Field of the Invention

The present invention relates to a refrigeration system.

Background of the Invention

In some refrigeration systems, a compressor may be used to drive a refrigerant around a circuit. The compressor can, however, be a significant source of noise.

Summary of the Invention

The present invention provides a refrigeration system comprising a thermal store, and a circuit around which a refrigerant circulates, the circuit comprising a compressor and a heat exchanger, wherein the heat exchanger exchanges heat between the refrigerant and the thermal store, and the thermal store surrounds at least part of the compressor to absorb noise generated by the compressor.

The thermal store may therefore perform two functions. Firstly, the thermal store may store heat transferred from the heat exchanger and thereby obviate the requirement to expel the heat immediately into the surrounding environment. For example, when used in an air conditioning system, the refrigeration system may cool the air within a room, with the extracted heat being stored in the thermal store. The heat stored by the thermal store may then be expelled at a later time when cooling is not required. For example, the refrigeration system may provide cooling during the day or when the room is occupied, and expel the heat stored in the thermal store overnight or when the room is unoccupied.

Secondly, the thermal store is made to surround at least part of the compressor. As a result, the thermal store performs a secondary function of absorbing noise generated by the compressor. Acoustic emissions from the refrigeration system may therefore be reduced without the requirement for separate noise-absorbing materials, such as acoustic foams. The thermal store may comprise a relatively dense material. As such, the thermal store may provide more effective absorption of lower frequency noise than that provided by, for example, acoustic foams.

The thermal store may comprise a phase change material. As a result, the thermal store is capable of storing and releasing relatively large amounts of heat for a given temperature range by taking advantage of the latent heat capacity of the phase change material. Accordingly, when providing cooling, the refrigeration system may store a greater amount of heat in the thermal store and thus provide cooling over a longer period of time. Furthermore, due to the relatively high density of the phase change material, the thermal store may absorb more of the noise (i.e. more of the sound power) generated by the compressor at lower frequencies. In some examples, the phase change material may have a melting point greater than an ambient temperature of the room. This then has the advantage that heat stored by the thermal store may be expelled to the room when cooling is not required. A relatively high melting point has the advantage of increasing the rate at which heat is expelled, and thus decreasing the time required to regenerate the thermal store. A relatively low melting point, on the other hand, has the advantage of improving the efficiency of the refrigeration system when cooling. A relatively good balance between these two competing factors may be achieved with a phase change material having a melting point of between 30 °C and 80 °C. In some examples, the phase change material may comprise an organic wax or inorganic salt hydrate.

The thermal store may have a thickness of between 20mm and 150mm. As a result, the thermal store may provide effective sound absorption. The thickness may be measured radially from a longitudinal axis of the thermal store.

The thermal store may form a sleeve that surrounds a major portion of the compressor. The thermal store may then act as an acoustic blanket to absorb a greater proportion of the noise generated by the compressor. Moreover, by forming a sleeve around the compressor, a more compact arrangement may be achieved. In contrast, with a conventional refrigeration system, the heat exchanger may be located remotely from the compressor. For example, the heat exchanger may be located on an external wall of the system or in a separate housing to that of the compressor. The thermal store may form a sleeve that surrounds at least 60 % of the compressor, when measured about a longitudinal axis of the sleeve, and more particularly at least 90 % of the compressor.

The thermal store may subtend a central angle of at least 180°. As a result, noise may be absorbed around at least one half of the compressor. Significant acoustic improvements may therefore be achieved without necessarily requiring the thermal store to surround wholly the compressor. For example, the refrigeration system may be sited adjacent a wall within a room. The refrigeration system may be oriented such that the covered portion of the compressor is directed towards the centre of the room, whereas the uncovered portion may be directed towards the wall. The thermal store may subtend a central angle of at least 270°, and more particularly at least 340°. As a result, noise generated by the compressor may be absorbed over a wider range of angles.

The thermal store may comprise a gap in the sleeve. The gap may provide space for locating other components of the refrigeration system, thereby providing a compact arrangement. For example, the refrigeration system may comprise a condensation collector that collects condensation that forms on components of the refrigeration system, and the condensation collector may be located in the gap.

The refrigeration system may comprise a chamber in which the compressor is located, and the thermal store may surround the chamber. This may result in the thermal store being spaced from the compressor, which may reduce the transmission of vibrations to the thermal store from the compressor. Additionally, this may reduce the transmission of heat between the thermal store and the compressor.

The thermal store may surround a side of the chamber and the refrigeration system may comprise a first plate for covering a top of the chamber and a second plate for covering a bottom of the chamber. Thereby the chamber may be at least partially acoustically sealed, which may further reduce the amount of compressor generated noise that reaches the user. The heat exchanger may be embedded within the thermal store. This may provide an arrangement with a large surface area between the heat exchanger and the thermal store and thereby provide efficient heat transfer between the heat exchanger and the thermal store. Additionally, by embedding the heat exchanger within the thermal store, the thermal store may absorb noise emitted from the heat exchanger. This noise may originate from the movement of refrigerant through the heat exchanger or from vibrations generated by the compressor and conducted to the heat exchanger via piping of the circuit.

