DE102022127011A1 - Heat pump device for energy-efficient generation of process heat, drying device for drying a material to be dried and method for operating a heat pump device - Google Patents

Abstract

The invention relates to a heat pump device for the energy-efficient generation of process heat with a cycle fluid, an evaporator and at least one compressor, wherein the heat pump device has a first mixing separator in a second pressure stage, wherein the first mixing separator has a first connection for the inlet of vaporous cycle fluid, a second connection for the inlet of condensed cycle fluid, a third connection for the outlet of vaporous cycle fluid and optionally a fourth connection for the outlet of condensed cycle fluid and a second compressor is connected downstream of the mixing separator, so that escaping vaporous cycle fluid can be compressed in the second compressor from the second pressure stage to a first pressure stage and by means of the one assigned heat exchanger, process heat can be indirectly transferred from the vaporous cycle fluid of the first pressure stage to the at least one assignable heat sink and that cycle fluid condensed in a heat exchanger can be returned to the first mixing separator of the second pressure stage, wherein at the same time vaporous cycle fluid can be introduced from the first compressor into the mixing separator. Furthermore, the invention relates to a drying device and a method for operating a heat pump device.

Description

The invention relates to a heat pump device for energy-efficient generation of process heat, wherein the heat pump device has a heat-absorbing side, a heat-emitting side, a cycle fluid and an evaporator on the heat-absorbing side, and a heat exchanger and at least one heat sink can be assigned to the heat pump device on the heat-emitting side and at least one heat source on the heat-absorbing side, wherein liquid cycle fluid can be indirectly heated by means of the evaporator by supplying heat from the assignable heat source and can be evaporated at a pressure stage of the evaporator, and the heat pump device has at least one first compressor downstream of the evaporator, wherein vaporous cycle fluid can be compressed from the pressure stage of the evaporator to a second pressure stage by means of the at least first compressor. The invention further relates to a drying device for drying a material to be dried by means of a heated process gas flow and a method for operating a heat pump device.

As part of the energy transition and the current crisis in the natural gas supply, heat pump processes are becoming the focus of process heat generation. While heat is released for heating residential buildings at low temperatures, industrial heating processes often require significantly higher temperatures. For example, industrial drying processes require large amounts of heat, sometimes several MW, at high temperatures of well over 100°C, often over 200°C, and are one of the largest emitters of greenhouse gases in the industrial sector, because the high temperatures have so far usually been generated by combustion processes of solid, liquid or gaseous carbon-based fuels.

When generating process heat through combustion processes or electrically operated heaters, a primary energy input is required that is approximately equal to the process heat output to be generated. In order to decarbonize industry, the primary energy input for process heat generation can and should be drastically reduced by replacing the combustion processes previously required to achieve temperatures of over 120°C to over 200°C as far as possible with suitable, highly efficient high-temperature heat pumps with electric drive.

In general, heat pump processes for generating process heat can be characterized by the heat output coefficient (COP _h) as the quotient of the heat output in relation to the drive power, the required temperatures on the heat-absorbing side, the achievable temperatures on the heat-emitting side and the resulting temperature difference between heat absorption and heat emission. To date, the development of economical high-temperature heat pumps with a heat emission above, for example, 200°C has failed due to the temperature difference being too low, since the heat output coefficient decreases with increasing temperature difference due to thermodynamic relationships in the cycle. To date, high heat source temperatures on the heat-absorbing side have therefore been required for the operation of known cycle processes, which in themselves already represent high-quality heat sources with a correspondingly uneconomically high price or are not even available as waste heat sources and would have to be manufactured.

The value of process heat is usually measured according to its temperature level. In a heat pump cycle, the temperature level of the released process heat is typically achieved by compressing the cycle fluid. The specific drive work for compression corresponds to the expenditure, i.e. the destruction, of specific exergy. The level of specific exergy destruction in turn correlates with the minimum compression pressure that must be overcome so that the cycle fluid reaches the required temperature so that process heat can be transferred at the highest sink temperature.

The higher the highest sink temperature required for transferring process heat, the more specific exergy is destroyed and the higher the quality of the process heat generated. Accordingly, less specific exergy destruction is required to generate lower quality process heat than for higher quality process heat.

Heat pump processes can generally be divided into cycle processes without phase transition of a cycle gas and those with phase transition of a cycle fluid from liquid to gaseous in an evaporator on the heat-absorbing side, whose phase transition from gaseous to liquid occurs either by condensation in a condenser or by transcritical cooling in gas coolers.

Well-known high-temperature heat pumps without phase transition of the always gaseous cycle gas, such as rotary heat pumps from the Austrian company ecop Technologies GmbH from AT-4531 Neuhofen an der Krems (www.ecop.at), which are based on a counterclockwise Joule process, can heat sink temperatures of 150°C with relatively high heat output coefficients of around 4, but only achieve a very low temperature lift of around 55 K and are limited in the maximum heat output per heat pump due to their design. A multi-stage arrangement of such heat pumps to overcome an overall higher temperature lift is possible in principle, but fails due to significantly higher drive power with constant heat output and significantly reduced heat output coefficients and the associated low economic efficiency as well as the design-related limitation of the heat output that can be absorbed and released.

In heat pump cycles with a phase transition of the cycle fluid through evaporation, a distinction is made between heat pumps with condensation of the cycle fluid on the heat-emitting side and transcritically operated heat pumps with heat emission from the cycle fluid in gas coolers at supercritical pressure, for example with carbon dioxide (CO ₂ ) as the cycle fluid. The main difference between the two types of cycle with phase transition of the cycle fluid is that transcritically operated heat pumps transfer the heat output over an extended temperature range and not on a fixed temperature plateau, as is the case with condensation operation.

In the recent past, transcritical CO ₂ heat pumps have already been built up to sink temperatures of around 120°C, for example in EP 2 321 589 B1 described. In transcritically operated heat pumps, no condensation pressure can be achieved because no actual condensation takes place in a condenser. Nevertheless, a correspondingly high compression pressure to a pressure above the critical point of the cycle fluid for the heat transfer in a gas cooler from the supercritical cycle fluid to the heat sink to be supplied with the highest temperature must be achieved by applying mechanical work for compression up to the corresponding supercritical pressure of the heat-emitting side.

One advantage of transcritically operated heat pumps with supercritically operated gas coolers is that the cycle fluid in the gas cooler can only cool down to a maximum temperature above the inlet temperature of the heat sink and then still has a portion of its specific heat, which may still be sufficient to supply process heat to lower-quality heat sinks. This means that mechanical work is required to exceed the highest sink temperature, but no additional mechanical compression would be required for the subsequent supply to an inferior heat sink.

However, this advantage of transcritically operated heat pumps is also a significant disadvantage, since the gradual cooling of the supercritical cycle fluid through heat transfer in the gas cooler requires that the respective heat sink still receives sufficient process heat despite this gradual loss of temperature of the cycle fluid, because for this to happen, the inlet and outlet temperatures of the respective heat sink must be sufficiently low. If this is not the case, and the lowest sink temperature is still too high to cool the supercritical cycle fluid to temperatures far enough below the critical point, the transcritical process becomes uneconomical or even impossible, since the evaporator on the heat-absorbing low-pressure side of the transcritical cycle often cannot be operated correctly.

Another significant disadvantage of transcritical heat pumps with the refrigerant CO ₂ for the purpose of being able to supply process heat to sink temperatures of well over 120°C is the high compression pressures of well over 120 bar in combination with the high CO ₂ outlet temperature of well over 130°C, which poses as yet unsolved technical problems for the design of the necessarily oil-lubricated piston or screw compressors with regard to material strength and temperature resistance of the oil. The mere fact that the compressors have to be oil-lubricated because the frictional forces at the high pressures are far too high for oil-free compression requires a functioning oil management system so that subsequent units, such as gas coolers and evaporators, are not impaired in their effectiveness by the oil content carried in the cycle fluid.

An additional disadvantage of known heat pumps with coolants such as CO ₂ , NH ₃ and hydrocarbons is their limitation when heating the heat-absorbing side of heat sources significantly above 30-40°C. If this is made possible with great technical effort, then the technically achievable temperature increase is so small that no high temperature for process heat transfer to a heat sink of over 200°C can be achieved on the heat-emitting side.

In a heat pump cycle with condensation of the cycle fluid, the condensation temperature to be achieved for heat transfer must be sufficiently higher than the temperature of the heat sink to be supplied with the highest sink temperature. The highest required sink temperature and the selection of the cycle fluid with regard to its pressure-dependent condensation temperature therefore determine the required pressure in the condenser, i.e. the heat exchanger of the heat sink in which heat is released from the heat pump cycle, whereby the condensation pressure must usually be achieved by compressing the cycle fluid using mechanical work, i.e. the destruction of exergy.

The current development of new heat pumps with condensation of a cycle fluid has for some time been focused on the investigation of suitable cycle fluids with regard to achieving the highest possible condensation temperature at the lowest possible condensation pressure on the heat-emitting side and, at the same time, a high evaporation pressure at the lowest possible evaporation temperature on the heat-absorbing side.

However, many fluids suitable as cycle fluids, also known as refrigerants, which come as close as possible to these properties, generally have very poor environmental compatibility in terms of climate damage, ozone depletion, flammability or toxicity. Therefore, natural refrigerants such as CO ₂ , NH ₃ and less harmful but flammable hydrocarbons are increasingly being used. However, it should be noted that almost all refrigerants currently approved by the EU Commission have strong to very strong restrictions on the level of the maximum sink temperature and the achievable temperature rise.

These include, for example, industrial heat pumps with the natural refrigerant NH ₃ , which are usually operated with condensation of the cycle fluid and were already referred to as high-temperature heat pumps in the past when the maximum possible sink temperature was, for example, 80-90°C. This is due to the critical point of NH ₃ , which is 132.35°C and 113.53 bar, which means that temperatures of 80-90°C are already just below the critical point. At 200°C, NH ₃ is already supercritical and would no longer be condensable, but could only be used in a transcritical process under very high pressures.

Due to the relatively low critical point of NH _3, a pressure of over 62.55 bar must be overcome for its condensation at 100°C, for example. This high pressure level requires a high specific compressor work and increases the operating costs for compression. In addition, it makes both the design of oil-lubricated compressors and the design of heat exchangers complex and expensive. In addition, the specific condensation enthalpy of NH ₃ at 100°C is only 715.7 kJ/kg, which means that the resulting heating performance figures are at an uneconomically low level. A temperature level of 100°C is also not sufficient for many industrial processes.

For CO ₂ the critical point is even lower, namely at 31.06°C and 73.83 bar, so that CO ₂ can no longer be condensed outside the two-phase region above 31.06°C, but can only be used for heat transfer in a transcritical process above the critical pressure. Due to the high pressures involved, similar restrictions apply to the design and operation as for NH ₃ .

The aforementioned cycle fluids are gaseous under standard conditions and are very well suited to generating process cooling due to high vapor pressures, for example at 0°C of 34.85 bar for CO ₂ and 4.29 bar for NH ₃ , which underlines their good suitability as refrigerants. An advantage for both fluids is their relatively low specific volume, for example at an evaporation temperature of 54°C, which is 57 l/kg for NH ₃ and 5.6 l/kg for CO _2. This allows relatively small apparatus sizes even with high evaporation rates. In the supercritical range, the specific volumes of both fluids are even less than 10-20 l/kg, which is very advantageous for the size of heat exchangers, but here again the high supercritical pressure comes into play, which makes the construction of apparatus complex.

