EP4160109A1 - Système de refroidissement et appareil de laboratoire doté d'un système de refroidissement - Google Patents

Système de refroidissement et appareil de laboratoire doté d'un système de refroidissement Download PDF

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Publication number
EP4160109A1
EP4160109A1 EP22196688.0A EP22196688A EP4160109A1 EP 4160109 A1 EP4160109 A1 EP 4160109A1 EP 22196688 A EP22196688 A EP 22196688A EP 4160109 A1 EP4160109 A1 EP 4160109A1
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EP
European Patent Office
Prior art keywords
refrigerant
compressor
pressure
temperature
designed
Prior art date
Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
Pending
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EP22196688.0A
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German (de)
English (en)
Inventor
Kai Peitzberg
Daniel Langer
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Thermo Electron LED GmbH
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Thermo Electron LED GmbH
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Publication of EP4160109A1 publication Critical patent/EP4160109A1/fr
Pending legal-status Critical Current

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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B9/00—Compression machines, plants or systems, in which the refrigerant is air or other gas of low boiling point
    • F25B9/10—Compression machines, plants or systems, in which the refrigerant is air or other gas of low boiling point with several cooling stages
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B1/00—Compression machines, plants or systems with non-reversible cycle
    • F25B1/10—Compression machines, plants or systems with non-reversible cycle with multi-stage compression
    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B04—CENTRIFUGAL APPARATUS OR MACHINES FOR CARRYING-OUT PHYSICAL OR CHEMICAL PROCESSES
    • B04B—CENTRIFUGES
    • B04B15/00—Other accessories for centrifuges
    • B04B15/02—Other accessories for centrifuges for cooling, heating, or heat insulating
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B40/00—Subcoolers, desuperheaters or superheaters
    • F25B40/02—Subcoolers
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B41/00—Fluid-circulation arrangements
    • F25B41/30—Expansion means; Dispositions thereof
    • F25B41/31—Expansion valves
    • F25B41/33—Expansion valves with the valve member being actuated by the fluid pressure, e.g. by the pressure of the refrigerant
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B43/00—Arrangements for separating or purifying gases or liquids; Arrangements for vaporising the residuum of liquid refrigerant, e.g. by heat
    • F25B43/006—Accumulators
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B9/00—Compression machines, plants or systems, in which the refrigerant is air or other gas of low boiling point
    • F25B9/002—Compression machines, plants or systems, in which the refrigerant is air or other gas of low boiling point characterised by the refrigerant
    • F25B9/008—Compression machines, plants or systems, in which the refrigerant is air or other gas of low boiling point characterised by the refrigerant the refrigerant being carbon dioxide
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2313/00—Compression machines, plants or systems with reversible cycle not otherwise provided for
    • F25B2313/001—Compression machines, plants or systems with reversible cycle not otherwise provided for with two or more accumulators
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2400/00—General features or devices for refrigeration machines, plants or systems, combined heating and refrigeration systems or heat-pump systems, i.e. not limited to a particular subgroup of F25B
    • F25B2400/13—Economisers
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2400/00—General features or devices for refrigeration machines, plants or systems, combined heating and refrigeration systems or heat-pump systems, i.e. not limited to a particular subgroup of F25B
    • F25B2400/23—Separators
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2600/00—Control issues
    • F25B2600/25—Control of valves
    • F25B2600/2509—Economiser valves
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2600/00—Control issues
    • F25B2600/25—Control of valves
    • F25B2600/2513—Expansion valves
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2700/00—Sensing or detecting of parameters; Sensors therefor
    • F25B2700/19—Pressures
    • F25B2700/195—Pressures of the condenser
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2700/00—Sensing or detecting of parameters; Sensors therefor
    • F25B2700/19—Pressures
    • F25B2700/197—Pressures of the evaporator
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2700/00—Sensing or detecting of parameters; Sensors therefor
    • F25B2700/21—Temperatures
    • F25B2700/2116—Temperatures of a condenser
    • F25B2700/21163—Temperatures of a condenser of the refrigerant at the outlet of the condenser
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2700/00—Sensing or detecting of parameters; Sensors therefor
    • F25B2700/21—Temperatures
    • F25B2700/2117—Temperatures of an evaporator
    • F25B2700/21174—Temperatures of an evaporator of the refrigerant at the inlet of the evaporator
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2700/00—Sensing or detecting of parameters; Sensors therefor
    • F25B2700/21—Temperatures
    • F25B2700/2117—Temperatures of an evaporator
    • F25B2700/21175—Temperatures of an evaporator of the refrigerant at the outlet of the evaporator

Definitions

  • the present invention relates to a cooling system that can be used to cool or regulate the temperature of a device.
  • the present invention relates to a cooling system for use in laboratory equipment.
  • laboratory equipment which includes, inter alia, centrifuges, incubators and biological safety cabinets.
  • a cooling system it is known to cool laboratory equipment, for example centrifuges, by means of a cooling system or to regulate the temperature of the laboratory equipment (for example the centrifuge).
  • Corresponding cooling systems usually have an evaporator, a compressor, a cooling component and an expansion device, which are connected to one another in a circuit via a line system.
  • Such a cooling system is operated with a refrigerant that undergoes changes of state in the circulatory system and therefore draws heat from the environment at one process point (typically at the evaporator) and gives off heat to the environment at another process point (typically at the cooling component).
  • the cooling component heat is extracted from the refrigerant in the system, i.e. the refrigerant is cooled (e.g. by a fan).
  • the cooling system is improved in several of these aspects.
  • the invention relates to a cooling system, the cooling system comprising an evaporator, a first compressor, a second compressor, a cooling component, an expansion device and a line system.
  • the Piping interconnects the evaporator, first compressor, second compressor, refrigeration component and expansion device.
  • the refrigeration system includes a refrigerant, where the refrigerant is carbon dioxide.
  • the first compressor and the second compressor are arranged in series with each other.
  • the refrigeration system can absorb heat at the evaporator and release heat at the refrigeration component.
  • the cooling system can in particular comprise a closed refrigerant circuit, so that the refrigerant circulates in the cooling system without material-bound exchange with the ambient atmosphere and/or a secondary refrigerant circuit.
  • the cooling component can be a heat exchanger which is designed to thermally couple the refrigerant to the ambient atmosphere in order to efficiently cool the refrigerant.
  • the refrigerant can be gaseous in the refrigeration component.
  • the cooling component can be a condenser, which is designed to convert the refrigerant into a liquid state.
  • the cooling system is multi-stage.
  • the advantage can be achieved that particularly high pressures can be achieved, or a high pressure can be achieved as energy-efficiently as possible.
  • the refrigerant may be sequentially compressed from a first pressure through at least one intermediate pressure to a final pressure.
  • a further compressor can be provided for each intermediate compression. This allows the compression process to be distributed over several compressors.
  • the expansion device can be designed in particular as an expansion valve.
  • the expansion valve can be designed to be controllable in order to implement controllable pressure regulation, in particular a pressure reduction.
  • the refrigerant can expand through the expansion valve.
  • the cooling component can be arranged downstream of the compressor and/or the further compressor in the direction of flow of the refrigerant.
  • the refrigerant is carbon dioxide (hereinafter also abbreviated as CO 2 or R744).
  • R744 may have low toxicity. Therefore, CO 2 is an alternative to ammonia in particular.
  • the refrigeration system can go through a transcritical cycle: the critical point can be exceeded. As a result, a pressure and/or a temperature at the compressor can be increased.
  • a global warming potential can be reduced compared to conventional refrigerants.
  • possible refrigerants can be flammable (e.g. A3 and A2L class refrigerants) or CO 2 , which is non-flammable, can be used.
  • the use of CO 2 can result in an operating pressure of up to 140 bar.
  • an increased operating pressure for example pressures greater than 60 bar, can occur.
  • CO 2 as a refrigerant differs from various cooling systems from the prior art, in which the refrigerant for such flammable refrigerants or hydrocarbons in which at least one hydrogen has been replaced by a halogen (e.g. so-called fluorinated and halogenated fluorocarbons or F gases) applied.
  • a halogen e.g. so-called fluorinated and halogenated fluorocarbons or F gases
  • a permissible ambient temperature for the operation of the laboratory device can be up to 40°C.
  • CO 2 can already be gaseous as a cooling medium in the cooling system.
  • CO 2 can be cooled via heat exchange with the surrounding atmosphere.
  • the CO 2 can remain gaseous. Therefore, a high pressure of at least 60 bar, preferably at least 70 bar, can be achieved in the cooling system.
  • a corresponding compression can be achieved by means of a two-stage compression.
  • embodiments of the invention thus provide a new cooling system for the efficient use of CO 2 as a refrigerant.
  • the refrigeration system may be configured to perform a transcritical vapor compression cycle.
  • the refrigerant can at least at times, or in parts of the cycle, above a critical point of the primary be refrigerant.
  • part of the cycle can occur at pressures above the critical point and another part of the cycle below the critical point.
  • the critical point can mark the upper limit for heat transfer processes based on evaporation or condensation. At temperatures and pressures above this critical point, it is no longer possible to clearly distinguish between liquid and vapor. All refrigerants have a critical point, but conventional refrigerants can always fall below this point in a typical refrigeration cycle.
  • the refrigeration circuit may have a specific cycle performance characteristic that corresponds to a preferred operating condition point at which the refrigeration system operates at optimal cycle efficiency.
  • the present cooling system can achieve the advantage that, if there is a deviation from this operating state point, it can be readjusted accordingly in order to achieve optimum system efficiency.
  • the control here includes in particular the adjustment of the local refrigerant temperature through internal heat transfer or secondary refrigerant flows. Internal heat transfer can be defined as heat conduction from a first cooling system section to a second cooling system section by thermal coupling of these cooling system sections. Auxiliary flows of refrigerant can be realized by additional line sections, which branch off refrigerant from a main circuit and feed it back to the main circuit at another point. This can realize a material-bound heat transfer.
  • the refrigerant compressed to a medium pressure can be fed to the second compressor, which compresses the refrigerant from the medium pressure to the high pressure.
  • the first compressor can be optimized for a first compression from low pressure to medium pressure and the second compressor can be optimized for compression from medium pressure to high pressure. Accordingly, each compression stage can be implemented with increased efficiency.
  • the first compressor and/or the second compressor can be embodied as fully hermetic compressors, which prevent the refrigerant from flowing into the surrounding atmosphere.
  • Embodiments of the invention can use selective heat transfer within the cooling system to regulate a temperature of the cooling medium in the different cooling system areas to a respective optimal value.
  • a pressure stroke can be achieved both process- and cost-efficiently through a two-stage compression.
  • a hot gas end temperature can be reduced.
  • regulation of the refrigerant temperature can be achieved without using a fluidly separated secondary refrigerant circuit.
  • a mixture of refrigerants from different process points of the refrigerant circuit can be used.
  • temperature regulation of the refrigerant by means of external devices can thus be dispensed with.
  • the present invention can achieve improved system security since crash safety of the laboratory device can be increased.
  • crash safety of the laboratory device can be increased.
  • the use of other refrigerants in particular flammable or toxic refrigerants (propane - R290, ammonia - R717) can be avoided.
  • a non-flammable refrigerant can increase device safety: The rotor of a centrifuge can penetrate a boiler wall and thus evaporator tubes carrying refrigerant. In comparison to flammable or toxic refrigerants, escaping CO 2 can be classified as a low safety risk.
  • CO 2 Due to high operating pressures, CO 2 can have both a high density during evaporation and a high volumetric cooling capacity during heat dissipation. As a result, the advantage of a reduced installation space can be realized in particular when the cooling system is integrated into a centrifuge. For example, an installation space volume of a cooling component, or a condenser, and/or an evaporator can be reduced.
  • the efficiency of the cooling system can be increased: the pressure ratio of low pressure to high pressure of the refrigerant can be limited. Furthermore, with the two-stage compression, the pressure ratio can be increased and the possible operating range can thus be expanded, in particular without external dissipation of heat on the basis of a secondary circuit.
  • the operating range can be determined in relation to a specific ambient temperature range for the operation of the laboratory device.
  • the cooling system can have a reduced device complexity and thus a lower susceptibility to errors and/or a reduced maintenance effort.
  • the cooling component can include a gas cooler and/or a condenser.
  • the gas cooler can provide the refrigerant in gaseous form and at a reduced temperature at an outlet of the cooling component.
  • the condenser can provide the refrigerant at the exit of the refrigeration component in a liquid state and at a reduced temperature.
  • the refrigeration system may be configured to perform a transcritical vapor compression cycle. Accordingly, for example in the second compression stage, a pressure and a temperature which exceed a respective critical value can be achieved by means of the second compressor.