The circuit may comprise a metering device and a further heat exchanger for exchanging heat between the refrigerant and a medium. This may enable the refrigeration system to operate a refrigeration cycle in which heat is transferred between the thermal store and the medium. For example, heat may be transferred between the refrigerant and the thermal store at the heat exchanger. The pressure of the refrigerant may then be reduced by the metering device, and heat may be transferred between the refrigerant and the medium at the further heat exchanger. Finally, the pressure of the refrigerant may be increased by the compressor, and heat may again be transferred between the refrigerant and the thermal store at the heat exchanger.

The refrigeration system may be operable in one of a first state and a second state. In the first state, the metering device may have a first restriction such that the thermal store is heated at the heat exchanger and the medium is cooled at the further heat exchanger. In the second state, the metering device may have a second, less restrictive restriction or is bypassed such that the thermal store is cooled at the heat exchanger and the medium is heated at the further heat exchanger. As a result, the refrigeration system may be employed in the first state to cool the medium, for example to cool the air in a room. The refrigeration system may then be used in the second state to cool the thermal store, for example to dissipate heat collected in the thermal store during the first state, such that the refrigeration system may be reused in the first state.

The refrigeration system may be portable. This may increase the utility of the refrigeration system as the user may be able to move the refrigeration system between locations where the refrigeration system is required. The refrigeration system may have a weight of less than 15 kg.

The present invention further provides a heating, ventilation and air conditioning (HVAC) system comprising a refrigeration system according to any one of the preceding paragraphs.

The present invention also provides a fan assembly comprising a refrigeration system according to any one of the preceding paragraphs, and an airflow generator for generating an airflow through the fan assembly, wherein the circuit of the refrigeration system comprises a further heat exchanger for exchanging heat between the refrigerant and the airflow.

By employing a refrigeration system that comprises a thermal store, a compact and self- contained fan assembly may be achieved. In particular, the fan assembly may cool the airflow, and the heat extracted from the airflow may be stored in the thermal store. The fan assembly may therefore be located within the room being cooled. By contrast, with a conventional air conditioning unit having a refrigeration system, the heat extracted from the cooled air is typically expelled to an area outside the room being cooled. As a result, the refrigeration system is typically larger and more complex.

The compressor may be located below the airflow generator. The compressor may be heavier than the airflow generator. By arranging the compressor below the airflow generator, a low centre of gravity may be achieved, which may improve the stability of the fan assembly.

The further heat exchanger may surround at least part of the airflow generator. As a result, a relatively compact arrangement for the refrigeration system and the airflow generator may be achieved. Additionally or alternatively, a straighter and less contorted path may be provided for the airflow between the further heat exchanger and the airflow generator. For example, a relatively straight radial path may be provided between the further heat exchanger and the airflow generator. By providing a straighter and less contorted path, a higher flow rate may be achieved for the airflow due to reduced pressure losses.

The fan assembly may comprise a housing within which the refrigeration circuit and the airflow generator may be housed. The housing may comprise a plurality of apertures through which the airflow may be drawn into the fan assembly by the airflow generator, and the apertures may surround the further heat exchanger. As a result, a large amount of air may be drawn into the housing which may increase the flow rate of the airflow emitted from fan assembly and thus the cooling provided by the fan assembly. Additionally, by having the apertures surround the further heat exchanger, a straighter and less contorted path for the airflow may be achieved between the inlet and the further heat exchanger. As a result, pressure losses may be reduced and thus a higher flow rate may be achieved.

The heat exchanger and the further heat exchanger may be cylindrical in shape and may be stacked vertically. By stacking the heat exchangers vertically, the footprint of the fan assembly may be reduced. This may increase the utility of the fan assembly by making it more easily accommodated within a domestic setting, e.g. a room. Additionally, by being cylindrical in shape, the heat exchangers may be located around other components of the fan assembly (e.g. the compressor or airflow generator) to provide a relatively compact arrangement.

The fan assembly may comprise a nozzle having an outlet through which the airflow is emitted from the fan assembly. Providing a nozzle may enable improved control over the direction of the emitted airflow. For example, the nozzle may be moveable or comprise moveable parts (e.g. slats or louvres) to change the direction of the airflow. This then enables the emitted airflow to be targeted in different directions.

The present invention additionally provides a product comprising a component and a thermal store, wherein the thermal store comprises a phase change material that undergoes a phase change in use, and the thermal store surrounds at least part of the component to absorb noise generated by the component. The thermal store may therefore perform two functions. Firstly, the thermal store may store or release thermal energy during use of the product and thereby obviate the requirement to expel the thermal energy immediately into the surrounding environment. Additionally, by comprising a phase change material that undergoes a phase change in use, the thermal store may be capable of storing and releasing relatively large amounts of thermal energy for a given temperature range by taking advantage of the latent heat capacity of the phase change material. Secondly, as the thermal store surrounds at least part of the component, the thermal store absorbs noise generated by the component. Acoustic emissions from the product may therefore be reduced without the requirement for separate noise-absorbing materials, such as acoustic foams.

The component may be a turbomachine. For example, the component may be a compressor, or an airflow generator comprising a motor and an impeller. A turbomachine may be a relatively noisy component of the product, and therefore the acoustic emissions from the product may be significantly improved by surrounding the turbomachine with the thermal store.

The phase change material may undergo a liquid-solid phase change in use. As a result, the phase change material is relatively dense in both states of phase in comparison to a phase change material which undergoes a liquid-gas phase change. As such, the thermal store may provide more effective absorption of noise from the component. Additionally, the design and manufacture of the thermal store may be simplified as the requirement to store high pressure gas may be removed.

