Method and correspondingly designed device for regulating a refrigeration circuit and compression refrigeration system

Abstract

The present invention relates to a method for operating a compression refrigeration machine (200) and a corresponding compression refrigeration machine (200), comprising a refrigerant, an evaporator (11, 240), a pressure boosting unit (12, 210), a condenser (13, 220), and a throttle element (15, 230). The method comprises calculating a control value for the throttle element (15, 230) as a function of process variables of the compression refrigeration machine (200) using a physical model containing coefficients, wherein the coefficients have at least one coefficient describing a universal physical dependency and one coefficient describing a tolerance-related dependency. The method further comprises the steps: a) adaptively correcting the coefficients describing tolerance-related dependencies, and b) adjusting the throttle element (15, 230) to the control value.

Description

[0001] The invention relates to a method and a correspondingly designed device for regulating a refrigeration circuit comprising a refrigerant, an evaporator, a pressure booster unit, a condenser, and a throttle device. The invention also relates to a corresponding compression refrigeration system.

[0002] In a compression refrigeration system, the refrigerant in the refrigeration circuit is essentially evaporated in the evaporator by removing heat from the medium to be cooled. The compressor increases the pressure and thus the temperature. The refrigerant is then liquefied again in the condenser, releasing heat. The throttle valve expands the refrigerant to the evaporation pressure.

[0003] Such compression refrigeration systems are used, for example, for heating rooms and preparing domestic water; both are referred to below as heat sinks.

[0004] The heat sink temperature is typically controlled by switching the compressor on and off or by modulating the compressor speed. Such methods are known, for example, from EP 1 355 207 A1 or DE 43 03 533 A1.

[0005] Furthermore, the control system is responsible for optimizing the efficiency of the evaporator and thus of the refrigeration circuit. The efficiency of the evaporator depends, among other things, on its filling level, i.e., on which part of the evaporator is filled with wet steam and which part of the evaporator is filled with superheated refrigerant. The higher the wet steam content, the lower the superheat and the better the efficiency.

[0006] However, if the entire evaporator is filled with wet steam and non-superheated wet steam reaches the compressor, this can lead to compressor damage. However, an insufficient refrigerant charge in the refrigeration circuit can also adversely affect the efficiency of the refrigeration circuit, making it impossible to ensure an efficiency-optimized wet steam charge level in the evaporator.

[0007] The preferred control variable for evaporator control is the refrigerant superheat at the evaporator outlet. This refrigerant superheat is preferably determined from the evaporator pressure p 0 and the temperature T 0h of the superheated refrigerant at the evaporator outlet. Temperature and pressure can be easily measured using suitable sensors. The difference between the evaporator outlet temperature T 0h and the evaporation temperature To, which is the temperature of the refrigerant during evaporation without superheat, is calculated and is the actual refrigerant superheat ΔT 0h-actual.

[0008] The evaporator superheat setpoint can be set as a fixed value for the refrigeration system. However, it is advantageous to adapt it to the operating point of the refrigeration system. This can be achieved using a characteristic curve map or automatic adaptation depending on dynamically changing variables in the refrigeration circuit. For example, if oscillations occur in the control loop, the superheat setpoint can be increased.

[0009] A superheat controller then determines the difference between the actual superheat value and the setpoint. The control variable, in this case the throttle valve, is adjusted based on the control deviation.

[0010] It has been shown that in practical operation, particularly with a wide range of permissible evaporator and condenser temperatures, the refrigeration circuit is exposed to widely varying operating conditions. From a control engineering perspective, the system to be controlled, i.e. the refrigeration circuit, varies greatly in gain and offset depending on the respective operating point. To set the target superheat, the control signal then also varies accordingly within a wide range. If such a refrigeration circuit is controlled, for example, with a conventional controller with preset controller parameters, precise control is not possible regardless of the respective refrigeration circuit operating point, since the controller does not adapt to the system, which varies depending on the operating point. Furthermore, in this case it is not possible to output an estimated control signal when the compressor starts up and at this time no superheat-relevant process data is yet available.

[0011] For example, from DE 10 2005 048 967 B4 a method for determining a control signal of the throttle element is known with the following steps: a) determining a first control value for the throttle element as a function of the deviation of an actual superheat of the refrigerant from a target superheat, b) determining the condenser pressure, c) measuring the evaporator pressure, d) forming a model that compares the refrigerant mass flow at the evaporator inlet with the refrigerant mass flow at the evaporator outlet, e) calculating a second control value for the throttle element based on the model from the evaporator pressure, the condenser pressure and refrigeration circuit-specific variables, f) determining a third control value for the throttle element by linking the first control value with the second control value and g) setting the throttle element to the third control value.

[0012] However, it has been found that such a method is too inaccurate for sensitive refrigeration circuits, for example those with an internal heat exchanger or a refrigerant with temperature glide such as R454C.

[0013] In particular, it has been found that the pilot control of the throttle device according to the known methods is too inaccurate for sensitive refrigeration circuits.

[0014] It is therefore an object of the present invention to provide a method and an associated compression refrigeration system which enables a more precise pre-control of the degree of opening of the throttle element in order to optimally control the superheating of the refrigerant at the evaporator outlet and thus achieve an optimization of the efficiency.

[0015] The object is achieved according to the invention by the features of claim 1.

[0016] Furthermore, the object is achieved by a correspondingly designed compression refrigeration system according to claim 9.

[0017] Preferred embodiments are defined in the subclaims.

[0018] The first control value for the throttle valve is determined from direct measured variables. For example, the first control value can be determined as a function of superheat. Preferably, a difference between the evaporator outlet temperature T 0h and the evaporation temperature To can be used to determine the superheat.

[0019] The evaporator pressure is a characteristic parameter of the refrigeration circuit, from which, like the condenser pressure, conclusions can be drawn about the condition of the refrigeration circuit. Based on fundamental equations that describe the refrigerant mass flow at the evaporator outlet and evaporator inlet, a model is developed according to the invention that generates a second control value for the throttle valve. If the first control value, which is determined from direct measured variables of the circuit, is linked to the second control value, a third control value is obtained for controlling the throttle valve, which optimally regulates the throttle valve.

[0020] The invention is therefore based on the assumption that the functions of the components in the refrigeration circuit, namely evaporator, compressor, condenser and throttle element, can be approximately described with the help of simplified physical description formulas.

[0021] From a few easily measurable process values, further, more difficult-to-determine process variables can then be calculated using the model, in particular the second control value of the throttle element. The control values, in particular the second control value, are preferably calculated by adaptively adjusting coefficients of the model. If this second value according to the invention is included as a basis for the inventive calculation of the third control value according to the invention, which is preferably determined at least partially from the first and second control values, for the superheat controller, the precalculated value advantageously results - the precalculated value being, for example, the control signal of the throttle element - as a well-approximated starting value for the control signal of the throttle element when the compressor starts.

[0022] When operating refrigeration circuits with internal heat exchangers, a refrigerant quantity adjusted to the operating point of the refrigeration circuit is stored in the heat exchangers. In particular, liquid or wet vapor refrigerant is stored in the low-pressure refrigerant paths of the evaporator and an internal heat exchanger (recuperator), depending on the operating point. Characteristic process variables are evaporation temperature, condensation temperature, and compressor speed.

