Technical Field

The invention relates to a system and a method for cryogenic refrigeration of a process medium. In particular, the invention relates to a counter flow heat exchanger configuration and pressure regulator arrangement to reduce exergetic losses in the system.

Technological Background

Refrigeration plants providing an isothermal load below the saturation temperature of the process medium at atmospheric pressure are commonly implemented by subcooling the supply flow by means of a counter flow heat exchanger configured as an evaporator. For example, for helium, the load may be provided below 4.4 K while the supply flow is generally provided above atmospheric pressure. In evaporator heat exchangers a part of the liquid phase from the supply flow above atmospheric pressure, e.g. between 1.05 and 1.50 bar, is supplied to a turbine, a control valve or similar expansion device and enters the heat exchanger and partly evaporates, wherein the evaporated gas is released into the heat exchanger in a warmer temperature level and the liquid is recirculated, i.e. the liquid phase exiting the evaporator heat exchanger reenters the evaporated heat exchanger at the entry thereof. Accordingly, when using helium as a process medium, a liquid phase temperature may be provided between e.g. 4.26 and 4.67 K while the temperature of the supply flow is between 4.3 and 4.7 K. The supply flow may then be further cooled in another downstream counter flow heat exchanger. Although the implementation of an evaporator heat exchanger may provide a pre-cooling of the supply flow, such implementation has several disadvantages. For example, exergetic losses occur due to turbine and heat exchanger inefficiencies. Such exergetic losses may cause over 95% of the irreversibilities occurring in a typical helium refrigeration coldbox. Furthermore, the refrigeration cycle comprises a large temperature factor, e.g. 300 K and 1.0 to 4.4 K for helium, such that a need for exergetic optimization exists to raise the efficiency of the system, e.g. Carnot efficiency, thereby reducing the power input to the process.

Furthermore, evaporator heat exchangers require recirculation of flash gas and evaporated gas on the atmospheric pressure level and furthermore require phase separators, both at the 4.5 K level in case of helium. Accordingly, a need exists to reduce the equipment count and size currently required when using evaporator heat exchangers. In addition, the different heat capacities of the process medium at various pressures below the inversion temperature cause a relatively high temperature difference at the warm end of the system, i.e. between the process medium exiting the system and the supply flow of the process medium entering the system. Such temperature differences generally cause irreversibilities in the system. Summary of the invention

It is an object of the present invention to provide an improved cryogenic refrigeration system and a corresponding cryogenic refrigeration method that reduces the above problems.

This object is achieved by the cryogenic refrigeration system comprising the features of claim 1 and the cryogenic refrigeration method comprising the features of claim 12. Preferred embodiments are provided in the dependent claims and by the specification and the Figures.

Accordingly, in a first aspect, a cryogenic refrigeration system is suggested, which comprises a conduit configured to provide a supply flow of a process medium and a counter flow heat exchanger, which is thermally coupled to a heat exchanger section of the conduit. The heat exchanger comprises an inlet at a cold end of the heat exchanger and an outlet at the warm end of the heat exchanger. The system furthermore comprises a first pressure regulator, which is in fluid communication with the conduit and is arranged downstream of the heat exchanger section, and a vessel, which is in fluid communication with the conduit and is arranged downstream of the first pressure regulator. The vessel is in fluid communication with the inlet of the heat exchanger and is configured to provide an evaporated gas flow from the process medium to the inlet of the heat exchanger. The conduit is free of any evaporation heat exchanger upstream of the heat exchanger section of the conduit.

Accordingly, by providing a cold counter flow heat exchanger comprising an evaporated gas flow with low specific enthalpy, the system does not require an evaporator to precool the supply flow. This is particularly advantageous when using helium, such that the system does not require an evaporating heat exchanger and phase separator at the 4.5 K level and furthermore no recirculation of flash gas or evaporated helium on the atmospheric pressure occurs. In addition, smaller equipment such as compressors and heat exchangers may be provided, such that the dimensions of the system may be reduced.

The cold end of the heat exchanger hence relates both to the evaporated gas having a lower temperature and latent heat before entering the heat exchanger via the inlet and the temperature of the process medium in the conduit directly downstream of the heat exchanger section. In this context the term“downstream” refers to the supply flow provided in the conduit and in relation to the initial entry of the supply flow into the system. Accordingly, the entry of the supply flow into the system occurs upstream of the heat exchanger section. The heat exchanger section may comprise only a part of the conduit, wherein the part of the conduit arranged upstream of the heat exchanger section and the parts of the conduit arranged downstream of the heat exchanger section and upstream of the first pressure regulator are arranged in parallel and adjacent to the outlet and in of the heat exchanger, respectively, to further improve the heat transfer efficiency. However, the heat exchanger section and the heat exchanger may also be configured such that the heat exchanger section essentially forms the conduit, e.g. the size and dimension of the fluid couplings between the various features are minimized. By the same token, the warm end of the heat exchanger relates to the evaporated gas that exits the heat exchanger and has absorbed heat from the process medium and may therefore be considered to comprise a warmer temperature and/or an increased latent heat with respect to the evaporated gas at the cold end of the heat exchanger. The warmed evaporated gas may then exit the system via the outlet at the warm end of the heat exchanger as an exhaust gas. For example, the exhaust gas may be directly released into the atmosphere or may be retained in the system for further purposes and applications.

Preferably, the heat exchanger is configured to provide a temperature factor of the evaporated gas at the pinch point of the heat exchanger relative to the process medium of the supply flow at the pinch point of the heat exchanger larger than 0.9 during normal operation of the cryogenic refrigeration system. Preferably, said temperature factor is larger than 0.98, such that temperature differences between the evaporated gas at the pinch point of the heat exchanger relative to the process medium of the supply flow at the pinch point of the heat exchanger are minimal and/or negligible, thereby not affecting the system.

Such temperature factor is possible since the system does not require an upstream evaporating heat exchanger, which generally provides temperatures of the process medium after passing the evaporating heat exchanger fixed at around e.g. 4.6K for helium, wherein the mass flows in a steady state process at the cold and warm ends are generally equal and constant. In contrast, the cold counter flow heat exchanger may provide the supply flow and the evaporating gas at a higher temperature level at the warm end of the heat exchanger having an increased heat capacity, such that temperature differences may be minimized.

The above-mentioned temperature factor FT of a counter flow heat exchanger can be expressed with the temperature of the cold stream TJx) (with 0 £ x £ L) and the temperature of the warm stream T _w (x) at the pinch point of the heat exchanger hence where the temperature difference of the two streams is the smallest.

Alternatively, or in addition, the heat exchanger comprises an NTU (Number of Transfer Units) configured to match a temperature of the evaporated gas with a temperature of the process medium at the warm end of the heat exchanger during normal operation of the cryogenic refrigeration system.

The implementation of a heat exchanger comprising the required NTU at least has the advantage that the system may be thermodynamically optimized while certain variables, e.g. heat exchanger parameters and boundary conditions, are not required or need not be known. Accordingly, the NTU configuration provides an alternative to the LMTD configuration to provide a thermally efficient cryogenic refrigeration system.