The phase change material may have a melting point of between 30 °C and 80 °C. As noted above, this range of melting points may provide a good balance between the competing factors of decreasing the time required to regenerate the thermal store and improving the efficiency of a refrigeration system of which the thermal store may form a part. The phase change material may be an organic wax or an inorganic salt hydrate. The thermal store may thereby comprise a relatively dense material. As such, the thermal store may provide more effective absorption of noise, particularly at lower frequencies, than that provided by, for example, acoustic foams.

The product may comprise a heat exchanger for transferring thermal energy to and from the thermal store. The heat exchanger may be embedded within the thermal store.

The thermal store and the heat exchanger may form part of a refrigeration system of the product, and the refrigeration system may be operable in a first state and a second state. In the first state, the heat exchanger may transfer thermal energy to the thermal store to warm the phase change material. In the second state, the heat exchanger may transfer thermal energy from the thermal store to cool the phase change material. As a result, the product may operate the refrigeration system in the first state to store thermal energy within the thermal store. The product may then operate the refrigeration system in the second state to release the thermal energy collected in the thermal store during the first state, such that the refrigeration system may be reused in the first state.

The component may be a compressor of the refrigeration system. As the compressor may be one of the loudest components of the refrigeration system, absorbing noise generated by the compressor may result in a significant reduction in the acoustic emissions of the refrigeration system.

The heat exchanger may exchange thermal energy between the thermal store and a refrigerant, and the product may comprise a further heat exchanger to exchange thermal energy between the refrigerant and a medium. As discussed above, this may enable the refrigeration system to operate a refrigeration cycle in which thermal energy is transferred between the thermal store and the medium. Thereby, the product may be used to heat or cool the medium, for example to heat or cool air in a room. The heat exchanger and the further heat exchanger may be cylindrical in shape and may be stacked vertically. As discussed previously, by stacking the heat exchangers vertically, the footprint of the product may be reduced. This may increase the utility of the product by making it more easily accommodated within a domestic setting, e.g. a room. Additionally, by being cylindrical in shape, the heat exchangers may be located around other components of the product to provide a relatively compact arrangement.

The thermal store may form a sleeve that surrounds a major portion of the component. As discussed previously, the thermal store may then act as an acoustic blanket to absorb a greater proportion of the noise generated by the component. Moreover, by forming a sleeve around the component, a more compact arrangement may be achieved.

Brief Description of the Drawings

Figure l is a front perspective view of a fan assembly;

Figure 2 is a rear perspective view of the fan assembly with the components of the fan assembly removed;

Figure 3 is a front perspective view of the fan assembly with components of the fan assembly removed;

Figure 4 is a block diagram of components of the fan assembly;

Figure 5 is a schematic of a refrigeration system of the fan assembly in a first state;

Figure 6 is a schematic of the refrigeration system in a second state;

Figure 7 is a vertical section through part of the fan assembly;

Figure 8 is a horizonal section through the fan assembly; and

Figure 9 is a perspective view of an alternative fan assembly with an alternative condensation collector visible.

Detailed Description of the Invention

The fan assembly 10 of Figures 1 to 4 comprises a nozzle 11 and a main body 15.

The nozzle 11 is attached to the main body 15 and comprises an inlet 12 for receiving an airflow from the main body, and an outlet 13 for emitting the airflow. In the example of Figure 1, the nozzle 11 is generally racetrack shaped, the inlet 12 comprises an opening in a base of the nozzle 11, and the outlet 13 comprises a pair of slots that each extend along straight portions of the nozzle 11. In some examples, the nozzle 11 may comprise slats, louvres or other means for changing the direction of the airflow emitted from the outlet 13. Thereby, the direction of the airflow may be changed without the need to rotate the nozzle 11 or main body 15.

The main body 15 comprises a housing 17, a pair of filter assemblies 18,19, an airflow generator 20, a refrigeration system 21, a condensation collector 77,79, and a control unit 23.

The housing 17 houses the filter assemblies 18,19, the airflow generator 20, the refrigeration system 21, the condensation collector 77,79, and the control unit 23. The housing 17 comprises an inlet 25 through which an airflow is drawn into the main body 15, and an outlet 26 through which the airflow is emitted into the nozzle 11. In the illustrated example, the housing 17 is cylindrical in shape, the inlet 25 comprises a plurality of apertures in a side wall of the housing 17, and the outlet 26 comprises an opening in a top wall of the housing 17.

Each of the filter assemblies 18,19 comprises a filter medium 24 supported by a frame 27, and a seal 28 provided around the perimeter of the frame 27. Each of the filter assemblies 18,19 is removably attached to a section of the housing 17, which in turn is removable from the main body 15. As a result, the sections of the housing may be removed, and the filter assemblies 18,19 removed from the sections for cleaning and/or replacement.

Each of the filter assemblies 18,19 is arcuate and subtends a central angle of roughly 180°. The filter assemblies 18,19 surround a heat exchanger 49 of the refrigeration system 21 (described below in more detail) and the airflow generator 20. In this example, the filter medium 24 is a HEPA filter medium that removes particulates, such as pollutants and bacteria, from the airflow. However, other or additional filter media could be employed, such as an activated carbon filter medium for removing undesirable gases, such as volatile organic compounds, from the airflow.

