.

[0080] Is the state of matter of the refrigerant when flowing into the recuperator 250 in a normal operating case, i.e. advantageously either saturated vapor with a low vapor content between 0 and 20% or, in particular, also advantageously already superheated refrigerant?

[0081] With advantageously saturated vapor, a refrigerant temperature is established which, according to the refrigerant's saturation vapor characteristic, is a function of the refrigerant pressure. If superheated refrigerant enters the system, the refrigerant temperature will at most reach a value corresponding to the inlet temperature of the heat source medium.

[0082] In order to transfer a sufficient amount of heat energy from the refrigerant of the high-pressure side refrigerant path to the refrigerant of the low-pressure side refrigerant path in the recuperator 250, it must be ensured that the temperature of the refrigerant of the high-pressure side refrigerant path at high pressure (HD) is greater than the temperature of the refrigerant of the low-pressure side refrigerant path at low pressure (LP) in as many surface elements of the transfer layer of the recuperator 250 as possible.

[0083] The corresponding temperatures of the heat sink system 400 of the vapor compression system 200 are higher in a heating case than the corresponding temperatures of the heat source such as the ground or the outside air.

[0084] The thermal energy Qi, which is transferred from the refrigerant at high pressure (HP) in the high-pressure refrigerant path to the refrigerant at low pressure in the low-pressure refrigerant path of the recuperator 250, causes a change in the state of matter of the refrigerant on the low-pressure side. The wet vapor fraction of the refrigerant flowing through the recuperator 250 at low pressure (LP) decreases during heat transfer to the refrigerant, and after complete evaporation, superheating of the refrigerant advantageously occurs.

[0085] The internal energy state of the refrigerant, upon exiting the low-pressure side path of the recuperator, is advantageously controlled depending on one or more of the following factors: wet steam fraction at the inlet to the recuperator 250, refrigerant mass flow, transferred heat power 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 of 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.

[0086] Advantageously, this ensures that, depending on the given operating conditions of the vapor compression circuit 200 and depending on the control variable "opening degree of expansion valve 230", a corresponding refrigerant state is established at the exit 252 from the recuperator 250 at low pressure LP.

[0087] In steady state, the control loop behavior of the "isolated" control loop at the low pressure (LP) of the refrigerant in the low-pressure-side path of the recuperator 250 exhibits a high steepness, while the internal energy state of the refrigerant at the inlet 251 to the low-pressure-side path of the recuperator 250 remains approximately constant. A change in the opening degree of the expansion valve of 1%, in particular relative to the temperature, advantageously results in a superheat change at the refrigerant outlet from the evaporator 230 of approximately 10 K or even more than 10 K.

[0088] Compared to the recuperator 250, a significantly higher heat transfer is advantageously achieved in the evaporator 240, resulting in a high driving temperature difference between the source medium and the refrigerant in the evaporator 240.

[0089] This state is advantageously set using at least one of the following time constants Z: An eleventh time constant, Z11, advantageously delays the mechanical change in the opening degree of the expansion valve 230 by limiting its travel speed. A twelfth time constant, Z12, advantageously delays the adjustment of a corresponding low-pressure (LP) when the opening degree of the expansion valve 230 changes due to the compressibility of the refrigerant vapor in the low-pressure path LP. A thirteenth time constant, Z13, is a thermal time constant of the heat transfer layer of the evaporator. Thus, a change in the evaporation pressure, and therefore the evaporation temperature, causes a delayed temperature change in the heat transfer layer, which often contains several kilograms of metal, and in the refrigerant of the high-pressure path HD of the recuperator 250. A fourteenth time constant, Z14, is advantageously determined or specified from delayed changes in the state of matter of the refrigerant when the evaporation temperature changes.A 15th time constant Z15 advantageously results from the transport of the refrigerant through the evaporator 240 at a finite flow velocity and is taken into account.

[0090] 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 delayed by at least two time constants Z, Z11, Z12, Z13, Z14, Z15, Ztotal after a change in the manipulated variable "opening degree of expansion valve".

[0091] After changing the control variable "opening degree expansion valve 230", a further delay in the corresponding refrigerant state change occurs due to the time behavior of the recuperator 250 when exiting the low-pressure side refrigerant path of the recuperator 250.

[0092] The time behavior of the recuperator 250 can be advantageously considered as the total recuperator time constant Zges, depending on the respective operating point of the steam compression circuit, in the range of approximately 1 minute to 30 minutes.

[0093] Advantageously, a weighted combination of compressor inlet superheat TÜE and evaporator outlet superheat TÜA is achieved, in particular by calculating the total control deviation using a weighted combination of the control deviation of the compressor superheat and the evaporator outlet superheat TÜA, which is fed into the controller 500 to control the steam compression circuit 200.

[0094] The compressor inlet superheat control TÜE is advantageously used as the main controlled variable, and the corresponding signal flows and signal processing take place in particular in the following process steps: Step 1

[0095] In a first process step, the process variables compressor inlet superheat TÜE are advantageously measured as the main control variable and the evaporator outlet superheat TÜA as an auxiliary variable.