The use of steam as a cycle fluid is already known, particularly in power plant technology, as a clockwise power process for generating electricity, while electrically driven heat pumps with steam as a cycle fluid in a counterclockwise working process have not yet been named as such. However, when so-called water heat pumps are currently discussed, this usually means heat pumps that heat a water flow at least on the heat-emitting side and possibly also cool a water flow on the heat-absorbing side, but their cycle fluid is not water, but another fluid.

Widely known as a type of heat pump, but usually not constructed in the form of a conventional cycle, is the compression of water vapor, for example by means of the frequently used vapor compression to raise the condensation temperature of water vapor in order to make it available for heat transfer at a frequently slightly higher temperature level, as is the case with modern evaporation processes with mechanical vapor compression. With a low temperature lift of just a few Kelvin, very high calculated heating output figures can be achieved, some of which are well over 20. However, this is a misleading assessment, since it is not a counterclockwise cycle process that is typical for heat pumps and is materially closed. In any case, a higher temperature lift of over 10-20 Kelvin usually requires a multi-stage structure, which significantly reduces the overall achievable heating output figure.

Compressors for water vapor are built in various designs as flow machines, for example on the one hand in the form of radial blowers for low compression rates below 2 and high volume flows at low final pressures below 5 bar or on the other hand as turbo compressors for medium compression rates up to about 3 and higher final pressures up to about 20 bar, as well as positive displacement machines in the design of rotary piston compressors for high volume flows at medium compression rates or piston machines for high compression rates up to 6 and particularly high final pressures up to over 70 bar.

For example, the company Spilling Technologies GmbH in Hamburg (www.spilling.de) describes such piston steam compressors for high outlet pressures of up to 70 bar, which, however, due to their design, cannot generally suck in vacuum steam below 100°C, but only steam well above atmospheric pressure. These piston steam compressors are now manufactured up to a pressure level of around 40 bar in a design and limited to a steam temperature of around 250°C even without oil lubrication of the pistons. To achieve this, a defined amount of water is injected into the sucked-in steam flow before it enters the cylinders to reduce the superheat in the compressor, so that supersaturated wet steam with a steam portion x _D and a liquid portion (1 - x _D ) is created before compression, with the liquid portion ensuring sufficient lubrication of the pistons during compression and evaporating due to the increase in temperature during polytropic compression, so that ideally saturated steam without any significant overheating is present at the outlet of the compressor.

The company Piller Blowers & Compressors GmbH in Moringen (www.piller.de) describes steam compressors in fan design, which are used for vapor compression in the vacuum range and are currently built up to outlet pressures of 5 bar overpressure. To achieve a higher temperature lift, these are connected in series in several stages. Due to the temperature-dependent material strength of the compressor impellers, it is necessary to inject a defined amount of water for cooling into the sucked-in steam flow before it enters the compressor impeller at the high circumferential speeds required for compression.

The company Boldrocchi Group S.r.l. in Biassono, Italy (www.boldrocchigroup.com) also describes steam blowers and turbo compressors in fan design for compression ratios between about 1.2-3.0 and final pressures of up to 100 bar, as well as multi-stage arrangements of different compressors to achieve a higher temperature lift. The Boldrocchi Group also describes the supersaturation of steam with injected water before it enters the compressor impeller to reduce superheating in the compressor.

The advantage of wet steam compression is the continuous cooling during compression by the evaporating liquid portion. Likewise, compression due to compensated superheating takes place at lower temperatures overall than would be the case with dry, superheated steam, which means that the specific volume during wet steam compression is lower in the superheating area than with uncooled compression of dry steam, which reduces the size of the compressors for the same mass flow.

A disadvantage of compressing wet steam, however, is that in addition to the mass flow of the vaporous steam portion x _D, the mass flow of liquid water (1 - x _D ) injected before compression, which evaporates during compression, accelerates to the high circumferential speed in the compressor impeller and must be compressed to the outlet pressure. In addition, this water portion of the wet steam, which evaporates to saturated steam, experiences an increase in its specific volume by a three-digit factor within the compressor impeller and a corresponding acceleration due to the increasing volume flow, which leads to a higher specific drive torque in relation to the mass flow.

A further disadvantage is that the water portion of the wet steam enters the compressor impeller in liquid form and hits the impeller structure as droplets, which leads to greater mechanical stress on the surfaces and higher fluid friction within the compressor impeller.

In addition, the injected water must be treated or at least decalcified to avoid limescale deposits in the compressor.

In particular, in a multi-stage compression arrangement for supplying a saturated steam mass flow at the outlet of the last compressor stage, each of the compressor stages must compress the saturated steam mass flow exiting from the previous stage and additionally the water mass flow injected in the respective stage, which is required to reduce the superheat.

The combination of these disadvantages increases the overall specific drive power of the compression in relation to the saturated steam mass flow delivered.

Another significant disadvantage of the known vapor compression or steam compression for providing higher-quality steam using low-temperature waste heat between 60-80°C, for example in the form of vacuum steam, is that the condensed steam when the condensation heat is transferred to a heat sink with a high sink temperature, i.e. the liquid steam condensate, is still at a temperature that is significantly too high to be able to be evaporated again in a materially closed cycle using the same low-temperature waste heat source between 60-80°C. Thus, in addition to the construction of a materially closed cycle, which could be referred to as a heat pump cycle, the known vapor compression lacks the necessary cooling of the condensed cycle fluid to the evaporation temperature of the heat-absorbing low-pressure side.

In addition, industrial processes often have several heat sinks that need to be supplied with process heat, and these are often even at different temperature levels. In order to supply such different heat sinks with a common heat pump process in condensation mode, the design of the process must also be based on the supply of the highest-value heat sink that has the highest required sink temperature.

A typical example of industrial processes with a high process heat requirement of sometimes several megawatts up to a high temperature level of over 200°C are large-scale drying processes, for example spray drying processes for producing powders. In such spray drying processes, large flows of used process gas enriched with water vapor usually result in large amounts of waste heat, which are of a similar magnitude to the heat requirement for heating the drying process. However, due to the nature of the process, this waste heat is consistently generated at a low temperature level, for example between 60-80°C.

If heating with process heat is to take place at a high sink temperature of, for example, over 200°C, then conventional heat pumps can only achieve a high heating output coefficient if a corresponding amount of heat is supplied to the heat pump cycle on the low-temperature side and the technically achievable temperature increase is not less than the difference between the highest required sink temperature and the lowest available temperature level of the low-temperature heat source.

In order to upgrade low-temperature waste heat flows from a temperature level frequently encountered in industry between 60-80°C using heat pump technology with an economically viable heat output factor of well over 2 to ideally over 3 to 5, and to be able to use it as high-quality process heat at a consistently high temperature level of over 200°C, a temperature increase of at least 120-140 Kelvin would be necessary. A heat pump technology that enables such a heat pump cycle is not yet available.

The object of the invention is to improve the state of the art.

The object is achieved by a heat pump device for energy-efficient generation of process heat, wherein the heat pump device has a heat-absorbing side, a heat-emitting side, a cycle fluid and an evaporator on the heat-absorbing side, and at least one heat exchanger and at least one heat sink can be assigned to the heat pump device on the heat-emitting side and at least one heat source on the heat-absorbing side, wherein liquid cycle fluid can be indirectly heated by means of the evaporator by supplying heat from the assignable heat source and can be evaporated at a pressure level of the evaporator, and the heat pump device has at least one evaporator has a first compressor downstream of the heat pump device, wherein vaporous cycle fluid can be compressed from the pressure stage of the evaporator to a second pressure stage by means of the at least first compressor, and the heat pump device has at least one first mixing separator in the second pressure stage, wherein the at least first mixing separator has a first connection for the inlet of vaporous cycle fluid, a second connection for the inlet of condensed cycle fluid, a third connection for the outlet of vaporous cycle fluid and optionally a fourth connection for the outlet of condensed cycle fluid and at least one second compressor is connected downstream of the mixing separator, so that vaporous cycle fluid emerging from the third connection can be compressed in the at least second compressor from the second pressure stage to a first pressure stage and process heat can be indirectly transferred from the vaporous cycle fluid of the first pressure stage to the at least one assignable heat sink by means of the at least one assigned heat exchanger and that condensed in the at least one heat exchanger Cycle fluid can be returned via the second connection into the at least first mixing separator of the second pressure stage, wherein at the same time vaporous cycle fluid from the at least first compressor can be introduced into the mixing separator via the first connection.

Thus, a high-temperature heat pump is provided in which, when the at least one mixing separator is arranged in the second pressure stage, process heat can be released in a temperature range of approximately 100 °C to 250 °C on the heat-emitting side by means of the at least one heat exchanger of the first pressure stage.

It is particularly advantageous that the heat pump device can be operated particularly preferably with water as a natural and completely environmentally friendly circulating process fluid.

Water is liquid under standard conditions and has the advantage of a high critical pressure of 221.2 bar and a high critical temperature of 374.15°C, whereby in the two-phase region, for example at a comparatively low pressure of 16 bar and a condensation temperature of over 201°C, a high condensation enthalpy of 1,933 kJ/kg can be used for heat transfer to a heat sink, while in comparison the specific condensation enthalpy of NH ₃ at only 100°C and already 62.55 bar is only 715.7 kJ/kg.

• - mindestens einer ersten Druckstufe mit einem Druck des dampfförmigen Kreisprozessfluids, welcher dem höchsten Kondensationsdruck des Kreisprozessfluids im Kreisprozess auf dessen wärmeabgebender Seite entspricht, sodass die Kondensationstemperatur des Kreisprozessfluids so weit oberhalb der höchsten erforderlichen Senkentemperatur der höchstwertigen zu versorgenden Wärmesenke liegt, dass dieses zur Wärmeübertragung vom kondensierenden Kreisprozessfluid auf die höchstwertige Wärmesenke ausreicht,
• - sowie mindestens einer zweiten Druckstufe, in welcher der mindestens eine Mischseparator angeordnet ist und welche einen Druck des dampfförmigen Kreisprozessfluids aufweist, welcher tiefer liegt als derjenige der ersten Druckstufe und höher liegt als derjenige Verdampfungsdruck des Kreisprozessfluids im Verdampfer, und
• - mindestens einer unteresten Druckstufe mit einem Druck des dampfförmigen Kreisprozessfluids entsprechend dem Verdampfungsdruck im Verdampfer, welche dem niedrigsten Kondensationsdruck des Kreisprozessfluids innerhalb des Kreisprozesses der Wärmepumpenvorrichtung entspricht.

Thus, a heat pump device is provided with a heat pump cycle process which is based on a cycle process with at least three pressure stages with different pressure levels, with

• - at least one first pressure stage with a pressure of the vaporous cycle fluid which corresponds to the highest condensation pressure of the cycle fluid in the cycle on its heat-emitting side, so that the condensation temperature of the cycle fluid is so far above the highest required sink temperature of the highest-value heat sink to be supplied that this is sufficient for heat transfer from the condensing cycle fluid to the highest-value heat sink,
• - and at least one second pressure stage in which the at least one mixing separator is arranged and which has a pressure of the vaporous cycle fluid which is lower than that of the first pressure stage and higher than the evaporation pressure of the cycle fluid in the evaporator, and
• - at least one lowest pressure stage with a pressure of the vaporous cycle fluid corresponding to the evaporation pressure in the evaporator, which corresponds to the lowest condensation pressure of the cycle fluid within the cycle of the heat pump device.

