EP3816543B1 - Method for controlling an expansion valve - Google Patents

Description

The present invention relates to a method for controlling an expansion valve of a refrigerant circuit with the features of the preamble of claim 1, a refrigerant circuit with the features of the preamble of claim 10 and a device with at least one such refrigerant circuit. WO-A-2019/020952 discloses a method and a refrigerant circuit with the pre-characterizing features of claims 1 and 10.

Refrigerant circuits known in the prior art, for example for heat pumps, refrigeration systems or air conditioners, include an evaporator, a compressor, a condenser, an expansion valve and a control device connected to the expansion valve in a signal-conducting manner for controlling the expansion valve. The evaporator, compressor, condenser and expansion valve are arranged one behind the other in series in a circulation direction of the refrigerant circuit and have a refrigerant flowing through them that circulates in the closed refrigerant circuit. A heat source acts on the evaporator in a known manner and causes heat to be introduced into the refrigerant in the evaporator and thus leads to an increase in the enthalpy of the refrigerant, so that the refrigerant vaporizes in the evaporator. The heat source can be the area around the evaporator, the ambient air of which surrounds the evaporator or is fed to the evaporator (for example in the case of an air heat pump). Another example of a heat source is water or another fluid, which is fed to the evaporator in a manner known per se via its own heating medium circuit, which is hydraulically decoupled from the refrigerant circuit and is therefore materially separate from it, in order to heat the refrigerant of the refrigerant circuit in the evaporator . In other words, the heat source is thermally connected to the evaporator and in the evaporator the refrigerant is supplied with heat from the heat source thermally connected to the evaporator (or its heat source medium, eg air or water) and the refrigerant evaporates while absorbing heat. In the compressor that follows in the direction of circulation (often also known as Compressor) the vaporized (i.e. gaseous) refrigerant is compressed, whereby the refrigerant is raised to a higher pressure and temperature level. The gaseous refrigerant is then passed on in the direction of the condenser with a correspondingly increased pressure and temperature. In the condenser (often also referred to as the liquefier), the gaseous, superheated refrigerant is cooled to a temperature at which the refrigerant liquefies and is thereby liquefied, with the release of heat. As it continues to flow through the refrigerant circuit, the liquefied refrigerant passes through the expansion valve, which represents a bottleneck in the refrigerant circuit. When passing this constriction in the form of the expansion valve, there is a rapid pressure drop in the refrigerant, since the refrigerant can expand after passing through the expansion valve. The drop in pressure is also accompanied by cooling of the refrigerant, which is fed back to the evaporator after the expansion valve and the cycle described starts again with at least partial evaporation of the refrigerant in the evaporator.

During the evaporation process, the refrigerant, which was previously brought to a low pressure level by the expansion valve, absorbs heat from the heat source (e.g. the environment). The refrigerant is (usually completely) vaporized and "superheated" by 5 to 15 K (degrees Kelvin). This so-called suction gas overheating (i.e. the increase in the gas temperature of the vaporized refrigerant above the saturation temperature) is required to protect the compressor from liquid hammer and lubricant-diluting aerosols. The suction gas overheating is therefore the temperature difference between the gas temperature of the vaporized refrigerant when it enters the compressor (so-called suction gas temperature) and the vaporization temperature. The evaporation temperature is the temperature at which the refrigerant can exist both as a liquid and as a gas and depends on the prevailing pressure. The evaporating temperature can be obtained from the pressure at a point between the valve outlet of the expansion valve and the Compressor inlet of the compressor can be calculated, or alternatively measured as a temperature after the expansion valve.

In most conventional heat pump and refrigeration systems - especially in small and medium-sized systems - so-called dry evaporation processes are used (simple dry evaporation).

The refrigerant is continuously expanded in the expansion valve, causing it to partially evaporate. The liquid-gas mixture then flows through the evaporator, in which heat from the heat source acting on the evaporator (or its heat source medium) is supplied to the refrigerant. The refrigerant initially evaporates essentially completely at constant pressure. After reaching the dew line of the refrigerant, the gaseous refrigerant is further heated approx. 5 to 15 K above the boiling temperature (suction gas overheating, so that the downstream compressor does not suffer any damage from liquid entry).

With conventional methods of controlling the expansion valve, the expansion valve controls the refrigerant mass flow and the pressure, so that the refrigerant at the compressor inlet has a certain suction gas superheat at all times. Insufficient or no suction gas superheat can cause damage to the compressor. In this case, the evaporation pressure must be reduced (ie the expansion valve closed). On the other hand, excessive suction gas superheat has a negative effect on the refrigeration cycle efficiency because the evaporating pressure is lower than necessary. With the help of electronic or thermal expansion valves, a fixed suction gas overheating (eg 5 K) is controlled in known control methods. The difference between the suction gas temperature (gas temperature of the vaporized refrigerant when it enters the compressor) and the vaporization temperature is used as the controlled variable.

Refrigerant circuits with a so-called internal heat exchanger or suction gas heat exchanger are also known, which are also operated with dry evaporation. A first fluid line of the internal heat exchanger is arranged between the condenser and the expansion valve (i.e. connects the condenser outlet with the valve inlet of the expansion valve) and a second fluid line of the internal heat exchanger is arranged between the evaporator and the compressor (i.e. connects the evaporator outlet with the compressor inlet). The refrigerant flowing through the first fluid line gives off heat to the refrigerant flowing through the second fluid line and thus heats the refrigerant before it enters the compressor.

The high-temperature liquid refrigerant exiting the condenser is routed through the internal heat exchanger (in its first fluid line) and is thereby cooled by a few Kelvin. This heat is used to further heat the already fully vaporized and slightly superheated refrigerant from the evaporator by passing it through the second fluid line of the internal heat exchanger. This means that the evaporation process can be operated with less overheating (< 5 K) without the compressor being damaged. The regulation of the expansion valve corresponds to that of the simple dry evaporation described above. In turn, the opening width of the expansion valve is controlled in order to maintain a certain suction gas superheat (difference between the suction gas temperature between the evaporator and the internal heat exchanger and the evaporating temperature).

A disadvantage of the known concept is that suction gas overheating (even if it is less) is necessary in the evaporator. This means that only small amounts of energy can be transferred in the internal heat exchanger. In addition, the suction gas temperature upstream of the compressor cannot be controlled, with excessively high suction gas temperatures at the compressor inlet leading to damage and overheating of the compressor. In addition, the temperature changes in the internal heat exchanger are strongly dependent on the operating conditions (e.g. partial load operation and pressure difference). For this reason, internal heat exchangers are usually only used in practice for small increases in the temperature of the refrigerant and the transfer surface is correspondingly small. So-called pipe-in-pipe heat exchangers or pipe spindles in liquid separators as combination devices are typical.

A refrigerant circuit can also each include more than one evaporator, internal heat exchanger, compressor or condenser. In the context of the present disclosure, the term “at least one” in connection with these components means that one instance or several instances of the respective component—arranged in parallel or one behind the other—is or are present. In the interest of easier readability, the components are often referred to in the singular below. In these cases, too, what is meant is that at least one instance of the designated component is present and several instances—arranged in parallel or one after the other—may be present. In the event that a refrigerant circuit includes several instances of a component (e.g. a refrigerant circuit with three evaporators and two compressors), the instances of the respective component are usually arranged in parallel (the three evaporators arranged in parallel would therefore represent the at least one evaporator and the two compressors arranged in parallel would represent the at least one compressor). There can also be use cases in which the instances of the respective component are arranged in series or mixed (some instances in parallel and some instances in series). It can also be provided that a refrigerant circuit includes more than one expansion valve. Provision can thus be made for two or more expansion valves to be present, which are arranged in parallel, with at least one of them being regulated. It is also possible that all expansion valves are controlled or that they are controlled in a staggered manner depending on the desired refrigerant mass flow.

The object of the invention is to avoid the disadvantages described above and to specify a method for controlling an expansion valve of a refrigerant circuit that is improved compared to the prior art and a refrigerant circuit that is improved compared to the prior art.

This object is achieved by a method having the features of claim 1 and by a refrigerant circuit having the features of claim 10. Advantageous embodiments of the invention are defined in the dependent claims.

The method according to the invention provides that the expansion valve is controlled as a function of a temperature difference between a heat source temperature of the heat source and the evaporation temperature of the refrigerant, which prevails in the area between the valve outlet of the expansion valve and the compressor inlet of the at least one compressor.

In contrast to conventional control methods, the suction gas overheating is not used as the controlled variable for controlling the expansion valve, but the temperature difference between a heat source temperature of the heat source and the evaporation temperature of the refrigerant, which prevails in the area between the valve outlet of the expansion valve and the compressor inlet of the at least one compressor, is used as the controlled variable. used. This ensures that the control system can react much more quickly.

The heat source temperature of the heat source can be the temperature of a heat source medium (eg air or water) of the heat source. However, the heat source temperature can also be a temperature that is dependent on a temperature of the heat source (or its heat source medium). For example, it can be a Act surface temperature of at least one evaporator, which changes depending on the temperature of the heat source or the heat source medium (eg ambient air that is fed to the evaporator or water of a heat medium circuit that is fed to the evaporator).

So the heat source temperature is a value that reflects the temperature of the heat source at the evaporator. The heat source temperature can be, for example, the inlet or outlet temperatures (e.g. if the heat source is water that is fed to the evaporator via a separate circuit) or surface temperatures on the evaporator (e.g. if the heat source is ambient air) as well as averaged or weighted values from these.