[0096] For this purpose, the evaporation temperature of the refrigerant at the respective measurement point is either directly measured using a temperature sensor positioned to detect a temperature corresponding to the refrigerant temperature in the wet vapor region, or indirectly measured using a pressure sensor that detects the refrigerant pressure of the refrigerant evaporating in the wet vapor region, and then calculates the evaporation temperature from the refrigerant-specific relationship between pressure and temperature in the wet vapor region.

[0097] Furthermore, the refrigerant temperature is recorded at the respective superheat measurement point, in particular at the evaporator outlet 242 and / or the compressor inlet 211, using temperature sensors 501 and 508. The temperature difference between the refrigerant at the respective measurement point and the evaporation temperature is then calculated, and this temperature difference value corresponds to the respective superheat of the refrigerant at the measurement point.

[0098] The input variables for the calculation in step 1 are then the compressor inlet superheat TÜE and the evaporator outlet superheat TÜA. Step 2

[0099] The process variables compressor inlet superheat TÜE and evaporator outlet superheat TÜA are advantageously calculated in a second step to form associated control deviations with their respective setpoints: The setpoint for the compressor inlet superheat TÜE is advantageously varied in the range between approximately 5 K and 20 K to ensure the permissible compressor operating range and the highest possible efficiency of the refrigeration circuit.

[0100] The setpoint for the evaporator outlet superheat TÜA is then varied in an optional operating mode so that, under steady-state conditions, it approximately corresponds to the established process value of the evaporator outlet superheat TÜA. This setpoint for the evaporator outlet superheat TÜA can be pre-calculated using a model based on an operating mode or operating point, depending on the evaporation temperature, the condensation temperature, the compressor output, a setpoint for the compressor inlet superheat TÜE, and / or component properties, and can be adaptively corrected.

[0101] The control deviation of the compressor inlet superheat T ÜE is then calculated by subtracting the setpoint value of the compressor inlet superheat T ÜE from the process value of the compressor inlet superheat T ÜE.

[0102] The control deviation of the evaporator outlet superheat TÜA is then calculated by subtracting the setpoint value of the evaporator outlet superheat TÜA from the process value of the evaporator outlet superheat TÜA. Step 3

[0103] In a third process step, the control deviation of the compressor inlet superheat T ÜE and the control deviation of the evaporator outlet superheat T ÜA are advantageously combined to form a total control deviation superheat.

[0104] The combination is achieved in particular by means of a weighted addition of the individual rule deviations.

[0105] The weighting influence is a measure of the proportional combination of the individual rule deviations and, in extreme cases, can result in the exclusive inclusion of only one individual rule deviation, but usually it results in the weighted inclusion of both individual rule deviations.

[0106] Advantageously, the weighting influence is estimated as a value between 0 and 1, i.e., 0 to 100%, and this value is included in the total control deviation according to the degree to which the compressor inlet superheat deviation T ÜE is incorporated, resulting in the following dependency for calculating the total control deviation: Gesamt − Regelabweichung Überhitzung = Gewichtungseinfluss * Regelabweichung Verdichtereintrittsüberhitzung + 1 − Gewichtungseinfluss * Regelabweichung Verdampferaustrittsüberhitzung

[0107] The value of the weighting influence can be advantageously varied depending on the operating mode and / or the operating point of the heat pump 100: During the operating mode transition between operation with compressor 210 switched off and operation with compressor 210 switched on in heating mode, due to the dynamic process value changes during the start-up of the steam compression system 200, it is advantageous to initially include only the evaporator outlet superheat TÜA in the overall control deviation. In particular, the weighting influence value is initially set to 0 or is advantageously below 20%. After a stabilization phase of the steam compression system 200, it is advantageous not to switch spontaneously to the weighting influence value designed for normal operation, but rather to implement a ramp-like transition. In this case, it is advantageous for the weighting influence value to be ramped up from the initial value of 0, or a value particularly below 20%, to the intended target value.This avoids inconsistencies in values ​​during spontaneous switching and thus prevents control oscillations. The target value of the weighting influence is advantageously adapted to the respective operating mode and operating point. Operating points characterized by an increased tendency to oscillate advantageously require a lower weighting of the control deviation of the compressor inlet superheat TÜE. In particular, this avoids a tendency to oscillate due to the signal behavior of the compressor inlet superheat TÜE, which is caused by the greater signal delay and steeper path gradient compared to the evaporator outlet superheat TÜA. Step 4:

[0108] In a fourth process 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, so that in the regulated case a control deviation of the superheat is set to approximately 0 Kelvin.

[0109] 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.