The seal 28 of each filter assembly 18,19 seals against the main body 15 and reduces potential leak paths, in which air is drawn into the main body 15 but bypasses the filter assemblies 18,19. Thereby the purity of the airflow emitted from the fan assembly 10 may be improved.

The airflow generator 20 comprises an impeller driven by an electric motor. The airflow generator 20 generates an airflow between the inlet 25 of the housing 17 and the outlet 26 of the main body 15. More particularly, the airflow is drawn into the housing 17 via the inlet 25 of the housing 17, whereupon the airflow is drawn through the filter assemblies 18,19 to remove particulates from the airflow. The airflow is then drawn over the heat exchanger 49 of the refrigeration system 21 to condition the airflow. The conditioned airflow then moves through the airflow generator 20, and is emitted from the main body 15 via the outlet 26.

By locating the filter assemblies 18,19 upstream of the heat exchanger 49, the filter assemblies 18,19 presents a restriction to the airflow that moves over the heat exchanger 49. As a result, the airflow may be more uniformly distributed over the heat exchanger 49, thereby improving the performance of the heat exchanger 49.

The refrigeration system 21, which is described below in more detail, is operable in one of two states to condition the airflow. In a first operating state, the refrigeration system 21 cools the airflow, and in a second operating state, the refrigeration system 21 warms the airflow.

The condensation collector 77,79 collects condensate that forms on the heat exchanger 49 and comprises a tray 77 and a bottle 79. The tray 77 is located beneath the heat exchanger 49 and acts to collect condensate that falls from the heat exchanger 49. The tray 77 has a sloped upper surface that guides the collected condensate to a drain 78 in the tray 77. The bottle 79 is located directly beneath the drain 78 such that condensate collected by the tray 77 drains into the bottle 79 via the drain 78.

In this example, the tray 77 has a generally circular shape with a pair of triangular holes for allowing components of the refrigeration system to pass through the tray 77.

The tray 77 and the bottle 79 are each removable from the fan assembly 10, e.g. in radial directions. The tray 77 and the bottle 79 may therefore be removed and emptied to remove collected condensate from the fan assembly 10. As a result, condensate can be collected and removed from the fan assembly 10 without requiring the fan assembly 10 to be connected to an external drain. The fan assembly 10 may therefore be self-contained, and thereby the portability of the fan assembly 10 may be improved. Additionally, moving the whole fan assembly 10 to remove condensate is not required. As the fan assembly 10 may be relatively heavy, this may improve the usability of the fan assembly 10. Furthermore, the tray 77 and the bottle 79 can be removed for cleaning, which may prevent of bacteria and malodour build up within the fan assembly 10.

The tray 77 is received within a slot in the main body 15 that is located inwardly of the filter assemblies 18,19. Accordingly, in order to remove the tray 77, the user must first remove one of the filter assemblies 18,19. The tray 77 is therefore located downstream of the seal 28 of the filter assembly 18 and thus does not present a leak path through which the airflow may bypass the filter assemblies 18,19.

The bottle 79 is received within a recess 80 in the main body 15. To facilitate the removal of the bottle 79, the bottle 79 comprises a slidable locking portion 82 which engages a slot 84 in the housing 17. To detach and attach the bottle 79, the user slides the locking portion 82 into and out of the slot 84. The bottle also comprises a spout 86 which engages with the drain 78 of the tray 77. In this example, the condensation collector 77,79 has a capacity of 400 mL. Specifically, the tray 77 has a capacity of 100 ml and the bottle 79 has a capacity of 300 mL. This may improve the usability of the fan assembly 10, as the condensation collector 77,79 may be emptied relatively infrequently. Indeed, the size of the condensation collector 77,79 may be sufficiently large to collect all condensation generated in a full day of use of the fan assembly 10. Equally, the condensation collector 77,79 may have a capacity of greater than 200 mL and realise the above benefit of relatively infrequent emptying.

Turning now to Figure 4, the control unit 23 comprises a controller 29, a wireless interface 31, and a temperature sensor 33.

The controller 29 is responsible for controlling the operation of the fan assembly 10. The controller 29 is connected to the airflow generator 20, the refrigeration system 21, the wireless interface 31, and the temperature sensor 33. The controller 29 controls the airflow generator 20 and the refrigeration system 21 in response to data received from the wireless interface 31 and the temperature sensor 33. For example, the controller 29 may power on and off the airflow generator 20, control the speed of the airflow generator 20, and/or control the operating state of the refrigeration system 21.

The wireless interface 31 receives command data from one or more remote devices 35. In examples, the remote devices 35 may comprise a user-operated device. For example, the remote devices 35 may comprise a dedicated remote control or a mobile device, such as a phone or tablet, running a suitable application. A user may then use the remote device to control remotely the operation of the fan assembly 10. For example, the device may be used to power on and off the fan assembly 10, control the speed and/or the direction of the airflow, as well as schedule operation of the fan assembly 10. In other examples, the remote devices 35 may comprise a room thermostat or other remote temperature sensor, which transmits temperature data to the wireless interface 31. The controller 29 may then operate the fan assembly 10 in response to changes in the temperature data. For example, the controller 29 may control the airflow generator 20 and/or the refrigeration system 21 such that a temperature within a room is maintained at a target temperature.