  • the cooling component can also be designed to convert the refrigerant from a gaseous phase into a liquid phase.
  • the cooling component is also designed to withstand pressures and temperatures above the critical point.
  • the refrigeration system may be configured to perform a subcritical vapor compression cycle. Accordingly, a pressure and a temperature, in particular on the second compressor, can also be regulated in some operating states in such a way that the critical point of CO2 is not reached or exceeded. With a design for a subcritical range, reduced requirements with regard to permissible pressures and/or temperatures can be used in order to use components that meet the reduced requirements but would be unsuitable for transcritical operation, for example. A cost reduction can be achieved as a result.
  • the cooling system can have a cooling capacity of 10 W to 100 kW, preferably 500 W to 10 kW. Accordingly, the refrigeration circuit can be scaled in a range from small mobile devices and laboratory benchtop devices to large industrial systems.
  • the refrigeration system can have a main circuit, which has the evaporator, the first compressor, the second compressor, the refrigeration component, the expansion device and at least part of the piping system. Furthermore, the refrigerant can be present in the main circuit. Most of the refrigerant flows through the main circuit. This can be defined as a flow of refrigerant that is greater than 50%wt (percent by weight) of the total refrigerant present in the refrigeration system. It can siblings Line sections may be provided which comprise a correspondingly smaller part of the refrigerant. In particular, the refrigerant can be routed via secondary line sections for internal temperature regulation.
  • the refrigeration component may be located downstream of the second compressor and upstream of the expander. Accordingly, the cooling component can reduce a temperature of the refrigerant before the refrigerant is supplied to the expander.
  • the cooling component can be embodied as a heat sink and extract heat from the refrigerant, which can be given off to the surrounding atmosphere or an external cooling system.
  • the refrigeration component may be configured to cool the refrigerant downstream of the second compressor. This allows the refrigerant to flow into the refrigeration component at a point in the cycle where the refrigerant temperature is at its maximum.
  • the refrigeration component can lower the enthalpy by lowering the refrigerant temperature. This process can be isobaric.
  • cooling to a temperature below 30°C, preferably below 31°C, can be achieved.
  • the refrigeration system can be set up in such a way that when the refrigerant leaves the first compressor at an outlet temperature, the refrigerant is fed to the second compressor at an inlet temperature which is lower than the outlet temperature.
  • a maximum temperature of the refrigerant in the second compressor, or at the outlet of the second compressor can be reduced, so that the thermal load on the second compressor can be reduced.
  • the enthalpy can also be reduced isobaric here.
  • the inlet temperature and the outlet temperature can differ by a temperature difference that is greater than 1K, preferably greater than 2K, more preferably greater than 3K.
  • the temperature difference can be in the range of 3K to 8K.
  • a heat sink can be provided between the compressors, which extracts heat from the refrigerant and transports it away from the cooling system.
  • a heat transfer to a central cooling system and/or a heat transfer to the surrounding atmosphere can be implemented.
  • the first compressor and/or the second compressor may be configured as scroll compressors, reciprocating compressors, screw compressors, rotary piston compressors, or a combination of the above. This allows a for the respective pressure range optimized compressor type can be used. In particular, a compressor type that differs from the second compressor can be used as the first compressor.
  • the cooling system may include a return section fluidly connected to the main circuit at a first connection point and at a second connection point.
  • the second junction may be located in the main loop downstream of the first compressor and upstream of the second compressor.
  • the efficiency of the cooling system can be increased via the return section.
  • a volumetric efficiency of the refrigeration system can be increased, so that a hot gas temperature at the exit of the second compressor is reduced.
  • the cooling capacity can be scaled with the capacity of the cooling system.
  • the volumetric efficiency can be a volumetric efficiency of the displacement of the compressors or the actual delivery volume of the compressors. High hot gas temperatures can influence, in particular reduce, the durability of a machine oil for lubricating the components. Therefore, with a temperature reduction, the durability of the components can be extended.
  • Refrigerant can be injected into the refrigerant flow in the main circuit in front of an inlet of the second compressor via the recirculation section.
  • the injected coolant can have a reduced temperature, in particular a lower temperature than the coolant of the main circuit at the injection point, here in particular at the second connection point.
  • a mixed flow with a reduced temperature can thus be provided by mixing the refrigerant at the injection point on the second compressor.
  • a heat load in a centrifuge can be dynamic, i.e. an evaporating temperature can vary during operation. Depending on the operating mode of the centrifuge, varying heat loads can occur. For example, different rotors can be used, and different setpoint speeds and/or setpoint temperatures can be set. With a variable post-injection, the hot gas temperature at the compressor outlet or at the outlet of the second compressor can be regulated in order to take account of the variable loads on the evaporator.
  • the cooling system can be set up in such a way that the refrigerant in the return section at the second connection point has a lower specific enthalpy than the refrigerant in the main circuit immediately upstream of the second connection point.
  • the enthalpy in the return section can be increased essentially isobaric and/or essentially isothermally.
  • the steam content can be changed by changing the enthalpy.
  • the steam content can be selected in such a way that the formation of droplets is prevented in order to protect the compressor.
  • steam with a small liquid content can be introduced.
  • the injection quantity can be chosen such that the enthalpy inflow makes it possible to achieve an increased vapor content.
  • the refrigeration system may include a heat exchanger having a primary side located in the main loop downstream of the refrigeration component.
  • the heat exchanger can be designed to cool the refrigerant in the main circuit.
  • the heat exchanger is designed to extract heat from the refrigerant downstream of the cooling component, more preferably at an outlet of the cooling component.
  • the heat exchanger can be used to optimize the energy use of the cooling system.
  • the cooling system can be used in devices for analyzing medical samples, in particular in centrifuges, which can be operated at an ambient temperature of up to 40°C. Depending on the temperature, the optimum high pressure can also increase. The optimum high pressure can depend on the COP. If the temperature and pressure values exceed a critical point of the refrigerant, the heat dissipation process can run transcritically as gas cooling. Transcritical gas cooling can take place isobaric. Liquefaction, on the other hand, can take place isobaric and isothermally to a large extent. Due to the increased pressure of a transcritical gas cooling, the drive power of the compressors can be increased.
  • the heat exchanger can be designed to provide the refrigerant at a predetermined temperature below an initial temperature of the cooling component at the first connection point and/or in the line section.
  • the line section may be delimited by the first connection point and an inlet of the expansion device.
  • the heat exchanger can isobarically extract heat from the refrigerant in the main circuit.
  • the refrigerant can pass through the heat exchanger in a transcritical, gaseous or liquid state.
  • the heat exchanger can be arranged upstream of the expansion device in the main circuit. As a result, the heat exchanger can contribute to the cooling of the refrigerant before it enters the expansion device.
  • the heat exchanger may include a secondary side located in the return section. Furthermore, the heat exchanger can be designed to absorb heat from the refrigerant by means of the primary side and to emit the absorbed heat to the refrigerant by means of the secondary side in order to heat the refrigerant in the return section.
  • a refrigerant flow in the direction of the second connection point can nevertheless have a lower temperature than a refrigerant flow of the main circuit, in particular than a refrigerant flow from the first compressor to the second compressor in the main circuit.
  • Even with a further heat exchanger arranged between the compressors and in front of the second connection point the temperature in the main circuit at the second connection point can always be higher than the temperature of a refrigerant flow from the return section to the second connection point.
  • a refrigerant temperature between the outlet of the primary side of the heat exchanger and the inlet of the secondary side of the heat exchanger can be essentially identical. Heat losses can occur here through the line and/or other components, which can produce a small temperature difference.
  • the secondary side can increase the enthalpy in the return section isobaric and/or isothermally.
  • the refrigerant can be in a wet vapor phase.
  • the refrigerant can be converted from the wet vapor phase into a gaseous phase at the second connection point. This can be realized in particular by mixing at the connection point.
  • the secondary side can be arranged in the return section upstream of the second connection point.
  • a partial refrigerant flow can flow from the first connection point through the secondary side to the second connection point.
  • the heat transfer of the heat exchanger can essentially take place internally, i.e. here heat can be conducted from a flow in the main circuit to a flow in the return section. Heat flow and mass flow can be regulated separately from each other.
  • the first compressor can be designed to compress the refrigerant from a primary pressure area in a secondary pressure area, wherein the secondary Pressure range in relation to the primary pressure range has higher pressures.
  • the temperature of the refrigerant and in particular the enthalpy can change between the compression stages.
  • the first compressor can be used to set a medium pressure for further compression by the second compressor.
  • a coolant temperature that is optimal for the second compressor or at least a sufficiently low one can be set between the first compressor and the second compressor.
  • the refrigerant temperature between the compressors can be realized by mixing refrigerant via the return section and/or by active cooling using a heat sink.
  • the second compressor may be configured to compress the refrigerant from the secondary pressure range to a tertiary pressure range, the tertiary pressure range having higher pressures relative to the secondary pressure range.
  • a pressure difference that the first compressor achieves can be smaller than a pressure difference that the second compressor achieves.
  • the second compressor can be designed for a higher inlet pressure than the first compressor.
  • suitable compressor capacities or suitable compressor types can be provided for the intended pressure ranges.
  • the cooling system can comprise a further expansion device which is arranged in the return section and is designed to lower the refrigerant from the tertiary pressure area into the secondary pressure area.
  • the pressure reduction can be implemented isenthalpically, so that both a pressure and a temperature of the refrigerant in the return section can be reduced. With the drop in temperature, the refrigerant can be transferred from a gaseous and/or transcritical phase to a wet vapor phase.
  • the further expansion device can be arranged upstream of the secondary side of the heat exchanger and/or downstream of the first connection point.
  • the secondary side can be thermally conductively coupled to the primary side.
  • the further expansion device can be designed to control a refrigerant flow into the return section.
  • the further expansion device can be designed as a valve, with a volume flow through the expansion device scaled to an opening degree of the expander.
  • a degree of opening of the expansion device can advantageously be controlled as a function of pressure and/or temperature.
  • the expansion device can be used to control a coolant flow through the return section as a function of a temperature at the outlet of the second compressor, or a hot gas temperature.
  • An inlet temperature at the inlet of the second compressor may scale with refrigerant flow through the recirculation section.
  • the refrigerant in the return section can have a lower temperature than the refrigerant between the compressors in the main circuit. Accordingly, a temperature at the inlet of the second compressor and thus also indirectly the temperature at the outlet of the second compressor can be regulated via the refrigerant flow through the recirculation section.
  • the cooling system can comprise a further heat exchanger which has a primary side which is arranged in the main circuit upstream of the expansion device and/or downstream of the cooling component.
  • a further heat exchanger which has a primary side which is arranged in the main circuit upstream of the expansion device and/or downstream of the cooling component.
  • a coefficient of performance can be increased through the use of post-injection and the heat exchanger. Furthermore, the refrigerant can be further supercooled upstream of the expansion device by the additional heat exchanger. As a result, a greater specific evaporation capacity can be achieved, so that a mass flow of the refrigerant and the resulting compressor speeds of the first compressor and/or the second compressor can be reduced. An evaporation capacity can be adjusted, in particular increased, with a post-injection via the return section.
  • the further heat exchanger can be arranged downstream of the primary side of the heat exchanger in the main circuit. Accordingly, the refrigerant in the main circuit can be cooled by the heat exchanger and the further heat exchanger.
  • the further heat exchanger can comprise a secondary side, which is arranged downstream of the evaporator and/or upstream of the first compressor in the main circuit. Furthermore, the further heat exchanger can be designed to absorb heat from the refrigerant by means of the primary side and to release the absorbed heat to the refrigerant by means of the secondary side in order to heat the refrigerant upstream of the first compressor. Accordingly, the further heat exchanger can implement a materially decoupled, internal heat transfer between different sections of the main circuit.
  • the heating of the refrigerant before the inlet of the first compressor can increase the temperature of the refrigerant, whereby in particular a wet vapor phase of the refrigerant can be overcome and the refrigerant is present in the gaseous phase at the inlet of the first compressor. Downstream of the process from the evaporator, the refrigerant can already be present at least partially in the gaseous phase. However, at least part of the refrigerant can still be in the liquid phase. For example, droplets of refrigerant may be suspended in a gaseous flow of refrigerant. These refrigerant droplets can damage the compressor.
  • the volumes of the refrigerant droplets can be reduced, or the droplets can be at least partially evaporated in the refrigerant flow.
  • the first compressor can thus be protected against liquid hammer and/or the wear can be reduced by a reduced number and/or a reduced size of the refrigerant droplets.
  • no additional energy may be necessary for the heat transfer between the primary side and the secondary side of the further heat exchanger. Accordingly, this heat transfer can increase the efficiency and/or durability of the cooling system.