Claims

1. Method for operating a heat pump (100) with a vapour compression system (200) in which a gaseous refrigerant is compressed from a low pressure ND to a high pressure HD by a compressor (210) controlled by means of a regulator (500), the refrigerant is driven through a condenser (220), in which it releases heating energy Qh to a heating medium located in a heat sink system (400), an internal thermal energy Qi of the refrigerant is exchanged in a recuperator (250) between the refrigerant flowing under high pressure HD from the condenser (220) to an expansion valve and the refrigerant flowing under low pressure ND from an evaporator (240) to the compressor (210), whereupon the refrigerant is fed in a high pressure flow direction SHD to the expansion valve (230) controlled by the controller (500), in which the refrigerant is expanded from high pressure HD to 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 QQ, including the method steps that a control variable RÜA of the evaporator discharge superheating TÜA is calculated, a control deviation of the evaporator discharge superheating TÜA is calculated with a target value ZÜA for an evaporator discharge superheating TÜA, a control variable RÜE of the compressor intake superheating TÜE is calculated, a control deviation of the compressor intake superheating TÜE is calculated with a target value ZÜE for a compressor intake superheating TÜE, characterised in that a control value R is calculated from a weighted influence of the control deviation of the evaporator discharge superheating TÜA with a weighted influence of the control deviation of the compressor intake superheating TÜE, and the expansion valve (230) is controlled on the basis of the control value R, wherein the control deviations of the evaporator discharge superheating TÜA and the compressor intake superheating TÜE are weighted with a parameter P to determine the control value R, and the parameter P is changed in a modulating manner during an operating phase after a commissioning phase, wherein weighting by means of a weighting coefficient W of the compressor intake superheating TÜE begins with an influence of less than 20 % compared to the evaporator discharge superheating TÜA, and during the commissioning phase, the weighting coefficient W of the compressor intake superheating TÜE is set to a value between 0 % and 20 % that is low in comparison with the operating phase, and after the commissioning phase has elapsed, it increases over time, in particular in a ramp-shaped manner, to a target value Z for the operating phase.

2. Method for operating a heat pump according to claim 1, including the method step that the evaporator discharge superheating TÜA is determined from an evaporator discharge temperature measured value TVa, which is measured with an evaporator discharge temperature sensor (508), and from a low pressure ND, which is measured with a low pressure sensor (502).

3. Method for operating a heat pump according to any one of the preceding claims, including the method step that the compressor intake superheating TÜE is determined from a compressor intake temperature TVE, measured with a compressor intake temperature sensor (501), and a low pressure ND, which is measured with a low pressure sensor (502).

4. Method for operating a heat pump (100) according to any one of the preceding claims including the method step that, in a time-limited commissioning phase after the compressor (210) has been switched on, the parameter value P is controlled in such a way that the weighting of the compressor intake superheating TÜE for determining the control value R has an influence of 0 % or close to 0 %, with the result that the evaporator discharge superheating TÜA is exclusively or significantly included in the control value R, wherein in a subsequent transition phase, the parameter P is controlled in such a way that the weighting of the control variable of the compressor intake superheating TÜE for determining the control value R is increased up to a target value and in the subsequent operating phase, the weighting of the compressor intake superheating TÜE reaches the target value.

5. Method for operating a heat pump according to any one of the preceding claims, including the method step that the control uses a compressor intake superheating TÜE as a main control variable, wherein the corresponding signal flows are processed in four method steps, wherein in a method step 1, the process variables of the compressor intake superheating TÜE are first determined as the main control variable with the evaporator discharge superheating TÜA as an auxiliary variable, whereby at a respective detection point in the wet vapour area, an evaporation temperature of the refrigerant is detected directly by means of a temperature sensor or indirectly by means of a pressure sensor, and at a respective superheating measuring point at the evaporator discharge (242) and at the compressor intake (211), the associated temperatures of the refrigerant temperature are detected by means of temperature sensors (501, 508), and wherein the temperature difference of the refrigerant at the respective superheating measuring point and the evaporation temperature is calculated, in a method step 2, the process variables of the compressor intake superheating TÜE and the evaporator discharge superheating TÜA are calculated to form associated control deviations with associated setpoints, wherein the setpoint for the compressor intake superheating TÜE, in particular to ensure the highest possible efficiency of the refrigerant circuit, is varied in the range between approx. 5 K to 20 K, and the setpoint value for the evaporator discharge superheating TÜA is then varied so that the setpoint value of the evaporator discharge superheating in the steady-state control case corresponds approximately to the resulting process value of the evaporator discharge superheating TÜA, after which the control deviation of the compressor intake superheating TÜE is then calculated by subtracting the setpoint value of the compressor intake superheating TÜE from the process value of the compressor intake superheating TÜE, whereby then the control deviation of the evaporator discharge superheating TÜA is calculated by subtracting the setpoint value of the evaporator discharge superheating TÜA from the process value of the evaporator discharge superheating TÜA, in a method step 3, the control deviation of the compressor intake superheating TÜE and the control deviation of the evaporator discharge superheating TÜA are advantageously combined to form an total control deviation of the superheating, taking into account the weighting influence, wherein the weighting influence is a measure of the proportional combination of the individual control deviations, and in a method step 4, the calculated total control deviation of the superheating is processed in the controller (500), which controls the corresponding actuators of the refrigerant circuit, in particular the expansion valve (230) with the adjustable degree of opening and / or the compressor (210) with adjustable compressor speed, in such a way that in the adjusted case a total control deviation of the superheating equal to approximately 0 Kelvin is set.