The temperature sensor 33 forming part of the control unit 23 monitors a temperature of the refrigeration system 21, discussed in more detail below, and outputs temperature data to the controller 29.

The control unit 23 may additionally comprise a user interface for controlling the operation of the fan assembly 10. For example, the user interface may comprise buttons, dials, a touchscreen or the like for powering on and off the airflow generator 20, as well as controlling the speed and/or direction of the airflow.

The fan assembly 10 is operable in one of a cooling mode and a regeneration mode.

In cooling mode, the controller 29 operates the refrigeration system 21 in the first state and operates the airflow generator 20 at a first speed. Thereby an airflow is drawn in through the inlet 25, through the filter assemblies 18,19, over the heat exchanger 49 of the refrigeration system 21 and emitted from the outlet 13 of the nozzle 11. As the refrigeration system 21 is operating in the first state, the airflow is cooled by the refrigeration system 21 and thus a cooled airflow is emitted from the nozzle 11. The speed of the airflow generator 20 in the cooling mode may be defined by the command data received by the controller 29. In this way, the speed of the airflow generator 20 may be controlled to achieve different cooling rates or profiles.

In regeneration mode, the controller 29 operates the refrigeration system 21 in the second state and operates the airflow generator 20 at a second speed. Again, an airflow is drawn in through the inlet 25, though the filter assemblies 18,19, over the heat exchanger 49 of the refrigeration system 21 and is emitted from the outlet 13 of the nozzle 11. As the refrigeration system 21 is operating in the second state, the airflow is warmed by the refrigeration system 21 and thus a warmed airflow is emitted from the nozzle 11. Regeneration mode is used to expel heat that was stored by the refrigeration system 21 during cooling mode. The speed of the airflow generator 20 in regeneration mode may be lower than that used in cooling mode, i.e. the second speed may be lower than the first speed. For example, the airflow generator 20 may operate at a relatively low or trickle speed in regeneration mode. As a result, the noise generated by the fan assembly 10 when operating in regeneration mode may be reduced.

The fan assembly 10 is intended to be used primarily to provide a cooled airflow. This cooled airflow may be used, for example, to cool a room. To achieve this, the fan assembly 10 operates in cooling mode. In cooling mode, as described above, the airflow is drawn over the heat exchanger 49, which extracts heat from the airflow, and the now cooled airflow is emitted from the nozzle 11. The extracted heat is stored within the refrigeration system 21. Cooling then continues until either cooling is no longer required (e.g. the fan assembly is turned off, or the temperature within the room has reached a target setpoint), or the maximum heat storage capacity of the refrigeration system 21 has been reached. As described below in further detail, the heat stored by the refrigeration system 21 may be sensed by the temperature data output by the temperature sensor 33, and the controller 29 may determine that the maximum heat storage capacity of the refrigeration system 21 has been reached when the temperature exceeds an upper threshold.

During periods when cooling is not required, or when the maximum heat storage capacity of the refrigeration system 21 has been reached, the fan assembly 10 may operate in regeneration mode. In regeneration mode, the fan assembly expels the heat that was stored during cooling. As a result, the fan assembly 10 is restored to a state in which cooling is possible. Regeneration mode may continue until either cooling is required or the full heat storage capacity of the refrigeration system 21 has been restored. As described below, the controller 29 may determine that the heat storage capacity of the refrigeration system 21 has been fully restored when the temperature drops below a lower threshold. As a warmed airflow is emitted from the fan assembly 10 when operating in regeneration mode, regeneration may occur at times when the room is unoccupied (or unlikely to be occupied) or at times when warming is actually desirable. For example, the fan assembly be scheduled to operate in cooling mode during the day, and regeneration mode during the night. In a further example, geofencing may be employed, and the fan assembly 10 may operate in regeneration mode when a user is no longer present in the room or building in which the fan assembly 10 is located.

Turning now to Figures 5 and 6, reference will now be made to the composition and operation of the refrigeration system 21. The refrigeration system 21 comprises a circuit 41 and a thermal store 43.

The circuit 41 comprises a series of pipes 45, a first heat exchanger 46, a compressor 47, a metering device 48, and a second heat exchanger 49.

The series of pipes 45 connect the compressor 47 to the first heat exchanger 46, the first heat exchanger 46 to the metering device 48, the metering device 48 to the second heat exchanger 49, and the second heat exchanger 49 to the compressor 47 such that a refrigerant can circulate around the circuit 41.

The first heat exchanger 46 is downstream of the compressor 47 and upstream of the metering device 48, and exchanges heat between the refrigerant and the thermal store 43. The second heat exchanger 49 is located downstream of the metering device 48 and upstream of the compressor 47, and exchanges heat between the refrigerant and the airflow moving through the fan assembly 10.

The compressor 47 drives the refrigerant around the circuit 41 in a direction indicated by the arrow in Figures 5 and 6. The refrigerant circulates from the compressor 47 to the first heat exchanger 46, from the first heat exchanger 46 to the metering device 48, from the metering device 48 to the second heat exchanger 49, and from the second heat exchanger 49 to the compressor 47. Depending on the state of operation, discussed subsequently, the compressor 47 may additionally compress the refrigerant.