[0023] The following relationship can be seen between the amount of refrigerant stored in these components and the operating point of the refrigeration circuit: (1) The lower the compressor speed, i.e. the speed of the pressure boosting unit, at constant evaporation and condensation temperatures, the more refrigerant is stored in the evaporator (and possibly also in the recuperator). This is due to the lower shear force, which, due to the refrigerant flow, drives refrigerant adhering to the surfaces of the heat exchanger towards the outlet connection, thus causing greater storage at low compressor speeds. Furthermore, with low heat energies transferred, the superheating distance in the heat exchanger is shorter, which leads to a lower gas fraction and an increased wet vapor fraction in the heat exchanger, and thus to greater storage. (2) The higher the temperature difference between the condensation temperature and the evaporation temperature, the more refrigerant is stored in the recuperator if the setpoint for the compressor inlet superheat is not varied depending on the operating point.

[0024] If a change in the operating point is requested with regard to the compressor speed, this also means that the optimal refrigerant charge in the components of the refrigeration circuit is adjusted to the changed operating point.

[0025] If the refrigerant flow rate is spontaneously adjusted by the expansion valve via a pilot control characteristic in relation to the current compressor speed, then in the idealized case there is no change in the filling quantity of the low-pressure side refrigerant paths of the evaporator and recuperator, at least due to the pilot control of the throttle device.

[0026] The result is that the refrigerant charge is not adjusted to the changed compressor speed when the compressor speed is changed, which • When the compressor speed increases, the refrigerant charge in the low-pressure side of the evaporator and recuperator is too high, which leads to undershooting of the superheat. • When the compressor speed decreases, the refrigerant charge in the low-pressure side of the evaporator and recuperator is too low, which leads to overshooting of the superheat.

[0027] According to the invention, the speed of the pressure boosting unit is now used to adjust the throttle element, so that the known disadvantages are avoided.

[0028] In a particularly preferred embodiment, the method according to the invention may comprise an expansion valve, a piston engine or a turbine as a throttle element.

[0029] Instead of the process variable of the evaporator pressure, the process variable of the evaporator inlet temperature can also be used for modeling if it is converted into the evaporator pressure using the refrigerant characteristic curve.

[0030] In the event of rapid disturbances in the system (rapid changes in the refrigeration circuit's operating point, e.g., due to temperature jumps), the second control signal according to the invention reacts immediately through adaptive adjustment. The control loop gain is defined by the advance calculation of the control value, and the controller can be adapted accordingly.

[0031] The advantages of this second control value according to the invention are that it reacts quickly to changes in ambient conditions, provides a good starting point when starting the compression refrigeration system, and serves as a reference for refrigerant deficiency detection, particularly through adaptive coefficient adjustment.

[0032] Based on the parameters determined in the process, it is also possible to detect a refrigerant shortage. This is detected if, during normal operation, the first control value, which is determined, for example, based on a universal physical parameter, particularly derived from the control deviation from superheat, exceeds a limit value for a parameterized period of time. Appropriate measures can then be initiated immediately to restore optimal operation of the compression refrigeration system as quickly as possible.

[0033] At system start-up, during the time window thereafter, and during emergency operation, the throttle valve can be set to the second control value. Immediately upon start-up, no suitable first control value—derived from the superheat control deviation—is available, so the third control value is calculated exclusively from the second control value. Preferably, immediately after start-up, the first control value is linked to the second control value to determine an optimized third control value.

[0034] A throttle valve offset, a refrigeration circuit-specific constant, and an exponent are included in the model as refrigeration circuit-specific variables. These are predefined and characteristic for each circuit, which simplifies integration into the model since they are only entered once.

[0035] In a preferred embodiment, the first control value and the second control value are linked by multiplication. This multiplicative link simplifies the operating point-dependent evaluation of the low-refrigerant detection. Furthermore, the multiplicative link takes into account the operating point-dependent system gain and results in a roughly constant gain throughout the entire control loop.

[0036] By preferably calculating the second control value for the throttle element based on the model of the evaporator pressure, the condenser pressure, and the speed of the pressure booster unit, particularly precise pre-control of the throttle element can be achieved. The additional control system then only has to compensate for small differences between the actual and target superheat, enabling stable operation of the refrigeration circuit. This is particularly advantageous in refrigeration circuits that operate with an internal heat exchanger or a refrigerant with temperature glide, as these react particularly sensitively to changes in the speed of the pressure booster unit.

[0037] The pressure boosting unit preferably comprises at least one variable-speed compressor. Different pilot control characteristics can be provided for single-compressor operation and dual-compressor operation. Particularly preferably, the pilot control characteristics of the degree of opening of the throttle element depend on the speed of the compressor(s) in operation.

[0038] Preferably, the speed of the pressure booster unit is included in the model in the form of an exponential dependence on the speed of the pressure booster unit.

[0039] An exponential function can particularly suitably approximate the physical behavior of the refrigeration circuit.

[0040] Preferably, the inclusion of the speed of the pressure boosting unit in the model is delayed, in particular by a first-order or further-order time element.

[0041] A disadvantage for the compensation of errors is the calculation of the pilot opening degree of the throttle valve based on a setpoint of the speed of the pressure booster unit, i.e., the value with which the control system controls the pressure booster unit. By including the delay, the delay of the actual value, which follows the setpoint, is advantageously taken into account.

[0042] The time delay between the actual value and the setpoint of the speed of the pressure booster unit (hereinafter referred to simply as compressor speed) caused, for example, by speed ramps of the pressure booster unit in the inverter causes • when a compressor speed increase is requested, the refrigerant charge quantity of the low-pressure side refrigerant paths of the evaporator and recuperator is increased in advance by leading the opening degree of the throttle element based on the compressor speed setpoint, which leads to an undershoot of the superheat, and • when a compressor speed reduction is requested, the refrigerant charge quantity of the low-pressure side refrigerant paths of the evaporator and recuperator is reduced in advance by leading the opening degree of the throttle element based on the compressor speed setpoint, which leads to an overshoot of the superheat.

[0043] To compensate for these dynamic processes in real refrigeration circuit systems, a dynamic model-based countermeasure is implemented, which adapts the compressor speed over time to the time constants of the refrigeration circuit and the delay times of the inverter communication.

[0044] Particularly preferably, the inclusion of the speed of the pressure booster unit in the model is based on a parameterizable transfer function. Particularly preferably, model parameters can be optimized during operation.

[0045] Preferably, the measured values / measurement parameters or the values ​​or parameters derived from them can be filtered, particularly to suppress rapid fluctuations.

[0046] The filtering can be implemented, for example, by a low-pass filter or an approximate low-pass filter.

[0047] Preferably, the method is carried out depending on the "heating" or "cooling" operating mode. The pre-control of the throttle valve's opening degree occurs in cooling mode as well as in heating mode, although other parameters may be suitable. Typically, only one compressor is operating in cooling mode, even in a system with multiple compressors.

[0048] Preferably, an offset of the throttle body, i.e. a relative degree of opening of the throttle body between the minimum stop and the beginning of the opening, is compensated by first subtracting the offset from the model calculations; then multiplying by the controller control factors; and then adding the offset again.

[0049] Preferably, the step of determining the first control value for the throttle element comprises the following substeps: measuring the evaporator pressure and the refrigerant temperature at the evaporator outlet; calculating the evaporation temperature from the evaporator pressure and refrigerant-specific data; determining an actual superheat of the refrigerant at the evaporator outlet from the difference between the refrigerant temperature and the evaporation temperature; determining the deviation of the actual superheat from a target superheat, and determining a first control value for the throttle element depending on the deviation of the actual superheat from the target superheat.

[0050] For simplicity, the evaporator pressure is also referred to as low pressure and the compressor pressure as high pressure.