The term“matches” here is to be understood as essentially matching said temperatures and hence also includes minimal differences, e.g. up to 0.05 K. For example, the area of the heat exchanger, e.g. the heat transfer area or length of the heat exchanger may be sized and dimensioned to provide the corresponding temperature range, wherein at least the mass flow and the heat capacity values at various temperatures of the process medium are considered to be known.

The NTU may be provided by a heat transfer area of the heat exchanger, preferably by a length of the heat exchanger. To provide a large heat transfer area the heat exchanger is preferably of a tube shape, coiled shape, and/or plate fin shape and at least partially surrounds a circumference of the conduit. This has the advantage that a large contact surface is provided and the area may be readily increased by increasing the length of the heat exchanger. For example, the heat exchanger may fully enclose the circumference of the conduit, wherein a longitudinal axis of the heat exchanger extending between the warm and cold end of the heat exchanger may coincide with a longitudinal axis of the conduit, e.g. a flowing direction of the process medium. However, instead an asymmetrical arrangement may also be provided, for example, wherein the longitudinal axis of the conduit is spaced apart from the longitudinal axis of the heat exchanger, e.g., in a lateral arrangement to said axis. Preferably, the heat exchanger may be configured as a plate fin heat exchanger, e.g. for larger systems or plants, or as a coil finned tube heat exchanger, e.g. for smaller systems or plants. During initial start-up of the system generally a normalization of the temperatures and pressures is required to provide a steady state, i.e. normal operation. By matching or minimizing the temperature difference of the process medium and the evaporated gas or exhaust gas at the warm end of the heat exchanger, the exergetic losses are reduced during normal operation. As such, the occurrence of irreversibilities in the system and the power input to the process are likewise reduced.

Furthermore, by providing a cold counter flow heat exchanger having a required NTU configuration, the system does not require an evaporator to precool the supply flow. This is particularly advantageous when using liquid helium, such that the system does not require an evaporating heat exchanger and phase separator at the 4.5 K level and furthermore no recirculation of flash gas or evaporated helium on the atmospheric pressure occurs. In addition, smaller equipment such as compressors and heat exchangers may be provided, such that the dimensions of the system may be reduced.

The increased heat transfer rate and hence cooling efficiency of the heat exchanger furthermore provides that the temperature of the supply flow at the warm end of the heat exchanger may be much higher than the saturation point, i.e. for helium and depending on the pressure of the process medium above 4.5 K and preferably as high as possible. However, said temperature range may be limited by real gas properties, such that, for e.g. helium, the temperatures are preferably between 4.5 and 20 K, more preferably between 8 and 15 K or between 10 and 13 K. Corresponding higher temperatures may be implemented for other process media, e.g. nitrogen. Accordingly, different supply pressures above atmospheric pressure may be provided. This not only reduces the operation costs of the system, but also is exergetic advantageous, as heat leaked into the process between the main refrigeration cycle and e.g. a load occurs on an increased temperature level and hence a higher capacity of the process medium.

Since the pressure influences or even determines the temperature and physical behavior of any devices or subsystems coupled to the vessel, e.g. a load such as a cryogenic user or

superconductor, when implemented as such, the pressure in the vessel is preferably maintained at a constant level. Accordingly, the linked saturation temperature for the process medium is generally also known.

Hence, in order to increase the efficiency of the cryogenic refrigeration system, the outlet of the heat exchanger may be coupled to a recuperation system, a compressor system, a vacuum pump, and/or a liquefaction system, which is configured to provide a constant pressure in the vessel. The matching temperatures of the exhaust gas and the supply flow at the warm end of the heat exchanger, in particular in the temperature range between 4.5 and 20 K, facilitates the conversion and recycling of the evaporated gas in the system. For example, a sub-atmospheric evaporated gas may be recuperated and/or a warm, cold, or mixed compression of a sub-atmospheric gas flow may be provided. Accordingly, the outlet of the heat exchanger may be intermittently coupled to the supply flow entry of the system, such that a closed cryogenic refrigeration system is provided.

Preferably, the process medium provided upstream of the first pressure regulator is a pressurized fluid, for example helium or nitrogen. However, different process media may be used. The provision of a liquid process medium at least has the advantage that the flow parameters within the heat exchanger section may be controlled and optimized and improved heat transfer between the process medium and the heat exchanger may be provided. For example, the supply flow may be configured to provide the required flow characteristics such as turbulence and boundary layers to increase the heat transfer. The supply pressure of the process medium in the conduit is thereby preferably maintained at a constant value to mitigate pressure fluctuations due to undesirable thermoacoustic oscillations, which e.g. could be caused by safety mechanisms such as safety valves to the 300K-level. Furthermore, by providing the process medium as a pressurized fluid, the heat capacity of the process medium may be varied by relaxation of the pressurized liquid in the first pressure regulator and/or adjusting the supply flow. In addition, the first pressure regulator may be configured to reduce the pressure of the process medium to provide a two-phase process medium flow downstream of the first pressure regulator. For example, the pressure reduction results in a reduced saturation temperature of the process medium, such that at least a part of the process medium is converted from the liquid phase to the gas phase. To adjust the pressure of the process medium, the pressure regulator preferably comprises a valve, expansion valve, and/or turbine. By providing the pressure regulator, both the specific enthalpy of the process medium and the mass flow of the liquid phase downstream of the pressure regulator may be adapted to adapt to variable conditions, e.g. due to different heat capacities at each pressure level and for each physical state.

Preferably, the vessel collects the liquid phase, wherein the vessel is thermally coupled to a load or wherein a load is disposed in the collected liquid phase of the vessel to provide an isothermal load. For example, the vessel may be dimensioned, such that the liquid phase immerses a load provided on the bottom of the vessel, e.g. to maximize a heat transfer area between the liquid phase and the load. Alternatively, the load may be thermally coupled to the vessel, for example, by means of fluid coupling and/or heat conducting surface. By the same token, the vessel may be dimensioned to at least partially enclose a load, wherein the liquid phase in the vessel may be either collected or circulated around at least a portion of the load. Furthermore, although for certain applications a liquid phase may be preferred, alternatively, the vessel may be dimensioned to generate and partially retain a sub-atmospheric evaporated gas, which may be used to isothermally cool the load. Preferably, the isothermal load is provided below the saturation temperature of the process medium at atmospheric pressure. The vessel may e.g. be configured as a cryostat or cryogenic user such as a superconductor.