In the area between the valve outlet and the compressor inlet, the refrigerant has essentially a constant pressure, as a result of which the evaporation temperature of the refrigerant, which is directly related to the pressure, is also essentially constant in this area.

The evaporation temperature of the refrigerant can be measured after the refrigerant has exited the valve outlet of the expansion valve or can be calculated from a pressure of the refrigerant at a point between the valve outlet and the compressor inlet with the help of the vapor pressure curve (also called boiling curve).

It is irrelevant how many compressors are used and how they are operated (e.g. electric or thermal). Likewise, the compressor or compressors can have different power levels or variable-power control.

In addition, the proposed control concept is independent of the heat source used (which acts on the evaporator) or heat sink (which refrigerant in or on the condenser extracts heat) and the refrigerant circuit can also contain other parts and components that have no significant influence on the functioning of the control strategy. Examples of this are sight glasses, collectors, filters, non-return valves, additional expansion valves, additional subcoolers, intermediate vapor injection systems or components that enable switching to reversible operation.

The refrigerant can exit the at least one evaporator partially evaporated, saturated or superheated.

The refrigerant circuit includes at least one internal heat exchanger. This heat exchanger is often also referred to as a suction gas heat exchanger.

A condenser outlet of the at least one condenser is connected to a first internal heat exchanger inlet of the at least one internal heat exchanger and a first internal heat exchanger outlet of the at least one internal heat exchanger is connected to a valve inlet of the expansion valve. The first fluid line runs between the first internal heat exchanger inlet and the first internal heat exchanger outlet. An evaporator outlet of the at least one evaporator is connected to a second internal heat exchanger inlet of the at least one internal heat exchanger and a second internal heat exchanger outlet of the at least one internal heat exchanger is connected to a compressor inlet of the at least one compressor. The second fluid line runs between the second internal heat exchanger inlet and the second internal heat exchanger outlet. The second fluid line is materially separate from the first fluid line, but thermally coupled or connected to the first fluid line, so that heat can be released from the refrigerant flowing through the first fluid line to the refrigerant flowing through the second fluid line in a manner known per se.

The at least one internal heat exchanger can be designed as a tube-in-tube heat exchanger, as a plate heat exchanger, as a tube bundle heat exchanger or the like.

If there is an internal heat exchanger, the refrigerant flows through the refrigerant circuit as follows: starting from the valve outlet of the expansion valve, the refrigerant is introduced into the evaporator, in which it evaporates completely or partially due to the effect of heat from the heat source that is thermally connected to the evaporator or acts on the evaporator becomes. After exiting the evaporator, the refrigerant flows through the second fluid line of the internal heat exchanger, in which the refrigerant is further completely evaporated and heated. After exiting the internal heat exchanger or its second fluid line, the refrigerant flows into the compressor, in which it is compressed and further heated. After exiting the compressor, the refrigerant flows through the condenser, in which it liquefies while releasing heat. After exiting the condenser, the refrigerant flows completely or partially through the first fluid line of the internal heat exchanger and ensures that the refrigerant flowing through the second fluid line is heated in the internal heat exchanger. After leaving the internal heat exchanger or its first fluid line, the refrigerant flows to a valve inlet of the expansion valve and after the refrigerant has left the valve outlet of the expansion valve, the cycle begins again.

The evaporation of the refrigerant in a tube (eg an evaporator designed as a tube evaporator) goes through several phases, with the tube wall temperature being indirectly proportional to the heat transfer coefficient. Starting with a completely liquid refrigerant, so-called nucleate boiling occurs first and then film evaporation. During nucleate boiling and film evaporation, the heat transfer coefficient is generally very high high. However, this changes when the so-called dryout point is reached (also known as a boiling crisis). The liquid film tears off the pipe wall and the heat transfer is essentially only given by convective gas transport. Individual liquid droplets are present as an aerosol in the refrigerant flow. As soon as these have completely evaporated, the overheating phase begins. In this phase, the refrigerant flow heats up, which also reduces the driving force of the heat transport (the temperature difference). The position of the dryout point depends on the flow speed, geometry/orientation and heat flow density, but is usually between approx. 70% and 90% gas mass fraction.

After the dryout point (in the direction of further evaporation), the heat transfer coefficient reduces by one to two orders of magnitude compared to film evaporation. The limitation of the heat transfer coefficient means that a large part of the evaporator's heat exchanger surface is required for complete evaporation after the dryout point and, above all, for superheating of the refrigerant. However, these two process steps only contribute to a fraction of the total energy input. About 80% to 90% of the heat is transferred to the refrigerant in the nucleate boiling and film evaporation region. In contrast, only about 5% to 15% of the heat is transferred during aerosol vaporization and less than 5% of the heat is transferred by suction gas superheat. Conversely, this means that a large part of the energy can be transferred with a fraction of the heat exchanger surface of the evaporator or alternatively with a significantly higher evaporation temperature (increase in efficiency). The heat transfer in the evaporator therefore has a significant influence on the efficiency of the refrigeration cycle process. If the evaporation process is operated with gas fractions below the dryout point (approx. 70% to 90% gas mass fraction), the heat transfer can be greatly improved.

The proposed method for controlling the expansion valve enables optimal utilization of the internal heat exchanger, while at the same time the control system can be kept stable. It is possible to increase the liquid content of the refrigerant in the evaporator and move the dry-out point from the evaporator to the internal heat exchanger. The overheating process is completely relocated and parts of the evaporation process are relocated to the internal heat exchanger. As a result, the entire heat exchanger surface of the evaporator can be used for the evaporation process, which leads to an increase in the evaporation temperature (and thus to an increase in efficiency). The internal heat exchanger can not only increase the temperature of the suction gas (the gaseous refrigerant when it enters the compressor), but also enable the wet vapor to be evaporated after the actual evaporator. Thus, the heat transfer in the evaporator is improved, which greatly increases the efficiency of the system.

Provision can preferably be made for the refrigerant circuit to include a first temperature sensor, with the first temperature sensor preferably being arranged in a heat source medium of the heat source or on the at least one evaporator, with the first temperature sensor measuring the heat source temperature and reporting it to the control device. The first temperature sensor can be arranged, for example, on at least one evaporator and can measure the temperature of the ambient air as the heat source medium. It is also conceivable that the first temperature sensor measures a surface temperature of the at least one evaporator, which is dependent on the temperature of the heat source medium. The first temperature sensor can also be arranged in a circulation line of a heat medium circuit, via which, for example, water or an antifreeze mixture is fed into the evaporator as a heat source medium.

According to a preferred embodiment, it can be provided that the refrigerant circuit comprises a second temperature sensor, which measures a refrigerant temperature of the refrigerant after the refrigerant has exited the valve outlet of the expansion valve and before the refrigerant has entered the at least one evaporator and reports it to the control device, with the temperature from the second Refrigerant temperature measured by the temperature sensor corresponds to the evaporation temperature. In the area between the valve outlet and the compressor inlet, the refrigerant has essentially a constant pressure, as a result of which the evaporation temperature of the refrigerant, which is directly related to the pressure, is also essentially constant in this area. The temperature of the refrigerant at the valve outlet of the expansion valve therefore reflects the evaporating temperature of the refrigerant. The refrigerant has the evaporation temperature in the entire area between the valve outlet and the entry into the evaporator. Only in the evaporator and the subsequent internal heat exchanger does the temperature of the refrigerant rise above its evaporation temperature. If the second temperature sensor is arranged between the valve outlet and the at least one evaporator, then it can measure the evaporation temperature of the refrigerant directly. In other words, the refrigerant temperature measured in the area between the valve outlet and the entry into the evaporator corresponds to the evaporation temperature of the refrigerant at the pressure conditions in this area.

In a particularly preferred embodiment, it can be provided that the refrigerant circuit includes a pressure sensor, with the pressure sensor measuring a refrigerant pressure of the refrigerant at a point between the valve outlet and the compressor inlet and reporting it to the control device, with the control device preferably determining the evaporation temperature from the refrigerant pressure. The evaporation temperature is the temperature at which the refrigerant changes from the liquid phase to the gaseous phase. The evaporation temperature depends on the pressure and can be calculated from the refrigerant pressure using the vapor pressure curve (also known as the boiling curve). be determined. Especially when determining the evaporation temperature from the pressure of the refrigerant after the evaporator, the proposed control is significantly faster than the conventional suction gas overheating control, since measuring the pressure, in contrast to measuring the suction gas temperature before the compressor, has no significant dead time.

It is provided that the heat source temperature of the heat source acting on the at least one evaporator and the evaporation temperature of the refrigerant in the area between the valve outlet and the compressor inlet are determined, with an actual heat source rating being determined from the temperature difference between the heat source temperature and the evaporation temperature, the actual heat source rating being determined by Regulation of an opening width of the expansion valve of a predetermined or specifiable target heat source degree is tracked. Provision can also be made for two or more expansion valves to be present which are arranged in parallel, at least one of which is regulated. It is also possible that all expansion valves are controlled or that they are controlled in a staggered manner depending on the desired refrigerant mass flow. For example, only one of the expansion valves can be regulated up to a first predetermined or predeterminable refrigerant mass flow, with the other expansion valves initially remaining closed. When the first predetermined or predeterminable refrigerant mass flow is reached, a further expansion valve can be regulated in order to be able to further increase the throughput of refrigerant. In this sense, further threshold values for the refrigerant mass flow can be predetermined or can be predetermined in order to achieve a desired staggering of the refrigerant mass flow by including further regulated expansion valves.