[0051] This is particularly advantageous for discontinuous operation, for example switching on the compressor or switching from defrosting to heating, since gradients occurring at low pressure are dampened during the calculation.

[0052] During the time period between power-on and the first start-up of the pressure booster unit, the low pressure or filtered low pressure is set equal to the measured or calculated low pressure. If the control is in "heating" mode in evaporator operation or in "heating" mode in recuperator operation, the low pressure or filtered low pressure can be calculated particularly easily and advantageously.

[0053] Preferably, the evaporator pressure is filtered for the calculation of the second control value, wherein the filter in particular contains a first-order low-pass filter.

[0054] In particular, separate filtering can be performed for the "heating" and "cooling" operating modes. The purpose of the filtering is to suppress controller oscillations, i.e., the coupling of the pilot control characteristic calculation to controller activity-related fluctuations in the low pressure and thus also the resulting fluctuations in the pilot control calculation.

[0055] A first-order low-pass filter is preferably used. In each current iteration step, the difference between the low pressure and the low pressure determined in a previous time step is determined. Depending on the iteration interval and a filter value, a portion of this difference is added to the filtered low pressure of the last iteration step to obtain the newly calculated, filtered low pressure.

[0056] Preferably, a separate filtered low pressure value is determined in an analogous manner for the "cooling" operating mode.

[0057] In addition, during operation in other operating modes, for example "heating" or "cooling", the value "learned" in the last operating period should be able to be immediately applied again when re-entering the corresponding operating mode.

[0058] Example: Operation in "Heating" mode with a low pressure of 4 bar; the filtered low pressure adjusts to 4 bar after an adjustment period. If the operating mode is changed to "Standby," "Cooling," "Defrost," etc., a different low pressure value is set. When the operating mode is switched back to "Heating," the filtered low pressure of 4 bar is again applied as the starting value for the pre-control calculation, which is a very suitable starting value, at least when the heat source temperature has changed little. If a readjustment were to take place based on the pressure operated in the previous "Standby," "Cooling," or "Defrost" operating mode, a significant deviation would occur, at least during the start-up phase.

[0059] For special operating modes, such as defrost or standby, the throttle valve can be set to a fixed value. Setting the throttle valve to predetermined values ​​in special operating modes is useful from a refrigeration engineering perspective to ensure efficient operation and condition the refrigeration circuit for resuming normal operation.

[0060] In a further preferred embodiment, the condenser temperature of the compression refrigeration system is measured and the condenser pressure is calculated therefrom.

[0061] In a preferred embodiment, the condenser pressure is measured.

[0062] The refrigeration circuit-specific constant is used as a characteristic value in the modeling process. It can be determined in laboratory tests for the respective system or system type or, preferably, adjusted during normal operation.

[0063] The process steps are executed whenever the refrigeration circuit is controlled for optimal superheating. This preferably occurs regularly, especially continuously, during operation of the compression refrigeration system. For example, regular control involves taking measurements at predetermined time intervals and updating the control loop settings.

[0064] In a particularly preferred embodiment, a heat pump is used as a compression refrigeration system.

[0065] The invention is illustrated below using an exemplary embodiment and with reference to the accompanying drawings. They show: Fig. 1 shows a schematic representation of a compression refrigeration system according to a first embodiment; Fig. 2 shows a representation of the flow diagram of the method according to the invention; Fig. 3 shows a further representation of the flow diagram of the method according to the invention; Fig. 4 shows a schematic representation of a refrigeration machine according to a second embodiment; and Fig. 5 shows a schematic representation of a further compression refrigeration system according to an embodiment.

[0066] A block diagram of a compression refrigeration system is shown in Fig. 1 according to a first exemplary embodiment. A refrigeration system consists of the components evaporator 11, compressor 12, condenser 13, and throttle element 15, which are connected by a piping system through which the refrigerant is conducted. The compressor 12 is an example of a pressure boosting unit.

[0067] In the Fig. 1 to 3In the illustrated embodiment, an expansion valve 15 is used as the throttle element 15. Alternatively, a piston engine or a turbine can be used as the throttle element.

[0068] By applying heat at a low temperature level, a medium with a low boiling point ("refrigerant," today mostly ozone-neutral CFCs or natural substances, carbon dioxide, propane, etc.) is evaporated in the evaporator 11. The gaseous phase is then compressed in a compressor 12 and thereby heated. Under high pressure, the working fluid releases its heat for use at the condenser 13 (heating water, air flow) and condenses in the process. Through an expansion valve 15, the working fluid re-enters the sub-circuit at low pressure and is again fed to the evaporator 11, at whose outlet the evaporator pressure is determined by the measuring unit 16.

[0069] The temperature difference between the heat source and the refrigerant enables a heat flow to the evaporator 11. The refrigerant vapor is then drawn in and compressed by the compressor 12. The temperature of the refrigerant is "pumped" above the temperature level of the heat distribution. At the condenser 13, a temperature difference again exists, and a heat flow occurs for heat distribution. The high-pressure refrigerant cools down again, condenses, and is expanded via an expansion valve 15. The entire process occurs again, thus forming a cycle.

[0070] In the context of the present application, a refrigerant with temperature glide, for example R454C or combinations thereof, is particularly preferred.

[0071] Another preferred refrigerant used in the present application is a flammable refrigerant, in particular propane.

[0072] According to the invention, the refrigeration machine additionally comprises a determination unit 21 for determining a first control value W1 for the expansion valve 15 as a function of the deviation of an actual superheat of the refrigerant from a target superheat. Furthermore, a unit 14 for determining the condenser pressure and a measuring unit 16 for measuring the evaporator pressure are provided.A unit 17 for forming a model which compares the refrigerant mass flow at the evaporator inlet with the refrigerant mass flow at the evaporator outlet, a computing unit 18 for calculating a second control value W 2 for the expansion valve 15 based on the model from the evaporator pressure, the condenser pressure and refrigeration circuit-specific variables, in particular also the compressor speed, a determination unit 19 for determining a third control value W 3 for the expansion valve 15 by linking the first control value W 1 with the second control value W 2 and an actuating unit 20 for setting the expansion valve 15 to the third control value W 3 is also provided.

[0073] During the process of controlling a compression refrigeration system, unit 14 determines the condenser pressure, and measuring unit 16 measures the evaporator pressure at the evaporator outlet. The evaporation temperature is determined from the evaporation pressure. The formula for the calculation is an approximation of the dependencies found through measurements for the refrigerant used.

[0074] The current actual superheat of the refrigerant can be derived from the evaporation temperature of the refrigerant and the evaporator outlet temperature. Control is preferably based on the actual superheat. By comparing the actual superheat with the target superheat, a controller determines a first control value W 1 for the expansion valve 15, to which the opening angle of the expansion valve 15 can be set, thus regulating the refrigerant flow in the circuit. In the context of the present invention, however, the first control value W 1 is not set directly, but is used to determine the third control value. If the actual superheat is greater than the target superheat, the actuator should open, i.e. the first control signal becomes larger. If the actual superheat is less than the target superheat, the actuator should close, i.e. the first control signal becomes smaller.The controller can be designed as a P, PI, I or PID controller.

[0075] During the method according to the invention, in addition to the first control value W 1 , a second control value W 2 and a third control value W 3 are determined. For this purpose, a model is created in unit 17 that compares the refrigerant mass flow at the evaporator inlet with the refrigerant mass flow at the evaporator outlet. In the computing unit 18, a second control value W 2 for the expansion valve 15 is calculated based on the model from the evaporator pressure, the condenser pressure, and refrigeration circuit-specific variables, in particular the compressor speed.