The evaporated gas from the process medium is preferably provided by a state of the two-phase process medium controlled by the pressure regulator, a pressure of the vessel, and the load, wherein the generated evaporated gas is a sub-atmospheric evaporated gas. Accordingly, the pressure regulator may adiabatically relax the process medium to provide a part of the process medium having a gas phase, wherein the state or specific enthalpy of the process medium downstream of the pressure regulator is dependent on a predefined expansion or pressure relaxation by the pressure regulator and a generally predefined state of the supply flow upstream of the pressure regulator, which is normally defined by the regulated constant supply pressure and a temperature slightly above the l-temperature, since lower temperatures are generally not reached in the heat exchanger due to the heat capacity peak around the l-line and the thermal conductivity increase. The pressure in the vessel is furthermore preferably remained at a constant level, such that the vessel configuration and pressure cause a further pressure drop of the process medium, such that evaporated gas at sub-atmospheric pressure is generated. The sudden expansion of the process medium in the vessel may further provide an evaporated gas and flash gas from the liquid phase resulting from the Joule-Thomson expansion. In addition, the generation of the sub- atmospheric evaporation gas is dependent on the load, which causes the liquid phase, preferably provided below the saturation temperature, to at least partially reach a temperature above the saturation temperature. The sub-atmospheric evaporated gas may subsequently enter the inlet of the heat exchanger to cool the supply flow in the heat exchanger section of the conduit. This at least has the advantage that the latent heat of the evaporated gas is at the lowest level in the system, such that improved heat absorption takes place within the heat exchanger. Furthermore, exergetic losses occurring within an evaporation heat exchanger are minimized using the sub- atmospheric evaporated gas as a coolant or refrigerant for the supply flow.

The cryogenic refrigeration system may further comprise a controller and at least one sensor in communication with the controller. Accordingly, the system may comprise at least one temperature sensor arranged upstream of the pressure regulator and downstream of the heat exchanger section, wherein the controller is configured to control the first pressure regulator based on the measured value of the at least one temperature sensor to control the state of the two-phase process medium. Alternatively, or in addition, the system may comprise at least one filling sensor arranged in the vessel and/or at least one flow sensor arranged downstream of the pressure regulator for measuring a mass flow of a liquid phase of the process medium to the load, wherein the controller is configured to control the pressure regulator to control the mass flow based on the measured value of the at least one filling sensor and/or the at least one flow sensor, and/or at least one pressure sensor arranged in communication with the vessel and a compressor system coupled to the outlet of the heat exchanger, wherein the controller is configured to control the pressure in the vessel by controlling the compressor system based on the measured value of the at least one pressure sensor. For example, as the temperature and pressure of the supply flow are generally regulated at a constant level and may hence be considered as fixed boundary conditions, a measured temperature deviation from a predefined temperature by the temperature sensor arranged downstream of the heat exchanger section and upstream of the pressure regulator may be corrected by accordingly adjusting the pressure regulator to control the state of the process medium, e.g. the specific enthalpy, downstream of the pressure regulator. As the pressure and the load in the vessel are considered to be constant, a change in the state of the two-phase process medium hence changes the volume flow of the sub-atmospheric evaporated gas entering the heat exchanger at the cold end. Accordingly, the measured temperature deviation of the process medium downstream of the heat exchanger section is corrected.

By the same token, a filling sensor may indicate an increased activity of a cryogenic load, such that an increased mass flow of the process medium to the load is required. Alternatively, or in addition, such indication may be provided by a flow sensor arranged downstream of the pressure regulator for measuring a mass flow of a liquid phase of the process medium to the load. Accordingly, the controller may adjust the pressure regulator to e.g. increase the mass flow according to the required isothermal load corresponding to the measured value of the filling sensor and/or the flow sensor. The controller may hence compensate the discrepancy between the mass flow needed to hold a predefined level in the liquid vessel, e.g. due to an increased evaporated gas phase provided by the load and a corresponding liquid phase deficit in the vessel, via the pressure regulator.

In addition, a feedback provided by a pressure sensor to the controller may indicate an undesirable pressure drop or overpressure in the vessel, which is preferred to be maintained at a constant pressure to provide continuous conditions and a predictable physical impact on the load coupled to or provided in the vessel. Accordingly, a compressor system coupled downstream of the vessel at the outlet of the heat exchanger may be adjusted to normalize the pressure of vessel and hence the process medium and evaporated gas to a tolerable predefined range.

Hence, the controller and the sensor arrangement provide for a feedback mechanism that provides a means to control the boundary conditions and parameters of the system within a predefined range.

The cryogenic refrigeration system may further comprise a control valve for controlling the mass flow of the supply flow, which is in communication with the controller and is arranged in parallel to and upstream of the first pressure regulator, wherein the controller is configured to control the mass flow of the supply flow via the control valve based on the measured value of the at least one temperature sensor, filling sensor, and/or flow sensor.

The control valve may hence be adjusted in response to system fluctuations, e.g. to adjust the liquid phase in the vessel and/or the volume of evaporated gas provided to the heat exchanger. The control valve may e.g. be configured to provide a partial bypassing of the supply flow to correct for an excess volume flow in the conduit, wherein the bypass may forward the excess volume flow to adjacent systems or may recollect said volume. By the same token, a parallel supply flow may compensate a deficit of the liquid phase in the vessel and may hence be partially fed to the supply flow via the parallel control valve. Alternatively, the supply flow may provide a volume flow which slightly exceeds the required volume flow to compensate for the occurrence of a deficit, wherein the parallel control valve continuously bypasses the excess supply flow to adjacent systems and does not bypass said excess in case of a detected deficit in the vessel.

For example, while maintaining a constant pressure of the supply flow, the controller may increase the volume and/or flow rate of the supply flow by accordingly adjusting the control valve, e.g. when the fill sensor indicates a reduced fill status of the liquid phase of the process medium in the vessel. Furthermore, the controller may adjust the flow rate of the supply flow, even when the fill status of the liquid phase indicates a normal range during normal operation, but an increased mass flow of evaporated gas is required. The controller may then control the first pressure regulator and the control valve, such that the volume of sub-atmospheric evaporated gas is increased while the level of the liquid phase of the process medium is remained constant, e.g. by adjusting the currently set value of the first pressure regulator and hence the specific state of the process medium, such that the pressure and hence the enthalpy of the cooled process medium is reduced while at the same time the volume flow or flow rate of the supply flow is increased by correspondingly adjusting the control valve. This results, given that the pressure and the load in the vessel are kept constant, in an increased gas phase in the two-phase process medium and a larger volume of sub-atmospheric evaporated gas while the volume of the liquid phase of the process medium collected in the vessel remains essentially unchanged.

The controller in the cryogenic refrigeration system may furthermore be configured to adjust the first pressure regulator to provide the process medium downstream of the heat exchanger section of the conduit at a temperature between the lambda point and the saturation temperature during normal operation of the cryogenic refrigeration system. Preferably, said temperature range is obtained upstream of the first pressure regulator, such that a gas phase of the process medium downstream of the first pressure regulator comprises a temperature within said range prior to entry into the vessel. The pressure and the load in the vessel are preferably remained constant, while the pressure provided in the vessel is lower compared with the pressure upstream of the pressure regulator. Accordingly, further relaxation of the process medium in the vessel due to a sudden volume expansion, may result in a further pressure drop, causing a further reduction of the latent heat and/or temperature of the evaporated gas due to e.g. a Joule-Thomson expansion and may hence provide an improved cooling of the supply flow by the heat exchanger. The fixed pressure of the process medium as a fixed boundary condition of the supply flow at a temperature between the lambda point and the saturation temperature downstream of the heat exchanger section and upstream of the first pressure regulator furthermore ensures that a stable physical state of the process medium is provided, such that heat transfer fluctuations are minimized.