Instead of the previously usual suction gas overheating, the so-called heat source gradient between the heat source temperature of the heat source and the evaporation temperature (e.g. the evaporator inlet temperature of the refrigerant after the refrigerant has exited the valve outlet of the expansion valve or determination via evaporation pressure) is used as a controlled variable. The respective current actual value of the heat source rating (actual heat source rating) is determined and tracked to a specified or specifiable setpoint (target heat source rating).

The heat source temperature can be measured in the heat source medium or on the evaporator (e.g. a surface temperature of the evaporator, an air temperature of the ambient air in the area of the evaporator or the water temperature of water supplied to the evaporator in a heat medium circuit when it enters the evaporator or when it leaves the evaporator). The evaporation temperature of the refrigerant can be measured, for example, at the evaporator inlet or can be calculated from a measured refrigerant pressure of the refrigerant before the refrigerant enters the at least one compressor.

The opening width of the expansion valve is changed continuously (time-continuously or time-discretely) in such a way that the actual degree of heat source matches the desired degree of heat source. In other words, the opening width of the expansion valve is regulated in order to achieve and/or maintain a predefinable or predefinable desired degree of heat source.

Provision can preferably be made for the control device to include a first control device, with the first control device determining a valve control value on the basis of a first control deviation between the target heat source degree and the actual heat source degree and reporting it to the expansion valve, with the expansion valve adjusting the opening width as a function of the valve control value.

The expansion valve can be a thermal valve or an electric or electronic valve, for example in the form of a stepping motor valve that changes the opening width with the help of an electromagnet.

The first controller can be a PID, PI, PD controller or the like.

The new control value for the expansion valve is generated from the comparison between the target value (target heat source rating) and the actual value (actual heat source rating). The opening width of the expansion valve controls the amount of refrigerant injected into the evaporator and thus has a direct influence on the evaporation pressure.

It is envisaged that the target heat source grade will be continuously adjusted.

In this way, the target heat source degree can be adjusted or set or specified continuously (time-continuously or time-discretely) so that on the one hand the compressor does not suffer any liquid hammer and on the other hand high suction gas temperatures upstream of the compressor are prevented.

It is provided that the control device comprises a further control device for preventing liquid refrigerant from entering the at least one compressor, with at least one measured or determined temperature of the refrigerant in the refrigerant circuit and/or at least one measured or determined pressure of the refrigerant in the refrigerant circuit being used the overheating state of the refrigerant is determined before or after the at least one compressor characterizing control actual value and the actual control value is tracked by controlling the desired heat source degree to a predetermined or predeterminable control setpoint.

The target heat source rating can be adjusted, for example, by determining the actual suction gas overheating of the refrigerant after the internal heat exchanger and before it enters the at least one compressor, with the target heat source rating being adjusted or set or changed depending on the actual suction gas overheating .

Provision can therefore preferably be made for a suction gas temperature of the refrigerant to be determined after the internal heat exchanger and before it enters the at least one compressor, with an actual suction gas overheating being determined from a temperature difference between the suction gas temperature and the evaporation temperature, the actual suction gas overheating being determined by controlling the setpoint -Heat source grade is tracked to a predetermined or specifiable target suction gas overheating.

It can be provided that the refrigerant circuit includes a third temperature sensor, which measures the suction gas temperature of the refrigerant after the internal heat exchanger and before entry into the at least one compressor and reports it to the control device, the control device including a second control device, the control device for determining the Actual suction gas superheat calculates the difference between the suction gas temperature and the evaporation temperature, with the second control device specifying the target heat source degree on the basis of a second control deviation between target suction gas superheat and actual suction gas superheat.

The second control device can in turn be a PID, PI, PD controller or the like.

The actual suction gas superheat is therefore the difference between the suction gas temperature and the evaporation temperature.

The target suction gas superheat can be a fixed stored value (e.g. 5 K) or can be dynamically specified as a variable depending on the operating conditions (e.g. 5 K at low evaporation temperatures and 10 K at high evaporation temperatures).

The second control device determines the desired degree of heat source and reports this to the first control device. For the first controller is thus the target heat source grade reported by the second control device is the target value for the control.

In other words, the second control device can ensure that the difference between the suction gas temperature and the evaporation temperature (evaporator inlet temperature) is regulated to the target value for superheating (target suction gas superheating) and thereby the target value for heat source grading (target heat source grading) continuously or discontinuously is adjusted.

The first control device can also be referred to as an inner cascade and the second control device can be referred to as an outer cascade. The basic principle of this control cascading is the division of the control system into an inner, very fast and precise control circuit (first control device) and an outer, more sluggish control circuit (second control device). The internal control loop regulates the expansion valve by comparing the heat source rating (comparison of the actual heat source rating with the target heat source rating). The outer control circuit adapts the target value of the heat source degree (target heat source degree) to the prevailing operating conditions by adjusting the overheating condition of the refrigerant upstream of the compressor. It regulates to the desired overheating state of the gas upstream of the compressor (target suction gas overheating) and dynamically specifies the target value in the form of the target heat source degree to the inner control loop. In principle, this results in an "approach" to the optimal operating conditions and at the same time a stable control for the inner control loop, which reacts quickly to short-term changes in operation.

As described above, instead of or in addition to suction gas superheat control as an external cascade, other concepts that fulfill the same task (preventing liquid refrigerant from entering the compressor) can alternatively be used as actual values, e.g additional control device for controlling the hot gas overheating. The hot gas overheating results from the temperature difference between the hot gas temperature (temperature at the compressor outlet) and the condensation temperature (condensing temperature of the refrigerant, which is calculated, among other things, via the pressure, measured at a point between the compressor outlet and the expansion valve inlet, using the vapor pressure curve of the refrigerant can be). A high discharge gas superheat is synonymous with a high suction gas superheat. The controller attempts to adjust a fixed or variable target hot gas superheat by adjusting the actual hot gas superheat. The desired hot gas overheating can be made dependent on the pressure difference (condensation pressure - evaporation pressure) and the compressor speed, for example. Another concept that can be used as an alternative to the suction gas overheating control is the control of the "minimum most stable signal". Only the suction gas temperature (temperature before the compressor inlet) is measured. As soon as this can no longer be kept stable, the minimum stable signal is reached. Any further increase in refrigerant flow through the expansion valve would cause liquid hammer in the compressor.

The outer cascade, which is used to determine the target heat source rating, does not necessarily have to consist of a classic control system. For example, it can also be provided that values for the overheating state of the refrigerant upstream of the compressor (ie the actual suction gas overheating) are continuously or discontinuously compared and the target heat source rating can be adjusted from the deviation.

In addition to the control cascade, further measured variables can optionally be implemented in the overall system (by adding further controller modules to the control device), for example to take into account the influence of various disturbance variables, such as compressor speed or capacity or subcooling temperature through a pre-control regulation. For example, the supercooling temperature (temperature of the refrigerant in front of the expansion valve), the compressor speed / compressor output or the fan speed in the form of a pilot control system (feed-forward) or a pilot control (feed-forward control) or another standard control method can also be implemented.

For example, when the compressor speed is reduced, the refrigerant mass flow and thus the opening width of the expansion valve can be reduced. In the control cascade already presented, this operational change is noticeable with a delay in an increase in the suction gas temperature in the outer cascade. To anticipate this, a change in the compressor speed can directly affect the setpoint of the brine grading (target brine grading). The same applies to the subcooling temperature and other influencing factors such as the speed of the heat source motor. The heat source motor is the device that transports the heat source medium of the heat source and brings it into thermal contact with the refrigerant in the evaporator (e.g. a fan for air as the heat source medium or a pump for water as the heat source medium).

Provision can therefore be made for the specified or specified target heat source grade to be changed by at least one change value, with the at least one change value depending on a temperature of the refrigerant upstream of the expansion valve and/or a compressor speed of the at least one compressor and/or a compressor output of the at least one compressor and/or a heat source motor speed of a heat source motor is determined. The heat source engine can generally be a turbomachine for the heat source medium of the heat source. For example, the heat source motor can be a fan that supplies the evaporator with ambient air as the heat source medium. The brine motor can also be a pump that supplies water or an antifreeze mixture as the brine to the evaporator.

According to a preferred embodiment, it can be provided that the refrigerant is only partially evaporated in the at least one evaporator, with the refrigerant being completely evaporated in the internal heat exchanger. The refrigerant, which is only partially evaporated in the evaporator, flows after exiting the evaporator through the second fluid line of the internal heat exchanger, in which the refrigerant is further fully evaporated and heated. This enables optimal utilization of the internal heat exchanger, while at the same time keeping the control system stable. The liquid content of the refrigerant in the evaporator is increased and the dryout point is moved from the evaporator to the internal heat exchanger. Parts of the evaporation process and the overheating process are therefore completely relocated to the internal heat exchanger. This allows the entire heat exchange surface of the evaporator to be used for the evaporation process before the dryout point, which leads to an increase in the evaporation temperature (and thus an increase in efficiency). The internal heat exchanger should not only increase the temperature of the suction gas, but also enable evaporation of the wet steam after the actual evaporator. Thus, the heat transfer in the evaporator is improved, which greatly increases the efficiency of the system.