The metering device 48 is operable in a restricted state and an unrestricted state. In the restricted state, the refrigerant flowing through the metering device 48 expands and the pressure and temperature of the refrigerant decreases. In the unrestricted state, the refrigerant flowing through the metering device 48 does not expand and the pressure and temperature of the refrigerant is unchanged. In this example, the metering device 48 comprises a variable expansion valve. In the restricted state, the variable expansion valve has a first restriction, and in the unrestricted state, the variable expansion valve has a second, less restrictive restriction. In other examples, the metering device 48 may comprise a capillary tube and the refrigeration system 21 may comprise a bypass valve for bypassing the metering device 48 in the second state.

The thermal store 43 stores thermal energy for transfer to and from the refrigerant in order to heat and cool the refrigerant. In this particular example, the thermal store 43 comprises a phase change material. This then has the benefit that the thermal store 43 can take advantage of the latent heat capacity of the phase change material to store more thermal energy for a given change in temperature. As a result, the refrigeration system 21 may provide cooling at the second heat exchanger 49 for a longer period. Nevertheless, the refrigeration system 21 may operate with a thermal store 43 that does not comprise a phase change material. The phase change material may have a melting point greater than the ambient temperature of the room. This then has the advantage that heat stored by the thermal store may be expelled to the room in regeneration mode. A relatively high melting point has the advantage of increasing the rate at which heat is expelled in regeneration mode, and thus decreasing the time required to regenerate the thermal store. A relatively low melting point, on the other hand, has the advantage of improving the efficiency of the refrigeration system in cooling mode. A relatively good balance between these two competing factors may be achieved with a phase change material having a melting point of between 30°C and 80°C. In some examples, the phase change material may comprise an organic wax or inorganic salt hydrate In addition to the functions described above, the controller 29 controls the compressor 47 and the metering device 48. For example, the controller 29 may power on and off the compressor 47, as well as control the state of the metering device 48 and the speed of the compressor 47 in response to control data received from the wireless interface 31 and the temperature sensor 33.

As discussed above, the refrigeration system 21 is operable in a first state and a second state.

In the first operating state, shown in Figure 5, the controller 29 moves the metering device 48 to the restricted state. As a consequence of the metering device 48 being in the restricted state, the pressure and temperature of the refrigerant flowing though the metering device 48 decreases. In this particular example, the refrigerant remains in the liquid state, but could conceivably undergo a phase transition from a liquid state to a liquid-vapour state. The refrigerant flowing through the second heat exchanger 49 is at a lower temperature than the airflow moving over the second heat exchanger 49. Consequently, the second heat exchanger 49 acts as an evaporator to cool the airflow, and heat and vaporise the refrigerant. The refrigerant therefore undergoes a phase transition from a liquid state to a vapour state. The refrigerant then flows from the second heat exchanger 49 to the compressor 47, whereupon the refrigerant is compressed to increase the pressure, and thus the temperature, of the refrigerant. The refrigerant then flows through the first heat exchanger 46, which exchanges heat between the refrigerant and the thermal store 43. The refrigerant flowing through the first heat exchanger 46 is at a higher temperature than the thermal store 43. As a result, the first heat exchanger 46 acts as a condenser to heat the thermal store 43, and cool and condense the refrigerant. The refrigerant therefore undergoes a phase transition from a vapour state to a liquid state. The refrigerant then flows to the metering device 48, and the cycle is repeated.

In the second operating state, shown in Figure 6, the controller 29 moves the metering device 48 to the unrestricted state. As a consequence of the metering device 48 being in the unrestricted state, the pressure and temperature of the refrigerant flowing though the metering device 48 is unchanged. In this particular example, the refrigerant is in a vapour state, but could conceivably be in a liquid-vapour or a liquid state. Refrigerant flowing through the second heat exchanger 49 is at a higher temperature than the airflow moving over the second heat exchanger 49. Consequently, the airflow is heated, and the refrigerant is cooled. In this particular example, the refrigerant is not cooled below its boiling point and thus the refrigerant does not condense or undergo a phase change. The refrigerant then flows from the second heat exchanger 49 to the compressor 47. Owing to the unrestricted state of the metering device 48, the compressor 47 does not compress the refrigerant. The refrigerant then flows through the first heat exchanger 46, which exchanges heat between the refrigerant and the thermal store 43. The refrigerant flowing through the first heat exchanger 46 is at a lower temperate than the thermal store 43. As a result, the thermal store 43 is cooled, and the refrigerant is heated. In this particular example, the refrigerant flowing through the first heat exchanger 46 is in a vapour state and does not therefore undergo a phase transition. The refrigerant then flows to the metering device 48, and the cycle is repeated.

As noted above, the fan assembly 10 is operable in one of two modes: cooling and regeneration. When operating in cooling mode, the controller 29 configures the refrigeration system in the first state. The controller 29 then monitors the temperature of the thermal store 43 (via the temperature sensor 33). In the event that the temperature of the thermal store 43 exceeds an upper threshold, the controller 29 powers off the airflow generator 20 and the refrigeration system 21 (i.e. the compressor 47), or alternatively switches from cooling mode to regeneration mode. The upper threshold may represent a temperature above which the refrigeration system 21 is no longer able to effectively or efficiently cool the airflow. In this regard, the efficiency of the refrigeration system 21 decreases as the difference in the temperatures of the two heat exchangers 46,49 increases. Alternatively, the upper threshold may represent a temperature above which the volume expansion of the thermal store 43 becomes excessive, or the temperature of the thermal store 43 becomes excessively hot, which may present a safety concern or may lead to adverse changes in the physical and/or chemical properties of the thermal store 43. Additionally or alternatively, the upper threshold may represent a temperature above which the pressure of the refrigerant becomes excessive.