Furthermore, the system may comprise at least one warm-end temperature sensor in

communication with the conduit and the outlet of the heat exchanger at the warm end of the heat exchanger, wherein the controller is configured to adjust the evaporative gas flow based on a temperature difference measured by the at least one warm-end temperature sensor by controlling the pressure regulator.

While the temperature of the supply flow at the warm end of the heat exchanger is generally considered a fixed boundary condition, the temperature measured by the sensor at the outlet at the warm of the heat exchanger may be dependent e.g. on the heat exchanger efficiency or provided cooling of the supply flow and hence the state of the process medium upstream of the pressure regulator as well as the cryogenic load or mass flow. Accordingly, to minimize the temperature difference detected at the warm end of the heat exchanger, the controller may increase the sub- atmospheric evaporated gas flow and/or the mass flow towards the load, as outlined in the above, e.g. by adjusting the pressure regulator and/or the control valve, preferably based on a measured temperature of the process medium by a temperature sensor provided upstream of the pressure regulator and downstream of the heat exchanger section.

In addition, the achieved temperature range of the liquid phase may not only be used to provide an isothermal load, but may also provide a liquid phase to be implemented in systems configured for studying e.g. molecular interactions and fluid characteristics, for example, to study the transition from helium-1 to helium 2 at the lambda point and the superfluidity or viscosity behavior of helium at supercritical temperatures.

The heat exchanger of the cryogenic refrigeration system may be configured as a plurality of heat exchanging modules, which are arranged in parallel and/or in series to the conduit. Preferably, a second pressure regulator in fluid communication with the conduit is arranged between each serially arranged heat exchanging module.

For example, the heat exchanger may comprise two heat exchanging modules that are arranged in series to the conduit, wherein between said heat exchanger modules, a second pressure regulator, e.g. a valve or expansion turbine, is arranged and in fluid communication with the conduit. This at least has the advantage that the supply flow after cooling by the first heat exchanger module may be throttled by an additional pressure regulator to an intermediate pressure level prior to cooling by the second heat exchanger module, thereby increasing the heat capacity and providing a gradual relaxation of the process medium. At the same time, the temperature level on the warm end of the first heat exchanger module may be increased with regard to a single heat exchanger configuration. Accordingly, the provision of a plurality of heat exchanger modules may further increase the efficiency of the process.

According to a further aspect of the invention, a method for providing a cryogenic refrigeration in a cryogenic refrigeration system is suggested, wherein the method comprises the steps of - providing a supply flow of a process medium in a conduit; cooling the supply flow in a counter flow heat exchanger; reducing the pressure of the supply flow by means of a pressure regulator; and receiving the supply flow in a vessel, wherein an evaporated gas flow from the process medium is used by the heat exchanger to cool the supply flow, wherein the cooling of the supply flow is provided free of any evaporating liquid phase.

Accordingly, the cooling of the supply flow or the process medium occurs by the gas flow with low enthalpy that has evaporated prior to entry into the heat exchanger. Hence, no liquid phase enters the heat exchanger, such that, contrary to an evaporating heat exchanger, no liquid phase is evaporated within the heat exchanger. This is particularly advantageous when using liquid helium, such that the system does not require an evaporating heat exchanger and phase separator at the 4.5 K level and furthermore no recirculation of flash gas or evaporated helium on the atmospheric pressure occurs. In addition, smaller equipment such as compressors and heat exchangers may be provided, such that the dimensions of the system may be reduced.

Furthermore, the method may comprise that a temperature factor of the evaporated gas at a warm end of the heat exchanger relative to the process medium of the supply flow at the warm end of the heat exchanger is provided by the heat exchanger, which is larger than 0.9 during normal operation of the cryogenic refrigeration system. Preferably, said temperature factor is larger than 0.98, such that temperature differences between the evaporated gas at the warm end of the heat exchanger relative to the process medium of the supply flow at the warm end of the heat exchanger are minimal and/or negligible, thereby not affecting the system.

Such temperature factor is possible since the system does not require an upstream evaporating heat exchanger, which generally provides temperatures of the process medium after passing the evaporating heat exchanger fixed at around e.g. 4.6K for helium, wherein the mass flows in a steady state process at the cold and warm ends are generally equal and constant. In contrast, the cold counter flow heat exchanger may provide the supply flow and the evaporating gas at a higher temperature level at the warm end of the heat exchanger having an increased heat capacity, such that temperature differences may be minimized.

Alternatively, or in addition, a temperature of the evaporated gas is matched to a temperature of the process medium at a warm end of the heat exchanger during normal operation of the cryogenic refrigeration system provided by an NTU configuration of the heat exchanger.

The implementation of a heat exchanger comprising the required NTU at least has the advantage that the system may be thermodynamically optimized while certain variables, e.g. heat exchanger parameters and boundary conditions, are not required or need not be known. Accordingly, the NTU configuration provides an alternative to the LMTD configuration to provide a thermally efficient cryogenic refrigeration system.

As outlined in the above, matching temperatures or a minimal temperature difference by implementation of a heat exchanger comprising the required NTU at least has the advantage that exergetic losses are reduced during normal operation. As such, the occurrence of irreversibilities in the system and the power input to the process are likewise reduced. The method furthermore preferably provides that the supply flow comprises a pressurized liquid, preferably liquid helium, wherein reducing the pressure of the supply flow by the pressure regulator provides a two-phase process medium flow downstream of the pressure regulator and wherein the evaporated gas in the vessel is provided at sub-atmospheric pressure. Providing the process medium as a pressurized liquid may facilitate the heat transfer in the heat exchanger section of the conduit and the handling of the process medium, e.g. providing the supply flow.

Preferably, the cooling of the supply flow provides the process medium between the lambda point and the saturation temperature downstream of the heat exchanger section of the conduit. As outlined in the above, such temperature range and having a fixed pressure as a boundary condition ensures that a stable physical state of the process medium is maintained and may hence reduce the occurrence of fluctuations in the system. At the same time, releasing the pressure of the process medium downstream of the heat exchanger section may then result in different physical states of the process medium, such that e.g. both a liquid phase and gas phase are obtained.

The method may furthermore provide a cryogenic refrigeration of a load. Accordingly, the vessel may collect the liquid phase of the process medium to refrigerate a thermally coupled load or a load disposed in the liquid phase of the process medium in the vessel, to provide an isothermal load.

To further optimize the efficiency of the cryogenic refrigeration method, the cooling of the supply flow may occur in series or in parallel by means of a plurality of heat exchanger modules arranged in series or in parallel. In such configuration, the pressure of the supply flow is preferably reduced between each serially arranged heat exchanger module by means of a second pressure regulator. The throttling of the process medium between the heat exchanger modules has the advantage that an intermediate pressure level is obtained and the heat capacity increased while furthermore a gradual relaxation of the process medium is provided. In addition, the serial cooling of the process medium provides that the temperature level on the warm end of the heat exchanger arrangement may be increased, such that the efficiency of the process is increased.