  • the further heat exchanger can cause a pressure loss, which can be overcome or compensated for by the compressor.
  • the further heat exchanger can be a line-to-line heat exchanger.
  • the primary side is advantageously coupled to the secondary side in a heat-conducting manner but not in a material-transferring manner.
  • the primary side can be coupled to the secondary side via a thermally conductive material.
  • an integral metal connection can be realized here.
  • the cooling system may include a liquid separator configured to separate the refrigerant in a liquid state.
  • the liquid separator in the main circuit can be downstream of the evaporator and/or located upstream of the first compressor.
  • the refrigerant can be in a wet-vapor phase, so that heat can be absorbed isobaric and isothermally by the refrigerant and an enthalpy increase is realized. This heat absorption can be used to cool laboratory equipment, in particular a rotor chamber of a centrifuge.
  • the refrigerant can be present at a phase transition between the wet vapor phase and the liquid phase.
  • the liquid collector can collect the still liquid parts of the refrigerant in order to provide the first compressor with a purely gaseous flow of refrigerant.
  • the liquid separator can be arranged upstream of the secondary side of the additional heat exchanger.
  • the secondary side of the further heat exchanger can form a type of refrigerant droplet filter that removes refrigerant droplets that pass through the liquid separator.
  • a heating capacity of the further heat exchanger can be used with the liquid separator in order to provide a pure gas phase of the refrigerant.
  • the further heat exchanger can be designed in such a way that a heating output is sufficient to remove droplets of refrigerant.
  • the heating capacity can be less than a capacity that can convert the entire refrigerant flow into the gas phase without a liquid separator.
  • the refrigerant can be efficiently provided in gaseous form to the first compressor through the combination of liquid separator and additional heat exchanger. Furthermore, the enthalpy can be increased isobaric via the secondary side of the further heat exchanger, so that the refrigerant moves away from the phase transition between the wet vapor phase and the gas phase.
  • the cooling system can include a filter drier, which can be designed to remove water from the refrigerant.
  • the water fractions can be separated and/or filtered.
  • the filter dryer can be arranged downstream of the cooling component and/or upstream of the expansion device in the main circuit.
  • the filter drier can advantageously be designed to bind moisture and/or acid present in the refrigerant.
  • the filter drier can filter dirt and/or other foreign bodies present in the refrigerant. The advantage can be achieved that acidification of a component oil can be suppressed and the compressors can thereby be protected.
  • the filter drier can be arranged between the primary side of the heat exchanger and the primary side of the further heat exchanger in the main circuit.
  • the refrigerant can pass through the filter drier in a liquid phase. Filtering, in particular upstream of the expansion device, can prevent the formation of frozen water in the line system, so that the flow of the refrigerant is not impeded.
  • the first compressor can also be protected from damage caused by ice particles.
  • the filter drier may be located downstream or upstream of the first junction. So that a partial flow of the refrigerant can be fed into the return section before filtering.
  • the refrigeration system may include a medium-pressure vessel, which may be configured to divide the refrigerant into a liquid phase and a gas phase, wherein the medium-pressure vessel may be located downstream of the refrigeration component and/or upstream of the expansion device in the main circuit.
  • the medium-pressure container can be a liquid separator in which a pressure within the secondary pressure range, or a medium pressure, prevails.
  • the liquid separator separates the liquid phase and the gas phase of the refrigerant at medium pressure.
  • the liquid can be used as a template for the expander, so that a flow of refrigerant flows from the liquid separator to the expander in the liquid phase.
  • the gas phase of the refrigerant can be fed to the second compressor via the return section.
  • the second compressor can advantageously be designed to suck off the refrigerant in the gas phase from the medium-pressure container.
  • This recirculated refrigerant gas can be mixed with refrigerant gas from the first compressor.
  • the recirculated refrigerant gas in the primary pressure region may be mixed with the refrigerant in the main loop. Whether the refrigerant is mixed in the secondary pressure section or in the primary pressure section may be determined by a type of the first compressor.
  • the container inlet can be arranged in the main circuit downstream of the heat exchanger and/or downstream of the filter dryer.
  • already filtered and/or cooled refrigerant can be provided to the medium-pressure container.
  • the first connection point can be formed by the medium-pressure container.
  • Refrigerant in the gas phase can be coupled into the recirculation section via the second container outlet. With the phase separation, the refrigerant in the gas phase can be introduced into the return section and the refrigerant in the liquid phase can correspondingly be fed to the expansion valve.
  • the refrigerant in the gas phase can have an increased enthalpy compared to the refrigerant in the liquid phase.
  • the first heat exchanger can bring about an additional heat input into the refrigerant in order to bring the refrigerant further into the gas phase from a phase transition from wet vapor to gaseous to a higher temperature. In this way, in particular, the liquid fraction of the refrigerant can be reduced upstream of the second compressor before it is injected into the main circuit.
  • the medium-pressure container can be designed to provide the refrigerant at the first container outlet, advantageously in the liquid phase, at least at a phase transition from the liquid phase to the wet vapor phase.
  • the enthalpy can be reduced to such an extent via the medium-pressure tank that the refrigerant changes from the wet vapor phase to the liquid phase. This enthalpy reduction can be realized isothermally.
  • the cooling system can include a high-pressure control device, which can be designed to reduce the pressure of the refrigerant, in particular to reduce the pressure from the tertiary pressure area into the secondary pressure area or to reduce the pressure within the tertiary pressure area.
  • the temperature of the refrigerant can change when there is a flow through the high-pressure control device.
  • the pressure control can be effected isenthalpic.
  • the refrigerant can be transferred from the liquid phase to the wet vapor phase.
  • the refrigeration system may include an enthalpy, Have pressure and / or temperature minimum. Accordingly, heat absorption at the evaporator can be maximized.
  • the high pressure control device may be located upstream of the intermediate pressure vessel, and/or downstream of the refrigeration component, downstream of the first heat exchanger, and/or downstream of the filter drier in the main loop.
  • the high-pressure control device can be arranged at the tank inlet of the medium-pressure tank in order to reduce an inlet pressure of the medium-pressure tank.
  • the high-pressure control device is advantageously designed to bring about a phase transition of the refrigerant from the liquid phase or a transcritical phase into the wet vapor phase.
  • the pressure change can be realized isenthalpic.
  • the high pressure control device may be controllable based on a pressure, in particular based on a pressure of the refrigerant downstream of the refrigeration component.
  • the pressure reduction can also be controlled by means of the high-pressure control device on the basis of a temperature, in particular a temperature at an outlet of the first heat exchanger.
  • a pressure at the outlet of the first heat exchanger can advantageously be used to control the high-pressure control device.
  • the high pressure control device may be controllable based on a pressure downstream of the first heat exchanger, based on a pressure upstream of the high pressure control device, and/or based on a pressure upstream of the filter drier.
  • the high-pressure control device can be controlled based on a maximum pressure of the refrigerant.
  • the expansion device can be an overheating control device and configured to regulate overheating of the refrigerant at the evaporator. This can achieve the advantage that a low-pressure liquid separator can be omitted. A liquid separator in the primary pressure area can therefore be dispensed with. In addition, there is no need to heat the refrigerant upstream of the first compressor. Accordingly, with a high-pressure control device and a medium-pressure container, the second heat exchanger can be omitted.
  • the overheating control device can be designed to control an injection temperature at the evaporator in such a way that it corresponds to the saturation temperature.
  • the expansion device can be designed to regulate a pressure of the refrigerant, in particular to reduce a pressure of the refrigerant from the secondary pressure area to the primary pressure area or to reduce it from the tertiary pressure area to the primary pressure area.
  • a pressure reduction from the tertiary pressure range into the secondary pressure range can be implemented upstream of the process by means of the high-pressure control device.
  • This can be a single-stage expansion, with a two-stage compression being realized by means of the first compressor and the second compressor.
  • the expansion device can be controllable based on a pressure, in particular based on a pressure of the refrigerant downstream of the evaporator and/or upstream of the first compressor.
  • the cooling system can have a maximum enthalpy and the refrigerant can have a minimum pressure.
  • the expansion device can be controlled as a function of a pressure to be achieved and/or a temperature to be achieved at the outlet of the evaporator. Accordingly, the heat absorption at the evaporator can be increased.
  • an inlet pressure and an inlet temperature can be set with the expansion device in such a way that the refrigerant can absorb the greatest possible amount of heat when flowing through the evaporator, or the greatest possible increase in enthalpy can be realized, so that the cooling capacity on the evaporator is as great as possible.
  • the evaporator can heat the refrigerant up to the phase transition from wet vapor phase to gas phase or beyond the phase transition.
  • the expansion device can be controllable based on a pressure upstream of the further heat exchanger. For example, a pressure between the evaporator or the expansion element and the first compressor can be detected in order to regulate an opening of the expansion device. In this way, the expansion device can be regulated independently of a heat input upstream of the first compressor, in particular independently of a heat input of the secondary side of the further heat exchanger.
  • the expansion device can be controllable on the basis of a parameter value of the medium-pressure container in order to control a flow of the refrigerant from the line section into the line section.
  • the intermediate pressure vessel may be a passive element, with refrigerant flow through the respective vessel exits by a pressure differential downstream towards the expander and/or a further pressure difference in the return section can be controlled.
  • the parameter can be a fill level, a pressure, a temperature and/or a state of aggregation of the refrigerant in the medium-pressure container.
  • the fill level can correspond to a volume of the refrigerant in the liquid phase in the medium-pressure container.
  • a ratio of liquid phase and gas phase in the intermediate pressure vessel can be regulated by an outflow of refrigerant into the return section and an outflow downstream to the expander.
  • a ratio of liquid phase and gas phase of the refrigerant in the medium-pressure container can be kept essentially constant.
  • the expansion device can be designed to detect the parameter value on the medium-pressure container.
  • the expansion device can thus set an overheating of the refrigerant at the outlet of the evaporator.
  • the superheat can range from 3K to 10K.
  • a pressure and/or a temperature on the medium-pressure container, in the medium-pressure container or in the line system in the immediate vicinity of the medium-pressure container can preferably be detected.
  • the combination of pressure and temperature can indicate the proportion of gaseous refrigerant; it can preferably be detected when only gaseous refrigerant is present.
  • the pressure in particular the pressure in the medium-pressure container, can be regulated in order to obtain at least part of the refrigerant in liquid form in the medium-pressure container.
  • the pressure can be regulated, for example, with a downstream valve, in particular by means of the expansion device.
  • a liquid portion of the refrigerant in the medium pressure vessel can be achieved by dropping the pressure (out of the transcritical range). If the refrigerant is in a transcritical state, the refrigerant can be converted into a liquid state or into the wet vapor phase by reducing the pressure. If the combination of pressure and temperature indicates only transcritical refrigerant in the container, the pressure can be reduced (opening of the downstream valve, closing of the upstream valve) in order to reach liquid or wet vapor in the container again, which is beneficial to the performance of the evaporator. If the condition of the gas in the container is subcritical, an increase in pressure is able to increase the liquid content in the container.
  • a cooling capacity at the evaporator can be increased by appropriate regulation.
  • the scheme can be adjusted by comparing the outlet temperature of the evaporator, the measured pressure and the resulting saturation temperature with the pressure difference between the Expansion device and the compressor inlet of the first compressor can be realized.
  • a liquid separator between the evaporator and the compressor can be omitted here.
  • the refrigeration system may include a second refrigeration component configured to cool the refrigerant and disposed downstream of the first compressor and/or upstream of the second compressor.
  • the second cooling component can form an intercooler. With the second cooling component, enthalpy can be removed isobaric from the refrigerant.
  • the cooling system can have a maximum enthalpy. The enthalpy can be reduced by the intermediate cooler and/or the injection of cooled refrigerant from the recirculation section.
  • the enthalpy at the inlet of the second compressor can be at a local minimum in the gas phase.
  • an enthalpy at an inlet of the first compressor can be higher than the enthalpy at the inlet of the second compressor. Even after compression by the second compressor, both the pressure and the temperature of the refrigerant can be higher than after compression by the first compressor. However, the enthalpy after the second compression can be lower than after the first compression.
  • the second cooling component can be designed to extract heat from the refrigerant and to release it to a central cooling system and/or the ambient atmosphere. Alternatively, the heat can be dissipated using a secondary cooling circuit.
  • a compression of the second compressor can be dependent on an inlet temperature of the refrigerant at the inlet of the second compressor.
  • an optimum high pressure of the refrigerant or a high pressure generated by the second compressor can be reduced.
  • a critical operating temperature of the second compressor can be prevented from being exceeded.