A key idea of the invention is that a reduction in the destruction of specific exergy in the total specific compressor work to be applied in relation to the specific usable process heat takes place at the temperature level of the highest required sink temperature in each case. This is achieved by successive enthalpy utilization of condensed cycle fluid after its condensation enthalpy at the condensation pressure of a first pressure stage has been made usable as process heat for a heat sink of the highest required sink temperature and the enthalpy contained in the condensed cycle fluid of this first pressure stage is partially fed to the vaporous cycle fluid of at least a second pressure stage in at least one mixing separator in the compression line. When the liquid cycle fluid, which is at approximately the condensation temperature of the condensed cycle fluid of the first pressure stage, passes into the at least one mixing separator of a second, i.e. a lower, pressure stage with a lower condensation temperature, the liquid cycle fluid cools down to the condensation temperature of the lower pressure stage and part of its enthalpy is used to evaporate part of the liquid cycle fluid. This evaporated portion is added to the already vaporous flow of cycle fluid in the compression train is supplied at this lower pressure level without compressor work, which increases the condensation enthalpy of the vaporous cycle fluid available at this pressure level, while the compressor work required up to this pressure level remains constant. This increases the heat output coefficient and improves energy efficiency.

• Eine „Wärmepumpenvorrichtung“ ist insbesondere eine Maschine, welche unter Aufwendung von technischer Arbeit thermische Energie aus einem Reservoir und/oder einer Wärmequelle mit niedriger Temperatur aufnimmt und zusammen mit einer Antriebsenergie als Nutzwärme auf ein zu beheizendes System und/oder eine Wärmesenke mit höherer Temperatur überträgt. Eine Wärmepumpenvorrichtung ist insbesondere derart eingerichtet, um einen Wärmepumpenkreisprozess an einem Kreisprozessfluid durchzuführen mit dem Ziel, Prozesswärme an eine Wärmesenke zu übertragen, indem Wärme von einer verfügbaren Wärmequelle mit niedriger Temperatur aufgenommen wird und die aufgenommene Wärme durch Temperaturerhöhung mittels mechanischer Arbeit auf mindestens eine für die Wärmesenke höchste erforderliche Senkentemperatur aufgewertet wird.

The following terminology is explained:

• A "heat pump device" is in particular a machine which, by expending technical work, absorbs thermal energy from a reservoir and/or a heat source with a low temperature and, together with drive energy, transfers it as useful heat to a system to be heated and/or a heat sink with a higher temperature. A heat pump device is in particular designed to carry out a heat pump cycle on a cycle fluid with the aim of transferring process heat to a heat sink by absorbing heat from an available heat source with a low temperature and upgrading the absorbed heat by increasing the temperature by means of mechanical work to at least one highest sink temperature required for the heat sink.

A "heat pump cycle" is in particular a cycle process in which a closed quantity of a cycle fluid is evaporated at a low pressure, in particular by supplying heat from a reservoir and/or a heat source, then vaporous cycle fluid is compressed to a high pressure, in particular by performing mechanical work or by other processes, such as mixing with vaporous cycle fluid of a higher pressure in a mixing separator, or a combination thereof, and in the process assumes a higher temperature, then is at least cooled or condensed by heat extraction from a heat sink at this high pressure and then is brought back to a low pressure before evaporation at least by throttling or heat extraction.

A “cycle fluid” is in particular a fluid which, under certain physical conditions, is either liquid, gaseous or both at the same time, the latter being referred to as a so-called two-phase state, which is represented in state diagrams within the so-called two-phase region between the boiling line and the condensation line, or is supercritical, which describes state conditions outside the two-phase region, i.e. either a pressure higher than the critical pressure of the fluid or a temperature higher than the critical temperature of the fluid. In addition, a quantity of liquid or vaporous cycle fluid of any pressure level can be removed from the cycle per unit of time for other thermal or material utilization, and at the same time the same quantity of liquid or vaporous cycle fluid can be fed back into the cycle at another point in the same unit of time. This can, for example, be vaporous cycle fluid which is used as a heating medium or motive steam for an external process and fed back into the cycle at another point, for example in liquid form as condensed cycle fluid. Likewise, a withdrawal point can be realized at which circulating process fluid is withdrawn, in particular sporadically or continuously, for cleaning purposes and cleaned circulating process fluid is fed back into the circulating process at a further feed point in order to control and/or adjust the quality of the circulating process fluid and its fill level within the heat pump device.

A “process heat” is in particular heat transferred by means of at least one heat exchanger of the first pressure stage to a heat sink or by means of another heat exchanger of a further pressure stage to a further heat sink.

A “heat sink” is in particular a reservoir and/or a flow with an inlet and an outlet of a fluid or heat transfer medium, the temperature of which is to be raised from an inlet temperature of the inlet to an outlet temperature of the outlet, wherein the target temperature to be achieved for the outlet temperature of the heat sink is the “highest required sink temperature” of a heat sink. A heat sink can be a heat transfer medium which is heated for the purpose of heating another fluid and/or material flow, in particular in an external device outside the heat pump device, so that, for example, a heat transfer medium is heated in a heat exchanger, which is used to heat an external process in another device, for example a dryer or its auxiliary media or its drying material. The heat sink can also be a process gas flow which is heated in a drying device for drying a moist material, so that a drying process is then operated with the process gas flow by the heated process gas transferring heat to a moist material. and thus transfers a material to be dried, the moisture contained therein evaporates and is possibly also removed from the material to be dried.

A "heat source" is in particular a reservoir whose temperature can be used to transfer heat and/or waste heat to a cyclic process fluid. A heat source can also be a flow with an inlet and an outlet of a fluid or heat transfer medium, the temperature of which can be reduced from a highest available temperature, i.e. the lowest available source temperature, from an inlet temperature of the inlet to an outlet temperature of the outlet, wherein in particular a target temperature for the heat transfer in an evaporator is the "lowest possible source temperature" of the outlet of a heat source. A heat source can also be a waste heat flow which is used to heat the evaporator of the heat pump device in order to supply the cyclic process with sufficient heat at a low pressure level of the cyclic process fluid so that sufficient process heat is provided in at least one heat exchanger. Likewise, a heat source can be a process gas stream that exits a drying device after drying a moist material, i.e. waste heat, and the heat content of the process gas stream is used for heat recovery via a heat exchanger of the heat source in order to heat the evaporator of the heat pump device, wherein drying devices can often have a high heat loss via a process gas stream that exits a drying device after drying a moist material and generally amounts to approximately the level of the process heat requirement of the drying system, but is regularly available at a significantly lower temperature level and can therefore possibly be used to heat the evaporator of the heat pump device. A heat source can also be a heat-emitting side, for example waste heat, of a refrigeration system used to generate process cooling, the heat emitted by which is used to heat the evaporator of the heat pump device. A heat source can also be a heat-emitting side of a refrigeration system used to generate process cooling, the heat emitted by which is used in a heat exchanger of the heat source to indirectly heat the evaporator of the heat pump device. In the case of refrigeration systems that frequently operate intermittently and are therefore subject to fluctuations in heat emission, a heat supply that is as constant as possible can also be provided to heat the evaporator of the heat pump device, in particular for compensation. Such compensation can be implemented in particular in the form of a combination of different waste heat flows, for example waste heat from a process gas flow in combination with waste heat from a refrigeration system or with other waste heat.

“Waste heat” refers in particular to a heat source whose heat is available at such a low temperature level that, in particular in the vicinity of the heat source, there is no heat sink with a lower maximum required sink temperature for using this heat as process heat or it can be used, or if this heat could only be used as process heat at an expense that is no longer economically justifiable.

A "heating performance coefficient" with the abbreviation COP _h (from the English "Coefficient of Performance - heat") is in particular the quotient of usable process heat per unit of time in relation to the mechanical work expended per unit of time and represents a key figure for the energy efficiency of a heat pump device when generating process heat.

A "heat exchanger" is in particular a device in which thermal energy, i.e. heat, is transferred from a material flow of higher temperature to another material flow of lower temperature, wherein the material flows are spatially and materially separated from one another by a wall of the heat exchanger. A heat exchanger is preferably an indirect heat exchanger. One of the material flows that flow through a heat exchanger can be, for example, the circulating process fluid or an inlet or outlet of a fluid or heat transfer medium of a heat sink or heat source. The heat pump device can also have two or more heat exchangers on the heat-absorbing side and/or heat-emitting side.

An "evaporator" is in particular a device or apparatus in which a liquid cycle fluid changes its state from liquid to gaseous, in particular through indirect heat transfer from a heat source, and thus undergoes a "phase transition" from liquid to gaseous, wherein the gaseous state of a cycle fluid is generally referred to as "vapor" and the state is referred to as "vaporous" if it is a cycle in which at least one phase transition between two states takes place, whereas a gaseous state indicates that no phase transition takes place. An evaporator is arranged in particular in the lowest and/or lowest pressure stage of the cycle.

A "compressor" (also called a compressor) is in particular a device or apparatus for compressing and/or increasing the pressure of a compressible fluid, for example a vaporous cycle fluid, in particular with the aim of achieving a higher pressure level in order to thereby raise the condensation temperature of the vaporous cycle fluid. This is also referred to as "upgrading" the heat content of the vaporous cycle fluid for use as "higher-quality" process heat. A compressor is in particular designed in the design of a turbomachine, such as an axial blower, a radial blower, a turbocompressor or a turbine. A compressor can also be designed in the design of a positive displacement machine, such as a piston compressor, a rotary piston compressor or a screw compressor. A compressor can be a thermal compressor, such as a vacuum steam jet pump, in which vaporous cycle fluid of a second or first pressure stage is used as motive steam, which reaches high speeds in the thermal compressor and thereby sucks in suction steam in the form of vaporous cycle fluid with a pressure lower than that of the motive steam, with motive steam and suction steam subsequently mixing to form a mixed steam with the pressure of a pressure stage that has a higher pressure than the suction steam and a lower pressure than the motive steam.

A “compression train” is in particular a sequence of compression processes of vaporous cycle fluid in at least two compressors connected directly or indirectly one after the other, regardless of the design, with the aim of compressing the vaporous cycle fluid from at least one lower pressure level to at least one higher pressure level.

"Superheating" is in particular a condition of a vaporous cycle fluid whose temperature is above the condensation temperature that correlates with its prevailing pressure and/or pressure level and is expressed in broadly simplified terms as a temperature difference in Kelvin. In contrast, the degree of superheating is correctly understood as the difference between the specific enthalpy of the vaporous cycle fluid and the specific saturated steam enthalpy in kJ/kg at the prevailing pressure.

A "pressure stage" is to be understood in particular as a synonym for a total pressure achieved at the outlet of a compressor, regardless of the design, after compression of vaporous cycle fluid to specifically indicate that a heat pump device involves a step-by-step compression of vaporous cycle fluid. This is accompanied by the corresponding boiling and condensation temperature of the cycle fluid at the pressure of this pressure stage.

A “first pressure stage” is to be understood in particular as the pressure stage with the highest condensation pressure of the cycle fluid within the heat pump device, which is required in order to be able to supply the heat sink with the highest required sink temperature with process heat by condensation, for example in a pressure range between approximately 1 bar for process heat of approximately 100°C to approximately 40 bar for process heat of approximately 250°C.