In order to ensure stable, incomplete evaporation in the evaporator with subsequent re-evaporation and overheating in the internal heat exchanger, the described refrigeration circuit design with an internal heat exchanger is required, with the internal heat exchanger being designed for a comparatively high transmission capacity, in contrast to the internal heat exchangers or suction gas heat exchangers that are customary in practice should be. A plate heat exchanger is preferably used for this. The described control strategy is required, which ensures a stable overheating condition directly before or (alternatively) directly after the compressor. The lower the superheated state of the refrigerant, the higher the liquid fraction of the refrigerant at the evaporator outlet.

In addition, a mechanical or gravimetric separation of the liquid and gaseous refrigerant in the refrigerant-carrying components between the evaporator inlet and the internal heat exchanger inlet should be prevented. As a result, the liquid content of the refrigerant changes continuously before it enters the internal heat exchanger, i.e. not suddenly. This condition is necessary for stable control.

Protection is also sought for a refrigerant circuit with the features of claim 10 and a device with at least one such refrigerant circuit. Advantageous embodiments are specified in the dependent claims.

The refrigerant circuit comprises at least one evaporator, at least one internal heat exchanger, at least one compressor, at least one condenser, an expansion valve and a control device connected to the expansion valve in a signal-conducting manner for controlling the expansion valve, in particular according to a method according to one of claims 1 to 9, wherein a first Fluid line of the at least one internal heat exchanger is arranged between the at least one condenser and the expansion valve and a second fluid line of the at least one internal heat exchanger is arranged between the at least one evaporator and the at least one compressor, the at least one evaporator, the second fluid line, the at least a compressor, the at least one condenser, the first fluid line and the expansion valve being arranged one behind the other in series in a circulation direction of the refrigerant circuit and through which a refrigerant can flow .

In the refrigerant circuit according to the invention, it is provided that the refrigerant circuit comprises a first temperature sensor connected to the control device in a signal-conducting manner, with a heat source temperature of a heat source acting on the at least one evaporator being able to be measured by the first temperature sensor and reported to the control device, wherein the first temperature sensor is preferably arranged in a heat source medium of the heat source or on the at least one evaporator, wherein the refrigerant circuit has a temperature determination device connected to the control device in a signal-conducting manner for determining the evaporation temperature of the refrigerant, which prevails in the area between the valve outlet of the expansion valve and the compressor inlet of the at least one compressor , Includes, wherein the control device controls an opening width of the expansion valve depending on a temperature difference between the heat source temperature and the evaporation temperature of the refrigerant in the area between the valve outlet and the compressor inlet. The evaporation temperature can either be calculated using the evaporation pressure at a point between the valve outlet of the expansion valve and the compressor inlet, or it can be measured as the temperature at the outlet of the refrigerant from the valve outlet of the expansion valve.

The heat source acting on the at least one evaporator can be the environment surrounding the evaporator or the air from which is supplied to the evaporator (e.g. in the case of an air heat pump). Another example of a heat source is water or another fluid, which is fed to the evaporator in a manner known per se via its own heating medium circuit, which is hydraulically decoupled from the refrigerant circuit and is therefore materially separate from it, in order to heat the refrigerant of the refrigerant circuit in the evaporator . In other words, the heat source is thermally connected to the evaporator, and in the evaporator, heat is supplied to the refrigerant from the heat source thermally connected to the evaporator, and the refrigerant evaporates while absorbing heat.

The refrigerant circuit comprises at least one internal heat exchanger, wherein heat is transferred from the refrigerant flowing through the first fluid line of the at least one internal heat exchanger to the refrigerant through the second Fluid line of at least one internal heat exchanger flowing refrigerant can be released.

The at least one internal heat exchanger - also referred to as a suction gas heat exchanger - can not only increase the temperature of the suction gas (the gaseous refrigerant when it enters the compressor), but also enable evaporation of the wet vapor after the actual evaporator. Thus, the heat transfer in the evaporator is improved, which greatly increases the efficiency of the system.

According to the invention, it can be provided that the temperature determination device comprises a second temperature sensor arranged between the valve outlet and the at least one evaporator, with the second temperature sensor being able to measure the evaporation temperature and report it to the control device. The second temperature sensor thus measures a refrigerant temperature of the refrigerant after the refrigerant has exited the valve outlet of the expansion valve and before the refrigerant has entered the at least one evaporator. In this range, the measured refrigerant temperature corresponds to the evaporation temperature of the refrigerant.

It can also be provided according to the invention that the temperature determination device comprises a pressure sensor arranged between the valve outlet and the compressor inlet, the pressure sensor being able to measure a refrigerant pressure of the refrigerant and to report it to the control device, the control device being able to determine the evaporation temperature from the refrigerant pressure.

It is inventively provided that the control device from the temperature difference between heat source temperature and evaporation temperature determines an actual heat source degree and the actual heat source degree by controlling the opening width of the expansion valve of a predetermined or specifiable target heat source degree tracked. It is also envisaged that the control device continuously adjusts the desired heat source degree. It is also provided that the control device comprises a further control device for preventing liquid refrigerant from entering the at least one compressor, the control device being based on at least one measured or determined temperature of the refrigerant in the refrigerant circuit and/or at least one measured or determined pressure of the refrigerant in the refrigerant circuit, an actual control value characterizing the overheating state of the refrigerant before or after the at least one compressor is determined, and the actual control value is tracked by controlling the target heat source grade to a predetermined or predeterminable control setpoint.

It can be provided that the control device includes a first control device that determines a valve control value in relation to the opening width on the basis of a first control deviation between the target heat source degree and the actual heat source degree and reports it to the expansion valve.

The expansion valve sets the opening width depending on the valve control value.

The expansion valve can be a thermal valve or an electric or electronic valve, e.g. in the form of a stepper motor valve that changes the opening width with the help of an electromagnet.

The first controller can be a PID, PI, PD controller or the like.

According to a particularly preferred embodiment variant, it can be provided that the refrigerant circuit includes a third temperature sensor, with a suction gas temperature of the refrigerant being measured by the third temperature sensor can be measured after the internal heat exchanger and before entering the at least one compressor and can be reported to the control device, the control device determining an actual suction gas overheating from a temperature difference between the suction gas temperature and the evaporation temperature, and the actual suction gas overheating by controlling the desired heat source grade to a predetermined or specifiable target -Suction gas overheating tracks.

To determine the actual suction gas overheating, the control device calculates the difference between the suction gas temperature and the evaporation temperature.

Provision can preferably be made for the control device to include a second control device which determines the target heat source degree on the basis of a second control deviation between the target suction gas superheat and the actual suction gas superheat and reports it to the first control device.

The second controller can be a PID, PI, PD controller or the like.

The proposed device can be, for example, a heat pump, a refrigeration system or an air conditioner.

Fig. 1

eine schematische Darstellung einer Vorrichtung mit einem Kältemittelkreislauf gemäß dem Stand der Technik,

Fig. 2

einen im Kältemittelkreislauf gemäß Figur 1 durchgeführten Kreisprozess in einem Druck-Enthalpie-Diagramm,

Fig. 3

eine schematische Darstellung einer Vorrichtung mit einem Kältemittelkreislauf umfassend einen internen Wärmetauscher gemäß dem Stand der Technik,

Fig. 4

einen im Kältemittelkreislauf gemäß Figur 3 durchgeführten Kreisprozess in einem Druck-Enthalpie-Diagramm,

Fig. 5

eine schematische Darstellung einer Vorrichtung mit einem Ausführungsbeispiel eines vorgeschlagenen Kältemittelkreislaufs,

Fig. 6

das Regelschema für die Regelung des Expansionsventils des Kältemittelkreislaufs gemäß Figur 5,

Fig. 7

einen im Kältemittelkreislauf gemäß Figur 5 durchgeführten Kreisprozess in einem Druck-Enthalpie-Diagramm,

Fig. 8

eine schematische Darstellung einer Vorrichtung mit einem weiteren Ausführungsbeispiel eines vorgeschlagenen Kältemittelkreislaufs,

Fig. 9

eine schematische Darstellung einer Vorrichtung mit einem weiteren Ausführungsbeispiel eines vorgeschlagenen Kältemittelkreislaufs,

Fig. 10

das Regelschema für die Regelung des Expansionsventils des Kältemittelkreislaufs gemäß Figur 9,

Fig. 11

eine schematische Darstellung einer Vorrichtung mit einem weiteren Ausführungsbeispiel eines vorgeschlagenen Kältemittelkreislaufs,

Fig. 12

eine schematische Darstellung einer Vorrichtung mit einem weiteren Ausführungsbeispiel eines vorgeschlagenen Kältemittelkreislaufs, und

Fig. 13

das um weitere Reglerbausteine ergänzte Regelschema für die Regelung des Expansionsventils des Kältemittelkreislaufs gemäß Figur 12.