  • the second cooling component may be located upstream of the joint.
  • the coolant can first be cooled by the second cooling component, and then further cooling can be realized by mixing with the coolant flow from the return section. This realizes a two-stage temperature reduction between the first compressor and the second compressor.
  • the second cooling component can be designed to cool the refrigerant when the refrigerant exceeds an ambient temperature in order to provide the refrigerant at an outlet of the second cooling component in gaseous form, within the secondary pressure range and at a reduced temperature.
  • a temperature reduction by means of the second cooling component can be controlled in such a way that a transition into the wet vapor phase is prevented.
  • the control also takes place in such a way that the combination cooling based on the second cooling component and the refrigerant supply via the recirculation section realizes a temperature of the refrigerant at the inlet of the second compressor above the wet vapor phase transition.
  • the second cooling component can be designed to cool the refrigerant to a predetermined temperature, such that the refrigerant has a temperature below a limit temperature downstream of the second compressor.
  • a limit value can be defined for an inlet temperature of the refrigerant at the second compressor, so that the second compression falls below a maximum temperature of the refrigerant.
  • the limit temperature can be determined in such a way that the refrigerant has a subcritical temperature.
  • the subcritical temperature can be set at the inlet or at the outlet of the second compressor.
  • the second cooling component can include a fan which is designed to generate an air flow on the second cooling component in order to extract heat from the refrigerant. This allows heat to be released into the surrounding atmosphere.
  • the second container outlet can be connected to the second cooling component via a line section.
  • a branch to the second compressor can be provided at the line connection, so that mixed refrigerant flows from the second cooling component and from the medium-pressure container to the second compressor.
  • a refrigerant flow from the return portion may be added.
  • a first flow of refrigerant may flow from the second refrigeration component to a further connection point and a second flow of refrigerant may flow from the second container outlet to the further connection point. Furthermore, the first refrigerant flow and the second refrigerant flow can become a combined refrigerant flow are mixed at the further connection point and the mixed flow of refrigerant can flow via the line section to the second connection point and/or to the second compressor. This can realize a three-stage temperature regulation of the refrigerant between the first compressor and the second compressor.
  • a first cooling can be achieved by means of the second cooling component, a second cooling can be achieved by means of an inflow of refrigerant from the medium-pressure container and a third cooling can be realized by means of an inflow of refrigerant via the return section.
  • This can be a maximum configuration of internal cooling, with the cooling stages being able to be designed redundantly.
  • a valve for pressure regulation can be provided both before and after the medium-pressure container.
  • redundant expansion can be realized from the tertiary pressure area via the secondary pressure area to the primary pressure area.
  • Both the expansion device and the high pressure control device can achieve superheat control.
  • a redundant high-pressure control can be provided here to prevent liquid hammer through the medium-pressure container and the high-pressure control device.
  • refrigerant can be injected from the return section into the line 204 via the second connection point.
  • a refrigerant flow through the return section can be regulated by means of the medium-pressure container or the further expansion device.
  • the heat exchanger can be configured to provide the refrigerant at the second connection point with a first specific enthalpy
  • the second cooling component can be configured to provide the refrigerant at the second connection point with a second specific enthalpy, the first specific enthalpy being smaller than that second specific enthalpy. This can ensure that with the injection via the return section, further cooling and no heating of the refrigerant can take place upstream of the second compressor.
  • a flow of refrigerant in the return section and another flow of refrigerant in the line may be mixed into a combined flow of refrigerant at the second junction.
  • the mixed flow of refrigerant may be provided to the second compressor.
  • At the second junction can be mixed of the refrigerant flows, depending on the volume fractions of the refrigerant flows, a temperature between the temperature of the refrigerant in the return section and the temperature of the refrigerant upstream of the second connection point in the main circuit can be adjusted.
  • at least a two-stage reduction in temperature can be achieved.
  • the second cooling component can be designed to cool the refrigerant based on an ambient temperature and/or based on a limit temperature of the evaporator.
  • An ambient temperature is advantageously lower than the temperature of the refrigerant in the second cooling component in order to prevent heat absorption from the ambient atmosphere at the second cooling component.
  • the higher an outlet temperature of the evaporator, the higher an ambient temperature can be at which heat can be given off to the ambient atmosphere at the second cooling component. Cooling, or enthalpy removal, by means of the second cooling component can be controlled in such a way that the formation of a liquid phase in the refrigerant flow to the second compressor is prevented.
  • a hot gas temperature for example a temperature of the refrigerant downstream of the first compressor
  • the evaporator outlet temperature can correspond to the evaporator inlet temperature.
  • the additional cooling component between the compressors can advantageously extract heat from the refrigerant.
  • a temperature at the outlet of the first compressor may be equal to or lower than the ambient temperature.
  • the temperature at the outlet of the first compressor can be below 30°C.
  • a displacement of the first compressor and a pressure in the primary pressure range can be selected such that the temperature at the outlet of the first compressor is always above ambient temperature.
  • a permissible maximum ambient temperature can be defined here: for example, permissible operation of the cooling system at an ambient temperature of up to 30° C., of up to 35° C. or of up to 40° C. can be defined.
  • the cooling system can include a circuit control, which is designed to control a flow of the refrigerant.
  • a circuit control which is designed to control a flow of the refrigerant.
  • a heat absorption in the evaporator, a heat dissipation at the cooling component and / or Ambient temperature changes to an equilibrium state of the refrigeration system, allowing the circuit controller to adjust the refrigerant flow accordingly to achieve optimum temperature and pressure levels.
  • the circuit control can be designed to regulate an opening of the expansion device on the basis of a pressure of the refrigerant, in particular a high pressure downstream of the cooling component and/or downstream of the second compressor.
  • a pressure and/or temperature sensor can be provided at an outlet of the cooling component and/or a pressure and/or temperature sensor at the outlet of the second compressor.
  • the circuit control can be designed to regulate an opening of the expansion device on the basis of a pressure of the refrigerant downstream of the heat exchanger.
  • a pressure and/or temperature sensor can be arranged at an outlet of the heat exchanger.
  • the temperature can be recorded upstream of the filter drier.
  • a controlled variable can be a pressure in the tertiary pressure range of the cooling system.
  • the controlled system can include the line system and the components of the cooling system downstream of the first compressor and/or the second compressor and upstream of the expansion device.
  • the expansion device can have a valve and be designed to adjust an opening of the valve in a range from completely closed to completely open step by step or continuously on the basis of a control actuation by the circuit control. With the opening of the expansion device, a volume flow of the refrigerant through the expansion device to the evaporator can be regulated.
  • the circuit control can be designed to regulate an opening of the expansion device on the basis of a temperature of the refrigerant upstream of the first compressor and/or downstream of the evaporator, in particular a temperature at an evaporator outlet.
  • expansion of the refrigerant process downstream of the expansion device can be controlled be.
  • a pressure reduction can be controlled, which can also bring about an isenthalpic temperature change.
  • the circuit controller can be designed to determine a predetermined temperature value and to regulate an opening of the expansion device when the predetermined temperature value is exceeded by a detected temperature, in particular an evaporator outlet temperature.
  • the predetermined temperature value can be an empirical and/or calculated media temperature.
  • the circuit control can be designed to regulate the opening of the further expansion device depending on a temperature of the refrigerant downstream of the first compressor and/or downstream of the second compressor, in particular a compressor outlet temperature or hot gas temperature, depending on the degree of opening of the further expansion device Relaxation of the refrigerant process downstream of the further expansion device is controllable.
  • the compressor outlet temperature or the hot gas temperature can be defined as the controlled variable
  • the first compressor and/or the second compressor can be defined as the controlled system.
  • a refrigerant flow through the recirculation section can thus be adjusted in order to set an optimal hot gas temperature.
  • a heat transfer of the heat exchanger from the main circuit into the return section can be controlled as a function of a degree of opening of the further expansion device.
  • the additional expansion device can advantageously set a pressure in the secondary pressure range in the return device. Depending on the degree of opening of the additional expansion device, a predetermined pressure value can be set.
  • the circuit control can be designed to control an opening of the high-pressure control device depending on a pressure of the refrigerant downstream of the second compressor and/or upstream of the high-pressure control device.
  • a pressure sensor can be provided at an outlet of the heat exchanger, at an outlet of the cooling component and/or at an outlet of the second compressor.
  • the circuit control can be designed to control the high-pressure control device as a function of a pressure downstream of the heat exchanger and/or upstream of the filter drier.
  • a detected pressure in the be defined as a controlled variable in the tertiary pressure range and a section of the cooling system in the tertiary pressure range can be defined as a controlled system.
  • this is a refrigerant section from the outlet of the second compressor to the inlet of the high-pressure control device or up to the medium-pressure vessel, or up to the expansion device.
  • the cooling system can have a compressor drive device which is designed to drive the first compressor and/or the second compressor.
  • a compressor drive device which is designed to drive the first compressor and/or the second compressor.
  • the advantage can be achieved that both the first compressor and the second compressor can be driven and controlled via a common drive device.
  • each of the compressors can be driven separately, that is to say they can be supplied with different drive powers.
  • the first compressor and the second compressor can be arranged in a common housing.
  • the compressor drive device may be a motor, which may be configured to drive the first compressor and drive the second compressor.
  • the compressor drive device can be designed to regulate a respective compressor speed of the first compressor and/or the second compressor as a function of a pressure of the refrigerant, in particular an evaporation pressure, and/or a temperature of the refrigerant, in particular an evaporation temperature, downstream of the expansion device, wherein in Depending on a compressor speed, a compression capacity of the first compressor and / or the second compressor is controllable.
  • a predetermined evaporation temperature ie a temperature during the isothermal enthalpy absorption in the evaporator
  • the evaporation pressure or the evaporation temperature can be defined as a controlled variable and the compressor output or speed can be defined as a controlled system.
  • the first compressor, the second compressor and/or the compressor drive device can each have a drive and be designed to gradually or continuously adapt a speed of the respective drive on the basis of control activation by the circuit control in a predetermined range from a minimum speed to a maximum speed. As a result, a particularly efficient control of the compressors can be implemented.
  • the circuit control can be designed to control the compressor speed on the basis of a predetermined evaporation temperature. For example, in order to lower the evaporation temperature, the compressor speed of the first compressor and/or the second compressor can be increased. In order to increase the evaporation temperature, the compressor speed of the first compressor and/or the second compressor can be reduced.
  • the circuit control can be designed to increase and/or decrease the compressor speed
  • the cooling component can include a fan which is designed to direct an air flow through the cooling component for cooling. As a result, the cooling component can be thermally coupled to the ambient atmosphere.
  • the refrigeration component may include a refrigerant-brine heat exchanger configured to transfer heat from the refrigerant to a brine.
  • the circuit control can be designed to regulate a fan speed of the fan as a function of a temperature of the refrigerant downstream of the cooling component, in particular a cooling component outlet temperature, a cooling capacity of the cooling component being controllable as a function of a fan speed.
  • Airflow to the cooling component can scale with fan speed.
  • a cooling capacity of the cooling component can also be scaled at least partially with the fan speed.
  • the fan can be designed to gradually or continuously adjust the fan speed on the basis of control activation by the circuit control in a predetermined range from a minimum speed to a maximum speed.
  • the circuit control can be designed to control the fan depending on a heat load in a plurality of stages, with the fan speed being zero in a first control stage, the fan rotating at a first speed greater than zero in a second control stage and/or in a third Control stage rotates at a second speed, the second speed being greater than the first speed.
  • the circuit control can be designed to control the pressure in the tertiary pressure range in such a way that cooling capacity is maximized.
  • an optimal high pressure can be set.
  • the optimum high pressure can depend on the ambient temperature and according to a cooling capacity of the cooling component be dependent.
  • Optimum high pressure may vary depending on subcritical or transcritical operation. For example, the optimal high pressure may be about 100 bar for a transcritical cycle and about 50 to 60 bar for a subcritical cycle.
  • the circuit control can be designed to control overheating at the evaporator by means of the expansion device in such a way that the overheating is minimal, the overheating preferably being at least 3 K at a compressor inlet of the first compressor.
  • the advantage can be achieved that the first compressor can be protected against liquid hammer caused by refrigerant in the liquid phase.
  • the process can essentially take place isothermally and isobaric up to the structural end of an evaporator tube of the evaporator. Overheating of the refrigerant can be realized at the constructive end of the evaporator. With a minimized overheating at the outlet of the evaporator, a cooling capacity can be maximized, the temperature at the inlet of the evaporator advantageously being equal to the saturation temperature.
  • the circuit control can be designed to regulate a respective compressor outlet temperature at the first compressor and/or at the second compressor by means of the further expansion device in such a way that the temperature at the compressor outlet falls below a predetermined limit.