A “pressure level of an evaporator” is to be understood in particular as the pressure level at which an evaporator of the heat pump device is operated in order to evaporate liquid cycle fluid with the supply of heat from a heat source, for example in a pressure range between approximately 40 hPa at a lowest source temperature of approximately 30°C to approximately 1.4 bar at a lowest source temperature of approximately 110°C.

A "second pressure stage" is to be understood in particular as a pressure stage below a first pressure stage and above the pressure stage of an evaporator. In particular, the at least one mixing separator is arranged in the second pressure stage.

A "third pressure level" is to be understood in particular as a pressure level below a second pressure level and above the pressure level of an evaporator. A "fourth pressure level" is to be understood in particular as a pressure level below a third pressure level and above the pressure level of an evaporator. A "fifth, sixth, seventh and optionally further pressure level" is to be understood in an analogous manner. The designation of the pressure levels serves in particular to distinguish the respective pressure present and does not represent a fixed order.

A "mixing separator" is in particular a device which is assigned to a specific pressure level with a pressure above the pressure level of an evaporator and less than and/or equal to the first pressure level, wherein a mixing separator encloses a volume and has at least one connection for the inlet for vaporous cycle fluid of the assigned pressure level, one connection for the inlet for liquid cycle fluid and one connection for the outlet for vaporous cycle fluid of the assigned pressure level. An essential function of a mixing separator is in particular the saturation of a flow of vaporous, superheated cycle fluid after compression by mixing with liquid cycle fluid. Another function of a mixing separator can also be the expansion and spontaneous evaporation of liquid cycle fluid, for example from a higher pressure stage, if its temperature is higher than the correlating condensation temperature of the associated pressure stage of the mixing separator.

In a further embodiment of the heat pump device, the at least first mixing separator has the fourth connection for the outlet of condensed cycle fluid, so that the escaping condensed cycle fluid can be returned directly or indirectly to the evaporator and/or fed to the at least second compressor.

As a result, at least the first mixing separator can have the connection for the outlet for liquid cycle fluid of the pressure stage assigned to the mixing separator, through which excess liquid cycle fluid possibly exits, which is not required or consumed for the saturation of vaporous cycle fluid entering the mixing separator superheated at an inlet connection or which does not evaporate by spontaneous evaporation when entering the mixing separator, and which then exits from the mixing separator at the corresponding condensation temperature of the relevant assigned pressure stage of the mixing separator.

The liquid cycle fluid emerging from at least the first mixing separator can be returned to the evaporator via a condensate separator. A partial flow of the liquid cycle fluid can also be fed to the second compressor via a condensate pump and/or a control valve. As a result, the liquid cycle fluid emerging from at least the first mixing separator can be used to generate wet steam.

In order to provide an interface to at least one heat sink, the heat pump device has at least one heat exchanger assigned to the first pressure stage and/or one heat exchanger assigned to the second pressure stage and/or one heat exchanger assigned below the second pressure stage on the heat-emitting side.

In a further embodiment, the heat pump device may comprise at least one heat source and/or at least one heat sink.

In order to provide an adaptable multi-stage, modular structure and to realize a subdivision of the compression, the heat pump device has a second mixing separator, a third mixing separator, a fourth mixing separator and/or optionally further mixing separators, wherein a further compressor is connected upstream of the respective mixing separator.

Thus, two or more mixing separators can be arranged between the evaporator and the first pressure stage, in particular connected in series.

Through a multi-stage structure and the division of the compression of the vaporous cycle fluid into several compression steps at individual pressure levels, i.e. from the lowest pressure of the cycle of a pressure level in an evaporator to the highest required pressure of the first pressure level, which build on each other, the desired temperature difference between the lowest available heat source and/or waste heat temperature and the highest required sink temperature can be configured very variably. This reduces the pressure ratio required per compression step, i.e. the quotient of the compression pressure achieved in relation to the inlet pressure before compression, whereby the superheating of the compressed cycle fluid, conventionally expressed in Kelvin above the condensation temperature at the compression pressure achieved, is significantly lower than with a single-stage compression from the lowest to the highest pressure in the cycle. This means that very high final pressures can be achieved, which, depending on the choice of cycle fluid and the location of its critical point and two-phase region, enable correspondingly high condensation temperatures and condensation enthalpies for the transfer of process heat to a heat sink. This means that, particularly with water as the cycle fluid, condensation conditions of just 40 bar up to around 250°C are possible.

Thanks to the variably scalable heat pump device and the widely configurable structure of the heat pump device, the area of application can be expanded to a temperature level of up to 250°C with a temperature lift of almost 200 Kelvin at heating performance factors of over 2.5, while maintaining the use of waste heat with temperatures below 60°C to 80°C for the generation of process heat, whereby process heat can also be released simultaneously at several different temperature levels between, for example, 100°C and 250°C.

This multi-stage design is advantageous for several reasons. Firstly, overheating during compression can be reduced, which does not represent a thermal advantage, but can, for example, make lubrication of pistons or other sliding surfaces of displacement bodies superfluous in high-pressure compressors based on the displacement principle, thus enabling an oil-free design of the compressor and a simplified construction, as well as expanding the area of application to include hygienic steam.

On the other hand, branches can be created in the cycle that allow the use of process heat at different temperature levels by not only generating the highest-value process heat for the highest required sink temperature, but also supplying heat sinks with a lower required sink temperature with process heat generated for this purpose at a lower-value condensation temperature of the cycle fluid at a second, third, fourth and/or further pressure level. This reduces the overall destruction of specific exergy, since less specific compression work has to be applied for lower-value process heat than for the highest-value process heat. By setting up a structure with more than three pressure levels and separate generation of different-value process heat at different temperature levels, the overall heating performance factor of the heat pump process is increased.

In a further embodiment of the heat pump device, a mixing separator or two or more mixing separators are connected upstream of the first mixing separator, wherein the vaporous cycle fluid of the respective upstream mixing separator can be fed to the subsequent mixing separator via a respective downstream compressor and/or the liquid cycle fluid from the respective downstream mixing separator can be returned to the upstream mixing separator at a higher pressure and/or a higher temperature than in the upstream mixing separator.

This provides a heat pump device in which vaporous cycle fluid of at least a second, third, fourth and/or further pressure stage condenses in at least one heat exchanger of at least a second heat sink and indirectly transfers heat to it, wherein a cycle fluid condensed on this second, third, fourth and/or further pressure stage in the heat exchanger of a second or further heat sink enters a mixing separator of a respectively lower pressure stage, i.e. a third, fourth, fifth and/or further pressure stage, with a pressure below the pressure of the cycle fluid condensing in the heat exchanger.

Thus, the heat pump device enables an extended heat pump cycle process, which is based on a cycle process with at least four pressure stages with different pressure levels, i.e. at least a first, at least a second, at least a third pressure level and at least one lowest pressure level of the heat pump cycle process at the pressure of an evaporator, wherein the heat pump device is easily scalable and variably configurable by adding further pressure levels, i.e. a fourth, fifth, sixth and/or subsequent pressure level with corresponding mixing separators and compressors, in order to adapt the overall achievable temperature rise as required.

In the heat pump device with step-by-step compression, the vaporous cycle fluid from a mixing separator of a pressure stage is compressed to a higher pressure stage for each compression step and is superheated due to polytropic compression. The superheat that is present in the vaporous cycle fluid after compression and then after entry into a mixing separator assigned to a pressure stage is used to partially evaporate liquid cycle fluid that enters the pressure stage reached by upstream compression via a connection for entry into the respective mixing separator. This evaporated portion is fed to the already vaporous flow of cycle fluid in the compression line at this pressure stage without compressor work, thereby increasing the total condensation enthalpy available at this pressure stage of the vaporous cycle fluid that is later available as process heat, while the compressor work remains constant until then. This increases the heating output coefficient.

A further advantage of the heat pump device with step-by-step compression is the use of cycle fluid to reduce the overheating that occurs due to compression, since the cycle fluid itself has already been treated or at least decalcified once for the initial filling of the heat pump device and is reused due to the closed cycle process, which completely avoids continuous water consumption through injection of tap water or other calcareous water as well as the unavoidable limescale deposits compared to the state of the art.

A further advantage of the heat pump device with step-by-step compression is the step-by-step cooling of condensed cycle fluid to complete the cycle by relaxation in lower pressure stages with full utilization of the enthalpy contained therein, whereby the latter is added as an additional saturated steam portion to the steam mass flow of the lower pressure stage without compressor work for the additional evaporation of liquid cycle fluid. This increases the heating performance coefficient even further.

A further advantage of the heat pump device with gradual cooling of condensed cycle fluid in individual pressure stages is the significantly increasing amount of vaporous cycle fluid during the gradual compression in the compression line along the pressure increase, which means that considerably less liquid cycle fluid has to be evaporated in the evaporator than has to be provided for heating heat sinks for condensation. As a result, only a significantly lower mass flow of vaporous cycle fluid has to be transported through the evaporator and compressed one after the other in the compressors of the lower pressure stages of the compressor line, which typically have the largest sizes due to the large specific volume of the vaporous cycle fluid at low pressure. This significantly reduces the size of these devices without any loss of performance.

In order to improve the return of the cycle fluid condensed in at least one heat exchanger from the first and/or further pressure stage, a condensate separator is or are connected downstream of the at least one heat exchanger and/or a condensate separator is or are connected downstream of the second heat exchanger and/or a condensate separator is or are connected downstream of the further heat exchangers for returning condensed cycle fluid from the first pressure stage to the first mixing separator and/or condensed cycle fluid from the second or a third pressure stage to an upstream mixing condenser, in particular a lower pressure stage.

In a further embodiment of the heat pump device, a condensate separator for returning the condensed cycle fluid to the respective upstream mixing separator or to the evaporator is or are arranged after the fourth connection for the outlet of condensed cycle fluid of the first mixing separator, an upstream mixing separator and/or the respective mixing separator.

In order to generate mixed steam of a pressure stage and to suck in suction steam from a lower pressure stage or the pressure stage of the evaporator, the first compressor is designed as a thermal compressor, so that by means of the thermal compressor the vaporous cycle fluid of the first pressure stage can be mixed as motive steam with the vaporous cycle fluid from the evaporator as suction steam and can be fed as compressed mixed steam to the first mixing separator or the first upstream mixing separator.

Accordingly, a compressor upstream of the mixing separator of the respective pressure stage is designed as a thermal compressor, so that by means of the thermal compressor the vaporous cycle fluid of a higher pressure stage can be mixed as motive steam with the vaporous cycle fluid of a lower pressure stage as suction steam and can be fed to the mixing separator of this pressure stage as compressed mixed steam.

In the thermal compressor, in particular a stream or partial stream of vaporous cycle fluid of a first or a further pressure stage can be used as motive steam, which reaches high speeds in the thermal compressor and thereby sucks in a suction steam in the form of vaporous cycle fluid with a pressure lower than that of the motive steam due to Venturi or Coanda effects, wherein the motive steam and suction steam subsequently mix to form a mixed steam with a pressure of a pressure stage that is higher than the pressure of the suction steam and lower than the pressure of the motive steam.

In a further embodiment of the heat pump device, a further compressor in the first pressure stage or the second pressure stage or a further pressure stage is arranged upstream of the thermal compressor, so that the compressed vaporous cycle fluid, in particular the vaporous cycle fluid originating from the first pressure stage and further compressed to the pressure of a motive steam pressure stage, can be used as motive steam.