Further details and advantages of the present invention are explained on the basis of the following description of the figures. show:

1

a schematic representation of a device with a refrigerant circuit according to the prior art,

2

one in the refrigerant circuit according to figure 1 cycle process carried out in a pressure-enthalpy diagram,

3

a schematic representation of a device with a refrigerant circuit comprising an internal heat exchanger according to the prior art,

4

one in the refrigerant circuit according to figure 3 cycle process carried out in a pressure-enthalpy diagram,

figure 5

a schematic representation of a device with an embodiment of a proposed refrigerant circuit,

6

according to the regulation scheme for the regulation of the expansion valve of the refrigerant circuit figure 5 ,

7

one in the refrigerant circuit according to figure 5 cycle process carried out in a pressure-enthalpy diagram,

8

a schematic representation of a device with a further embodiment of a proposed refrigerant circuit,

9

a schematic representation of a device with a further embodiment of a proposed refrigerant circuit,

10

according to the regulation scheme for the regulation of the expansion valve of the refrigerant circuit figure 9 ,

11

a schematic representation of a device with a further exemplary embodiment of a proposed refrigerant circuit,

12

a schematic representation of a device with a further exemplary embodiment of a proposed refrigerant circuit, and

13

the control scheme for controlling the expansion valve of the refrigerant circuit, which has been supplemented by additional controller modules figure 12 .

figure 1 shows a schematic representation of a device 19 with a refrigerant circuit 2 according to the prior art and figure 2 shows a cycle process carried out in the refrigerant circuit 2 in a pressure-enthalpy diagram or log-ph diagram.

The device 19 can be, for example, a heat pump, a refrigeration system or an air conditioner. The refrigerant circuit 2 comprises an evaporator 3, a compressor 4, a condenser 5, an expansion valve 1 and a control device 6, which is connected to the expansion valve 1 via a signal line 20 in a signal-conducting manner, for controlling the expansion valve 1.

The evaporator 3, the compressor 4, the condenser 5 and the expansion valve 1 are arranged in series in a circulation direction Z of the refrigerant circuit 2 and have a refrigerant K flowing through them, which circulates in the closed refrigerant circuit 2 in the circulation direction Z. A heat source 8 acts on the evaporator 3 in a known manner and leads to an increase in the enthalpy of the refrigerant K in the evaporator 3 , so that the refrigerant K is at least partially evaporated in the evaporator 3 . The heat source 8 can be ambient air that surrounds the evaporator 3 or is supplied to the evaporator 3 (for example in the case of a device in the form of an air heat pump). Another example of a heat source 8 is water or another fluid, which is fed to the evaporator 3 in a manner known per se via its own heat medium circuit, which is hydraulically decoupled from the refrigerant circuit 2 and is therefore materially separate from it, in order to cool the refrigerant K of the refrigerant circuit 2 to heat in the evaporator 3. In other words, the heat source 8 is thermally connected to the evaporator 3, and in the evaporator 3, heat is supplied to the refrigerant K from the heat source 8 thermally connected to the evaporator 3, and the refrigerant K evaporates while absorbing heat. In the compressor 4 (often also referred to as a compressor) following the evaporator 3 in the direction of circulation Z, the heated and at least partially evaporated (i.e. gaseous) refrigerant K is compressed, whereby the refrigerant K is raised to a higher pressure and temperature level. The gaseous refrigerant K is then forwarded in the direction of the condenser 5 with a correspondingly increased pressure and correspondingly increased temperature. In the condenser 5 (often also referred to as the condenser), the gaseous, overheated refrigerant K is cooled to a temperature at which the refrigerant K liquefies, and heat is thereby given off to a heat sink (not shown in detail) (e.g. ambient air or a condenser 5 connected circuit) liquefied. As it continues to flow through the refrigerant circuit 2, the liquefied refrigerant K passes through the expansion valve 1, which has a constriction in the Refrigerant circuit 2 represents. When this constriction in the form of the expansion valve 1 is passed, there is a rapid drop in pressure in the refrigerant K, since the refrigerant K can relax after passing through the expansion valve 1 . The drop in pressure is also accompanied by a cooling of the refrigerant K, which is fed back to the evaporator 3 after the expansion valve 1 and the cycle described starts again with at least partial evaporation of the refrigerant K in the evaporator 3 .

In the cycle process shown in the form of a so-called dry evaporation process, the refrigerant K is continuously expanded in the expansion valve 1, as a result of which it partially evaporates. The refrigerant K in the form of a liquid-gas mixture then flows through the evaporator 3, whereby the remaining liquid is first completely evaporated and finally overheated by 5 to 15 K (so-called suction gas overheating) before the gaseous refrigerant K reaches the compressor 4. The compressor 4 increases the pressure of the gaseous refrigerant K. In the condenser 5, the refrigerant K is liquefied by dissipating heat.

figure 2 shows an example of a cycle process C in the refrigerant circuit 2 according to FIG figure 1 in the well-known log-ph diagram. The specific enthalpy E (energy content of the refrigerant K) is plotted on the x-axis and the logarithmically scaled pressure P is plotted on the y-axis. To the left of the bell-shaped curve, the refrigerant K is liquid, to the right of it (i.e. to the right of the dew line T) it is completely gaseous. In between, the gas content increases continuously from left to right. The cycle process C is indicated by dashed lines and includes the process steps C1, C2, C3 and C4. If heat is supplied to the evaporator 3 (from the heat source 8 acting on the evaporator 3 or thermally connected to the evaporator 3), the refrigerant K initially evaporates completely at constant pressure in the evaporator 3 (process step C1). After reaching the dew line T, the then completely gaseous refrigerant K is further heated by approx. 5 to 15 K above the boiling point. This so-called suction gas overheating is necessary so that the compressor 4 does not suffers liquid shocks. In the compressor 4, the pressure and temperature of the refrigerant K are increased (process step C2). In the condenser 5, the refrigerant K condenses at a constant pressure while releasing heat (process step C3). In the expansion valve 1, the pressure of the refrigerant K drops (process step C4) and the cycle process C begins again with the process step C1.

In conventional control methods, the expansion valve 1 is controlled in order to achieve a specified desired value for the suction gas overheating. A second temperature sensor 13 and a third temperature sensor 16 are provided to determine the actual value of the suction gas overheating and are connected to the control device 6 in a signal-conducting manner. The second temperature sensor 13 records the temperature of the refrigerant K before it enters the evaporator 3 and reports this temperature to the control device 6 via a second sensor line 22. The third temperature sensor 16 records the temperature of the refrigerant K at the evaporator outlet before it enters the compressor 4 and reports this temperature to the control device 6 via a third sensor line 23. The control device 6 determines the actual value of the suction gas overheating by measuring the temperature difference between the temperature of the refrigerant K before it enters the compressor 4 (suction gas temperature) and the evaporation temperature (e.g. measured by the temperature of the Refrigerant K before entering the evaporator 3) is calculated. The expansion valve 1 is controlled via the signal line 20 in such a way that an opening width of the expansion valve 1 is adjusted so that the actual value of the suction gas overheating is regulated to the target value of the suction gas overheating. A (eg electronic or thermal) expansion valve 1 can thus be used to regulate to a fixed suction gas overheating (eg 5 K). The difference between the suction gas temperature and the evaporation temperature is used as the controlled variable. In other words, the expansion valve 1 regulates the refrigerant mass flow and the pressure, so that the refrigerant K has a specific suction gas overheating at the compressor inlet. Too small or no suction superheat can cause compressor 4 damage. In this case, the evaporation pressure must be reduced (ie the expansion valve 1 closed). On the other hand, excessive suction gas superheat has a negative effect on the refrigeration cycle efficiency because the evaporating pressure is lower than necessary.

figure 3 shows a device 19 according to FIG figure 1 , wherein the refrigerant circuit 2 additionally includes a heat exchanger 9 in the form of a so-called internal heat exchanger or suction gas heat exchanger and the third temperature sensor 16 is arranged between the evaporator 3 and the internal heat exchanger 9 and thus measures the suction gas temperature of the refrigerant K at the evaporator outlet. A first fluid line 10 of the internal heat exchanger 9 is arranged between the condenser 5 and the expansion valve 1 and a second fluid line 11 of the internal heat exchanger 9 is arranged between the evaporator 3 and the compressor 4, with heat from the refrigerant K flowing through the first fluid line 10 can be delivered to the coolant K flowing through the second fluid line 11 .

Specifically, a condenser outlet 24 of the condenser 5 is connected to a first internal heat exchanger inlet 25 of the internal heat exchanger 9 and a first internal heat exchanger outlet 26 of the internal heat exchanger 9 is connected to a valve inlet 27 of the expansion valve 1 . The first fluid line 10 runs between the first internal heat exchanger inlet 25 and the first internal heat exchanger outlet 26. An evaporator outlet 28 of the evaporator 3 is connected to a second internal heat exchanger inlet 29 of the internal heat exchanger 9 and a second internal heat exchanger outlet 30 of the internal heat exchanger 9 is connected to a compressor inlet 31 of the Compressor 4 connected. The second fluid line 11 runs between the second internal heat exchanger inlet 29 and the second internal heat exchanger outlet 30. The second fluid line 11 is materially separate from the first fluid line 10, however thermally coupled or connected to the first fluid line 10 so that heat can be released from the refrigerant K flowing through the first fluid line 10 to the refrigerant K flowing through the second fluid line 11 in a manner known per se.