When operating in regeneration mode, the controller 29 configures the refrigeration system in the second state. The controller 29 again monitors the temperature of the thermal store 43. In the event that the temperature drops below the lower threshold, the controller 29 powers off the airflow generator 20 and the refrigeration system 21, or alternatively switches from regeneration mode to cooling mode. As noted, the efficiency of the refrigeration system 21 increases as the difference in the temperatures of the heat exchangers 46,49 decreases. The lower threshold may therefore represent a temperature below which the refrigeration system 21 is again able to effectively or efficiently cool the airflow. Where the thermal store 43 comprises a phase change material, the upper and lower thresholds may be respectively greater and lower than the melting point of the phase change material. For example, where the phase change material has a melting point of 46°C, the upper threshold may be 48°C and the lower threshold may be 44°C.

Turning now to Figures 7 and 8, reference will now be made to how the various components of the fan assembly 10 are packaged.

The main body 15 comprises a tank 50 that contains the thermal store 43. The tank 50 is generally cylindrical in shape, but comprises a gap such that the tank 50 (and thus the thermal store 43) is c-shaped in cross-section. The tank 50 comprises an inner wall 51 and an outer wall 53 that are arranged concentrically. The inner wall 51 subtends a central angle of 360°, whilst the outer wall 53 subtends a central angle of 340°. Radial walls then extend between the ends of the outer wall 53 and the inner wall 51. The tank 50 is enclosed at the top and bottom by top and bottom walls 57,59.

The tank 50 partly defines a chamber 67 within which the compressor 47 is located. The compressor 47 is therefore partially surrounded by the thermal store 43. In the example of Figures 7 and 8, the thermal store 43 may be regarded as a sleeve that surrounds a major portion of the compressor 47. This then has two benefits. Firstly, a relatively large thermal store 43 may be packaged within the main body 15 in a relatively compact manner. Secondly, the thermal store 43 may absorb noise generated by the compressor 47.

The first heat exchanger 46 is embedded within the thermal store 43 and comprises piping 65 through which the refrigerant flows. Owing to spacing required for turns in the piping 65, the thermal store 43 does not completely surround the compressor 47 but instead subtends a central angle of about 340° about the compressor 47. Conceivably, the thermal store 43 could completely surround the compressor 47. However, this would then result in a portion of the thermal store 43 that is not in direct thermal contact with the first heat exchanger 46. Although this may help further absorb noise from the compressor 47, the additional material may not necessarily increase the heat storage capacity of the thermal store 43. By omitting this additional material, the cost and weight of the fan assembly 10 may be reduced. Moreover, advantage may be taken of the gap in the thermal store 43 by using it to locate another component of the fan assembly 10. In this example, the recess 80 extends into the gap such that the bottle 79 of the condensation collector 77,79 is located in the gap. Furthermore, in this example, the conditioned airflow is emitted from the front of the fan assembly 10 whereas the gap is located towards the rear of the fan assembly 10. As a result, any noise that may escape through the gap may be directed away from a user of the fan assembly 10.

The thermal store 43 has a height at least equal to that of the compressor 47. As a result, noise emitted in a sideways direction from the compressor 47 may be absorbed by the thermal store 43. Furthermore, the thermal store 43 has a thickness of around 60mm measured radially from a longitudinal axis of the thermal store 43. A relatively thick thermal store 43 has the advantage of absorbing more of the noise generated by the compressor 47, as well as increasing the thermal mass of the thermal store 43. A relatively thin thermal store 43, on the other hand, has the advantage of reducing the cost and weight of the fan assembly 10, which in turn may improve the portability of the fan assembly 10. A thermal store having a thickness of between 20mm and 150mm provides a good balance between these competing factors. The thermal store 43 comprises a high-density medium (e.g. a solid, liquid, or solid-liquid phase change material). As a result, relatively good sound absorption may be achieved at the lower frequencies typically generated by the compressor 47.

The first heat exchanger 46 is embedded within the thermal store 43. Consequently, the first heat exchanger 46 may be said to be cylindrical or annular in shape. By providing a thermal store 43 and/or a heat exchanger 46 that is cylindrical in shape, a relatively compact arrangement may be achieved. In particular, the thermal store 43 and/or the first heat exchanger 46 may surround one or more other components of the fan assembly 10, such as the compressor 47 in this instance.

The compressor 47 is located towards a base of the main body 15. The compressor 47 is a relatively heavy component of the fan assembly 10 and thus locating the compressor 47 towards the base of the main body 15 provides a lower centre of gravity which may improve the stability of the fan assembly 10. The compressor 47 is located within the chamber 67, bounded by the inner wall 51 of the tank 50. The top and bottom of the chamber 47 are then bounded by a first plate 73 and a second plate 75. The first plate 73 has one or more openings through which pipes 45 of the circuit 41 pass into and out of the chamber 67. By providing plates 73,75 directly above and below the compressor 47, noise emitted from the compressor 47 in the vertical directions may be absorbed.