In order to realize wet steam compression or dry compression, a further mixing separator is arranged downstream of the first mixing separator in the second pressure stage, whereby the further mixing separator is arranged in the first pressure stage.

The downstream mixing separator in the first pressure stage can have a downstream compressor or it can be free from a compressor

In a further embodiment, the heat pump device has a control and/or regulating device for controlling and/or regulating components of the heat pump device and optionally the at least one heat source and/or the at least one heat sink.

This enables the control and/or regulation of the heat pump device and its process heat supply with a view to achieving the highest possible energy efficiency due to the lowest possible drive power by controlling pressures and temperatures of individual pressure stages by controlling compressors and controlling the heating output of the heat source as well as controlling flows of condensed cycle fluid at the inlet to individual mixing separators and/or compressors. Thus, at least one sub-process contained in or connected to the heat pump device or the entire process of the heat pump device can be controlled and regulated by at least one control and/or regulating device.

In order to return the condensed cycle fluid in the form of droplets to the respective mixing separator, the second connection for the inlet of condensed cycle fluid has a spraying device for spraying the liquid cycle fluid entering the first mixing separator or the respective mixing separator.

A spraying device can be, for example, a spray nozzle. As a result, the incoming liquid cycle fluid is sprayed into the mixing separator in the form of droplets, creating a pressure loss, whereby a portion of the liquid cycle fluid entering the second connection evaporates depending on its temperature and overheating of the vaporous cycle fluid entering the first connection is reduced.

In a further aspect of the invention, the object is achieved by a drying device for drying a material to be dried by means of a heated process gas flow, wherein the drying device has a heat pump device as described above, so that the process gas flow can be heated as a heat sink on the heat-emitting side of the heat pump device.

Thus, process heat at a high temperature level of up to 250°C can be used to heat a process gas stream of the drying device.

• - Beheizen eines flüssigen Kreisprozessfluids durch indirekte Wärmezufuhr von einer Wärmequelle und Verdampfen des flüssigen Kreisprozessfluids in einem Verdampfer auf einer Druckstufe des Verdampfers,
• - Verdichten des dampfförmigen Kreisprozessfluids mittels eines ersten Verdichters auf eine zweite Druckstufe,
• - Zuführen des verdichteten dampfförmigen Kreisprozessfluids in mindestens einem Mischseparator einer zweiten Druckstufe,
• - Zuführen des aus dem mindestens einem Mischseparator austretenden dampfförmigen Kreisprozessfluids in mindestens einem zweiten Verdichter und Verdichten des dampfförmigen Kreisprozessfluid von der zweiten Druckstufe auf eine erste Druckstufe,
• - Übertragen einer Prozesswärme von dem verdichteten, dampfförmigen Kreisprozessfluid der ersten Druckstufe auf mindestens eine zuordenbare Wärmesenke,
• - und optional Rückführen des kondensierten Kreisprozessfluids der ersten Druckstufe in den mindestens einen Mischseparator.

In an additional aspect of the invention, the object is achieved by a method for operating a heat pump device for energy-efficient generation of process heat, in particular in a temperature range between 100 °C and 250 °C, by means of a previously described heat pump device, with the following steps:

• - Heating a liquid cycle fluid by indirect heat supply from a heat source and evaporating the liquid cycle fluid in an evaporator at a pressure level of the evaporator,
• - compressing the vaporous cycle fluid by means of a first compressor to a second pressure stage,
• - feeding the compressed vaporous cycle fluid in at least one mixing separator to a second pressure stage,
• - feeding the vaporous cycle fluid emerging from the at least one mixing separator into at least one second compressor and compressing the vaporous cycle fluid from the second pressure stage to a first pressure stage,
• - Transferring process heat from the compressed, vaporous cycle fluid of the first pressure stage to at least one assignable heat sink,
• - and optionally returning the condensed cycle fluid of the first pressure stage to the at least one mixing separator.

In a further embodiment of the process, water, an alcohol and/or a water-soluble organic substance is or are used as the circulating process fluid.

Preferably, the heat pump device is operated with water (chemical formula H ₂ O) as a natural and completely environmentally friendly cycle fluid. Likewise, the heat pump device can also be operated with an alcohol and/or an aqueous solution of an alcohol as a cycle fluid. Likewise, an aqueous solution of a water-soluble organic substance and/or one or more organic substances can also be used as a cycle fluid. Likewise, any combination of water, alcohol and/or water-soluble organic substance can be used as a cycle fluid.

For example, methanol, ethanol or propanol can be used as alcohol and an ester and/or an ether can be used as organic substance.

In a further embodiment of the method, in a stream of the condensed cycle fluid emerging from the at least one mixing separator, in particular a Pressure level higher than the pressure level of the evaporator, a pressure increase can be carried out by means of a condensate pump, then this stream can be mixed with a stream of liquid circulating process fluid which has previously been indirectly heated in a heat exchanger of a heat source, and the mixed streams can be used for indirectly heating the evaporator before, in particular, the mixed stream is returned to the evaporator.

In a further embodiment of the method, a pressure increase is carried out in a stream of the condensed cycle fluid emerging from the at least one mixing separator by means of at least one condensate pump, and finally this stream is introduced into vaporous cycle fluid from the at least one mixing separator to generate cycle fluid wet steam before the cycle fluid wet steam is compressed to the first pressure stage in the downstream compressor.

• 1 eine stark schematische Darstellung einer Wärmepumpenvorrichtung mit einem Verdampfer, einem ersten Verdichter, einem Mischseparator einer zweiten Druckstufe, einem zweiten Verdichter und einem Wärmeübertrager einer ersten Druckstufe,
• 2 eine stark schematische Darstellung einer Alternative der in 1 gezeigten Wärmepumpenvorrichtung mit einer Beheizung eines Verdampfers über einen Kreislauf durch einen Wärmeübertrager einer Wärmequelle,
• 3 eine stark schematische Darstellung einer Alternative der in 2 gezeigten Wärmepumpenvorrichtung mit einem zusätzlichen Mischseparator nach einer Verdichtung auf eine erste Druckstufe,
• 4 eine stark schematische Darstellung einer Alternative der in 3 gezeigten Wärmepumpenvorrichtung und einem mit Treibdampf aus einer ersten Druckstufe versorgten thermischen Verdichter zur Verdichtung von dampfförmigem Kreisprozessfluid von der Druckstufe des Verdampfers auf eine zweite Druckstufe,
• 5 eine stark schematische Darstellung einer Alternative der in 4 gezeigten Wärmepumpenvorrichtung und einem mit Treibdampf aus einer Treibdampf-Druckstufe versorgten thermischen Verdichter zur Verdichtung von dampfförmigem Kreisprozessfluid von der Druckstufe des Verdampfers auf eine zweite Druckstufe,
• 6 eine stark schematische Darstellung einer weiteren Alternative der in 3 gezeigten Wärmepumpenvorrichtung mit einer zusätzlichen dritten Druckstufe,
• 7 eine stark schematische Darstellung einer Alternative der in 6 gezeigten Wärmepumpenvorrichtung mit einer zusätzlichen Übertragung von Prozesswärme einer zweiten Druckstufe in einem Wärmeübertrager einer zweiten Wärmesenke,
• 8 eine stark schematische Darstellung einer weiteren Alternative der in 6 gezeigten Wärmepumpenvorrichtung mit einer zusätzlichen vierten Druckstufe,
• 9 eine stark schematische Darstellung einer Alternative der in 8 gezeigten Wärmepumpenvorrichtung mit einem mit Treibdampf aus einer Treibdampf-Druckstufe versorgten thermischen Verdichter zur Verdichtung von dampfförmigem Kreisprozessfluid von der Druckstufe des Verdampfers auf eine vierte Druckstufe,
• 10 eine stark schematische Darstellung einer weiteren Alternative der in 8 gezeigten Wärmepumpenvorrichtung mit einer zusätzlichen Übertragung von Prozesswärme einer zweiten Druckstufe in einem Wärmeübertrager einer zweiten Wärmesenke,
• 11 eine stark schematische Darstellung einer Alternative der in 10 gezeigten Wärmepumpenvorrichtung mit einer Übertragung von Prozesswärme einer dritten anstelle einer zweiten Druckstufe in einem Wärmeübertrager einer zweiten Wärmesenke, und
• 12 eine schematische Darstellung eines Mischseparators einer Wärmepumpenvorrichtung.

The invention is explained below using exemplary embodiments.

• 1 a highly schematic representation of a heat pump device with an evaporator, a first compressor, a mixing separator of a second pressure stage, a second compressor and a heat exchanger of a first pressure stage,
• 2 a highly schematic representation of an alternative to the 1 shown heat pump device with heating of an evaporator via a circuit through a heat exchanger of a heat source,
• 3 a highly schematic representation of an alternative to the 2 shown heat pump device with an additional mixing separator after compression to a first pressure stage,
• 4 a highly schematic representation of an alternative to the 3 shown heat pump device and a thermal compressor supplied with motive steam from a first pressure stage for compressing vaporous cycle fluid from the pressure stage of the evaporator to a second pressure stage,
• 5 a highly schematic representation of an alternative to the 4 shown heat pump device and a thermal compressor supplied with motive steam from a motive steam pressure stage for compressing vaporous cycle fluid from the pressure stage of the evaporator to a second pressure stage,
• 6 a highly schematic representation of another alternative of the 3 shown heat pump device with an additional third pressure stage,
• 7 a highly schematic representation of an alternative to the 6 shown heat pump device with an additional transfer of process heat of a second pressure stage in a heat exchanger of a second heat sink,
• 8th a highly schematic representation of another alternative of the 6 shown heat pump device with an additional fourth pressure stage,
• 9 a highly schematic representation of an alternative to the 8th shown heat pump device with a thermal compressor supplied with motive steam from a motive steam pressure stage for compressing vaporous cycle fluid from the pressure stage of the evaporator to a fourth pressure stage,
• 10 a highly schematic representation of another alternative of the 8th shown heat pump device with an additional transfer of process heat of a second pressure stage in a heat exchanger of a second heat sink,
• 11 a highly schematic representation of an alternative to the 10 shown heat pump device with a transfer of process heat of a third instead of a second pressure stage in a heat exchanger of a second heat sink, and
• 12 a schematic representation of a mixing separator of a heat pump device.

One in 1 The heat pump device 1 shown has an evaporator 200 with heating via an inlet 14 and an outlet 15 of a heat source, a compressor 401 from the pressure stage of the evaporator 200 to a second pressure stage, a mixing separator 400 of a second pressure stage, a compressor 501 from a second to a first pressure stage, a heat exchanger 520 of a first heat sink for transferring process heat of a first pressure stage and an inlet 10 and an outlet 11 of a dryer device 5 as a heat sink. The heat pump device 1 has a control and regulation unit 2 for controlling and regulating the functions and components of the heat pump device 1.

The mixing separator 400 of the second pressure stage has a connection 406 for the inlet of vaporous cycle fluid, a connection 407 for the inlet of liquid cycle fluid, a connection 408 for the outlet 408 of vaporous cycle fluid and a connection 409 for the outlet of liquid cycle fluid. A spraying device for spraying the liquid cycle fluid is integrated in the connection 407 for the inlet of liquid cycle fluid in order to spray the incoming liquid cycle fluid into the mixing separator 400 in the form of droplets while building up a pressure loss, wherein a portion of the liquid cycle fluid entering at connection 407 evaporates depending on its temperature and in the process overheating of the vaporous cycle fluid entering at connection 406 is reduced.