Brief description of the drawings

The present disclosure will be more readily appreciated by reference to the following detailed description when being considered in connection with the accompanying drawings in which:

Figure 1 is a schematic view of a heat exchanger, a vessel, and a pressure regulator in a cryogenic refrigeration system;

Figure 2 is a schematic view of the embodiment according to Figure 1 configured to provide the process medium in predefined physical states; Figure 3A is a schematic cross-sectional view of a tubular heat exchanger;

Figure 3B is a schematic top view of the tubular heat exchanger according to Figure 3A seen from the cold end of the heat exchanger;

Figure 4 is a schematic view of a cryogenic refrigeration system having a controller and a load; Figure 5 is a schematic view of the cryogenic refrigeration system according to Figure 4 with a further controller configuration;

Figure 6A is a schematic view of a cryogenic refrigeration system having a serial heat exchanger and pressure regulator arrangement; Figure 6B is a schematic view of the cryogenic refrigeration system according to Figure 6A, comprising a further parallel heat exchanger arrangement.

Detailed description of preferred embodiments

In the following, the invention will be explained in more detail with reference to the accompanying figures. In the Figures, like elements are denoted by identical reference numerals and repeated description thereof may be omitted in order to avoid redundancies.

In Figure 1 a cryogenic refrigeration system 1 is schematically shown in operation using a process medium. In order to provide refrigeration, a supply flow 10 of a process medium is provided in the conduit 2. Although the process media may comprise various compounds and may furthermore be provided in different physical states, the process medium in the exemplary embodiment according to Figure 1 comprises pressurized liquid helium. The liquid helium is hence at the pressure above atmospheric pressure, preferably between 1.5 and 10 bar, more preferably between 1.5 and 8.0 bar.

All features of the system 1 and in particular the conduit 2 are thermally isolated, such that the amount of heat entering and leaving the system 1 is considered to be zero or negligible. The cryogenic refrigeration system 1 comprises a counter flow heat exchanger 3, which is thermally coupled to a heat exchanger section 2A of the conduit 2, such that the supply flow 10 is cooled by means of the counter flow heat exchanger 3. After cooling by the heat exchanger 3 the supply flow 10 arrives at a first pressure regulator 4, which is in fluid communication with the conduit 2 and is arranged downstream of heat exchanger section 2A of the conduit 2. In this context the term “downstream” refers to the supply flow 10 provided in the conduit 2 and in relation to the initial entry of the supply flow 10 into the system 1 . Accordingly, the entry of the supply flow 10 into the system 1 occurs upstream of the heat exchanger section 2A.

The first pressure regulator 4 is provided as an expansion valve or valve arrangement. By means of the first pressure regulator 4 the pressure of the process medium in the supply flow 10 is reduced to a pressure slightly above atmospheric pressure, e.g. 1.05 to 1.2 bar. The supply flow 10 then flows into a vessel 5, which is in fluid communication with the conduit 2 and is hence arranged downstream of the first pressure regulator 4. Although the fluid communication between the first pressure regulator 4 and the vessel 5 is depicted in figure 12 to comprise a conduit, e.g. an outlet of the first pressure regulator 4 and/or a corresponding inlet of the vessel 5, the fluid communication may also be provided by coupling a downstream end of the first pressure regulator 4 directly to a corresponding opening or coupling element of the vessel 5.

The vessel 5 comprises a constant pressure, which is lower than the pressure upstream of the vessel 5, and is configured to collect a liquid phase and provide an evaporated gas from the process medium. The evaporated gas is generated depending on the state of the process medium downstream of the first pressure regulator 4, e.g. the specific enthalpy, any boundary activities or implementations of the vessel 5, e.g. a load (not shown), and the pressure in the vessel, which is remained constant. Due to a sudden volume increase in the vessel 5 compared to the volume of the process medium downstream of the first pressure regulator 4, the process medium is further relaxed downstream of the first pressure regulator 4. For example, the vessel 5 is sized and dimensioned to promptly expand the process medium. The sudden volume increase of the process medium in the vessel 5 results in a rapid pressure reduction of the process medium, such that a gas phase or flash gas is generated, which comprises a sub-atmospheric pressure, i.e. below 1.0 bar. In this Joule-Thomson expansion, the temperature of the sub-atmospheric evaporated gas may remain constant or is slightly reduced while the latent heat of the evaporated gas is reduced. In addition, as outlined in the above, an implementation of the vessel 5 may cause the liquid phase in the vessel also to provide an evaporated gas. Accordingly, the sub-atmospheric evaporated gas 12 is then provided to an inlet 34 of the heat exchanger 3 to serve as a coolant or refrigerant for the supply flow 10 of the process medium. The inlet 34 of the heat exchanger 3 may be either directly coupled to the vessel 5 or may be fluidly connected to an outlet of the vessel 5 by means of a conduit or tube section.

As the latent heat and temperature of the sub-atmospheric evaporated gas 12 are considered to be at its lowest in the system 1 at the inlet 34 of the heat exchanger 3, this region is considered the cold end 30 of the heat exchanger 3. During cooling of the supply flow 10 of the process medium in the heat exchanger section 2A by the sub-atmospheric evaporated gas 12 in the heat exchanger 3, the sub-atmospheric evaporated gas 12 absorbs heat from the supply flow 10 of the process medium, such that the outlet 36 of the heat exchanger 3 is considered to be a warm end 32 of the heat exchanger 3. Accordingly, the sub-atmospheric evaporated gas 12 flows from an inlet 34 at the cold end 30 of the heat exchanger 3 to an outlet 36 at a warm end 32 of the heat exchanger 3, thereby absorbing heat from the supply flow 10 of the process medium and transitioning from a cold sub-atmospheric evaporated gas 12 to a warm sub-atmospheric evaporated gas 12, and leaves the system 1 at the outlet 36 as an exhaust gas 14. Although the cryogenic refrigeration system 1 requires a normalization and stabilization of the temperatures in the system 1 during start up or an initial phase of operation, the temperature of the process medium at various points or locations in the system 1 is considered to be constant and predictable during normal operation. Accordingly, the process medium in the vessel 5 may be used to provide isothermal conditions, e.g. an isothermal load (not shown).

The conduit 2 is free of any evaporation heat exchanger upstream of the heat exchanger section 2A of the conduit 2. Accordingly, by providing a cold counter flow heat exchanger 3 comprising an evaporated gas flow with low specific enthalpy, the system does not require an evaporator to precool the supply flow 10. Furthermore, the cold counter flow heat exchanger 3 may provide the supply flow 10 and the evaporated gas 12 at a higher temperature level at the warm end 32 of the heat exchanger 3 having an increased heat capacity, such that temperature differences may be minimized.