[0076] In particular, the model additionally includes a speed of the compressor 12, which is particularly preferably implemented as an inverter-controlled, speed-regulated compressor. In other words, the second control value W 2 is influenced by the speed of the compressor 12. This is preferably an exponential dependency, with the base of the exponential function, in particular, depending on the compressor speed.

[0077] The determination unit 19 links the first control value W 1 with the second control value W 2 and in this way determines a third control value W 3 , to whose value the expansion valve 15 is set by means of the control unit 20.

[0078] In Fig. 2 shows how a control loop for evaporator superheating can be operated taking into account the pre-calculated control signal variable.

[0079] In block B1, the sensor signals from the refrigeration circuit are preprocessed and evaluated. For example, the sensor signals are filtered for interference (e.g., 50 Hz hum) using a low-pass filter, and the sensor time constants are compensated. Furthermore, the actual superheat is calculated from the evaporator outlet temperature and evaporator pressure, and the condenser pressure is calculated from the condenser temperature.

[0080] The input signals of block B1 are the evaporator pressure po, the compressor inlet temperature t ν1 , the evaporator outlet temperature t 02 , the condenser outlet temperature t c2 and the speed N of the compressor.

[0081] If no recuperator is installed in the refrigeration circuit, both temperatures (compressor inlet temperature t ν1 and evaporator outlet temperature t 02 ) are equal because the evaporator outlet is connected directly to the compressor inlet. If a recuperator is installed, it increases the refrigerant temperature as it passes through by dissipating heat, and the superheat can be controlled either before or after the recuperator, depending on the refrigeration circuit control design.

[0082] In block B3, the second control signal for the expansion valve is then calculated in advance using the process values ​​from block B1 using the refrigeration model.

[0083] In block B2, a signal oscillation detection is carried out, and together with block B5, the operating point of the refrigeration circuit is evaluated using the process values ​​from block B1 and a corresponding target superheat is determined.

[0084] In block B4, a controller, the control deviation of the superheat (the difference between the actual superheat ΔT Actual and the target superheat ΔT Target ) is fed in, and a control signal influenced by the control deviation is output. The first control value is calculated in this process step.

[0085] The second control signal is then combined with the first control signal, influenced by the control deviation, using the refrigeration model to form a total control signal. This is preferably done by multiplication. In this case, the factor formed by the controller output is 1, assuming no control deviation is present.

[0086] If a control deviation of the superheat occurs, the factor formed by the controller output is not equal to 1, and the predicted control signal is corrected accordingly using the refrigeration model. However, other mathematical combinations such as addition or weighting are also possible.

[0087] The precalculated control signal passes through block B6 for further processing. Here, the third control signal is adjusted, for example, to the control range limits of the expansion valve, and the control signal rise is also limited to avoid "overtaxing" the time constant of the refrigeration circuit. This means that, from a control engineering perspective, it is not necessary if the control speed of the control element exceeds the time constant of the refrigeration circuit by a multiple (for example, by a factor of 100). In this case, in the case of very short-term disturbances (EMC, measurement signal fluctuations, etc.), a very short-term fluctuating control signal would be calculated, which would be completely damped by the time constant of the refrigeration circuit, but would still load the control element.

[0088] This is particularly advantageous because the valve cannot be adjusted infinitely quickly and the refrigeration process cannot react arbitrarily quickly. Furthermore, Block 6 limits the control signal to the physical control range of the valve.

[0089] In block B7, the signal to be passed on to the actuator as a control signal is selected depending on the operating state. In control mode, the mathematically linked and limited control signal is passed on, as already explained. In particular, the third control signal, which results from the linking of the first and second control signals. Other operating modes include pump-down mode, an existing fault or defrost mode, as well as start-up or emergency operation. For example, the throttle element can be set to the second control value when the system starts up, during the time window thereafter, and during emergency operation. Immediately upon start-up, there is no suitable first control value - derived from the control deviation of the superheat - so the third control value is calculated exclusively from the second control value.

[0090] In special operating modes such as defrosting or standby, a fixed value is advantageously passed on to the actuator.

[0091] Block B8 is an evaluation unit used to evaluate the first control signal W1. If a refrigerant deficiency is detected, it is evaluated whether the first control signal exceeds a parameterized value (e.g., a value of >>1) for a minimum period of time in the control mode. In this case, a refrigerant deficiency is detected, displayed, and, if necessary, modified processing of the third control signal W3 is initiated in block B7, for example, emergency operation.

[0092] M denotes the actuator of the expansion valve, which is coupled to it.

[0093] In Fig. 3The flow diagram of the method according to the invention is shown schematically. The process variables used in the calculations are the evaporator pressure po, the condenser pressure pc, the compressor speed N, and the associated temperature variables. The compressor speed N can be either the absolute compressor speed or a relative compressor speed N rel relative to a maximum compressor speed of the compressor.

[0094] Based on the evaporator pressure po and the condenser pressure pc, a factor for the volumetric efficiency λ is calculated and fed into the lower block as an output, corresponding to the arrow on the right side of the figure to the lower block, as a relative mass flow. The lower block then has, for example, the relative compressor speed (or the compressor speed itself) as an input and calculates the mass flow itself from the exponents of the compressor speed, the relative mass flow as volumetric efficiency from the above block, and po, ​​and passes this on to the module at the bottom left as indicated.

[0095] As an example, simplified dependencies for the advance calculation of the control signal for an expansion valve of a refrigeration circuit of a compression refrigeration machine are described below.

[0096] The model is based on the physical background that in a refrigeration circuit in steady state at constant ambient conditions, the refrigerant mass flow at the evaporator inlet (from the expansion valve into the evaporator) is equal to the refrigerant mass flow at the evaporator outlet (from the evaporator to the compressor).

[0097] For modeling purposes, the two refrigerant mass flows are equated with their respective influencing variables measured in the refrigeration circuit. Furthermore, physical dependencies in the compressor and expansion valve are incorporated into the modeling.

[0098] The inventive concept is based on the fact that the influence of the compressor speed on the pre-control cannot be neglected, but must be taken into account. For this reason, the model contains an explicit, particularly exponential and delayed, dependence on the compressor speed.

[0099] The mass flow at the evaporator outlet depends on the compressor's delivery behavior. In the case of a variable-speed compressor, this is determined primarily by the compressor's speed, but also by the refrigerant pressures on the high- and low-pressure sides of the refrigeration circuit, as well as the resulting volumetric efficiency. The factor const 1 parameterizes the design-specific delivery rate for the refrigerant used in the compressor. This refers to a characteristic operating point; deviations are tolerated for other operating points, which can usually be found in a compressor data sheet or determined through laboratory measurements.

[0100] The formula for calculating the compressor intake mass flow from the evaporator pressure po and the condenser pressure pc, taking into account the volumetric efficiency curve, is: Ansaugmasse Verdichter = p o ⋅ 0,95 − p c p 0 ⋅ 0,1 ⋅ N Verdichter N max const 1 ⋅ const 2

[0101] The compressor type-dependent "compressor speed exponent", designated as "const 1" in the formula, describes the "curvature" of the relationship between refrigerant flow rate and compressor speed with otherwise constant process values ​​low pressure (LP); high pressure (HD).