At the lowest pressure of the cycle in the evaporator 200, usually a partial vacuum below atmospheric pressure, and at the same time at the lowest temperatures of the cycle, liquid cycle fluid returned from a mixing separator 400 of the second pressure stage is heated and evaporated by a heat source by means of inlet 14 and outlet 15 and then vaporous cycle fluid is compressed as dry vapor from the pressure stage of the evaporator in a compressor 401 to a second pressure stage and enters the mixing separator 400 of the second pressure stage superheated.

Vaporous cycle fluid emerges from the mixing separator 400 as dry vapor of a second pressure stage and, before compression in a compressor 501, is supersaturated with injected condensate and thus liquid cycle fluid via a condensate pump 250 and is thereby converted into wet steam with a portion of liquid cycle fluid, such that during the subsequent compression to a first pressure stage in the compressor 501, this liquid portion evaporates and dry-saturated vaporous cycle fluid of a first pressure stage emerges from the compressor 501. In a heat exchanger 520 of a heat sink, process heat at the condensation temperature of a first pressure stage is transferred from the vaporous cycle fluid to the heat sink with an inlet 10 and an outlet 11, whereby the cycle fluid condenses. Liquid cycle fluid exits the heat exchanger 520 at approximately condensation temperature and at a first pressure level via a condensate separator 522 and is introduced into the mixing separator 400 of the second pressure level, where it expands at its second pressure level and cools down to its condensation temperature of the second pressure level. In the process, a portion of the cycle fluid introduced in liquid form from the heat exchanger 520 evaporates and forms vaporous cycle fluid at the second pressure level. Another portion of the cycle fluid introduced in liquid form from the heat exchanger 520 evaporates due to the degree of superheating of the vaporous cycle fluid entering the mixing separator 400 superheated from the compressor 401. The remaining portion of the cycle fluid introduced from the heat exchanger 520 that is introduced into the mixing separator 400 in liquid form is considered excess liquid cycle fluid and exits the mixing separator 400 at the pressure and condensation temperature of the second pressure stage before being reintroduced into the evaporator 200.

In a 2 In the alternative shown, a heat pump device 1 has an indirect heating of the evaporator 200 via a circuit through a heat exchanger 220 of a heat source (14, 15), wherein, in contrast to the representation in 1 a flow of condensed cycle fluid from a mixing separator 400 of a second pressure stage is introduced into a flow of liquid cycle fluid, which was previously indirectly heated in the heat exchanger 220 of the heat source, and the mixture of both flows is used to indirectly heat an evaporator 200. A condensate pump 201 is used to overcome a pressure difference from the outlet from the second pressure stage to the circuit with a circulation pump 222.

Furthermore, the 2 shown heat pump device 1 in contrast to the representation in 1 designed to introduce a cycle fluid condensed at a second pressure level from a mixing separator 400 of a second pressure level into vaporous cycle fluid from a second pressure level to generate cycle fluid wet steam before this wet steam is compressed in a compressor 501 to a first pressure level. In this case, a condensate pump 201 conveys the amount of condensed cycle fluid that exits from the mixing condenser 400 of the next higher pressure level above the pressure level of the evaporator 200. At the same time, a condensate pump 201 provides a sufficient injection pressure of liquid cycle fluid of a second pressure level before compression 501 to a first pressure level. Otherwise, the 2 The heat pump device 1 shown is operated as described above.

One in 3 The alternative heat pump device 1 shown in FIG. 1 has, in contrast to the illustration in 2 that vaporous cycle fluid after a compression 501 from a second pressure stage to a first pressure stage and condensed cycle fluid of a second pressure stage are simultaneously introduced into a mixing separator 500 of a first pressure stage in order to reduce superheating of the vaporous cycle fluid during compression without prior injection of liquid cycle fluid and to provide saturated vapor at a first pressure stage at the outlet 508 from the mixing separator 500.

In a minimal design of the heat pump device 1 with at least three pressure stages according to 1 , 2 or 3 and the preferred use of water as the cycle fluid, the evaporator 200 can be operated using waste heat at temperatures of, for example, 60-80°C at an evaporation temperature of, for example, 54°C, which corresponds to an evaporation pressure of the cycle fluid of 150 hPa. This evaporation pressure defines the operating conditions of a heat-absorbing side of the cycle and corresponds to the lowest condensation pressure of the cycle fluid within the cycle, which at the same time corresponds to the lowest pressure level of the heat pump device 1.

Starting from this lowest pressure level in the evaporator 200, the vaporous cycle fluid, for example water, is compressed to a second pressure level in at least one compressor 401. The pressure of the second pressure level depends on the achievable compression ratio of the selected compressor 401, which can be in a range of 1.2-6.0, for example. Depending on the compressor design, the vaporous cycle fluid is heated to varying degrees during compression.

Based on an evaporation pressure of the cycle fluid in the evaporator 200 of 150 hPa and a compression ratio in the compressor 401 of, for example, 3.0, the pressure of the second pressure stage would be calculated to be 450 hPa. For the cycle fluid water, the boiling and condensation temperature of the second pressure stage corresponding to 450 hPa would be 78.7°C.

If one compares these two pressure stages (150 hPa; 54.0°C) and (450 hPa; 78.7°C), for example, with an isentropic compression with n = κ = 1.333 from an evaporator pressure stage (150 hPa; 54.0°C) to a second pressure stage with 450 hPa, which corresponds to a compression ratio of 3.0, then the temperature after the isentropic compression would be T _V2 = 157.4°C, which represents a superheat of T _V2 - T _D2 = 157.4°C - 78.7°C = 78.7 K compared to the corresponding condensation temperature of the second pressure stage of T _D2 = 78.7°C.

This positive temperature difference of 78.7 K makes it possible for heat to be transferred from a steam compressed to a pressure level p _V2 and thus superheated to a liquid cycle fluid whose temperature corresponds at most to the corresponding boiling temperature T _D2 of the pressure level p _D2 .

This circumstance is used to transfer the superheat enthalpy resulting from the compression at a pressure level to the evaporation of liquid cycle fluid at the same pressure level and thus to generate a higher proportion of vaporous cycle fluid at this pressure level without this additional vaporous portion having to be compressed to this pressure level by performing mechanical work, i.e. the destruction of exergy.

This is achieved by the vaporous cycle fluid compressed in a compressor 401 to a second pressure stage entering the mixing separator 400 of the second pressure stage at a connection 406 for the entry of vaporous cycle fluid and being brought there into direct contact with liquid cycle fluid of the second pressure stage, which is located in the mixing separator 400 of the second pressure stage.

Within the mixing separator 400 of the second pressure stage, a quasi-stationary equilibrium is formed due to the direct contact of vaporous and liquid cycle fluid and due to rapidly occurring heat and mass transfer processes through simultaneous condensation of vaporous and evaporation of liquid cycle fluid, which is supported, for example, by a cyclone effect in order to separate vaporous cycle fluid from liquid and at the same time to allow it to exit separately from the mixing separator 400 of the second pressure stage at separate connections 408, 409, in particular vaporous cycle fluid at at least connection 408 and liquid cycle fluid at at least connection 409.

The vaporous cycle fluid which exits at at least one connection 408 of the mixing separator 400 of the second pressure stage is then compressed in at least one compressor 501 from the second to a first pressure stage, wherein vaporous cycle fluid of this first pressure stage passes through the heat-emitting side of the heat pump device 1 and indirectly transfers process heat to at least one heat sink (10, 11) in a heat exchanger 520 and condenses in the process.

The transfer of process heat of the heat pump device 1 therefore takes place at least by condensation of the vaporous cycle fluid at the condensation pressure of the first pressure stage and the corresponding condensation temperature. The pressure of the first pressure stage depends on the achievable compression ratio of the selected compressor 501, which can be in a range of 1.2-6.0, for example.

Based on a pressure of a second pressure stage of 450 hPa and a compression ratio in the compressor 501 of, for example, 4.0, the pressure of the second pressure stage is then 1,800 hPa. For the circulating process fluid water, the boiling and condensation temperature of the second pressure stage corresponding to 1,800 hPa is therefore 116.9°C.

During the transfer of process heat on the heat-emitting side of the heat pump device 1, vaporous cycle fluid of the first pressure stage condenses at 116.9°C in at least one heat exchanger 520 of the heat sink (10, 11), wherein condensed cycle fluid of the first pressure stage exits the heat exchanger 520 at a temperature that is not significantly below the corresponding condensation temperature of the first pressure stage and enters a mixing separator 400 of a second pressure stage at at least one connection 407, wherein the corresponding boiling and condensation temperature of the second pressure stage at approximately 78.7°C is significantly below the corresponding boiling and condensation temperature of the first pressure stage of approximately 116.9°C.

Due to this temperature difference when the condensed cycle fluid of the first pressure stage enters the mixing separator 400 of the second pressure stage, the pressure of which, as described, is lower than the pressure of the first pressure stage by the reciprocal of the compression ratio of the compressor 501, a spontaneous evaporation of a further portion of liquid cycle fluid can occur.

In such a spontaneous evaporation, the excess enthalpy of the liquid cycle fluid entering at a higher temperature can be used as evaporation enthalpy until it is cooled to the corresponding boiling and condensation temperature of the second pressure stage. With the described temperature difference of 116.9°C - 78.7°C = 38.2 K and an assumed specific heat capacity of 4.186 kJ/kg·K of the liquid cycle fluid, a specific evaporation enthalpy of about 2,313 kJ/kg is required for isobaric evaporation at the pressure of the second pressure stage of 450 hPa. Based on the mass flow of the liquid cycle fluid entering the connection 407 for entry into the mixing separator 400 of the second pressure stage from the condensed cycle fluid of the first pressure stage, the ratio of the specific enthalpies results in a mass ratio of (4.186 kJ/kg·K·38.2 K)/(2,313 kJ/kg) = 0.069 kg/kg of additionally vaporized cycle fluid, without this additional vaporous portion having to be compressed in a previous compression, such as in a compressor 401, by applying mechanical work to the second pressure stage.

The portion of liquid cycle fluid which enters the mixing separator 400 of the second pressure stage at a connection 407 and is not evaporated to reduce overheating or by spontaneous evaporation is referred to as the excess portion of liquid cycle fluid of the second pressure stage and exits the mixing separator 400 at least at the connection 409, from where at least a portion is passed on directly or indirectly into the evaporator 200.

In the following, the previously described minimum configuration of a heat pump device 1 according to the 1 to 3 with the cycle fluid water in terms of the achievable heat performance coefficient COP _h and the thus achievable energy efficiency.

The specific condensation enthalpy corresponding to the first pressure stage (1,800 hPa; 116.9°C) is 2,213 kJ/kg as a measure of the process heat that can be transferred to a heat sink (10, 11) at the corresponding condensation temperature of 116.9°C on the heat-emitting side of the heat pump device 1. For a process heat transfer of 1 MW output, a mass flow of 1,628 kg/h of circulating process fluid of the first pressure stage condensing in the heat exchanger 520 is therefore required, which then enters the mixing separator 400 of the second pressure stage as liquid circulating process fluid at about 116.9°C at a connection 407.