The liquid refrigerant K emerging from the condenser 5 at a high temperature level is routed through the internal heat exchanger 9 and is thereby cooled by a few Kelvin. This heat is used to further heat the already fully evaporated and slightly overheated refrigerant K from the evaporator 3 . The evaporation process can thus be operated with less overheating (<5 K) without the compressor 4 being damaged. The suction gas temperature of the refrigerant K is measured with the third temperature sensor 16 between the evaporator 3 and the internal heat exchanger 9 . The evaporation temperature of the refrigerant K can be measured at the inlet of the evaporator 3 using the second temperature sensor 13 . The regulation of the expansion valve 1 corresponds to that of simple dry evaporation (see figure 1 ). The opening width of the expansion valve 1 is therefore in turn regulated in order to maintain a specific suction gas superheat (temperature difference between the suction gas temperature and the evaporation temperature).

figure 4 shows an example of a cycle process C in the refrigerant circuit 2 according to FIG figure 3 in the log ph diagram. In comparison with the cyclic process C of figure 2 (Refrigerant circuit 2 without internal heat exchanger 9) it can be seen that in process step C1 the overheating of the completely gaseous refrigerant K takes place after reaching the dew line T in the internal heat exchanger 9 (in its second fluid line 11) and accordingly in process step C3 the last cooling of the refrigerant K before the subsequent entry into the expansion valve 1 also takes place in the internal heat exchanger 9 (in its first fluid line 10).

figure 5 shows a device 19 with an exemplary embodiment of a proposed refrigerant circuit 2. The structure and connection of the expansion valve 1, evaporator 3, internal heat exchanger 9, compressor 4 and condenser 5 correspond to that in FIG figure 3 refrigerant circuit 2 shown. The refrigerant circuit 2 comprises a temperature determination device 18, which is connected to the control device 6 in a signal-conducting manner, for determining an evaporator inlet temperature of the refrigerant K after the refrigerant K has exited from a valve outlet 7 of the expansion valve 1. In the example shown, the temperature determination device 18 comprises a second temperature sensor 13 , wherein the evaporation temperature (corresponds to the evaporator inlet temperature) can be measured by the second temperature sensor 13 and can be reported to the control device 6 via a second sensor line 22 . The proposed refrigerant circuit 2 also includes a first temperature sensor 12, which is connected to the control device 6 in a signal-conducting manner and is arranged in a heat source medium of the heat source 8 or on the at least one evaporator 3, with the first temperature sensor 12 receiving a heat source temperature of a heat source 8 acting on the at least one evaporator 3 can be measured and reported to the control device 6 via a first sensor line 21 . The control device 6 is configured to control an opening of the expansion valve 1 depending on a temperature difference between the heat source temperature and the evaporation temperature. The control device 6 determines an actual heat source degree IW from the temperature difference between the heat source temperature and the evaporation temperature and tracks the actual heat source degree IW by controlling the opening width of the expansion valve 1 to a predetermined or specifiable target heat source degree SW. For this purpose, the control device 6 includes a first control device 15, not shown in detail here, which is configured to determine a valve control value in relation to the opening width on the basis of a first control deviation between the target heat source degree SW and the actual heat source degree IW and the expansion valve 1 via a signal line 20 to report.

figure 6 shows schematically the control scheme for the control of the expansion valve 1 of the refrigerant circuit 2 according to figure 5 . Instead of the suction gas overheating as in conventional control methods, the heat source gradient (difference between the heat source temperature and the evaporation temperature) is used as the control variable. On the basis of a first control deviation between the setpoint heat source degree SW and the actual heat source degree IW, the first control device 15 determines a valve control value V in relation to the opening width of the expansion valve 1 and reports this via the signal line 20 to the expansion valve 1, which represents the controlled system in the control scheme. A new actual heat source degree IW results from a changed opening width of the expansion valve 1, which is fed back in the control scheme to determine the first control deviation. The target heat source grade SW can be specified as a fixed value (fixed stored value). The actual heat source degree IW is determined by the control device 6 by calculating the temperature difference between the heat source temperature reported by the first temperature sensor 12 and the evaporator inlet temperature reported by the second temperature sensor 13 (corresponds to the evaporation temperature).

The control is to be explained using the following example. The target heat source degree SW should have a value of 5 K, the evaporation temperature should be -5 °C, the heat source temperature (e.g. air temperature) should be 1 °C and the actual value of the opening width of expansion valve 1 should be 40% at the beginning of the control . With these exemplary values, the actual heat source rating IW has a value of 6 K (heat source temperature minus evaporation temperature), ie the evaporation temperature could be raised by 1 K, which increases the refrigeration circuit efficiency. In the first control device 15, the deviation between the setpoint heat source degree SW and the actual heat source degree IW is processed, for example in a PID controller, and a new valve control value V for the expansion valve 1 is generated therefrom. In this case the expansion valve 1 opens to 42%, for example, so that more refrigerant K flows into the evaporator 3 and the pressure and thus the evaporation temperature rise. This reduces the actual heat source rating IW to 5.8 K and a new control cycle begins.

figure 7 shows an example of a cycle process C in the refrigerant circuit 2 according to FIG figure 5 in the log ph diagram. In comparison with the cyclic process C of figure 4 it can be seen that in this case significantly larger proportions of the process steps C1 and C3 take place in the internal heat exchanger 9. Since the dryout point in the proposed refrigerant circuit 2 is strongly shifted in the direction of the internal heat exchanger 9, the internal heat exchanger 9 not only increases the temperature of the suction gas, but also enables the wet vapor to be evaporated after the actual evaporator 3. Overall, this allows the refrigerant circuit to be 2 operate much more efficiently.

figure 8 shows a device 19 with a further exemplary embodiment of a proposed refrigerant circuit 2. In contrast to the refrigerant circuit 2 according to FIG figure 5 In this case, temperature determination device 18 includes a pressure sensor 14, with pressure sensor 14 being able to measure a refrigerant pressure of refrigerant K at a point between valve outlet 7 and compressor inlet 31 and reporting this to control device 6 via a pressure sensor line 32, with control device 6 being able to determine the evaporation temperature from the refrigerant pressure is. The expansion valve 1 is controlled in the same way as in the exemplary embodiment according to FIG Figures 5 and 6 .

figure 9 shows a device 19 with a further embodiment of a proposed refrigerant circuit 2. The refrigerant circuit 2 corresponds to the refrigerant circuit 2 of FIG figure 8 , supplemented by additional sensors and controller modules. Specifically, the refrigerant circuit 2 shown also includes a third temperature sensor 16, which is located between the internal heat exchanger 9 and the compressor 4 and thus measures the suction gas temperature of the refrigerant K after the internal heat exchanger 9 and before it enters the compressor 4 and reports it to the control device 6 via a third sensor line 23 . Contrary to what is shown, the temperature determination device 18 can also include a second temperature sensor 13 for directly determining the evaporation temperature from the evaporator inlet temperature (see figure 5 ).

The control device 6 includes a second control device 17, not shown in detail here. To determine the actual suction gas overheating IS, the control device 6 calculates the difference between the suction gas temperature reported by the third temperature sensor 16 and the evaporation temperature determined by the temperature determination device 18, and the second control device 17 gives the basis a second control deviation between a predefined or predefinable target suction gas superheat SS and the actual suction gas superheat IS, the target heat source degree SW, which is fed to the first control device 15 as a reference variable. In other words, the actual suction gas overheating IS is tracked by controlling the setpoint heat source degree SW to a predetermined or predeterminable setpoint suction gas overheating SS.

figure 10 shows schematically the control scheme for the control of the expansion valve 1 of the refrigerant circuit 2 according to figure 9 . The control scheme shows a 2-stage control cascade, in which the first control device 15 represents the inner cascade (inner control circuit) and the second control device 17 represents the outer cascade (outer control circuit). The inner cascade corresponds to the control scheme of figure 6 . On the basis of a second control deviation between the target suction gas superheat SS and the actual suction gas superheat IS, the second control device 17 specifies the target heat source degree SW, which is fed to the first control device 15 as a reference variable. As described above, the first control device 15 determines a valve control value V in relation to the opening width of the Expansion valve 1 and reports this via the signal line 20 to the expansion valve 1, which represents the controlled system in the inner cascade. A changed opening width of the expansion valve 1 results in a new actual heat source degree IW, which is fed back in the inner cascade to determine the first control deviation. A change in the opening width of the expansion valve 1 causes a changed refrigerant mass flow and thus a changed pressure and a changed temperature of the refrigerant K upon entry into the evaporator 3, which represents the controlled system of the outer cascade with the subsequent internal heat exchanger 9. After the refrigerant K has exited the internal heat exchanger 9, it has a new actual suction gas overheating IS, which is fed back in the outer cascade to determine the second control deviation.

The basic principle of this control cascading is the division of the control system into an inner, very fast and precise control circuit (first control device 15) and an outer, more sluggish control circuit (second control device 17). The inner control loop regulates the expansion valve 1 by comparing the heat source rating (comparison of the actual heat source rating IW with the target heat source rating SW). The outer control loop adjusts the target value of the heat source degree (target heat source degree SW) to the prevailing operating conditions by adjusting the overheating state of the refrigerant K upstream of the compressor 4 . It regulates to the desired superheating state of the gas upstream of the compressor 4 (target suction gas superheat SS) and dynamically specifies the target value in the form of the target heat source degree SW for the inner control loop. In principle, this results in an "approach" to the optimal operating conditions and at the same time a stable control for the inner control loop, which reacts quickly to short-term changes in operation. The input setpoint value for the outer cascade in the form of the setpoint suction gas superheat SS is intended to ensure on the one hand that the compressor 4 does not suffer any liquid hammer and on the other hand high suction gas temperatures prevent before the compressor 4. The setpoint suction gas superheat SS can be a permanently stored value or can be specified dynamically as a variable depending on the operating conditions.