[0102] This characteristic curve is usually such that, on the one hand, the ratio between mass flow rate and compressor speed drops at (very) low compressor speeds, because then the influence of leaks between sealing component pairs (piston / cylinder; rolling piston / sealing land; valve / valve seat) is more significant because the residence time is longer

[0103] On the other hand, the relationship between mass flow rate and compressor speed typically decreases at (very) high compressor speeds, because then the influence of pressure drops within the compressor (which are quadratic to the mass flow) is more significant, and also because the mass inertia for opening valves has a flow-reducing effect at faster cycles.

[0104] The characteristic curve is thus curved in such a way that the relationship between mass flow rate and compressor speed reaches its maximum, especially at medium speeds. This relationship can be fairly well approximated by an exponent over the relative compressor speed.

[0105] Alternatively, a functional dependency with an integrated component of the relative compressor speed as a basis and a constant, e.g. const 1 , as an exponent can be included.

[0106] The volumetric efficiency described below is a compressor-type-dependent geometric property that primarily affects the dead space of the compression chamber and is largely independent of the compressor speed. Liefergrad Verdichter = 0,95 − p c p 0 ∗ 0,1

[0107] The mass flow at the evaporator inlet depends on the mass throughput at the expansion valve. This is largely determined by the refrigerant pressures on the high- and low-pressure sides, as well as the mean opening cross-section of the expansion valve. In electronic expansion valves, the opening cross-section is controlled by a control or regulation system. The mass throughput of the expansion valve is parameterized for the refrigerant used in the factor const 2. This refers to a characteristic operating point; deviations are tolerated for other operating points.

[0108] The formula for calculating the mass flow at the nozzle of the expansion valve from the evaporator pressure po, the condenser pressure pc and the nozzle cross-section of the expansion valve, as well as a refrigeration system-specific constant const 3, is: Massenstrom Düse = p c − p 0 ⋅ Querschnitt Düse ⋅ const 3

[0109] In a refrigeration circuit, in steady state and under constant ambient conditions, the refrigerant mass flow at the evaporator inlet is equal to the refrigerant mass flow at the evaporator outlet. This results in: Massenstrom Düse = Ansaugmasse Verdichter

[0110] Equating the formulas for the mass flows and resolution according to the nozzle cross-section as the control variable results in: Querschnitt Düse = p 0 p c − p 0 ⋅ Liefergrad Verd . ⋅ N Verd . N max const _ Verd ⋅ const

[0111] The relationship between nozzle cross-section and control signal for an expansion valve with a conical nozzle needle is: Stellgröße Expansionsventil _ rel = 1 − 1 − Querschnitt Düse _ rel

[0112] The following describes how the nozzle cross-section can be replaced by a setting step depending on an exemplary valve characteristic curve with offset. Schritt EXV = 1 − 1 − p 0 p c − p 0 ⋅ Liefergrad Verd . ⋅ N Verd . N max const Verd ⋅ const + Offset EXV

[0113] The compressor volumetric efficiency factor and the relationship between nozzle cross-section and control signal for an expansion valve can be approximately integrated into Exp valve characteristic and const: Schritt EXV = p 0 p c − p 0 Exp ⋅ N Verd . N max const Verd ⋅ const + Offset EXV

[0114] The exponent then integrates the non-linearities of the following dependencies: • The dependence between (mass flow rate / compressor speed) as a function of the compressor speed and • The dependence on (expansion valve opening degree / mass flow rate) as a function of the mass flow rate.

[0115] This method of separating both exponents offers the developmental advantage that the exponents can be empirically optimized individually by separately varying operating parameters.

[0116] The exponent over the pressures is optimized by measurements with pressure variations at constant compressor speed, the exponent over the relative compressor speed when the latter varies.

[0117] The evaporator pressure and the condenser pressure are measured as process variables in the refrigeration circuit. In another embodiment of the invention, the condenser pressure can be calculated from the condenser temperature using refrigerant data.

[0118] The model includes the following fixed variables: the exponent Exp, the offset, and the refrigeration circuit-specific constant const, whereby these fixed variables depend on the respective components of a refrigeration circuit. The expansion valve offset, which describes the number of adjustment steps until the first opening, is included as a fixed variable. The exponent represents both the function of the nozzle cross-section via the control level and the function of the compressor's volumetric efficiency. The exponential function formed by the exponent approximates the refrigeration circuit component-specific functions.

[0119] The model is parameterized using a single refrigeration-circuit-dependent constant, const. This parameter represents the sum of the parameters in the compressor, condenser, expansion valve, and evaporator, which is determined by laboratory measurements or calculations. As a further advantageous embodiment, the refrigeration-circuit-specific constant, const, can be adapted during refrigeration circuit operation so that the calculation of the expansion valve steps becomes increasingly accurate based on the refrigeration circuit model.

[0120] A further advantageous embodiment of the method is to adapt the refrigeration circuit-specific constant const, determined, for example, in laboratory tests, during operation in such a way that the control signal obtained with the help of the refrigeration model, including the constant const, optimally adapts to the refrigeration process. In this case, the controller corrections required due to a control deviation are included in block B4 of the Figure 2minimal, the control is very precise.

[0121] Furthermore, the behavior of the closed control loop can be used to determine whether there is a refrigerant shortage. The relationships described in the refrigeration model are based on the assumption that a sufficient amount of refrigerant is available to operate the refrigeration circuit. If refrigerant escapes, e.g., due to leaks, or if the refrigeration circuit is insufficiently filled before commissioning or after component replacement, a control variable for the expansion valve that differs from the refrigeration model is required to adjust the superheat at certain operating points.

[0122] During operation, this manifests itself in the fact that the control signal (block B3) (w2) specified by the refrigeration model must be corrected to a greater extent by the controller (block B4). This, in turn, means that a much larger control signal is required to set the target superheat than predicted. This means that with multiplicative linking of the control signals, the controller output signal is significantly greater than 1 when the control loop is in steady state.

[0123] Particularly advantageous is an embodiment of the method according to the invention in which a lack of refrigerant is detected and corresponding measures are triggered when, in the steady state of the control loop in control operation, the control signal, which is the output signal w1 of the superheat controller, is detected above a specified value for a specified time.

[0124] The expansion valve in a preferred embodiment can be adjusted to each of the three control values ​​depending on the operating mode in order to optimally adapt the functioning to the respective operation.

[0125] Although according to the Fig. 1 to 3 a compression refrigeration system has been described, the principle of the first embodiment can also be applied to an absorption refrigeration machine.

[0126] Fig. 5 shows schematically and exemplarily a heat pump 100. The heat pump 100 essentially consists of a vapor compression system 200 forming a compression refrigeration system, which contains the following components: A compressor 210 for compressing the superheated refrigerant, a condenser 220, with a refrigerant-side condenser inlet 221 and a condenser outlet 222 for transferring 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 a building heating system or a system for hot water heating, advantageously a refrigerant collector 260, which is used as a refrigerant reservoir to compensate for different refrigerant quantity requirements depending on the operating conditions, a throttle element 230 designed as an expansion valve for expanding the refrigerant, an evaporator 240, with an evaporator inlet 241, for transferring source energy 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 thermal energy QQ from the ground or an air system which absorbs thermal energy QQ from the ambient air and transfers it to the vapor compression system 200 or any other heat source, a recuperator as an example of an internal heat exchanger 250 which is intended to transfer internal thermal energy Q i between the refrigerant flowing from the condenser 220 to the expansion valve 230 to the refrigerant flowing from the evaporator 240 to the compressor 210 and a refrigerant, in particular a flammable refrigerant or in particular a refrigerant mixture of at least two substances or two refrigerants, which flows in a flow direction S HD and S ND through the vapor compression circuit 200, wherein in the vapor compression circuit 200 refrigerant vapor is brought to a high pressure HD by the compressor 210 and is led to a condenser 220,wherein a high-pressure path with the high-pressure flow direction S HD is formed from the compressor 210 to the expansion valve 230. After the expansion valve 230 to the compressor 210, a low-pressure path with a low-pressure flow direction S ND of the refrigerant is formed, in which the evaporator 240 is located.