  • chilled refrigerant from the recirculation section may be injected to lower a compressor exit temperature.
  • the circuit control can be designed to continuously control a first compressor speed of the first compressor and a second compressor speed of the second compressor. Furthermore, the circuit control can be designed to regulate the first compressor speed and the second compressor speed independently of one another. As a result, an optimal compression of the refrigerant can be achieved by the respective compressor and, in particular, an optimal high pressure can be set. Vibrations can be reduced thanks to the independent, infinitely variable control of the compressors.
  • the cooling system is used in a centrifuge, transmission of vibrations to samples can be reduced, so that the centrifugation quality can be increased. Furthermore, improved user perception can be achieved through reduced noise emission or vibration of the centrifuge.
  • the first compressor and the second compressor can be driven for a predetermined time range, for example 60 s to 120 s after a starting process with a reduced acceleration.
  • a predetermined time range for example 60 s to 120 s after a starting process with a reduced acceleration.
  • An improved centrifugation quality, an improved centrifugate and an improved user perception can thus be achieved.
  • the first compressor and the second compressor can be formed by two compression chambers in a compressor module.
  • the first compressor and/or the second compressor can each have a predetermined starting speed, so that the compressors only start when a predetermined frequency and/or voltage threshold is exceeded. This means that the compressor starts up with a low threshold or excitation speed.
  • the circuit control can be designed to accelerate the first compressor and/or the second compressor in a starting phase with a respective predetermined acceleration value to a respective predetermined setpoint speed.
  • An acceleration can preferably be less than or equal to 8 revolutions/s 2 .
  • the predetermined setpoint speed can form a limit value for distinguishing between a reduced acceleration and an increased acceleration.
  • a reduction in vibrations in the laboratory device can advantageously be achieved with a reduced acceleration.
  • the circuit control can be designed, after a first acceleration phase with a reduced acceleration, in particular with an acceleration less than or equal to 8 revolutions/s 2 , in a second acceleration phase the first compressor and/or the second compressor with an acceleration greater than 8 revolutions/s 2 to accelerate.
  • the advantage can be achieved that an oil flow, in particular an oil flow of the compressor oil, can also be accelerated.
  • the cooling system is designed to operate the first compressor and/or the second compressor at a first rotational speed and at a second rotational speed, which is greater than the first rotational speed, but the cooling system is further designed so that rotation speeds between the first and the second rotation speed are not assumed (ie only instantaneously during a speed change). In other words, operating speeds can be avoided. In this way, in particular, resonance excitations at such speeds can be avoided.
  • the circuit control can be designed to control a speed of the fan depending on an acceleration of the first compressor and/or an acceleration of the second compressor, in particular to control it continuously.
  • the fan speed may be regulated such that as the speed of the first compressor and/or the second compressor increases, the fan speed may increase.
  • a thermal load on the refrigeration component may be directly proportional to the compressor speed of the first compressor and/or the second compressor.
  • a control characteristic of the fan may lead a control characteristic of the first compressor and/or the second compressor.
  • the fan can be started with a gently rising characteristic curve.
  • the first compressor and/or the second compressor can also be started with a comparable gently rising characteristic curve.
  • the characteristic curve of the fan can run ahead of the characteristic curve of the compressors in order to generate a power reserve, i.e. before the compressors are accelerated, the fan can first be accelerated and then the compressors can catch up.
  • the advantage can be achieved that a load-adapted control of the fan reduces the primary energy to be used.
  • the noise emissions from the cooling system or the laboratory device can also be reduced.
  • the circuit control can be designed to regulate a fan speed of the fan as a function of an ambient temperature, in particular proportional to the ambient temperature, the ambient temperature being detectable on the laboratory device, preferably on an air outlet of the cooling component.
  • the fan on the cooling component With a reduced ambient temperature, for example below 22°C, the fan on the cooling component can run at reduced fan speed. If the ambient temperature is increased, for example above 24°C, the fan speed may be increased compared to operation at a standard temperature, for example 23°C.
  • a temperature window can be defined around the ideal ambient temperature. At a predetermined ambient temperature, for example 23° C., the temperature window can be 2 K, for example: 1 K up and 1 K down. The fan can run at an advantageous optimum speed within the temperature window.
  • the circuit control can be used to continuously regulate the speed up or down from this optimum speed.
  • the performance of the cooling component can be increased in order to take into account an increased ambient temperature.
  • the increased ambient temperature can a lower temperature difference between the cooling medium, for example an ambient air flow, and refrigerant and thus lead to lower performance. This power loss can be compensated for by increasing the fan speed.
  • the circuit control can be designed to control the expansion device as a function of an outlet temperature of the cooling component and/or an ambient temperature in order to set a pressure in the tertiary pressure range, in particular an optimal high pressure.
  • a cooling capacity can be increased device-specifically, in particular centrifugally-specifically, for special environmental but also application conditions.
  • the expansion device may comprise a thermostatic valve or an electronic expansion valve.
  • An electronic expansion valve can be controlled by the circuit control.
  • a thermostatic valve can be operated independently depending on a temperature value.
  • the closed-loop control can be designed to detect a change in the ambient temperature and to control the expansion device, the compressor speeds and/or a fan speed based on the change in the ambient temperature.
  • By controlling the fan and linking it to the compressors it is possible to react to changes in the ambient temperature.
  • An increased ambient temperature, and thus a higher optimal high pressure can be counteracted by closing the expansion device, increasing the speed of the compressors, or reducing the speed of the fan.
  • the high pressure, or the pressure at the outlet of the second compressor can be increased and thus optimal performance can be achieved.
  • the invention relates to a laboratory device with a cooling system.
  • the cooling system can be designed as described.
  • the laboratory device can be a centrifuge, an incubator, and/or a biological safety cabinet.
  • the cooling system can advantageously cool a component of the laboratory device, for example a cooling chamber, a sample chamber and/or a centrifuge rotor. Temperature control can also be implemented here, so that the component can be regulated to a predetermined temperature.
  • the laboratory equipment can also be a mixing tank, a stirred tank, a reactor and/or general temperature-controlled laboratory equipment, such as a refrigerator or Freezer, especially for biological or chemical samples.
  • the cooling system can be integrated into the respective laboratory device or connected to the cooling system via a line system for heat exchange.
  • the laboratory device can be a tabletop device or a stand-alone device.
  • the cooling system can advantageously be integrated into a housing of the laboratory device.
  • the cooling system can be designed to provide a cooling capacity independently of external cooling circuits in the laboratory device.
  • the laboratory device can include a rotor vessel on which the evaporator is arranged.
  • the evaporator may be thermally coupled to the rotor bowl to extract heat from the rotor bowl. In this way, cooling of the rotor chamber and in particular of biological samples arranged in the rotor chamber can be realized.
  • the evaporator can include an evaporator winding which is arranged on an outside of the rotor vessel, the evaporator winding being formed by a circumferential pipe.
  • the evaporator coil can run in a spiral around the rotor shell.
  • a contact surface of the evaporator winding on the rotor tank can be maximized in order to achieve thermal coupling between the evaporator winding and the rotor tank.
  • the evaporator winding can have a shape that is flattened on at least one side, in particular a D-shape, in order to form a flat side, the flat side bearing against the outer surface of the rotor tank in order to form a surface contact.
  • a heat flow from the rotor tank to the evaporator winding can advantageously be increased.
  • the evaporator coil can have an outer tube diameter in a range from 5 mm to 20 mm, preferably 10 mm or 16 mm. Furthermore, the evaporator winding can have a wall thickness of 0.5 mm to 5 mm, preferably a wall thickness of 1 mm.
  • a surface contact of the evaporator winding with the rotor tank can advantageously be increased.
  • the number of turns of the evaporator winding on the outer surface of the rotor tank can be increased. As a result, with the rotor tank surface remaining the same, greater coverage of the rotor tank surface can be achieved by the evaporator winding. Spaces between turns of the evaporator coil may be reduced.
  • this allows a one smaller temperature difference between samples in the rotor vessel and an evaporation temperature can be achieved.
  • the possibility of using smaller outer tube diameters, in particular using 10 mm or 12 mm tubes instead of 16 mm tubes, can be linked to the use of CO 2 as a refrigerant. With the reduction in tube cross-section, a pressure drop across the evaporator may be increased.
  • the heat transfer area can be maximized and the pressure drop minimized.
  • An increase in the contact area between the evaporator winding and the rotor tank can result in a change in the temperature difference between the inside of the tube and the inside of the tank.
  • a temperature on the inside of a boiler can approach the evaporation temperature as a result of the increase in surface area.
  • a lower control temperature, in particular a lower boiler temperature and thus better sample cooling can be achieved.
  • the reduction in pipe diameter can realize a cost advantage.
  • the rotor vessel may have a side shell surface, a bottom shell surface and a bottom surface, the side shell surface being cylindrical and the bottom shell surface having a curved profile and being formed to connect the side shell surface to the bottom surface.
  • the line of the evaporator winding is pressed into a D-shape on the side jacket surface.
  • the evaporator winding can be arranged on the side jacket surface, the bottom jacket surface and/or the bottom surface.
  • the contact area between the evaporator winding and the rotor tank can advantageously be increased in order to maximize heat transfer from the rotor tank to the evaporator.
  • the evaporator winding can form a surface contact on the lateral jacket surface.
  • the surface contact can be a continuous surface, with respective contact surfaces of the individual windings of the evaporator winding being arranged next to one another in such a way that the windings are in positive contact with one another and no free space is formed.
  • the clearance may have a substantially triangular shape.
  • a surface area of the free space can scale with the outer tube diameter of the evaporator coil and correspondingly be proportionally smaller with a reduced outer tube diameter.
  • the laboratory device can include a user interface, which is designed on the basis of a user input, a target temperature and/or a rotor speed of an in rotor arranged in the rotor tank to the circuit control.
  • the target temperature can be a default value for a control temperature, in particular a boiler temperature.
  • the control temperature can be chosen according to the design of the laboratory equipment.
  • the circuit control can determine the respective compressor speed of the first compressor and/or the second compressor, the fan speed of the fan on the cooling component and/or a Regulate the degree of opening of the expansion device.
  • the circuit control can be designed to compare a setpoint temperature input with a detected temperature of the rotor bowl and to determine a temperature difference. Furthermore, the circuit control can be designed to adjust the compressor speed, adjust the fan speed and/or adjust the degree of opening of the expansion device when the temperature difference exceeds a differential threshold value.
  • the differential threshold value can be in the range of 1 K to 10 K, preferably the differential threshold value is 5 K.
  • the compressor speed, the fan speed and/or the degree of opening of the expansion device can preferably each be reduced or increased by 20%.
  • the percentage change can relate to a respective maximum value, i.e. a compressor final speed, a fan final speed and a maximum opening angle.
  • the circuit control can also be designed to detect whether an activation triggered by the setpoint temperature input, in particular a corresponding opening, of the expansion device reduces or eliminates overheating upstream of the first compressor, the activation can be suppressed on the basis of the setpoint temperature input.
  • the circuit control can be designed to detect overheating and to interrupt the activation if the overheating is less than 1 K.
  • the circuit control can be designed to control the expansion device or the corresponding valve in such a way that the corresponding valve closes completely. As a result, the optimum high pressure can advantageously be set.
  • the circuit control can be designed in particular, the opening degree change of the expansion device with a time delay, advantageously with an offset of 30 s to change the compressor speed.
  • the circuit control can also be designed to detect a temperature difference between a cooling component outlet temperature at an outlet of the cooling component and the ambient temperature and, when a temperature difference of +3 K is reached, to prevent an adjustment of the fan speed, which is based on a changed setpoint temperature input.
  • the circuit control is designed to change the fan speed from a temperature difference of +5 K. If the 20% change in fan speed does not result in an improvement in a temperature difference range of ⁇ 5 K, the fan speed can be adjusted again. Only then does the change take place on the next control level.
  • the circuit control can be designed to adjust the compressor speed as a first control stage, adjust the fan speed as a second control stage and adjust the degree of opening of the expansion device as a third control stage. In this case, the circuit control can be designed to carry out the control stages in the order of the first control stage, the second control stage and the third control stage.
  • a further control cycle with a respective 20% level can be carried out by means of the circuit control at the end of the control cycle with a respective 10% level.
  • the 10% level can also refer to a respective maximum value, i.e. a compressor final speed, a fan final speed and a maximum opening angle.
  • a final control cycle can be performed with a control level in the range of 1% to 5%, preferably 2%.
  • the area of application of a corresponding system can be expanded.
  • Such an extension can result from the special two-stage design in conjunction with R744 (i.e. CO2).