To deliver 1,628 kg/h of vaporous cycle fluid of the first pressure stage, compression by means of compressor 501 from a second pressure stage (450 hPa; 78.7°C) to the first pressure stage (1,800 hPa; 116.9°C) is required. In principle, a distinction can be made between dry compression and wet steam compression. In wet steam compression according to 1 or 2 the exiting mass flow already corresponds to the required delivery mass flow of 1,628 kg/h of vaporous cycle fluid of the first pressure stage, since exactly enough liquid cycle fluid is sprayed at the inlet of the compressor 501 to reduce the overheating that occurs during compression. In the case of dry compression, the overheating of the compression in the subsequent mixing separator 500 of the first pressure stage would have to be reduced according to 3 by spraying approximately 138 kg/h of liquid cycle fluid from the mixing separator 400 of the second pressure stage into the mixing separator 500 of the first pressure stage and, due to the overheating of the dense 501 dry compressed steam. These 138 kg/h of liquid cycle fluid therefore do not need to be compressed in compressor 501 by applying mechanical work to the first pressure stage. This means that the mass flow through compressor 501 only needs to be 1,628 kg/h - 138 kg/h = 1,490 kg/h. The electrical power consumption of compressor 501 is around 131.6 kW.

These 1,490 kg/h to be compressed in the compressor 501 exit from the connection 408 of the mixing separator 400 of the second pressure stage. The 1,628 kg/h of liquid cycle fluid of the first pressure stage condensed in the heat exchanger 520 enter via the connection 407 of the mixing separator 400, of which around 0.069 kg/kg evaporate through spontaneous evaporation, so around 112 kg/h of vaporous cycle fluid of the second pressure stage are formed through spontaneous evaporation. Of the remaining approximately 1,516 kg/h of liquid cycle fluid in the mixing separator 400, around 87 kg/h are required to reduce the superheat from compression to the second pressure stage, so that the mass flow of vaporous cycle fluid in the compressor 401 must be around 1,490 kg/h - 112 kg/h - 87 kg/h = 1,291 kg/h. The electrical power consumption of the compressor 401 is around 81.5 kW.

These 1,291 kg/h of vaporous cycle fluid to be compressed in the compressor 401 exit from the evaporator 200 and are previously evaporated there at the pressure stage (150 hPa; 54.0°C) of the evaporator 200 by means of indirect heating with an output of around 848 kW. This heating output is made up of around 34 kW of heat recovery from the liquid cycle fluid exiting the mixing separator 400 at connection 409 and around 811 kW of waste heat from a heat source (14, 15).

In the alternative of the heat pump device 1 in a minimal configuration with only a few main components according to 3 A total electrical power consumption of around 131.6 kW + 81.5 kW + 1.9 kW = 215 kW is used for the compression of vaporous cycle fluid in the compressors (401, 501) and for various pumps (201, 222, 251) for liquid cycle fluid to deliver 1,000 kW of process heat. The heat output coefficient COP _h is therefore 4.65 at a temperature difference of 116.9°C - 60°C = 56.9 K.

One in the 1 to 11 The compression shown by means of the corresponding compressors 301, 311, 401, 501, 601 within the respective alternative of the heat pump device 1 according to the invention represents in general terms that a compression by means of the corresponding compressors 301, 311, 401, 501, 601 of vaporous cycle fluid takes place by performing mechanical work either in a compressor in the design of a turbomachine, such as an axial fan, a radial fan, a turbocompressor or a turbine, or that the compression of vaporous cycle fluid takes place by performing mechanical work in a compressor in the design of a positive displacement machine, such as a piston compressor, a rotary piston compressor or a screw compressor.

One in 4 The alternative of the heat pump device 1 shown has, in contrast to the 3 that the compression by means of a compressor 401 of vaporous cycle fluid takes place in a thermal compressor, such as a vacuum steam jet pump, in which vaporous cycle fluid of a first pressure stage is used as motive steam, which reaches high speeds through thermal compression and thereby sucks in a suction steam in the form of vaporous cycle fluid with a pressure lower than that of the motive steam due to Venturi or Coanda effects, wherein the motive steam and suction steam then mix to form a mixed steam with the pressure of a second pressure stage, which has a higher pressure than the suction steam and a lower pressure than the motive steam.

One in 5 The alternative of the heat pump device 1 shown has, in contrast to the 4 that vaporous cycle fluid from a first pressure stage is further compressed in a compressor 601 and this compressed vaporous cycle fluid is used as motive steam, which reaches high speeds in a thermal compression and thereby sucks in a suction steam in the form of vaporous cycle fluid with a pressure lower than that of the motive steam due to Venturi or Coanda effects, wherein the motive steam and suction steam then mix to form a mixed steam with the pressure of a second pressure stage, which has a higher pressure than the suction steam and a lower pressure than the motive steam. In addition, in 5 shown that a cycle fluid condensed at the second pressure stage from the mixing separator 400 of a second pressure stage is introduced into vaporous cycle fluid from a first pressure stage to generate cycle fluid wet steam before this wet steam is compressed in the compressor 601 to a motive steam pressure stage. At the same time, a condensate pump 260 provides a sufficient injection pressure of liquid cycle fluid of a second pressure stage before the Compression by means of compressor 601 to a motive steam pressure stage.

One in 6 The alternative of the heat pump device 1 shown has, in contrast to the 3 that a further third pressure stage with a mixing separator 300 and a compressor 301 has been added and liquid cycle fluid from a mixing separator (400, 500) of a second or first pressure stage is introduced into the further mixing separator (300, 400) of a second pressure stage, the pressure and temperature of which are lower than the pressure and temperature of the introduced liquid cycle fluid. In this case, a condensate pump 201 conveys the amount of condensed cycle fluid which in this case and in contrast to the representation in the 1 to 5 exits the mixing condenser 300 of the next higher pressure level above the pressure level of the evaporator 200. At the same time, compression from a low to a second pressure level can take place either in the compressor 301 in the design of a turbomachine or displacement machine, or analogously to the representation in 4 or 5 in the design of a thermal compressor.

In a further alternative of a heat pump device 1, the performance coefficient can be increased by increasing the number of pressure stages to, for example, at least 4 pressure stages according to 6 For example, if instead of a first pressure stage (1,800 hPa; 116.9°C), a second pressure stage (450 hPa; 78.7°C) and a pressure stage of the evaporator 200 (150 hPa; 54.0°C) similar to the representations in the 1 to 3 If the gradation is reduced to, for example, a first pressure stage (1,800 hPa; 116.9°C), a second pressure stage (450 hPa; 78.7°C), a third pressure stage (225 hPa; 62.6°C) and a pressure stage of the evaporator 200 (150 hPa; 54.0°C), this results in a heating performance coefficient COP _h = 4.71 at a temperature difference of 116.9°C - 60°C = 56.9 K, since smaller differences between the pressures of the pressure stages mean that more exergy can be recovered from the condensed cycle fluid of higher stages.

In a further alternative of a heat pump device 1, the temperature lift can be increased by increasing the number of pressure stages at the same time, for example according to 6 the compression ratios of individual compressors (301, 401, 501) are increased. If, for example, a first pressure stage (4,044 hPa; 144.0°C), a second pressure stage (674 hPa; 88.9°C), a third pressure stage (225 hPa; 62.6°C) and a pressure stage of the evaporator 200 (150 hPa; 54.0°C) are set up, the result is a heating performance coefficient COP _h = 3.44 with a temperature difference of 144.0°C - 60°C = 84.0 K.

In a further alternative of a heat pump device 1, the performance coefficient can be increased by using the same temperature lift and the same number of pressure stages, for example at least four pressure stages according to 6 a distribution of the transfer of process heat to a first and a second heat sink with different sink temperatures according to 7 If all process heat is kept at a uniform sink temperature according to 6 For example, a heat output coefficient COP _h = 3.44 can be achieved at a temperature difference of 84.0 K.

When dividing the process heat according to another alternative of the heat pump device 1 according to 7 In such a way that approximately half is transferred at a first pressure stage and the other half at a second pressure stage by means of the heat exchangers 420, 520, a heating performance coefficient of at least COP _h = 4.35 can be achieved, for example, without having to reduce the maximum temperature lift of 84.0 K.

One in 8th The alternative of the heat pump device 1 shown has, in contrast to the illustration in 6 that a further pressure stage with a mixing separator 310, a compressor 311 and a condensate separator 312 was added as a third pressure stage, whereby the third pressure stage consists of 6 to a fourth pressure stage, and liquid cycle fluid from a mixing separator 400 of a second pressure stage is introduced into the mixing separator 310 assigned to the third pressure stage, the pressure and temperature of which are lower than the pressure and temperature of the introduced liquid cycle fluid of the pressure stage above.

One in 9 The alternative heat pump device shown in Fig. 1 has, in contrast to the illustration in 8th that a compression from one pressure stage of the evaporator 200 to a fourth pressure stage in a compressor 301 can also be carried out in the design of a thermal compressor, as analogous to the illustration in 4 or 5 and also applies analogously to compression processes between other pressure levels.

Based on the 1 to 9 It is clear from the analogies shown that a heat pump device 1 according to the invention with regard to the number of pressure stages built up, including the required components and devices per pressure stage, as well as with regard to the selection of the design of compressors, is fundamentally can be carried out variably. For example, starting from a second pressure stage, a compression of vaporous cycle fluid to motive steam can be carried out by means of a compressor 601, as in 9 shown by taking vaporous cycle fluid from a mixing separator 400 of a second pressure stage and compressing it as motive steam in a compressor 601 for the operation of a steam jet vacuum pump, with which vaporous cycle fluid of any lower pressure stage or, as in 9 shown, from the evaporator 200 is compressed to a fourth pressure stage.

One in 10 The alternative of the heat pump device 1 shown has, in contrast to the illustration in 8th that vaporous cycle fluid of a second pressure stage condenses in at least one heat exchanger 420 of a heat sink with at least one inlet 12 and at least one outlet 13 and indirectly transfers heat to these, wherein a cycle fluid condensed at a second pressure stage in the heat exchanger 420 of the heat sink enters a mixing separator 310 of a third pressure stage with a pressure below the pressure of the cycle fluid condensing in the heat exchanger 420 via at least one connection 317.

One in 11 The alternative of the heat pump device 1 shown has, in contrast to the illustration in 10 that vaporous cycle fluid of a third pressure stage condenses in at least one heat exchanger 420 of a heat sink with at least one inlet 12 and at least one outlet 13 and indirectly transfers heat to these, wherein a cycle fluid condensed at a third pressure stage in the heat exchanger of a heat sink enters a mixing separator 300 of a fourth pressure stage with a pressure below the pressure of the cycle fluid condensing in the heat exchanger 420 via at least one connection 307.

Based on the illustrations in the 10 and 11 it is shown that at least a second temperature level is provided for a further process heat transfer to at least one further heat sink at the level of the condensation temperature of a second or third pressure stage in at least one heat exchanger 420 of a heat sink with at least one inlet 12 and at least one outlet 13. This ensures that a process heat transfer at at least a second temperature level already enables a process heat supply up to a temperature close to the condensation temperature of that second or third pressure stage, whereby up to that point only the mechanical work for the compression up to that second or third pressure stage needs to be expended.

For a remaining process heat transfer up to a temperature close to the condensation temperature of a first pressure stage, only the difference in mechanical work for compression from the second or third pressure stage already reached to that first pressure stage would then have to be expended, but with a significantly lower specific process heat requirement and lower specific mechanical work for compression. By grading the provision of process heat in this way, the specific exergy destruction in relation to the total process heat transferred can be advantageously reduced, which further increases the overall heating performance factor.