[0127] 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.

[0128] In the Fig. 5In the example shown, heat pump 100 is shown as a brine heat pump. Of course, similar considerations and advantages can be achieved with air-to-water heat pumps. In particular, with air-to-water heat pumps, a fan / ventilator is arranged as the heat source instead of the brine circuit with brine pump 330.

[0129] The compressor 210 serves to compress the superheated refrigerant from an inlet connection 211 to a compressor outlet pressure Pva at a compressor outlet temperature Tva 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 a variable speed. The compression unit can be designed as a rolling piston unit, scroll unit or otherwise. At the compressor outlet 212 the compressed superheated refrigerant at the compressor outlet pressure Pva is at a higher pressure level, in particular a high pressure HD, than at the inlet connection 211 with a compressor inlet pressure Pve, in particular a low pressure ND, at a compressed inlet temperature Tvε, which describes the state at the inlet connection 211 in a compression chamber.

[0130] In the condenser 220, the transfer of heat energy Q HD from the refrigerant of the vapor compression system 200 to a heating medium of the heat sink system 400 takes place. First, the desuperheating of the refrigerant takes place in the condenser 220, whereby superheated refrigerant vapor transfers part of its heat energy to the heating medium of the heat sink system 400 by reducing its temperature.

[0131] After the refrigerant vapor has been deheated, further heat transfer Q HD advantageously occurs in the condenser 220 through condensation of the refrigerant during the phase transition from the gas phase of the refrigerant to the liquid phase of the refrigerant. In this process, further heat Q HD is transferred from the refrigerant in the vapor compression system 200 to the heating medium of the heat sink system 400.

[0132] The high pressure HD of the refrigerant that occurs 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 operation of the compressor 210.

[0133] The heating medium, in particular water, is conveyed by means of a heating medium pump 410 through the heat sink system 400 in a direction SW through the condenser 220, whereby the heat energy Q HD is transferred from the refrigerant to the heating medium.

[0134] The downstream receiver 260 stores refrigerant exiting the condenser 220, which, depending on the operating point of the vapor compression circuit 200, should not be fed into the circulating refrigerant. If more refrigerant is fed from the condenser 220 than is passed through the expansion valve 230, the receiver 260 fills; otherwise, it empties or empties.

[0135] In the subsequent recuperator 250, which can also be referred to as an internal heat exchanger, internal heat energy Q i is transferred from the high-pressure refrigerant (HD), which flows from the condenser 220 to the expansion valve 230 in a high-pressure flow direction S HD , to the low-pressure refrigerant (ND), which flows from the evaporator to the compressor in a low-pressure flow direction S ND . In this case, the refrigerant flowing from the condenser to the expansion valve 230 is advantageously subcooled.

[0136] First, the refrigerant flows into the expansion valve through an expansion valve inlet 231. In the expansion valve 230, the refrigerant pressure is throttled from the high pressure (HD) to the low pressure (ND) by the refrigerant advantageously passing through a nozzle arrangement or throttle with an advantageously variable opening cross-section, whereby the low pressure advantageously corresponds approximately to the suction pressure of the compressor 210. Instead of an expansion valve 230, any other desired pressure-reducing device can also be used. Pressure-reducing pipes, turbines, or other expansion devices are advantageous.

[0137] The degree of opening of the expansion valve 230 is adjusted by an electric motor, typically embodied as a stepper motor, which is controlled by the control unit or regulator 500. The low pressure ND at the expansion valve outlet 232 of the refrigerant from the expansion valve 230 is controlled such that the resulting low pressure ND of the refrigerant during operation of the compressor 210 approximately corresponds to the evaporation pressure of the refrigerant at the heat source temperature T WQ . Advantageously, the evaporation temperature of the refrigerant will be a few Kelvin below the heat source temperature T WQ so that the temperature difference drives heat transfer.

[0138] In the evaporator, evaporation heat 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 transfers the source energy QQ to the vapor compression system 200.

[0139] The refrigerant flowing into the evaporator 240 reduces its wet vapor content by absorbing heat QQ as it flows through the evaporator 240 and advantageously leaves the evaporator 240 with a low wet vapor content or advantageously also as a superheated gaseous refrigerant. The heat source medium is conveyed through the heat source medium path of the evaporator 240 by means of a brine pump 330 in brine-water heat pumps or an outside air fan in air-to-water heat pumps, whereby the heat energy QQ is extracted from the heat source medium as it flows through the evaporator.

[0140] 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, wherein the refrigerant flowing from the evaporator 240 to the compressor 210 is further superheated.

[0141] This superheated refrigerant, which exits the recuperator 250 with a superheat temperature T Ke, is directed to the refrigerant inlet connection 211 of the compressor 210.

[0142] The recuperator 250 is used in the vapor compression circuit 200 to increase the overall efficiency as a quotient of the delivered heating power QH and the absorbed electrical power P e to drive the compressor motor.

[0143] 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.

[0144] The internal energy state of the refrigerant upon entering the evaporator 240 is reduced by this heat removal Q i , so that the refrigerant can absorb more heat energy QQ from the heat source 300 at the same evaporation temperature level.

[0145] Subsequently, after the evaporator outlet 242 from the evaporator 240, the heat energy Q i extracted in the high-pressure path is fed back to the refrigerant in the low-pressure path at low pressure ND and at a low-pressure temperature T Va in the recuperator 250. The supply of energy advantageously reduces the wet vapor content to a state without a wet vapor content, and then superheating occurs through further energy supply.

[0146] Furthermore, the following sensors are advantageously arranged to detect the operating state of the vapor compression system 200, with which a model-based pilot control is implemented, in particular to safeguard the operating conditions of the vapor compression system 200.

[0147] On the one hand, the process values ​​recorded by sensors advantageously provide safeguards with regard to permissible operating ranges of the components, such as in particular the compressor 210. On the other hand, model-based pre-controls, in particular a speed of the compressor 210 and / or a valve opening degree of the expansion valve, are carried out based on the sensor data, so that the controller only has to carry out minor corrections to compensate for a control deviation that is nevertheless smaller due to the pre-control: A high-pressure sensor 503 is advantageous for detecting the high pressure HD of the refrigerant at the compressor outlet 212 or between the compressor outlet 212 and the expansion valve inlet 231, a hot gas temperature sensor 504 is advantageous for detecting a hot gas temperature T HG of the refrigerant at the compressor outlet 212 or in the refrigeration circuit section between the compressor outlet 212 and the condenser inlet 221, an internal temperature sensor 506 is advantageous for detecting the internal temperature T Ie of the refrigerant between the high-pressure side internal recuperator outlet 252 of the refrigerant from the recuperator 250 and the expansion valve inlet 231. The internal temperature is advantageously also referred to as "recuperator outlet temperature high pressure path" and advantageously a recuperator internal temperature sensor 505.The recuperator internal temperature sensor 505 advantageously detects the condenser outlet temperature T FA of the refrigerant in the flow direction at the condenser outlet or the high-pressure side recuperator inlet and therefore the condenser outlet temperature T FA is advantageously measured by the recuperator internal temperature sensor 505. .