  • R744 i.e. CO2
  • maximum operating pressures of 75 bar can be achieved, which corresponds to around 35 °C.
  • the upper application limit for a hermetic, single-stage compressor is the transcritical range.
  • Embodiments of the present technology overcome this by using the refrigerant CO2 in conjunction with two compressors connected in series so that there is no subcritical regime limitation and the system can also be operated in the transcritical regime.
  • the achievable pressure difference from the lower stage to the upper stage may be too low in relation to evaporation pressures in a range from 10 bar to 30 bar.
  • the high pressure can depend on the ambient temperature.
  • a low low pressure can be achieved even at elevated ambient temperatures, in particular at ambient temperatures above 30°C.
  • the degree of delivery can also be determined by a volume flow-related quality level.
  • the volume flow-related quality level describes the re-expansion into the design-related damage or dead space (e.g. distance between pistons and cylinder cover).
  • the actual volume flow decreases due to reverse expansion into the clearance volume.
  • the higher the high pressure the more refrigerant with higher enthalpy expands back into the dead space (when the piston starts to move down and the suction cycle starts). This refrigerant expands as the piston moves down.
  • the refrigerant sucked in from the low-pressure side like that re-expanded refrigerants mix.
  • the enthalpy is increased at the start of compression and the compressor outlet temperature rises. The latter is shown by a parallel shift of the compression line towards higher enthalpies in a log(p)-h diagram.
  • the degree of delivery can also be determined by a degree of wall quality.
  • the degree of wall quality describes the extent to which gas flowing in from the low-pressure side is heated before compression by secondary effects, such as heating on a cylinder wall, steam friction, etc. This warming is preferably kept low because it reduces the density and thus the actual volume flow.
  • the degree of delivery can also be determined by the degree of nonchalance. Permeability losses increase with increasing compression end pressure and decrease with increasing compressor speed, since less time can then be available for vapor exchange.
  • volume flow-related quality level and the wall quality level correlate negatively with the pressure ratio, the degree of permeability correlates with the compression end pressure.
  • the polytrope ratio can be defined as follows k k ⁇ 1 ⁇ ln T 2 T 1 / ln P 2 p 1 . 1 indicates a value before the compressor and 2 indicates a value after the compressor decrease per compression with a smaller pressure ratio.
  • the increase in entropy with a compression is higher than with a double compression, due to the above-mentioned mechanical and refrigeration phenomena, as well as the material behavior as a result of the lower pre-temperatures and pressures.
  • the intake conditions in low-pressure and medium-pressure can be so different that the division into two compressors is advantageous:
  • the volume at the compressor outlet is very high, i.e. a large displacement is required here. Due to the reduced pressure difference, the drive power can be lower than it would have to be with a single-stage compression. The same considerations apply to the second stage, only here the volume is smaller and the drive power is larger.
  • the optimal high pressure in the refrigeration circuit (high pressure value for optimal refrigeration capacity), which is largely dependent on the ambient temperature (or the temperature at the gas cooler outlet), can only be achieved through the two-stage compression at higher ambient temperatures.
  • a single-stage compressor may never or only temporarily reach the optimum high pressure economically and with great efficiency. This means that in order to achieve maximum cooling capacity with a centrifuge and its operating conditions (sometimes up to 40 °C), a two-stage cooling circuit is advantageous.
  • the maximum stroke between low and high pressure of current single-stage compressors is usually 40 to 60 bar in fully hermetic configuration (where the transcritical area is only temporarily used), whereby the use of fully hermetic compressors is advantageous because a semi-hermetic compressor can be used in the interior where centrifuges are predominantly operated, be critical (damaging physiological effects of CO 2 on the human organism).
  • embodiments of the technology are particularly directed to the cooling system being used in a centrifuge. Such use can also be advantageous in terms of thermal load.
  • a centrifuge differs from other applications in terms of heat load and therefore benefits from two-stage compression.
  • the heat load in a centrifuge is dynamic (unlike other compression-expansion refrigeration cycle applications), while the application in a refrigerator or Freezer can be considered more as static.
  • the operator can only leave the door open and close the refrigerator with a delay.
  • the evaporation temperature is always within a relatively narrow range. If the door is left open, the required cooling capacity may fluctuate briefly.
  • the heat loads can be permanently variable due to the different rotors used (type, speed, target temperature setting by the operator, imbalance, loading).
  • the use of variable post-injection to regulate the hot gas temperature at the compressor outlet is therefore advantageously necessary in order to take into account variable loads on the evaporator.
  • a soft start or a separate start can already be provided.
  • the centrifugation quality can be increased in this way.
  • an R744 compressor offers advantages in terms of start-up performance. Thanks to the stepless control of both compressors independently of each other, vibrations that can affect the samples and the centrifugation quality or user perception can be reduced.
  • a soft start i.e. operating the two-stage compressor for 60-120 seconds after the start-up process with a low acceleration, has a positive effect on the centrifugation quality, the centrifugate and user perception (especially with two compression chambers in one compressor).
  • the compressor there can be a fixed starting speed for the compressor, which only allows the compressor to start when a certain frequency or voltage threshold is exceeded. This means that the compressor starts up smoothly (low threshold or excitation speed) and then starts up at low speed.
  • the acceleration of the speed should ideally be small (e.g. 8/s per 10 seconds or less, i.e. an acceleration of 8 revs/s per second is the threshold acceleration rate for distinguishing between smooth and fast acceleration of the compressor).
  • a lower acceleration rate results in less vibration in the device.
  • embodiments of the technology allow for increasing operator comfort via stepless fan control based on compressor acceleration rate.
  • the fan speed can be controlled (higher compressor speed means more load on the gas cooler or condenser).
  • the advantage of this type of control is the load-adapted use of the fan, which results in lower primary energy use.
  • the closed-loop control may further achieve the benefit of reducing high-to-intermediate pressure ratio overshoots. With a power reduction of the fan, noise emissions from the fan can advantageously be reduced.
  • Embodiments of the technology described also allow the refrigeration cycle operating conditions to be adjusted as a function of the ambient temperature (depending on variable thermal load of the rotor and the site conditions).
  • a fan speed can be related to the measured ambient temperature (around the device) at the gas cooler outlet.
  • the cooling device can in particular comprise a fan and the fan speed designates the speed of this fan. If the ambient temperature is lower (e.g. below 22°C), then the fan on the gas cooler can run more slowly. When the ambient temperature is higher (e.g. higher than 24°C), the fan may spin faster than at 23°C.
  • the window around the ideal ambient temperature of 23°C can be 2 Kelvin, for example, 1 Kelvin up and 1 Kelvin down, in which the fan runs at an ideal speed. Depending on the fluctuation in the ambient temperature, this speed can be deviated from steplessly upwards or downwards. In this way, the performance of the gas cooler is gradually increased to take account of the increased ambient temperature.
  • the increased ambient temperature leads to a lower temperature difference between the cooling medium and refrigerant and thus to lower performance.
  • the increased fan speed with a deviation of 23°C can compensate for this.
  • the optimal high pressure for the refrigeration capacity at the evaporator can be set by an electronic or thermostatic expansion valve located in the return section. This can increase load- and thus centrifuge-specific performance for special environmental and application conditions. This means that the ambient temperature of a centrifuge also depends on the waste heat (rotor-specific) and the installation location. It is thus possible to react to any changes in the ambient temperature by controlling the fans and linking them to the compressors or to the compressor. An increased ambient temperature (and thus a higher optimal high pressure) can be countered by the The valve is closed, the compressor speed is increased or the fan speed is reduced. With these measures, the high pressure can be increased and thus the optimal performance can be achieved.
  • the system can be operated with a control system that is adapted to the two-stage CO2 refrigeration system.
  • the two-stage CO2 refrigeration system refers to the refrigeration system in which CO2 is used as the refrigerant and in which the two compressors are arranged in series.
  • the control system can be used in a centrifuge that has the appropriate cooling system. A relatively precise temperature control can be achieved in this way.
  • the valve With an increase in temperature (ie an increase in the setpoint temperature), the valve can be opened and with a decrease in temperature (ie a decrease in the setpoint temperature), the valve can be closed. Overheating can be reduced by opening the valve. Accordingly, a mass flow and thus a cooling capacity can be increased.
  • this control process for the valve is not carried out (e.g. if the superheat is already ⁇ 1 Kelvin).
  • the degree of opening of the valve can be changed with an offset of, for example, 30 seconds to change the compressor speed.
  • the speed is not changed.
  • the fan speed is only changed from a deviation of 5 Kelvin in the gas cooler outlet temperature. If the 20% change does not result in an improvement in the deviation range of ⁇ 5 Kelvin, the fan speed is changed again. Only then does the change take place at the next change level.
  • control temperature deviation is > 3 Kelvin (after the upper control cycle has expired)
  • the process from above is repeated, with adjustments in 10% increments (from the maximum opening degrees or component speeds).
  • the target temperature can be gradually approached.
  • the last iteration loop can be lowered to an extent of 5%. If correspondingly fast controllers and components are used, this threshold can be set to 2% if necessary. In this regard, a trade-off can be made between switching frequency and control quality, since frequent switching can cause greater wear on components.
  • the two-stage compression according to the invention can possibly lead to relatively high hot gas end temperatures, or a high temperature at the outlet of the second compression stage. Therefore, in embodiments of the invention, means are provided which reduce this temperature - for example the return section or the second cooling component.
  • Embodiments of the invention enable the efficient use of CO 2 as a refrigerant in a refrigeration system.
  • the cooling system is designed for heat dissipation in a laboratory device 300, specifically in a centrifuge.
  • an operating mode of the cooling system can be optimized by means of heat flows and/or refrigerant flows in addition to the main circuit, and the efficiency of the cooling system can correspondingly be increased.
  • FIG variants A and B of the Figures 1 to 6 each form a unit consisting of a schematic circuit diagram and a corresponding log-p enthalpy diagram.
  • process points which describe a state of the refrigerant are identified in the circuit diagram and in the log-p enthalpy diagram as a number with a rectangular marking.
  • On the x-axis shows the specific enthalpy of the refrigerant used (CO 2 ) and the pressure on the y-axis, with the y-axis being logarithmic.
  • the graph has a bell-shaped line whose left portion is referenced 802 and right portion is referenced 804 .
  • the refrigerant is present as wet vapor - i.e. in a mixture of the liquid and the gaseous Condition.
  • the line area 802 is also referred to as the boiling line.
  • the line area 804 is also referred to as the dew line.
  • Boiling line 802 and dew line 804 meet at the critical point. Above this point (i.e. at a pressure that exceeds the critical pressure) there can be no difference be made between the gas and liquid phase, so that this area is also referred to as the transcritical area.
  • FIG. 1B 12 is a diagram showing a refrigeration process according to an embodiment of the invention represented by straight lines. This refrigeration process can in the embodiment of the Fig. 1a be used.
  • the cooling system can have a plurality of sensors 60 which are each designed to detect a pressure and/or a temperature. Accordingly, the temperature and pressure within the system 10 can be determined at different locations.
  • a two stage compression refrigeration system is shown in Figure 1A and 1B .
  • Cooling system 10 is also referred to simply as system 10 .
  • the system 10 has the following: an evaporator 11, a first compressor 12, a second compressor 14, a first cooling component 16 and a first expansion device 18, which can be designed, for example, as an expansion valve 18 and is also referred to simply as an expansion valve 18 below .
  • the first compressor 12 and the second compressor can be arranged in one housing, in particular in a common housing.
  • a common drive can be provided for the first compressor 12 and the second compressor 14, which is designed to drive both the first compressor 12 and the second compressor 14.
  • the cooling system 10 has a line system 20, which includes a plurality of lines 20', 21', where the character 'stands for a number, which connect further components of the system 10 to one another.
  • the system 10 forms a first circuit in which the evaporator 11, the first compressor 12, the second compressor 14, the refrigeration component 16 and the expansion valve 18 are connected to one another in this order, with the evaporator 11 following this order again being connected to the expansion valve 18 connected - this circuit is also called the main circuit.
  • the first compressor 12 and the second compressor 14 are arranged between the evaporator 11 and the cooling component 16 and are provided in series with each other.
  • the system 10 has a filter drier 34 which is arranged between the first cooling component 16 and the expansion valve 18 .
  • the filter drier 34 can be arranged upstream of the expansion device 18 in the main circuit.
  • upstream and downstream are used in various places in this document. For example, looking at Fig. 1a it should be understood that the coolant is counterclockwise, so for example the expander 18 is located downstream of the filter drier 34 .