• Soll-Wärmequellengrädung SW = 5 K
• Soll-Sauggasüberhitzung SS = 10 K
• Sauggastemperatur (dritter Temperatursensor 16) = 10 °C
• Verdampfungstemperatur (Temperaturermittlungsvorrichtung 18) = -5 °C
• Wärmequellentemperatur, z.B. Lufttemperatur (erster Temperatursensor 12) = 0 °C
• Istwert Öffnungsweite Expansionsventil 1 = 40 %
• Gasgehalt am Austritt aus dem Verdampfer 3 = 85 %
  1. 1) Die Ist-Wärmequellengrädung IW beträgt: Wärmequellentemperatur - Verdampfungstemperatur = 5 K. Das entspricht der Soll-Wärmequellengrädung SW, somit ist der innere Regelkreis eingeregelt.
  2. 2) Der äußere Regelkreis vergleicht die Soll-Sauggasüberhitzung SS mit der Ist-Sauggasüberhitzung IS = Sauggastemperatur - Verdampfungstemperatur = 15 K. Die Sauggastemperatur vor dem Verdichter 4 ist somit 5 K wärmer als benötigt. Das heißt der Kältemittelstrom kann erhöht, sprich das Expansionsventil 1 geöffnet werden, um somit die Verdampfungstemperatur zu erhöhen bzw. die Sauggastemperatur zu verringern.
  3. 3) In der zweiten Regeleinrichtung 17 wird aufgrund der Abweichung der Sauggasüberhitzung die Soll-Wärmequellengrädung SW auf 4,8 K reduziert.
  4. 4) Um die Ist-Wärmequellengrädung IW an die neue Soll-Wärmequellengrädung SW anzupassen wird der Ventilstellwert V des Expansionsventils 1 angepasst und das Expansionsventil 1 geöffnet. Dadurch steigt der Verdampfungsdruck und damit die Verdampfungstemperatur, in dem Beispiel auf -4,8 °C.
  5. 5) Die veränderte Verdampfungstemperatur führt zu einem verringerten Wärmestrom im Verdampfer 3, wodurch weniger Kältemittel K im Verdampfer 3 verdampft. Der Gasgehalt am Verdampferaustritt bzw. Eintritt in den internen Wärmetauscher 9 sinkt auf 83 %.
  6. 6) Im internen Wärmetauscher 9 muss somit mehr Kältemittel K verdampft werden. Da die übertragene Energiemenge in etwa gleich bleibt, reduziert sich die Sauggastemperatur. Die neue Ist-Sauggasüberhitzung IS beträgt z.B. 10 K, somit ist auch der äußere Regelkreis eingeregelt und das System ist vollständig eingeregelt.

The regulation is to be explained using the following example, which is based on the following specifications and assumptions:

• Target heat source rating SW = 5 K
• Target suction gas superheat SS = 10 K
• Suction gas temperature (third temperature sensor 16) = 10 °C
• Evaporation temperature (temperature detection device 18) = -5 °C
• Heat source temperature, eg air temperature (first temperature sensor 12) = 0 °C
• Actual opening width of expansion valve 1 = 40%
• Gas content at the outlet of the evaporator 3 = 85%
  1. 1) The actual heat source rating IW is: Heat source temperature - evaporation temperature = 5 K. This corresponds to the target heat source rating SW, so the inner control circuit is regulated.
  2. 2) The outer control circuit compares the target suction gas superheat SS with the actual suction gas superheat IS = suction gas temperature - evaporation temperature = 15 K. The suction gas temperature upstream of the compressor 4 is therefore 5 K warmer than required. This means that the refrigerant flow can be increased, ie the expansion valve 1 can be opened, in order to increase the evaporation temperature or reduce the suction gas temperature.
  3. 3) In the second control device 17, the setpoint heat source degree SW is reduced to 4.8 K due to the deviation in the suction gas overheating.
  4. 4) In order to adapt the actual heat source degree IW to the new target heat source degree SW, the valve control value V of the expansion valve 1 is adjusted and the expansion valve 1 is opened. This increases the evaporation pressure and thus the evaporation temperature, in the example to -4.8 °C.
  5. 5) The changed evaporation temperature leads to a reduced heat flow in the evaporator 3, as a result of which less refrigerant K evaporates in the evaporator 3. The gas content at the evaporator outlet or entry into the internal heat exchanger 9 drops to 83%.
  6. 6) In the internal heat exchanger 9, more refrigerant K must therefore be evaporated. Since the amount of energy transferred remains roughly the same, the suction gas temperature is reduced. The new actual suction gas overheating IS is, for example, 10 K, so the outer control loop is also adjusted and the system is completely adjusted.

figure 11 shows a device 19 with a further embodiment of a proposed refrigerant circuit 2. The refrigerant circuit 2 corresponds to the refrigerant circuit 2 of FIG figure 9 , but here the temperature determination device 18 includes a second temperature sensor 13 for direct measurement of the evaporation temperature and wherein the refrigerant circuit includes 2 additional sensors. Specifically, a second pressure sensor 33 for determining the pressure of the refrigerant K after it exits the compressor 4 and before it enters the expansion valve 1 and a fourth temperature sensor 34 for determining the temperature of the refrigerant K after it exits the compressor 4 and before it enters the condenser 5 provided. The signals of the second pressure sensor 33 are supplied to the control device 6 via a second pressure sensor line 35 and the signals of the fourth temperature sensor 34 are supplied via a fourth sensor line 36 . As an outer cascade of the control cascade, the hot gas overheating can be controlled on the basis of the hot gas temperature (determined by the fourth temperature sensor 34) compared to the condensation temperature (determined from the vapor pressure curve by measuring the pressure from the second pressure sensor 33) to specify the target heat source degree. The hot gas superheat control behaves similarly to the suction gas superheat control. A slight hot gas superheat leads to Liquid slugging in the compressor 3, excessive hot gas overheating to loss of efficiency. The hot gas overheating is adjusted to a fixed or changeable target hot gas overheating. A changeable desired hot gas overheating can be dependent on the evaporation pressure, the condensation pressure and the compressor speed, for example.

figure 12 shows a device 19 according to FIG figure 11 , supplemented by an additional valuation method and additional controller modules. A further sensor 37 is specifically provided for determining the output and/or speed of the compressor 4 . The signals of the sensor 37 are supplied to the control device 6 via a further sensor line 38 .

The compared to the control scheme of figure 10 other controller blocks are in the schematic control scheme of figure 13 shown. The added controller modules are a first pre-control 39 and a second pre-control 40. The first pre-control 39 can take into account the temperature of the refrigerant K at the inlet to the expansion valve 1, and the second pre-control 40 can set a compressor speed and/or Compressor performance of the compressor 4 (determined by the other sensor 37) are taken into account.

To simplify the presentation, the proposed refrigerant circuits were each shown with an evaporator, internal heat exchanger, compressor and condenser. However, a proposed refrigerant circuit can also include more than one evaporator, internal heat exchanger, compressor or condenser. In the event that a proposed refrigeration cycle includes multiple instances of a component (e.g., a refrigeration cycle with three evaporators and two compressors), the instances of the respective component are typically arranged in parallel. Provision can also be made for a proposed refrigerant circuit to include more than one expansion valve. So it can be provided that two or more expansion valves are present, which are parallel are arranged, at least one of which is regulated as proposed. It is also possible that all of the expansion valves are controlled as proposed, or that they are controlled in a staggered manner, as proposed, depending on the desired refrigerant mass flow.

Reference list:

1

expansion valve

2

Refrigerant circulation

3

Evaporator

4

compressor

5

capacitor

6

control device

7

valve outlet

8th

heat source

9

internal heat exchanger

10

first fluid line of the internal heat exchanger

11

second fluid line of the internal heat exchanger

12

first temperature sensor

13

second temperature sensor

14

pressure sensor

15

first control device

16

third temperature sensor

17

second control device

18

temperature detection device

19

contraption

20

signal line

21

first sensor line

22

second sensor line

23

third sensor line

24

condenser output

25

first internal heat exchanger inlet

26

first internal heat exchanger outlet

27

valve inlet

28

evaporator outlet

29

second internal heat exchanger inlet

30

second internal heat exchanger outlet

31

compressor inlet

32

pressure sensor line

33

second pressure sensor

34

fourth temperature sensor

35

second pressure sensor line

36

fourth sensor line

37

another sensor

38

further sensor line

C

cycle process

C1-C4

Process steps of the cycle process

E

specific enthalpy

K

refrigerant

P

Print

T

dew line

V

valve control value

Z

circulation direction

SW

Target Heat Source Grading

IW

Actual Heat Source Grading

ss

Target Suction Superheat

IS

Actual Suction Superheat

Claims (13)