[0148] The following sensors are particularly advantageous for carrying out the method according to the invention: A low-pressure sensor 502 for detecting the low pressure ND of the refrigerant at the compressor inlet 211, or between the expansion valve 230 and the compressor inlet 211, an evaporator outlet temperature sensor 508 for detecting the evaporator outlet temperature T Va of the refrigerant at the evaporator outlet 242 or between the evaporator outlet 242 and the low-pressure side inlet of the refrigerant into the recuperator inlet 251 of the recuperator 250 and a low-pressure temperature sensor 501 advantageously measures a compressor inlet temperature or 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.

[0149] The process variable that significantly influences the overall efficiency of the vapor compression circuit 200, as the quotient between the heat output QH transferred by the vapor compression circuit 200 and the electrical power P e absorbed by the compressor 210, is the superheat of the refrigerant at the compressor inlet 211. However, to maintain permissible compressor operating conditions, it is advantageous to adhere to restrictions regarding the permissible superheat range of the refrigerant at the compressor inlet. Excessively low superheats, in particular, endanger the lubricating properties of the machine oil; excessive superheats, in particular, result in excessively high hot gas temperatures.

[0150] The superheat describes the temperature difference between the measured compressor inlet temperature T KE of the refrigerant and the evaporation temperature of the refrigerant at saturated vapor.

[0151] According to the invention, the compressor inlet superheat is preferably controlled in such a way that no condensate forms on components of the refrigeration circuit. For example, the refrigerant separator, which is to be protected, is located at the compressor inlet 211 on the compressor. This separator is difficult to enclose, so the temperature here should be kept high enough to prevent condensation. The problem of condensation does not usually occur on the high-pressure side.

[0152] If - for the purpose of a numerical example - an evaporation temperature level of approximately -10 °C is assumed and the temperature at the brine inlet 330 is approximately -10 °C, at the brine outlet 310 approximately -13 °C and at the compressor inlet 5 °C, the superheat would be 15 K.

[0153] Room temperature sensors and room humidity sensors are advantageous in many systems, allowing for precise determination of the condensation conditions of the air. For example, at 21 °C and 60% relative humidity, the condensation temperature is in the range of 13 °C. Therefore, as long as the pipe temperature is above 13 °C plus a buffer, e.g., 1 K, no condensation occurs.

[0154] Based on this numerical example, which is of course not limiting, a compressor inlet temperature of 5 °C is reached for a superheat of 15 °C. This is below the 13 °C specified as the condensation temperature for the current ambient conditions. Therefore, condensation is occurring. If the compressor inlet temperature is to be at least 14 °C, i.e., the condensation temperature plus the buffer, the superheat must be increased by 9 °C, i.e., a superheat of 24 °C must be maintained.

[0155] Limit values, in particular for superheating, determine the permissible superheating range of the components at the compressor inlet 211 depending on the operating point. Furthermore, there are also dependencies between the compressor inlet superheat T ÜE and the overall efficiency of the vapor compression circuit 200 or between the compressor inlet superheat T ÜE and a stability S of a control value R advantageous when controlling the compressor inlet superheat.

[0156] Advantageously, the degree of opening of expansion valve 230 is used as a control value for controlling the compressor inlet superheat T ÜE . The influence of the degree of opening of expansion valve 230 on the compressor inlet superheat T ÜE is as follows: Expansion valve 230 acts as a nozzle with an electric motor-adjustable nozzle cross-section, in which a needle-shaped nozzle needle is typically threaded into a nozzle seat by means of a stepper motor.

[0157] 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 degree of opening and advantageously a constant dependent on the refrigerant and a geometry of the expansion valve 230.

[0158] Since, when the heating medium temperature Tws is assumed to be constant at an operating point, the corresponding low pressure ND of the refrigerant upon entry into the expansion valve 230 can also be assumed to be constant, the degree of opening of the expansion valve 230 significantly influences only the low pressure ND, i.e. the outlet pressure from the expansion valve 230.

[0159] If the degree of opening of the expansion valve 230 is reduced, less refrigerant passes through the expansion valve 230 at constant high pressure HD and initially constant low pressure ND. However, since the compressor 210 continues to initially deliver the same refrigerant mass flow, less refrigerant is supplied through the expansion valve 230 in the high pressure flow direction S HD than is extracted by the compressor 210.

[0160] Since refrigerant vapor is a compressible medium, the low pressure ND on the low pressure side of the vapor compression circuit 200 then drops. As the low pressure ND drops, the mass flow of refrigerant through the compressor 210 drops approximately proportionally, since its delivery rate can be approximately described as volume / time, due in particular to the piston strokes, and a correspondingly reduced low pressure value ND is established, at which the refrigerant mass flow supplied through the expansion valve 230 is equal to the refrigerant mass flow discharged by the compressor 210.

[0161] If the degree of opening of the expansion valve 230 is increased, more refrigerant passes through the expansion valve 230 at constant high pressure HD and initially constant low pressure ND. However, since the compressor 210 continues to initially deliver the same refrigerant mass flow, more refrigerant is supplied to the low pressure side ND of the refrigeration circuit through the expansion valve 230 than is extracted by the compressor 210. Since the refrigerant vapor is a compressible medium, the low pressure ND on the low pressure side of the vapor compression circuit 200 increases. As the low pressure ND increases, the mass flow rate of the compressor 210 increases approximately proportionally, since its delivery rate can be approximately described as volume / time, and a correspondingly increased low pressure ND is established, at which the refrigerant mass flow supplied through the expansion valve 230 is equal to the refrigerant mass flow discharged by the compressor 210.

[0162] 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 heat transfer resistance of a heat transfer layer of the evaporator 240.

[0163] The heat transfer resistance between the heat source medium path of the evaporator and the refrigerant path of the evaporator is assumed to be approximately constant in a respective vapor compression circuit 200. Therefore, the magnitude of the heat transfer performance in the evaporator 240 depends significantly on the integral of the temperature differences of all surface elements of the heat transfer layer.

[0164] In order to be able 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 of the transfer layer of the heat exchanger, here the evaporator 240, as possible is greater than the temperature of the refrigerant at the respective surface element.

[0165] If the state of aggregation of the refrigerant as it flows through the evaporator 240 is saturated vapor, a refrigerant temperature is established which, due to the saturation vapor characteristic curve as a material property of the refrigerant, is a function of the low pressure ND of the refrigerant. Thus, by controlling the low pressure ND or an evaporation pressure, the evaporation temperature of the refrigerant as it flows through the recuperator 250 can be indirectly controlled.

[0166] The heat energy QQ , which is transferred from the heat source system to the refrigerant flowing through the evaporator 240, influences the state of aggregation of the refrigerant.

[0167] The wet vapor fraction in the saturated refrigerant vapor decreases at constant low pressure during heat transfer to the refrigerant. In the case of incomplete evaporation, the wet vapor fraction and thus the internal energy state of the refrigerant upon exiting the heat exchanger is a function of: Wet vapor fraction upon entry into the evaporator 240, refrigerant mass flow, transferred heat output QQ , and an enthalpy difference in the wet vapor region at the respective low pressure ND, whereby the refrigerant has a material constant as a function of pressure.

[0168] For complete evaporation, additional energy is supplied in the recuperator 250 to superheat the refrigerant beyond the saturated vapor state.

[0169] With the method, a corresponding refrigerant state is set at the outlet from the evaporator 240 under given operating conditions of the vapor compression circuit 200 as a function of the manipulated variable "degree of opening of the expansion valve 230".