  • the process used is a cyclic process. After the refrigerant has left the expansion device 18, after a certain time it will have run through the evaporator 11, the two compressors 12 and 14 and the cooling component 16 and then come back to the filter drier 34.
  • the process connection between the filter drier 34 and the expansion device 18 is shorter than that between the expansion device 18 (via the other elements) and the filter drier 34 shorter process-technical connection between two elements considered.
  • the expansion device 18 is located downstream of the filter drier 34 and the evaporator 11 is located downstream of the expansion device 18 .
  • the second compressor 14 located downstream of the evaporator 11 .
  • the first compressor 12 is arranged in this sense, for example, between the evaporator 11 and the second compressor 14 .
  • a first compression can be achieved by means of the first compressor 12 and a second compression can be achieved by means of the second compressor 14 in order to bring the refrigerant from a low pressure at the outlet of the evaporator 11 to a high pressure at the outlet of the second compressor 14 .
  • the cooling component 16 implements an enthalpy reduction in order to dissipate heat absorbed at the compressor 11 from the system 10 . With this cooling, the refrigerant can change from a gas phase to a liquid phase.
  • the expansion device 18 can now by a Pressure reduction transfer the refrigerant from the liquid phase to the wet vapor phase and make it available to the evaporator for heat absorption.
  • Process point 1 is downstream of the evaporator 11 and upstream of the first compressor 12. At this process point 1, the refrigerant is in gaseous form, at a relatively lower temperature and at a relatively low pressure.
  • the refrigerant is gaseous, at an intermediate pressure and at an intermediate temperature.
  • the refrigerant is at a high pressure and at a high temperature.
  • the refrigerant is transcritical at this process point 3 . It is pointed out, however, that this is not necessary and that the refrigerant in process point 3 can also be present in gaseous form.
  • the process point 3 (or generally the process point between the second compressor 14 and the cooling component 16) is shown in the diagram Figure 1B with respect to the pressure so arranged below the critical pressure, and this also applies to the other embodiments.
  • the refrigerant is cooled by means of the cooling component 16, so that it is present at a high pressure and a low temperature at the process point 4, which lies between the cooling component 16 and the expansion device 18.
  • the refrigerant at process point 4 is transcritical.
  • the refrigerant it is also possible here for the refrigerant to be present in a different phase at this process point 4 (or generally at the process point between the cooling component 16 and the expansion device 18), in particular as wet steam.
  • the phase diagram is as follows Figure 1B such that the pressure at the corresponding process point (here: process point 4) is lower than the critical pressure, and this possibility also exists in the embodiments described below.
  • the refrigerant can then be expanded by means of the expansion device 18 so that it is present between the expansion device 18 and the evaporator 11 at the process point 5 as wet vapor, at low pressure and at a low temperature.
  • the refrigerant can then be evaporated by means of the evaporator 11 so that it is present between the evaporator 11 and the first compressor 12 at process point 1 at a low pressure and at a low temperature.
  • the refrigerant is in gaseous form at process point 1.
  • FIG. 2A and 2 B shown embodiment is compared to the cooling system according to Figures 1A and 1B supplemented by the following (ie it additionally has): a liquid separator 30 which is arranged between the evaporator 11 and the first compressor 12 .
  • the system 10 also includes a return section 40 fluidly connected to the circuit at two connection points 42 and 44, also referred to simply as joints.
  • the first connection point 42 is provided between the first refrigeration component 16 and the expansion valve 18 and the second connection point 44 is provided between the first compressor 12 and the second compressor 14 .
  • the recirculation section 40 also has an expansion device 46 (which is also referred to as an expansion valve 46) and runs through a heat exchanger 48, which is also run through by the circuit described and in particular by a line 208 that connects the cooling component 16 to the expansion valve 18.
  • an expansion device 46 which is also referred to as an expansion valve 46
  • a heat exchanger 48 which is also run through by the circuit described and in particular by a line 208 that connects the cooling component 16 to the expansion valve 18.
  • the system 10 comprises a further heat exchanger 50 through which the line 208 flows on the one hand and through which the line 202 flows on the other hand, which connects the evaporator 11 to the first compressor 12 .
  • the further expansion device 46 can be regulated in particular as a function of a high pressure and/or a hot gas temperature at an outlet of the second compressor 14 .
  • a refrigerant flow can thus be reduced or increased through the return section 40 in order to correspondingly decrease or increase the hot gas temperature.
  • the additional expansion device 46 can implement post-expansion of the refrigerant in the high-pressure area.
  • the mixing at the second connection point 44 and/or the flow through the heat exchanger 48 can increase the enthalpy in order to avoid or at least avoid wet suction, ie suction of refrigerant with a partially liquid phase and/or refrigerant in the wet vapor phase before the second compressor 14 reduce the likelihood of wet pick-up.
  • the expansion device 18 can be regulated depending on a pressure and/or a temperature at an outlet of the heat exchanger 48 .
  • this pressure at the outlet of the heat exchanger can be essentially equal to a pressure at the inlet of the expansion device 18 .
  • phase diagram of the Figure 2B can first of all refer to the description of the phase diagram Figure 1B are referred to, whereby it should be understandable for the person skilled in the art that the process points (taking into account the additions below) essentially correspond to the following: process point Figure 1B process point Figure 2B 1 1 2 2 3 4 4 9 5 10
  • the embodiment of Figures 2A , 2 B has a recirculation section 40 in addition to the main circuit.
  • this embodiment also has, among other things, a heat exchanger 48 which is arranged downstream of the cooling component 16 .
  • the refrigerant downstream of the cooling component 16 is further cooled by means of the heat exchanger 48 , so that downstream of it (at process point 6 ) it is at an even lower temperature than at process point 5 .
  • a further heat exchanger 50 is arranged in the main circuit further downstream from the process, which further cools the refrigerant so that it is transported downstream thereof, at process point 9 ( Figure 2A / 2 B ), at an even lower temperature.
  • the refrigerant can then be expanded again—as described above—by means of the expansion valve 18 and is thus present as wet vapor at process point 10 .
  • the refrigerant can in turn be evaporated by means of the evaporator 11 .
  • the evaporation will not be complete, so that the refrigerant will flow directly downstream of the evaporator (at process point 11, Figure 2A / 2 B ) is near the dew line and liquid components are present in the refrigerant.
  • Such liquid can be separated by means of the liquid separator 30 arranged downstream of the evaporator 11, and additional energy can be introduced into the refrigerant between the evaporator 11 and the first compressor 12 by means of the heat exchanger 50 already described, so that the refrigerant is present in gaseous form at process point 1.
  • a return section 40 is arranged between the connection points 42 and 44 .
  • an expansion device 46 which can be configured as an expansion valve, for example. Downstream of this at process point 7, the medium-pressure refrigerant is present as wet vapor at a relatively low temperature.
  • the refrigerant can be heated by the heat exchanger 48 . In the embodiment shown here, the refrigerant is heated up to the dew point, but it is also possible that the refrigerant is present downstream of the heat exchanger 48 (ie at process point 8) as wet vapor or in gaseous form. Overall, the system is operated in such a way that the specific enthalpy that results from the mixing of the refrigerants at connection 44 (i.e. at process point 3) is suitable for the further cycle process.
  • the use of the recirculation section 40 therefore makes it possible overall to provide a post-injection of refrigerant and to regulate the hot gas temperature at the outlet of the second compressor 14 via this. This allows variable heat loads to be taken into account.
  • the COP (Coefficient of Performance) can be increased by using post-injection and one or more internal heat exchangers.
  • the COP can be defined via the ratio of a cooling capacity to an electrical power, in particular an electrical power consumed.
  • process point 4 in the log(p)-h diagram ( Figure 3A / 3B ) can be shifted further to the left, ie in the direction of lower enthalpy.
  • the refrigerant through the heat exchanger 48 at process point 7 has a lower enthalpy than at process point 6, so that Refrigerant according to the isenthalpic (i.e. in Figure 3B : vertical) expansion between process points 11 and 12 at process point 12 there is a lower enthalpy than in the case that no heat exchanger is provided.
  • the heat exchanger 48 can be configured as an economizer heat exchanger, for example. It is arranged in front of the valve 18 and can thus further supercool the refrigerant. This can lead to a greater specific evaporation capacity q and thus to a smaller required mass flow and the resulting lower compressor speeds. It is therefore possible to use the post-injection to adapt to the required evaporation capacity and thus increase it.
  • the system 10 further includes a second refrigeration component 32 disposed between the first compressor 12 and the second compressor 14 .
  • FIGS. 10a and 10b a refrigeration process is shown, which is carried out by the refrigeration system 10 according to the Figure 3A embodiment shown can be used.
  • the reference numbers 1 to 13 shown in rectangular boxes in FIGS. 10a and 10b correspond here.
  • process point condition of the refrigerant 1 gaseous, low pressure, low temperature 2 gaseous, medium pressure, medium temperature 3 gaseous, medium pressure, low temperature (less than 2) 4 gaseous, medium pressure, low temperature (less than 3) 5 transcritical, high pressure, high temperature 6 transcritical, high pressure, low temperature (less than 5) 7 transcritical, high pressure, low temperature (less than 6) 8th Wet steam, medium pressure, low temperature 9 gaseous, medium pressure, low temperature 10 transcritical, high pressure, low temperature (as in 7) 11 transcritical, high pressure, low temperature (less than 10) 12 Wet steam, low pressure, low temperature 13 gaseous, low pressure, low temperature
  • the refrigerant at process point 5 can also be gaseous.
  • the refrigerant at process points 6, 7, 10 and 11 can also be liquid.
  • the refrigerant at process point 1 i.e. upstream of the first compressor 12
  • the refrigerant is in gaseous form, at a low pressure and at a low temperature.
  • the refrigerant is compressed in the first compressor 12 so that it is present in gaseous form and at a medium pressure downstream of the first compressor 12 (ie at process point 2). Compression also heats the refrigerant so that it is at an intermediate temperature.
  • the refrigerant is cooled so that downstream thereof (at process point 3) it is gaseous, at medium pressure and at a low temperature (i.e. at a lower temperature than at process point 2), for example around 28°C.
  • the cooling component 32 can be provided in particular at higher evaporation temperatures.
  • This refrigerant is mixed with refrigerant from the recirculation section 40, the refrigerant from the recirculation section 40 being even colder, so that the refrigerant at process point 4, which is between the connection 44 and the second compressor 14, is gaseous, at medium pressure and at a low temperature is present, whereby this temperature is even lower than at process point 3.
  • the refrigerant is further compressed.
  • the compression takes place here in this way (see Figure 3b ) that the refrigerant can be compressed to a pressure beyond the critical pressure.
  • the refrigerant Downstream of the second compressor 14 at the process point 5, the refrigerant is therefore transcritical and at a high pressure (in embodiments of the invention it is also possible for the refrigerant at the process point 5 to be in gaseous form).
  • the refrigerant was also heated by the compression, so that it is also at a high temperature.
  • the refrigerant is cooled by means of the cooling component 16 . Therefore, downstream of the refrigeration component, at process point 6, the refrigerant is transcritical (liquid in other possible embodiments), at high pressure and at low temperature (which temperature is lower than at process point 5).
  • the coolant can be further cooled by means of a heat exchanger 48 . Therefore, downstream of the heat exchanger 48, at process point 7, the refrigerant is transcritical (also liquid in some embodiments), at a high pressure and at a low temperature (which temperature is even lower than at process point 6).
  • the refrigerant is routed downstream of process point 6 through the filter dryer 34, which has no or negligible influence on the state variables of the refrigerant, so that the state downstream of this dryer 34, at process point 10, is at least essentially the same is identical to the condition at process point 7.
  • a minimal pressure loss can occur at the filter drier 34, but this is negligible in relation to pressure changes through the expansion device 11 and the compressors 12 and 14.
  • the filter drier 34 can also be arranged downstream of a further heat exchanger 50 in the main circuit.
  • the refrigerant is further cooled downstream of the filter drier 34 by a heat exchanger 50 so that it is present downstream of the heat exchanger 50 (at process point 11) in transcritical (also gaseous in some embodiments), at high pressure and at a low temperature (which is even lower than at process points 7 and 10).
  • the refrigerant is passed through the expansion valve 18, thereby being expanded and transformed into the wet vapor. Downstream of the process from the expansion valve 18 (that is to say at the process point 12), the refrigerant is therefore present as wet vapor, at a low pressure and at a low temperature.
  • the refrigerant is then evaporated by the evaporator 11, ie ideally it is completely converted into the gaseous state.
  • the refrigerant is therefore in gaseous form (or according to FIG. 10b at the transition between wet vapor and gaseous form), at low pressure and at low temperature.
  • the evaporator can at least partially only bring the refrigerant closer to the dew line 804, respectively bring it exactly up to the dew point.