1. Method for controlling an expansion valve (1) of a refrigerant circuit (2) comprising at least one evaporator (3), at least one internal heat exchanger (9), at least one compressor (4), at least one condenser (5), the expansion valve (1) and a control device (6), connected to the expansion valve (1) in a signal-conducting manner, for controlling the expansion valve (1), wherein a first fluid line (10) of the at least one internal heat exchanger (9) is arranged between the at least one condenser (5) and the expansion valve (1) and a second fluid line (11) of the at least one internal heat exchanger (9) is arranged between the at least one evaporator (3) and the at least one compressor (4), wherein a refrigerant (K) circulates in the refrigerant circuit (2), wherein the refrigerant (K) flows, in a circulation direction (Z) of the refrigerant circuit (2) starting from a valve outlet (7) of the expansion valve (1), through the at least one evaporator (3), the second fluid line (11) of the at least one internal heat exchanger (9), the at least one compressor (4), the at least one condenser (5), the first fluid line (10) of the at least one internal heat exchanger (9) and the expansion valve (1), wherein the refrigerant (K) is at least partially evaporated in the at least one evaporator (3) through the input of heat into the refrigerant (K) by a heat source (8) acting on the at least one evaporator (3), wherein the refrigerant (K) flowing through the first fluid line (10) releases heat to the refrigerant (K) flowing through the second fluid line (11) and the enthalpy of the refrigerant (K) is thus increased before entry into the at least one compressor (4), characterized in that the expansion valve (1) is controlled in dependence on a temperature difference between a heat source temperature of the heat source (8) and the evaporation temperature of the refrigerant (K), which prevails in the region between valve outlet (7) of the expansion valve (1) and compressor inlet (31) of the at least one compressor (4), wherein the heat source temperature of the heat source (8) acting on the at least one evaporator (3) and the evaporation temperature of the refrigerant (K) are determined in the region between valve outlet (7) and compressor inlet (31), wherein an actual heat source temperature difference (IW) is determined from the temperature difference between heat source temperature and evaporation temperature, wherein the actual heat source temperature difference (IW) is adjusted to a predefined or predefinable target heat source temperature difference (SW) by controlling an opening width of the expansion valve (1), wherein the target heat source temperature difference (SW) is continuously adjusted, wherein the control device (6) comprises a further control mechanism for preventing the entry of liquid refrigerant (K) into the at least one compressor (4), wherein, from at least one measured or determined temperature of the refrigerant (K) in the refrigerant circuit (2) and/or at least one measured or determined pressure of the refrigerant (K) in the refrigerant circuit (2), a control actual value characterizing the superheated state of the refrigerant (K) before or after the at least one compressor (4) is determined and the control actual value is adjusted to a predefined or predefinable control target value by controlling the target heat source temperature difference (SW).
2. Method according to claim 1, characterized in that the refrigerant circuit (2) comprises a first temperature sensor (12), wherein the first temperature sensor (12) is preferably arranged in a heat source medium of the heat source (8) or on the at least one evaporator (3), wherein the first temperature sensor (12) measures the heat source temperature and reports it to the control device (6).
3. Method according to claim 1 or 2, characterized in that the refrigerant circuit (2) comprises a second temperature sensor (13), which measures a refrigerant temperature of the refrigerant (K) after the refrigerant (K) exits the valve outlet (7) of the expansion valve (1) and before the refrigerant (K) enters the at least one evaporator (3) and reports it to the control device (6), wherein the refrigerant temperature measured by the second temperature sensor (13) corresponds to the evaporation temperature.
4. Method according to one of claims 1 to 3, characterized in that the refrigerant circuit (2) comprises a pressure sensor (14), wherein the pressure sensor (14) measures a refrigerant pressure of the refrigerant (K) at a point between valve outlet (7) and compressor inlet (31) and reports it to the control device (6), wherein the control device (6) preferably determines the evaporation temperature from the refrigerant pressure.
5. Method according to one of claims 1 to 4, characterized in that the control device (6) comprises a first control mechanism (15), wherein the first control mechanism (15) determines a valve set point (V) on the basis of a first control deviation between target heat source temperature difference (SW) and actual heat source temperature difference (IW) and reports it to the expansion valve (1), wherein the expansion valve (1) sets the opening width in dependence on the valve set point (V).
6. Method according to one of claims 1 to 5, characterized in that a suction gas temperature of the refrigerant (K) after the internal heat exchanger (9) and before entry into the at least one compressor (4) is determined, wherein an actual suction gas superheating (IS) is determined from a temperature difference between suction gas temperature and evaporation temperature, wherein the actual suction gas superheating (IS) is adjusted to a predefined or predefinable target suction gas superheating (SS) by controlling the target heat source temperature difference (SW).
7. Method according to claim 6, characterized in that the refrigerant circuit (2) comprises a third temperature sensor (16), which measures the suction gas temperature of the refrigerant (K) after the internal heat exchanger (9) and before entry into the at least one compressor (4) and reports it to the control device (6), wherein the control device (6) comprises a second control mechanism (17), wherein, to determine the actual suction gas superheating (IS), the control device (6) calculates the difference between the suction gas temperature and the evaporation temperature, wherein the second control mechanism (17) predefines the target heat source temperature difference (SW) on the basis of a second control deviation between target suction gas superheating (SS) and actual suction gas superheating (IS).
8. Method according to one of claims 1 to 7, characterized in that the predefined or predefinable target heat source temperature difference (SW) is changed by at least one change value, wherein the at least one change value is determined in dependence on a temperature of the refrigerant (K) before the expansion valve (1) and/or a compressor speed of the at least one compressor (4) and/or a compressor performance of the at least one compressor (4) and/or a heat source motor speed of a heat source motor.
9. Method according to one of claims 1 to 8, characterized in that the refrigerant (K) is only partially evaporated in the at least one evaporator (3), wherein the refrigerant (K) is completely evaporated in the internal heat exchanger (9).
10. Refrigerant circuit (2) comprising at least one evaporator (3), at least one internal heat exchanger (9), at least one compressor (4), at least one condenser (5), an expansion valve (1) and a control device (6), connected to the expansion valve (1) in a signal-conducting manner, for controlling the expansion valve (1), in particular in accordance with a method according to one of claims 1 to 9, wherein a first fluid line (10) of the at least one internal heat exchanger (9) is arranged between the at least one condenser (5) and the expansion valve (1) and a second fluid line (11) of the at least one internal heat exchanger (9) is arranged between the at least one evaporator (3) and the at least one compressor (4), wherein the at least one evaporator (3), the second fluid line (11), the at least one compressor (4), the at least one condenser (5), the first fluid line (10) and the expansion valve (1) are arranged one behind the other in series in a circulation direction (Z) of the refrigerant circuit (2) and a refrigerant (K) can flow through them, characterized in that the refrigerant circuit (2) comprises a first temperature sensor (12), connected to the control device (6) in a signal-conducting manner, wherein a heat source temperature of a heat source (8) acting on the at least one evaporator (3) can be measured by the first temperature sensor (12) and reported to the control device (6), wherein the first temperature sensor (12) is preferably arranged in a heat source medium of the heat source (8) or on the at least one evaporator (3), wherein the refrigerant circuit (2) comprises a temperature-determining device (18), connected to the control device (6) in a signal-conducting manner, for determining the evaporation temperature of the refrigerant (K) which prevails in the region between valve outlet (7) of the expansion valve (1) and compressor inlet (31) of the at least one compressor (4), wherein the temperature-determining device (18) comprises a second temperature sensor (13) arranged between the valve outlet (7) and the at least one evaporator (3), wherein the evaporation temperature can be measured by the second temperature sensor (13) and reported to the control device (6), and/or the temperature-determining device (18) comprises a pressure sensor (14) arranged between valve outlet (7) and compressor inlet (31), wherein a refrigerant pressure of the refrigerant (K) can be measured by the pressure sensor (14) and reported to the control device (6), wherein the evaporation temperature can be determined by the control device (6) from the refrigerant pressure, wherein the control device (6) controls an opening width of the expansion valve (1) in dependence on a temperature difference between the heat source temperature and the evaporation temperature of the refrigerant (K) in the region between valve outlet (7) and compressor inlet (31), wherein the control device (6) determines an actual heat source temperature difference (IW) from the temperature difference between heat source temperature and evaporation temperature and adjusts the actual heat source temperature difference (IW) to a predefined or predefinable target heat source temperature difference (SW) by controlling the opening width of the expansion valve (1), wherein the control device (6) comprises a first control mechanism (15), which determines a valve set point (V) relating to the opening width on the basis of a first control deviation between target heat source temperature difference (SW) and actual heat source temperature difference (IW) and reports it to the expansion valve (1), wherein the control device (6) continuously adjusts the target heat source temperature difference (SW), wherein the control device (6) comprises a further control mechanism for preventing the entry of liquid refrigerant (K) into the at least one compressor (4), wherein the control device (6) determines a control actual value characterizing the superheated state of the refrigerant (K) before or after the at least one compressor (4) from at least one measured or determined temperature of the refrigerant (K) in the refrigerant circuit (2) and/or at least one measured or determined pressure of the refrigerant (K) in the refrigerant circuit (2) and adjusts the control actual value to a predefined or predefinable control target value by controlling the target heat source temperature difference (SW).
11. Refrigerant circuit according to claim 10, characterized in that the refrigerant circuit (2) comprises a third temperature sensor (16), wherein a suction gas temperature of the refrigerant (K) after the internal heat exchanger (9) and before entry into the at least one compressor (4) can be measured by the third temperature sensor (16) and reported to the control device (6), wherein, from a temperature difference between suction gas temperature and evaporation temperature, the control device (6) determines an actual suction gas superheating (IS) and adjusts the actual suction gas superheating (IS) to a predefined or predefinable target suction gas superheating (SS) by controlling the target heat source temperature difference (SW).
12. Refrigerant circuit according to claim 11, characterized in that the control device (6) comprises a second control mechanism (17), which determines the target heat source temperature difference (SW) on the basis of a second control deviation between target suction gas superheating (SS) and actual suction gas superheating (IS) and reports it to the first control mechanism (15).
13. Device (19), in particular heat pump or refrigeration system or air-conditioning unit, with at least one refrigerant circuit (2) according to one of claims 10 to 12.

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