[0170] In the steady state, the control system behavior of the "isolated" control system "Evaporator 240" has a moderate slope.

[0171] Advantageously, a refrigerant, in particular a refrigerant mixture, is used that exhibits a temperature glide; in particular, R454c is advantageously used. Advantageously, with a refrigerant mixture with a temperature glide, a relative opening degree change of the expansion valve actuator of 1% rel. is set with a superheat change at the refrigerant outlet from the evaporator, usually with a superheat change of approximately less than 1 K.

[0172] Flammable refrigerants are also advantageously used. In the event of deviations in the control variable of the expansion valve from usual control variables, in particular deviations of the control variable of the expansion valve from a usual or predetermined range of the control variable of the expansion valve, the controller advantageously compares the bandwidth with the current value of the control variable of the expansion valve and activates a safety action if the current value of the control variable is not within the usual or predetermined range or is outside the limit values. A safety action is advantageously the interruption of the pump output of a circulating pump in a heating circuit or an interruption of the flow of heating medium through the heating circuit by means of blocking devices such as valves, check valves or ball valves, as well as an alarm indication via sound, on a display or via a network on a mobile device.

[0173] After flowing through the evaporator 240, the refrigerant enters the low pressure path of the recuperator 250 at low pressure ND.

[0174] The aggregate state of the refrigerant when flowing into the recuperator 250 in a normal operating case is therefore advantageously either saturated vapor with a low vapor content between 0 and 20% or, in particular, also advantageously already superheated refrigerant.

[0175] With advantageously saturated vapor, a refrigerant temperature is established which, due to the saturation vapor characteristic of the refrigerant, is a function of the refrigerant pressure. When superheated refrigerant enters, the refrigerant temperature will reach a maximum value corresponding to the inlet temperature of the heat source medium.

[0176] In order to be able to transfer a sufficient amount of thermal energy from the refrigerant of the high-pressure side refrigerant path to the refrigerant of the low-pressure side refrigerant path in the recuperator 250, it must be ensured that the temperature of the refrigerant of the high-pressure side refrigerant path at high pressure HD in as many surface elements of the transfer layer of the recuperator 250 as possible is greater than the temperature of the refrigerant of the low-pressure side refrigerant path at low pressure ND at the respective surface element.

[0177] In a heating case, the corresponding temperatures of the heating system 400 of the vapor compression system 200 are higher than the corresponding temperatures of the heat source such as the ground or the outside air.

[0178] The thermal energy Q i , which is transferred from the refrigerant at high pressure HD of the high-pressure side refrigerant path to the refrigerant at low pressure in the low-pressure side refrigerant path of the recuperator 250, influences the state of aggregation of the refrigerant on the low-pressure side. The wet vapor fraction of the refrigerant flowing through the recuperator 250 on the low-pressure side at low pressure ND decreases during heat transfer to the refrigerant, and after complete evaporation, the refrigerant is advantageously superheated.

[0179] The internal energy state of the refrigerant, upon exiting the low-pressure path of the recuperator, is advantageously controlled depending on one or more of the following factors: Wet steam fraction upon entry into the recuperator 250, refrigerant mass flow, transferred heat output Q i , which is advantageously controlled depending on the temperature difference between the temperature of the refrigerant at high pressure HD in the high-pressure side refrigerant path and the temperature of the refrigerant in the low-pressure side refrigerant path at low pressure ND, and / or an enthalpy difference in the wet steam region at the respective low pressure ND.

[0180] This advantageously results in a corresponding refrigerant state being established at the outlet 252 from the recuperator 250 at low pressure ND, depending on the given operating conditions of the vapor compression circuit 200 and depending on the manipulated variable "degree of opening of expansion valve 230".

[0181] In the steady state, with regard to the control path slope of the "isolated" control path at the low pressure ND of the refrigerant in the low-pressure side path of the recuperator 250, a control path behavior with a high slope results, with an approximately constant internal energy state of the refrigerant at the inlet 251 into the low-pressure side ND path of the recuperator 250. With a particularly relative opening degree change of 1% of the expansion valve, a superheat change at the outlet of the refrigerant from the evaporator 230 of advantageously about 10 K or even more than 10 K is set.

[0182] Compared to the recuperator 250, a significantly higher heat transfer takes place in the evaporator 240 and thus a high driving temperature difference between the source medium and the refrigerant in the evaporator 240.

Claims

1. A method for operating a compression refrigeration machine (200), having a refrigerant, an evaporator (11, 240), a pressure boosting unit (12, 210), a condenser (13, 220) and a throttle element (15, 230), comprising the steps of: a) calculating a control value for the throttle element (15, 230) as a function of process variables of the compression refrigeration machine (200) using a physical model containing coefficients and characterized in that the coefficients have at least one coefficient describing a universal physical dependency and one coefficient describing a tolerance-related dependency, and the method further comprises the steps of: b) adaptively correcting the coefficients describing tolerance-related dependencies and c) setting the throttle element (15, 230) to the control value.

2. The method according to claim 1, wherein the coefficients describing universal physical dependencies are not adaptively corrected.

3. Method according to one of the preceding claims, wherein the control value for the throttle element (15, 230) is calculated on the basis of the model from an evaporator pressure, a condenser pressure and a speed of the pressure increasing unit (12, 210) as process variables.

4. Method according to one of the preceding claims, wherein an adaptation speed of the adaptive correction of the coefficients is dependent on operating conditions, in particular on a degree of opening of the throttle element (15, 230).

5. Method according to one of the preceding claims, wherein an adaptation speed is different for at least two of the coefficients, in particular increases for one of the coefficients with increasing degree of opening of the throttle element (15, 230) and decreases for another of the coefficients with increasing degree of opening of the throttle element (15, 230).

6. Method according to one of the preceding claims, wherein the determination of the control value of the throttle element (15, 230) comprises the determination of a first further and a second further control value.

7. The method according to claim 6, wherein the first further control value for the throttle element (15, 230) is determined as a function of the deviation of the actual superheat from the target superheat.

8. The method according to claim 6 or 7, wherein the second further control value for the throttle element (15, 230) is calculated on the basis of the model from an evaporator pressure, a condenser pressure and a speed of the pressure increasing unit (12, 210) as process variables.

9. Compression refrigeration machine with a refrigerant, an evaporator (11, 240), a pressure boosting unit (12, 210), a condenser (13, 220) and a throttle element (15, 230), further comprising: - a determination unit (21) for determining a first further control value for the throttle element (15, 230) as a function of the deviation of an actual superheat of the refrigerant from a target superheat; - a unit for determining the condenser pressure; - a measuring unit for measuring the evaporator pressure; - a unit (17) for forming a model which compares the refrigerant mass flow at the evaporator inlet with the refrigerant mass flow at the evaporator outlet; - a computing unit (18) for calculating a second further control value W2 for the throttle element (15, 230) based on the model from the evaporator pressure, the condenser pressure and refrigeration circuit-specific variables;- a determination unit (19) for determining a control value W3 for the throttle element (15, 230) by linking the first control value W1 with the second control value W2; and - an adjustment unit (20) for setting the throttle element (15, 230) to the third control value W3.; 10. Compression refrigeration machine according to claim 9, wherein the pressure increasing unit (12, 210) is a compressor.

11. Compression refrigeration machine according to claim 9 or 10, wherein the throttle element (15, 230) is an expansion valve or a piston engine or a turbine.

12. Compression refrigeration machine according to claim 9 to 11, wherein the refrigeration circuit-specific variables include a compressor speed.

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