  • the liquid components of the refrigerant can be separated by a liquid separator 30 and/or the further heat exchanger 50 and/or converted into gaseous refrigerant.
  • the liquid separator 30 can be provided downstream of the evaporator 11 and at least partially remove liquid components in order to ensure that these do not get into downstream components (and in particular not into the compressors 12, 14). For example, liquid components can be reduced in such a way that the formation of droplets at the compressor inlet is prevented or at least reduced.
  • the refrigerant Down-process from the evaporator 11 and upstream from the first compressor 12, the refrigerant can also pass through the already described heat exchanger 50 and be heated thereby, so that downstream thereof, at process point 1, the refrigerant is gaseous, at low pressure and at low temperature (but warmer than at process point 13) is present, which closes the process cycle. Particularly in the event of incomplete evaporation in the evaporator, it can thereby be additionally ensured that no or as little refrigerant as possible in the liquid state reaches the compressor 12 arranged downstream of the process.
  • a further return section 40 can also be provided in embodiments of the invention. As can be seen in FIG. 10a, this recirculation section 40 can be connected via connections 42 and 44 to the main circuit. Connection 42 is herein provided downstream of heat exchanger 48 and upstream of expansion valve 18, and in the embodiment shown upstream of filter drier 34. Connection 44 is provided between first compressor 12 and second compressor 14.
  • Refrigerant is thus supplied to the return section in the state at process point 7, this refrigerant being transcritical at a high pressure and at a low temperature.
  • the refrigerant passes through an expansion valve 46, by means of which the pressure of the refrigerant is reduced towards the medium pressure.
  • the refrigerant is present as wet vapor, at an intermediate pressure and at a low temperature.
  • the refrigerant passes through the heat exchanger 48, which heats the refrigerant.
  • the heating is such that the refrigerant evaporates, i.e. it is converted into the gaseous state, or at least it is brought to the dew point (i.e. to the transition between wet vapor and gas). Warming up to just below the dew line is also possible.
  • the mixing enthalpy from process point 3 and process point 9 can ideally be in the gaseous range.
  • the heat exchangers 32, 48 can be designed accordingly and a degree of opening of the further expansion valve 46 can be adjusted accordingly to regulate the ratio of the mass flow in the return section 40 and the mass flow in the main circuit. Process downstream of the heat exchanger 48 (i.e.
  • the refrigerant is again in gaseous form (or at the transition from wet vapor to gas, but possibly also as wet vapor), at medium pressure and at low temperature, since the energy input via the heat exchanger is required for the phase transition becomes.
  • the refrigerant at process point 9 has a lower specific enthalpy than at process point 3, so that the mixture of these refrigerants (see process point 4) leads to a temperature drop in relation to process point 3.
  • the first connection point can be supplemented or replaced by a medium-pressure tank 70 .
  • a high-pressure control device 74 can be arranged in the main circuit upstream of an inlet 71 of the medium-pressure container 70 , which device can be designed in particular as an expansion valve.
  • the high pressure can be lowered to a medium pressure by means of the high pressure control device 74 .
  • a transition from the liquid phase or transcritical phase to the wet vapor phase can be implemented here.
  • Liquid refrigerant can be fed to the expansion device 18 via a first outlet 72 of the medium-pressure container 70 and/or at least partially gaseous refrigerant or refrigerant in the wet vapor phase can be introduced into the recirculation section 40 via a second outlet 73 of the medium-pressure container 70 .
  • the refrigerant can be drawn off from the second compressor 14 via the return section 40 .
  • the refrigerant can be mixed with low-pressure refrigerant in the gas phase.
  • the second connection point 44 can couple the recirculation section 40 upstream of the first compressor 12 into the main circuit.
  • a second heat exchanger 50 and/or a further cooling component 32 can be omitted with the use of a medium-pressure container 70 .
  • the expansion device 18 can be controlled on the basis of an evaporator outlet temperature, which is detected, for example, by a sensor in the line section 202 . As a result, the expansion device 18 can regulate overheating of the refrigerant.
  • the embodiment according to Figures 5A and 5B is based on the embodiment according to FIG Figure 4A and is additionally supplemented by the second cooling component 32 .
  • a first cooling of the refrigerant can be achieved between the compressors 12, 14 by the second cooling component 32.
  • a further cooling of the refrigerant can then be achieved by mixing the refrigerant at the connection point 44 with cooled refrigerant from the return section.
  • the second cooling component 32 can be provided with a fan, so that the cooling capacity of the second cooling component 32 can be regulated via a fan speed.
  • the medium-pressure container 70 can be connected via a line segment 216, which connects the second container outlet 73 to a further connection point 52, to the main circuit downstream of the evaporator 11, advantageously downstream of the liquid separator 30, downstream of the first compressor 12 and/or downstream of the second cooling component 32 .
  • a first mixed flow of refrigerant may flow from the further junction 52 to the second junction 44 .
  • connection point 42 is arranged downstream of the process point 6, where the refrigerant flow splits.
  • a high-pressure control device 74 e.g., an expansion valve
  • the connection point 42 is arranged downstream of the process point 6, where the refrigerant flow splits.
  • a high-pressure control device 74 e.g., an expansion valve
  • the devices 74, 46 expand the refrigerant equally, so that the process points located downstream of these devices in Figure 6B both denoted 8b, the process point downstream of the device 46 in Figure 6A designated as 8b', in Figure 6B but coincides with process point 8b.
  • this is only an example and that the devices 74 and 46 can also be designed differently in other embodiments.
  • the refrigerant flow can be divided into two flows (process points 7 and 8a).
  • the expanded refrigerant flow in the process point 8b' which is then heated via the heat exchanger 48 to the process point 8c, can be in a wet vapor phase (process point 8c).
  • the refrigerant is preferably completely in a gaseous state at process point 8a, or on the saturation line 804.
  • the state of the refrigerant can be determined by the temperature of the heating refrigerant between process points 5 and 6, and the mass flow ratios between process points 5 and 6, as well as 8b and 8c can be regulated.
  • Heat flow into the recirculation section 40 may be proportional to a thermal interface area of the heat exchanger 48 .
  • the process point 8c can be variable and, for example, also coincide with the process point 8a.
  • the refrigerant can be partially liquid at the further connection point.
  • a corresponding temperature equalization can be implemented via the return section 44 in order to provide a completely gaseous refrigerant at the inlet of the second compressor 14 .
  • Three pressure levels can be distinguished in the cooling system. 7 separates with the line AA' a primary low-pressure area from a secondary medium-pressure area.
  • the low-pressure area is delimited by the outlet of the expansion device 18 and the compressor inlet of the first compressor 12 .
  • a tertiary high-pressure area begins with the outlet of the second compressor 14 and extends to the respective pressure reducer in the form of the expansion device 18 and/or the further expansion device 46.
  • the recirculation section can have an extended line segment downstream of a secondary side of the heat exchanger 48.
  • the cooling system 10 can be used in a centrifuge 300 to cool a centrifuge bowl 301 ( 8 ).
  • an evaporator winding 302 of the evaporator 11 can be arranged with a relatively large surface contact on an outer wall of the centrifuge bowl in order to maximize heat transfer from the rotor interior into the evaporator winding 302 .
  • the windings of the evaporator winding 302 can be at least partially pressed onto the rotor tank 301 or at least partially pressed into a flattened shape.
  • flattened windings of the evaporator winding 302 can be arranged flush with one another on a straight section of the centrifuge bowl.
  • CO 2 as a refrigerant
  • the evaporator coil 302 can have a tube with an outer tube diameter of 16 mm and a wall thickness of 1 mm, with each coil having a contact length in the x-direction of 11.6 mm and with the evaporator coil 302 in a vertical section in the X-direction being 10 contact windings, resulting in a total length of 116 mm in the X direction ( Figure 9A ).
  • the evaporator winding 302 can have a tube with an outer tube diameter of 10 mm and a wall thickness of 1 mm, each winding having a contact length in the x-direction of 8.9 mm and the evaporator winding 302 in a vertical section in the x-direction 14 contact windings, resulting in a total length of 124.6 mm in the X direction ( Figure 9B ).
  • the temperature at the edge of the tube can be 0.58 K lower due to the increase in surface area, so that the boiler temperature can also be reduced.
  • the lower boiler temperature results in better sample cooling. Reducing the pipe diameter can also result in material savings.
  • steps are mentioned in this document, it should be noted that the order in which the steps are mentioned in this text may be random. That is, the order in which the steps are presented may be random unless otherwise specified or obvious to those skilled in the art. That is, if in the present document e.g. B. it is stated that a method comprises steps (A) and (B), this does not necessarily mean that step (A) occurs before step (B), but it is also possible that step (A) (at least in part ) is carried out simultaneously with step (B) or that step (B) occurs before step (A). Furthermore, when it is said that a step (X) precedes another step (Z), this does not mean that there is no step between steps (X) and (Z).
  • step (X) before step (Z) includes the situation that step (X) is performed directly before step (Z), but also the situation that (X) before one or more steps (Y1), .. . followed by step (Z).
  • step (X) before step (Z) includes the situation that step (X) is performed directly before step (Z), but also the situation that (X) before one or more steps (Y1), .. . followed by step (Z).

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  • Chemical Kinetics & Catalysis (AREA)
  • Power Engineering (AREA)
  • Fluid Mechanics (AREA)
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  • Other Air-Conditioning Systems (AREA)
  • Heat-Exchange Devices With Radiators And Conduit Assemblies (AREA)
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  • Devices For Use In Laboratory Experiments (AREA)
EP22196688.0A 2021-09-30 2022-09-20 Système de refroidissement et appareil de laboratoire doté d'un système de refroidissement Pending EP4160109A1 (fr)

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DE102021125446.9A DE102021125446A1 (de) 2021-09-30 2021-09-30 Kühlsystem und Laborgerät mit Kühlsystem

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Citations (8)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
EP0295377A2 (fr) * 1987-06-16 1988-12-21 Maschinenfabrik Berthold Hermle Aktiengesellschaft Procédé et dispositif de régulation de température en particulier pour machines frigorifiques centrifuges
US5509881A (en) * 1994-07-07 1996-04-23 Beckman Instruments, Inc. Centrifuge rotor identification and refrigeration control system based on windage
US6923011B2 (en) * 2003-09-02 2005-08-02 Tecumseh Products Company Multi-stage vapor compression system with intermediate pressure vessel
US7096679B2 (en) * 2003-12-23 2006-08-29 Tecumseh Products Company Transcritical vapor compression system and method of operating including refrigerant storage tank and non-variable expansion device
EP3015791A1 (fr) * 2014-10-29 2016-05-04 Eppendorf Ag Centrifugeuse dotée d'un circuit de refroidissement de compresseur et procédé de fonctionnement d'une centrifugeuse dotée d'un circuit de refroidissement de compresseur
US20170209874A1 (en) * 2014-07-24 2017-07-27 Andreas Hettich Gmbh & Co. Kg Centrifuge
EP3479903A1 (fr) * 2017-11-06 2019-05-08 Sigma Laborzentrifugen GmbH Centrifugeuse
US20210252526A1 (en) * 2018-06-15 2021-08-19 Eppendorf Ag Temperature-controlled centrifuge with crash protection

Patent Citations (8)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
EP0295377A2 (fr) * 1987-06-16 1988-12-21 Maschinenfabrik Berthold Hermle Aktiengesellschaft Procédé et dispositif de régulation de température en particulier pour machines frigorifiques centrifuges
US5509881A (en) * 1994-07-07 1996-04-23 Beckman Instruments, Inc. Centrifuge rotor identification and refrigeration control system based on windage
US6923011B2 (en) * 2003-09-02 2005-08-02 Tecumseh Products Company Multi-stage vapor compression system with intermediate pressure vessel
US7096679B2 (en) * 2003-12-23 2006-08-29 Tecumseh Products Company Transcritical vapor compression system and method of operating including refrigerant storage tank and non-variable expansion device
US20170209874A1 (en) * 2014-07-24 2017-07-27 Andreas Hettich Gmbh & Co. Kg Centrifuge
EP3015791A1 (fr) * 2014-10-29 2016-05-04 Eppendorf Ag Centrifugeuse dotée d'un circuit de refroidissement de compresseur et procédé de fonctionnement d'une centrifugeuse dotée d'un circuit de refroidissement de compresseur
EP3479903A1 (fr) * 2017-11-06 2019-05-08 Sigma Laborzentrifugen GmbH Centrifugeuse
US20210252526A1 (en) * 2018-06-15 2021-08-19 Eppendorf Ag Temperature-controlled centrifuge with crash protection

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