In particular, the heat exchanger 3 of the system 1 is configured such that during normal operation the temperature of the exhaust gas 14 matches the temperature of the supply flow 10 of the process medium at the warm end 32 of the heat exchanger 3. The term“matches” here is to be understood to also include minimal differences, e.g. up to 0.5 K, preferably between 0.05 and 0.2 K. This matching minimal difference of said temperatures is achieved by the configuration of the heat exchanger 3, wherein the corresponding NTU or the heat transfer rate is accordingly adapted. For example, the area of the heat exchanger 3, e.g. the heat transfer area or length of the heat exchanger 3 may be sized and dimensioned to provide the corresponding temperature range, wherein at least the mass flow and the heat capacity values at various temperatures of the process medium are considered to be known. For example, the heat transfer area of the heat exchanger 3 may be sized to provide the required NTU to provide a sufficient cooling of the process medium such that the process medium downstream of the heat exchanger section 2A of the conduit 2 and upstream of the first pressure regulator 4 is provided above the lambda point at a temperature of 2.14 to 2.40 K while at the same time providing a temperature of the exhaust gas 14 matching the temperature of the process medium upstream of the heat exchanger section 2A at the warm end 32 of the heat exchanger 3 between 4.5 and 20 K or even higher, preferably around 12 K. The corresponding NTU of the heat exchanger 3 may hence be optimal for said temperature ranges of liquid helium. However, the NTU may be adapted for other temperature ranges and/or compounds and may furthermore provide an excess to accommodate system fluctuations or variating needs, e.g. of a load to be cooled by the system 1.

The cryogenic refrigeration system 1 according to Figure 2 largely corresponds to the embodiment depicted in Figure 1 . Again, the process medium is provided by a supply flow 10 in the conduit 2 and is cooled by the heat exchanger 3 as described in the above. In addition, the heat transfer area of the heat exchanger 3 is adapted to provide a heat transfer rate, which provides a cooling of the process medium resulting in a cooled process medium 1 1 , e.g. comprising a temperature just above the lambda point and below the saturation temperature of the corresponding pressure of the supply flow 10, for example between 2.14 and 2.40 K. The pressure of the cooled process medium 1 1 is then reduced by the first pressure regulator 4 or expansion valve, to obtain a two-phase process medium 13. In other words, the pressurized liquid helium in the supply flow 10 is first cooled by the heat exchanger 3 to a predetermined temperature and is subsequently depressurized to provide a process medium comprising a liquid and gas phase. The configuration of the vessel 5 is such that the liquid phase 15 of the two-phase process medium 13 is collected upon entry into the vessel 5 while at the same time the configuration, e.g. the dimensioning and the constant pressure in the vessel 5, causes the generation of sub-atmospheric evaporated gas 12, depending on the respective state of the two-phase process medium 13. The sub-atmospheric evaporated gas 12 then flows into the heat exchanger 3 via an inlet 34 at a cold end of the heat exchanger 3 to cool the supply flow 10. The sub-atmospheric evaporated gas 12 leaves the heat exchanger 3 at a warm end 32 of the heat exchanger 3 and exits the system 1 via an outlet 36 as exhaust gas 14.

Accordingly, the cryogenic refrigeration system 1 according to Figure 2 is optimized to both provide a sufficient cooling of the supply flow 10 by the sub-atmospheric evaporated gas 12 and a sufficient amount of the liquid phase 15 of the process medium at a required temperature, e.g. for further refrigeration requirements, by means of a corresponding depressurization of the supply flow 10 to provide a two-phase process medium 13, a configuration and constant pressure of the vessel 5, and a configuration of the heat exchanger 3, e.g. by a corresponding NTU or heat transfer rate.

In Figures 3A and 3B the counter flow heat exchanger 3 is schematically shown in further detail. The process medium is provided by means of the supply flow 10 in the conduit 2. The heat exchanger 3 comprises a tube shape, which surrounds the circumferential area of the conduit 2 forming a heat exchanger section 2A. Although the heat exchanger 3 is depicted comprising a cylindrical form and fully surrounding the conduit 2, other shapes and configurations are possible. However, in any case the NTU of the heat exchanger 3 is predefined to accordingly cool the supply flow 10 and minimize a temperature difference of the exhaust gas 14 and the supply flow 10 at the warm end 32 of the heat exchanger 3.

As shown in Figure 3A, the heat exchanger section 2A of the conduit 2 linearly traverses the heat exchanger 3 from the warm end 32 to the cold end 30 of the heat exchanger 3 and comprises a substantially straight configuration. However, other configurations that increase the heat transfer rate or are thermodynamically efficient are possible, for example, a meandering, sinusoidal, or coiled shape of the conduit 2. While traversing the heat exchanger 3 the supply flow 10 is cooled by the heat exchanger 3 by means of the sub-atmospheric evaporated gas 12 entering the heat exchanger 3 at the cold end 30 via an inlet 34.

The cooling of the supply flow 10 is provided by the sub-atmospheric evaporated gas 12, which is distributed through the heat exchanger 3 by means of a spirally formed heat exchanger element 38. Accordingly, the spirally formed heat exchanger element 38 traverses the heat exchanger 3 in a counter flow direction of the conduit 2, wherein the sub-atmospheric evaporated gas 12 absorbs the heat from the supply flow 10 provided in the thermally coupled heat exchanger section 2A of the conduit 2, either through direct contact or thermal coupling by means of a heat conducting material. At the warm end 32 of the heat exchanger 3 the sub-atmospheric evaporated gas 12 then exits the heat exchanger 3 via an outlet 36 as an exhaust gas 14.

The inlet 34 and the outlet 36 of the heat exchanger 3 are arranged in parallel and adjacent to the conduit 2 at the cold end 30 and the warm end 32 of the heat exchanger 3, respectively. This configuration is also shown in Figure 3B, which shows the heat exchanger 3 from a perspective in flowing direction of the sub-atmospheric evaporated gas 12 and in counter flowing direction of the cooled process medium 1 1 at the cold end 30 of the heat exchanger 3. Although the conduit 2 and the inlet 34 of the heat exchanger 3 are arranged adjacently in a vertical orientation, any orientation that is perpendicular to an extending direction of the heat exchanger 3 or the spirally formed heat exchanger element 38 or a substantial lateral arrangement may be provided. By the same token, the spirally formed heat exchanger element 38 may be arranged adjacent to the conduit 2 within the heat exchanger 3 to provide a direct heat transfer between the spirally formed heat exchanger element 38 and the conduit 2. Accordingly, the heat exchanger 3 may be alternatively dimensioned to comprise a smaller size in a radial direction.

However, other configurations of the heat exchanger 3 may be provided. For example, the heat exchanger 3 may be configured as a plate fin heat exchanger, e.g. for larger systems or plants, or as a coil finned tube heat exchanger, e.g. for smaller systems or plants. In a plate fin heat exchanger, the heat exchanger comprises a plurality of compartments that are arranged adjacently and in countercurrent orientation to each other and wherein said compartments comprise either the sub-atmospheric evaporated gas or the supply flow. When implementing the heat exchanger 3 as a coil finned tube heat exchanger on the other hand, the sub-atmospheric evaporated gas may be guided along the conduit 2 comprising the supply flow 10 in a coiled fashion, wherein the coiled arrangement furthermore comprises a plurality of loop sections that extend radially outward, thereby defining a plurality of fins.