In an alternative (not shown) of a heat pump device 1 with a further subdivision into at least 7 pressure stages, i.e. a first pressure stage (15,550 hPa; 200.0°C), a second (3,390 hPa; 142.6°C), a third (1,300 hPa; 107.0°C), a fourth (441 hPa; 78.2°C), a fifth (260 hPa; 65.8°C), a sixth (180 hPa; 57.7°C) and a pressure stage of the evaporator 200 (150 hPa; 54.0°C), a heating performance coefficient COP _h = 2.48 can be achieved over a temperature rise of just 200.0°C - 60°C = 140.0 K. Here, too, the performance coefficient can be increased by dividing up the transfer of process heat with the same temperature rise and the same number of pressure stages. If the distribution is such that about half of the process heat is provided at the first pressure stage (15,550 hPa; 200.0°C) and the other half at the third pressure stage (1,300 hPa; 107.0°C), the heating performance factor increases by around +40% to over COP _h = 3.40.

By definition, an electrical power consumption equal to the product of process heat and the reciprocal of the heating performance coefficient is required to provide process heat. This means that with a heat pump device 1, an electrical power consumption of less than 0.3 kWel/kW of process heat is required to provide process heat at a temperature level of, for example, 200°C, which equates to a saving of over 70% in primary energy consumption and a 100% reduction in fossil process heat generation using the heat pump device 1.

The heat pump devices 1 have a high energy efficiency and are flexible, modularly expandable and adaptable to a wide variety of requirements regarding required sink temperatures and/or other parameters.

List of reference symbols

1

Heat pump device

2

Control and regulation unit

5

Drying device

10

Inlet of a first heat sink

11

Process of a first heat sink

12

Inlet of a second heat sink

13

Process of a second heat sink

14

Supply of a heat source

15

Process of a heat source

200

Evaporator

201

Condensate pump

202

Control valve

220

Heat exchanger of a heat source

222

Circulation pump

250

Condensate pump for spraying in front of a compressor

251

Condensate pump for spraying into a mixing separator

252

Control valve

260

Condensate pump for spraying in front of a motive steam compressor

300

Mixing separator of a third pressure stage

301

Compressor to the pressure of a third pressure stage

306

Connection for the inlet of vaporous cycle fluid

307

Connection for the inlet of condensed cycle fluid

308

Connection for the outlet of vaporous cycle fluid

309

Connection for the outlet of condensed cycle fluid

310

Mixing separator of a further pressure stage

311

Compressor to the pressure of another pressure stage

312

Condensate separator of another pressure stage

316

Connection for the inlet of vaporous cycle fluid

317

Connection for the inlet of condensed cycle fluid

318

Connection for the outlet of vaporous cycle fluid

319

Connection for the outlet of condensed cycle fluid

320

Heat exchanger of a heat sink of a third pressure stage

322

Condensate separator of a third pressure stage heat exchanger

400

Mixing separator of a second pressure stage

401

Compressor to the pressure of a second pressure stage

402

Condensate separator of a second pressure stage

406

Connection for the inlet of vaporous cycle fluid

407

Connection for the inlet of condensed cycle fluid

408

Connection for the outlet of vaporous cycle fluid

409

Connection for the outlet of condensed cycle fluid

420

Heat exchanger of a heat sink of a second pressure stage

422

Condensate separator of a heat exchanger of a second pressure stage

500

Mixing separator of a first pressure stage

501

Compressor to the pressure of a first pressure stage

502

Condensate separator of a first pressure stage

506

Connection for the inlet of vaporous cycle fluid

507

Connection for the inlet of condensed cycle fluid

508

Connection for the outlet of vaporous cycle fluid

509

Connection for the outlet of condensed cycle fluid

520

Heat exchanger of a heat sink of a first pressure stage

522

Condensate separator of a heat exchanger of a first pressure stage

601

Compressor for generating motive steam

QUOTES INCLUDED IN THE DESCRIPTION

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

Cited patent literature

• EP 2321589 B1 [0010]

Claims (18)

Heat pump device (1) for energy-efficient generation of process heat, wherein the heat pump device (1) has a heat-absorbing side, a heat-emitting side, a cycle fluid and an evaporator (200) on the heat-absorbing side, and at least one heat exchanger (520) and at least one heat sink (10, 11) can be assigned to the heat pump device (1) on the heat-emitting side and at least one heat source (14, 15) on the heat-absorbing side, wherein liquid cycle fluid can be indirectly heated by means of the evaporator (200) by supplying heat from the assignable heat source (14, 15) and can be evaporated at a pressure stage of the evaporator (200), and the heat pump device (1) has at least one first compressor (401) downstream of the evaporator (200), wherein vaporous cycle fluid from the pressure stage of the evaporator (200) can be compressed to a second pressure stage, characterized in that the heat pump device (1) has at least one first mixing separator (400) in the second pressure stage, wherein the at least first mixing separator (400) has a first connection (406) for the inlet of vaporous cycle fluid, a second connection (407) for the inlet of condensed cycle fluid, a third connection (408) for the outlet of vaporous cycle fluid and optionally a fourth connection (409) for the outlet of condensed cycle fluid and at least one second compressor (501) is connected downstream of the mixing separator (400), so that vaporous cycle fluid emerging from the third connection (408) can be compressed in the at least second compressor (501) from the second pressure stage to a first pressure stage and by means of the at least one associated heat exchanger (520) can be indirectly separated from the vaporous cycle fluid of the first pressure stage Process heat can be transferred to the at least one assignable heat sink (10, 11) and the cycle fluid condensed in the at least one heat exchanger (520) can be returned via the second connection (407) to the at least first mixing separator (400) of the second pressure stage, wherein at the same time vaporous cycle fluid from the at least first compressor (401) can be introduced into the mixing separator (400) via the first connection (406). Heat pump device (1) according to Claim 1 , characterized in that the at least first mixing separator (400) has the fourth connection (409) for the outlet of condensed cycle fluid, so that the escaping condensed cycle fluid can be returned directly or indirectly to the evaporator (200) and/or fed to the at least second compressor (501). Heat pump device (1) according to Claim 1 or 2 , characterized in that the heat pump device (1) has at least one heat exchanger (520) assigned to the first pressure stage and/or one heat exchanger (420) assigned to the second pressure stage and/or one heat exchanger (320) assigned below the second pressure stage on the heat-emitting side. Heat pump device (1) according to one of the preceding claims, characterized in that the heat pump device (1) has at least one heat source (14, 15) and/or at least one heat sink (10, 11). Heat pump device (1) according to one of the preceding claims, characterized in that the heat pump device (1) has a second mixing separator, a third mixing separator, a fourth mixing separator and/or optionally further mixing separators (300, 310), wherein a further compressor (301, 311) is connected upstream of the respective mixing separator (300, 310). Heat pump device (1) according to Claim 5 , characterized in that a mixing separator or two or more mixing separators (300, 310) is or are connected upstream of the first mixing separator (400), wherein the vaporous cycle fluid of the respective upstream mixing separator (300, 310) can be fed to the subsequent mixing separator (310, 400) via a respective downstream compressor (311, 401) and/or the liquid cycle fluid from the respective downstream mixing separator (310, 400) can be returned to the upstream mixing separator (300, 310) at a higher pressure and/or a higher temperature than in the upstream mixing separator (300, 310). Heat pump device (1) according to one of the preceding claims, characterized in that the at least one heat exchanger (520) is followed by a condensate separator (522) and/or the second heat exchanger (420) is followed by a condensate separator (422) and/or the further heat exchanger (320) is followed by a condensate separator (322) for returning condensed cycle fluid from the first pressure stage to the first mixing separator (400) and/or condensed cycle fluid from the second or a third pressure stage to an upstream mixing condenser (300, 310). Heat pump device (1) according to one of the preceding claims, characterized in that that a condensate separator (312, 402, 502) for returning the condensed cycle fluid to the respective upstream mixing separator (300, 310, 400) or to the evaporator (200) is or are arranged downstream of the fourth connection (409) for the outlet of condensed cycle fluid from the first mixing separator, an upstream mixing separator and/or the respective mixing separator (300, 310, 400, 500). Heat pump device (1) according to one of the preceding claims, characterized in that the first compressor (301, 401) is designed as a thermal compressor, so that by means of the thermal compressor the vaporous cycle fluid of the first pressure stage can be mixed as motive steam with the vaporous cycle fluid from the evaporator (200) as suction steam and can be fed as compressed mixed steam to the first mixing separator (400) or the first upstream mixing separator (300). Heat pump device (1) according to Claim 9 , characterized in that a further compressor (601) in the first or second pressure stage or a further pressure stage is arranged upstream of the thermal compressor, so that the compressed, vaporous cycle fluid can be used as motive steam. Heat pump device (1) according to one of the preceding claims, characterized in that a further mixing separator (500) is connected downstream of the first mixing separator (400) in the second pressure stage, wherein the further mixing separator (500) is arranged in the first pressure stage. Heat pump device (1) according to one of the preceding claims, characterized in that the heat pump device (1) has a control and/or regulating device (2) for controlling and/or regulating components of the heat pump device (1) and optionally the at least one heat source (14, 15) and/or the at least one heat sink (10, 11). Heat pump device (1) according to one of the preceding claims, characterized in that the second connection (407) for the entry of condensed cycle fluid has a spraying device for spraying the liquid cycle fluid entering the first mixing separator or the respective mixing separator. Drying device for drying a material to be dried by means of a heated process gas stream, wherein the drying device comprises a heat pump device (1) according to one of the Claims 1 until 13 so that on the heat-emitting side of the heat pump device (1) the process gas stream can be heated as a heat sink (10, 11). Method for operating a heat pump device (1) for energy-efficient generation of process heat, in particular in a temperature range between 100°C and 250°C, by means of a heat pump device (1) according to one of the Claims 1 until 13 , with the following steps: - heating a liquid cycle fluid by indirect heat supply from a heat source (14, 15) and evaporating the liquid cycle fluid in an evaporator (200) at a pressure level of the evaporator (200), - compressing the vaporous cycle fluid by means of a first compressor (301, 401) to a second pressure level, - feeding the compressed vaporous cycle fluid in at least one mixing separator (400) to a second pressure level, - feeding the vaporous cycle fluid emerging from the at least one mixing separator (400) in at least one second compressor (501) and compressing the vaporous cycle fluid from the second pressure level to a first pressure level, - transferring process heat from the compressed, vaporous cycle fluid of the first pressure level to at least one assignable heat sink (10, 11), - and optionally returning the condensed cycle fluid of the first pressure stage into the at least one mixing separator (400). Procedure according to Claim 15 , characterized in that water, an alcohol and/or a water-soluble organic substance is or are used as the circulating process fluid. Method according to one of the Claims 15 or 16 , characterized in that a pressure increase is carried out in a stream of the condensed cycle fluid emerging from the mixing separator (300, 400) by means of a condensate pump (201), then this stream is mixed with a stream of liquid cycle fluid which was previously indirectly heated in a heat exchanger (220) of a heat source (14, 15), and the mixed streams are used for indirectly heating the evaporator (200). Method according to one of the preceding claims, characterized in that a pressure increase is carried out in a stream of the condensed cycle process fluid emerging from the at least one mixing separator (300, 400) by means of at least one condensate pump (201), then this stream is converted into vaporous cycle process fluid from the at least one Mixing separator (400) for generating cycle fluid wet steam before the cycle fluid wet steam is compressed to the first pressure stage in the downstream compressor (501).

Images