A further embodiment of the cryogenic refrigeration system 1 is shown in Figure 4. Figure 4 essentially corresponds to the system 1 according to Figure 2, such that like features and functions are not discussed in further detail. The system 1 comprises a controller 7, which is in

communication with the first pressure regulator 4 and is configured to control the first pressure regulator 4 in order to relax or expand the cooled process medium 1 1 to provide a two-phase process medium 13 downstream of the first pressure regulator 4. To appropriately adjust the pressure of the cooled process medium 1 1 , the controller 7 is in communication with a temperature sensor 70 which is in communication with the conduit 2 and the outlet 36 of the heat exchanger 3 at the warm end 32 of the heat exchanger 3. Said sensor 70 hence provides an actual temperature of the supply flow 10 entering the system 1 and the exhaust gas 14 exiting the system 1 via the outlet 36. The measured values of the sensor 70 are provided to the controller 7, wherein the controller 7 controls the first pressure regulator 4 based at least on the measured values of the sensor 70, the state of the two-phase process medium 13, and the pressure in the vessel 5.

Although the system 1 is generally designed for specific boundary conditions and the status of the system 1 is maintained constant, the provision of the controller 7 and the temperature sensor 70 allow the system 1 to react to or prevent minor fluctuations in the system 1 , e.g. by adjusting the volume flow of the sub-atmospheric evaporated gas 12. The volume flow of sub-atmospheric evaporated gas 12 is dependent on the state of the two-phase process medium 13 and the pressure in the vessel 5, which is maintained at a constant level by a compressor (not shown) in communication with the vessel 5 at a downstream end, e.g. downstream of the outlet 36. As both the temperature and the pressure of the supply flow 10 are fixed boundary conditions and the cooling efficiency of the heat exchanger 3, and therefore the state of the cooled process medium 1 1 , is generally known, the state or specific enthalpy of the two-phase process medium may be controlled by adjusting the pressure regulator 4. For example, the controller 7 may adjust the first pressure regulator 4 to further reduce the pressure of the cooled process medium 1 1 , when an undesirable temperature difference between the exhaust gas 14 and the supply flow 10 is measured, e.g. when the measured temperature of the exhaust gas 14 is higher than the temperature of the supply flow 10, such that the two-phase process medium 13 is relaxed and/or the gas phase is increased and hence, at a constant vessel pressure, a larger volume flow of sub- atmospheric evaporated gas 12 is provided to the heat exchanger 3. Accordingly, an improved cooling of the supply flow 10 may be provided while at the same time absorbed heat in the sub- atmospheric evaporated gas 12 levels out the temperature difference between the exhaust gas 14 and the supply flow 10 at the warm end 32 of the heat exchanger 3. Provided in the liquid phase 15 of the process medium collected in the vessel 5 is a load 6. The load also affects the volume flow of sub-atmospheric evaporated gas 12 as, depending on the activity of the load 6, the liquid phase 15 may partially attain a temperature above the saturation temperature and hence enter the gas phase. In order to maintain an isothermal load 6, the controller 7 may hence accordingly adjust the first pressure regulator 4 to e.g. compensate for a loss of liquid phase 15. For example, the controller 7 may adjust the pressure and hence the specific enthalpy of the two-phase process medium 13 by controlling the first pressure regulator 4 to increase the liquid phase 15 of the two-phase process medium 13 to be collected in the vessel 5 and to compensate for an increased amount of sub-atmospheric evaporated gas 12 and a loss of the liquid phase 15 in the vessel 5. By the same token, the change in mass flow to the load 6 may be detected by a change in temperature, which is measured by the temperature sensor 70 and may be provided to the controller 7 as feedback.

In addition to the temperature sensor 70, the embodiment according to Figure 5 comprises a fill sensor 72 and a pressure sensor 74 disposed in the vessel 5 that are in communication with the controller 7. Accordingly, the controller 7 controls the first pressure regulator 4by adjusting the pressure of the cooled process medium 1 1 based on a fill status measured by the fill sensor 72 in the vessel 5. For example, an increase in the activity of the load 6 may reduce the fluid level of the liquid phase 15 of the process medium, which is detected by the fill sensor 72 and indicates to the controller that a deficit of the liquid phase 15 is present in the system 1 . The controller 7 may then control the first pressure regulator 4 to accordingly adjust the state of the two-phase process medium 13 and hence the liquid phase 15 provided to the vessel 5.

In addition, the controller 7 is in communication with a control valve 20, which is arranged in parallel to and upstream of the pressure regulator 4. The control valve 20 is configured as a three-way- valve and connects the conduit 2 to a parallel system. Should the fill sensor 72 indicate a deficit or excess of the liquid phase 15 of the process medium in the vessel 5, the controller 7 may control the control valve 20 to accordingly adjust the mass flow while retaining a constant pressure and temperature of the supply flow. Alternatively, or in addition, such indication may be provided by a flow sensor 76 in communication with the controller 7 and provided downstream of the pressure regulator 4 and indicating a mass flow to a load 6 The pressure in the vessel 5 is furthermore maintained at a constant level by a compressor (not shown) in communication with the vessel 5 at a downstream end, e.g. downstream of the outlet 36. The pressure in the vessel 5 is measured by the pressure sensor 74. Should a pressure deviation from a predefined range or threshold occur, said pressure sensor 74 provides a feedback to the controller 7, which accordingly adjusts the pressure via the downstream compressor. Furthermore, a temperature sensor 70 is provided, which is arranged downstream of the heat exchanger section 2A and upstream of the pressure regulator 4 and is in communication with the controller 7. As the temperature and pressure of the supply flow 10 are generally regulated at a constant level and may hence be considered as fixed boundary conditions, a measured temperature deviation from a predefined temperature may be corrected by accordingly adjusting the pressure regulator 4 to control the state of the process medium, e.g. the specific enthalpy, downstream of the pressure regulator 4. As the pressure and the load 6 in the vessel 5 are considered to be constant, a change in the state of the two-phase process medium 13 hence changes the volume flow of the sub-atmospheric evaporated gas 12 entering the heat exchanger 3 at the cold end 30. Accordingly, the measured temperature deviation of the process medium downstream of the heat exchanger section 2A is reduced.

Although the load 6 may be disposed in the liquid phase 15 of the process medium in the vessel 5, the load 6 may also be provided outside of the vessel 5, as depicted in Figure 5. The volume flow entering and exiting the vessel 5 is hence not affected by the dimensions of the load 6 while a thermal coupling between the vessel 5 and the load 6 provides a similar refrigeration of the load 6, e.g. to provide an isothermal load 6. The thermal coupling may be provided by either a direct contact between the outer surface of the vessel 5 and the load 6 or by means of e.g. a fluid coupling such as a check valve.

The heat exchanger 3 may comprise various configurations to provide the required temperature factor at the warm end of the heat exchanger, e.g. by a corresponding NTU or heat transfer rate.

For example, the heat exchanger 3 may comprise a plurality of counter flow heat exchanger modules 3A, 3B, 3C, which are arranged in series and/or in parallel, as shown in the embodiments according to Figures 6A and 6B. In Figure 6A the heat exchanger comprises two heat exchanging modules 3A and 3C that are arranged in series. The serial heat exchanger modules 3A, 3C are fluidly coupled to each other and thermally coupled with the conduit 2, comprising the process medium.

In operation, the sub-atmospheric evaporated gas 12 enters the second serial heat exchanger module 3C at a cold end 30 and traverses said heat exchanger module 3C, thereby absorbing heat from the process medium in the conduit 2. The sub-atmospheric evaporated gas exiting the second serial heat exchanger module 3C hence comprises a different latent heat and/or temperature compared with the sub-atmospheric evaporated gas 12 provided in the inlet 34 and is hence considered a warmed sub-atmospheric evaporated gas 17. The warmed sub-atmospheric evaporated gas 17 then enters the first serial heat exchanger module 3A and exits the system 1 as an exhaust gas 14 at the warm end 32 via an outlet 36. While the warmed sub-atmospheric evaporated gas 17 absorbs heat in the first serial heat exchanger module 3A, the process medium in the supply flow 10 is accordingly cooled, such that the process medium in the conduit 2 arriving at the second serial heat exchanger module 3C is considered to be a subcooled process medium 16. Subsequent cooling of the subcooled process medium 16 by the second serial heat exchanger module 3C then results in the cooled process medium 1 1 downstream of the second serial heat exchanger 3C.

The system 1 furthermore comprises a pressure regulating arrangement comprising a first pressure regulator 4A and a second pressure regulator 4B that are in fluid communication with the conduit 2. The first pressure regulator 4A is arranged downstream of the second serial heat exchanger module 3C and upstream of the first pressure regulator 4A to adjust a pressure of the process medium 1 1 provide a two-phase process medium 13 downstream of the first pressure regulator 4A. The second pressure regulator 4B is arranged between the first and second heat exchanger modules 3A, 3C. this arrangement provides that the pressure of the process medium or the pressurized liquid may be adjusted or reduced after subcooling of the process medium and prior to the cooling by the second heat exchanger module 3C to provide the cooled process medium 1 1 , wherein the subcooled process medium 16 may be provided as a liquid or as a two-phase process medium. Accordingly, the system 1 is configured to optimally use the different heat capacity values of the process medium for different temperatures and pressures, thereby providing an NTU of the heat exchanger to match the temperatures of the exhaust gas 14 and the supply flow 10 at the warm end 32. A combination of a parallel and serial arrangement of counter flow heat exchanger modules is shown in figure 6B. In addition to the first and second heat exchanger modules 3A, 3C, the system 1 comprises a parallel heat exchanger module 3B, such that the first serial heat exchanger module 3A and the parallel heat exchanger module 3B are arranged in parallel. In order to provide such arrangement of the cryogenic refrigeration system 1 , the vessel 5 is fluidly coupled via an inlet 34 to a cold end 30 of the second heat exchanger module 3C to provide the sub-atmospheric evaporated gas 12 exiting the vessel 5. After traversing the second heat exchanger module 3C, the warmed sub-atmospheric evaporated gas is then divided or split into a first and second parallel warmed sub- atmospheric evaporated gas 17A, 17B and introduced into the first serial exchanger module 3A and the parallel heat exchanger module 3B, respectively, using parallel fluid couplings. The warmed sub-atmospheric evaporated gas 17A, 17B subsequently exits the respective first serial exchanger module 3A and the parallel heat exchanger module 3B as a first and second exhaust gas 14A, 14B, respectively, wherein the first and second exhaust gas 14A, 14B are coupled to the outlet 36 at the warm end 32 and are combined to provide exhaust gas 14 exiting the system 1 via the outlet 36. In order to provide the parallel cooling, the conduit 2 is split into two parallel sections that are thermally coupled to the parallel exchanger modules 3A, 3B at a point just before the first serial exchanger module 3A and the parallel heat exchanger module 3B. The parallel heat exchanger modules 3A, 3B hence provide a subcooling of the process medium as described in further detail for the embodiment according to Figure 6A. The parallel sections of the conduit 2 are then merged again downstream of the parallel heat exchanger modules 3A, 3B and prior to entry into the second pressure regulator 4B. Downstream of the second pressure regulator 4B the process medium is further cooled by the second serial heat exchanger module 3C and passes the first pressure regulator 4A prior to entry into the vessel 5, as described in relation to Figure 6A. According to the embodiment of Figure 6B, the second serial heat exchanger module 3C comprises a tube shape surrounding the circumference of the conduit 2, while the parallel heat exchanger modules 3A, 3B are depicted to be thermally coupled to the parallel sections of the conduit 2 in an adjacent manner. However, configurations other than those depicted are possible, e.g. a plurality of tubular heat exchanger modules and/or heat exchanger modules only partially surrounding the circumference of the conduit 2 may be provided. Furthermore, the conduit sections and the fluid couplings are adjacently arranged to each other to both increase thermal efficiency and reduce the dimensions and size of the system 1. However, it will be understood that other configurations, wherein e.g. the conduit sections and the fluid couplings are at least partially space apart, may also be provided. In particular, further possible configurations of the heat exchanger as described in view of Figures 3A and 3B, i.e. plate fin heat exchanger modules or coil finned tube heat exchanger modules, may also be implemented.

It will be obvious for a person skilled in the art that these embodiments and items only depict examples of a plurality of possibilities. Hence, the embodiments shown here should not be understood to form a limitation of these features and configurations. Any possible combination and configuration of the described features can be chosen according to the scope of the invention.

List of reference numerals

1 Cryogenic refrigeration system

10 Supply flow of a process medium

1 1 Cooled process medium

12 Sub-atmospheric evaporated gas

13 Two-phase process medium

14 Exhaust gas

14A First parallel exhaust gas

14B Second parallel exhaust gas

15 Liquid phase of process medium

16 Subcooled process medium

17 Warmed sub-atmospheric evaporated gas

17A First parallel warmed sub-atmospheric evaporated gas

17B Second parallel warmed sub-atmospheric evaporated gas

2 Conduit

2A Heat exchanger section

20 Control valve

3 Counter flow heat exchanger

3A First serial counter flow heat exchanger module

3B Parallel counter flow heat exchanger module

3C Second serial counter flow heat exchanger module

30 Cold end of heat exchanger

32 Warm end of heat exchanger

34 Inlet

36 Outlet

38 Spirally formed heat exchanger element

4 First pressure regulator

4A First pressure regulator

4B Second pressure regulator

5 Vessel

6 Load

7 Controller

70 Temperature sensor

72 Fill sensor 74 Pressure sensor

76 Flow sensor