The metering device 48 is located within the chamber 67. Thereby, noise generated by the metering device 48 (e.g. due to movement of the refrigerant through the metering device 48) may be absorbed by the thermal store 43 and plates 73,75. However, it is conceivable that the metering device 48 may be located outside of the chamber 67.

The airflow generator 20 is located above the compressor 47, towards the top of the main body 15. The airflow generator 20 is located centrally along a longitudinal axis of the main body 15. The second heat exchanger 49 is cylindrical or annular in shape and is positioned vertically above the thermal store 43 and the first heat exchanger 46. In this example, the first heat exchanger 46 and the second heat exchangers 49 are concentric. The second heat exchanger 49 then surrounds a lower part of the airflow generator 20. By stacking the heat exchangers 46,49 vertically in this way, and by locating components of the fan assembly 10 (e.g. airflow generator 20 and compressor 47) within the interior space defined by the heat exchangers 46,49, a relatively compact arrangement may be achieved. In particular, the footprint of the fan assembly 10 may be reduced.

The second heat exchanger 49 surrounds a major portion of the airflow generator 20. In this example, the second heat exchanger 49 subtends a central angle of roughly 300° about the airflow generator 20. In other examples, the second heat exchanger 49 may subtend a larger or smaller central angle. As the central angle decreases, however, the area of the second heat exchanger decreases. Accordingly, the second heat exchanger 49 may subtend a central angle of at least 270°. As a result, the height and footprint of the fan assembly 10 may be reduced whilst achieving a good area for the second heat exchanger 49.

The tray 77 of the condensation collector 77,79 is spaced from, and extends broadly parallel to, an inlet 22 of the airflow generator 20 such that the tray 77 may be considered to face the inlet 22 of the airflow generator 20. An axial gap 90 (i.e. measured in a direction normal to the inlet 22 or parallel to the rotational axis of the airflow generator 20) exists between the inlet 22 of the airflow generator 20 and the tray 77. In this example, the tray 77 has a sloped upper surface, so the axial gap 90 is measured between a centre of the inlet 22 and the tray 77. The axial gap 90 has a value of 35mm and the inlet 22 of the airflow generator 20 has a diameter of 68mm. Therefore a ratio of the axial gap 90 to the diameter of the inlet 22 is around 1 : 1.9. The applicant has observed that, if this ratio is too low, say below 1 :2.3, the airflow between the tray 77 and the inlet 22 of the airflow generator 20 may become pinched, leading to an increase in the velocity and distortion of the airflow. This increase in velocity and distortion results in several negative effects. Firstly, the noise generated by the airflow increases, which may be undesirable for a user of the fan assembly 10. Secondly, the airflow over the second heat exchanger 49 may become non-uniform. Specifically, the flow rate of the airflow over the lower part of the second heat exchanger 49 may be higher, whilst the flow rate of the airflow over the upper part of the second heat exchanger 49 may be lower. As a result, the performance of the second heat exchanger 49 decreases. By having a relatively high ratio, i.e. one that is at least 1 :2.3, the uniformity of the airflow over the second heat exchanger 49 may be improved, which may improve the performance of the second heat exchanger 49. Additionally, the noise produced by the airflow generator 20 may be reduced. As this ratio increases, however, the overall height of the fan assembly 10 is likely to increase. Accordingly, the ratio of the axial gap to the diameter of the inlet may be between 1 :2.3 and 1 : 1, and more particularly between 1 :2 and 1 : 1.1. Moreover, the axial gap 90 itself may be between 30mm and 70mm. This then provides a relatively good balance between the competing needs to reduce the height of the fan assembly 10 whilst also preventing pinching of the airflow.

The second heat exchanger 49 has a height 92 of 115mm. As a result, a ratio of the axial gap 90 to the height 92 of the second heat exchanger 49 is around 1 :3.3. The applicant has observed that, if this ratio is too low, say below 1 :4, the airflow over the second heat exchanger 49 may be non-uniform. Specifically, the flow rate of the airflow over the lower part of the second heat exchanger 49 may be higher, whilst the flow rate over the upper part of the second heat exchanger 49 may be lower. As this ratio increases, however, the overall height of the fan assembly 10 is likely to increase. A ratio of between 1 :4 and 1 :3, and more particularly between 1 :3.8 and 1 :3.1, may provide a good balance between the competing needs to reduce the height of the fan assembly 10, whilst also providing uniform airflow over the second heat exchanger 49.

The fan assembly 10 additionally comprises a radial gap 94 between the airflow generator 20 and the second heat exchanger 49 of 30mm. Specifically, the radial gap 94 is in a direction parallel to the inlet 22 of the air flow generator 20 and is measured in a plane of the inlet 22 of the airflow generator 20. The fan assembly 10 therefore has a ratio of the radial gap 94 to the height 92 of the second heat exchanger 49 of around 1 :3.3. For relatively low ratios, the applicant has observed that the airflow moving over the second heat exchanger 49 may become non-uniform, which may result in a reduction in the performance of the second heat exchanger 49. However, relatively high ratios, on the other hand, are likely to increase the footprint of the fan assembly 10. Having a ratio of between 1 :5 and 1 :3 may provide a good balance between the competing needs of improving the uniformity of the airflow over the second heat exchanger 49, and providing a compact arrangement. Additionally, having a radial gap 94 of between 20mm and 50mm may also provide a good balance between the competing needs in a fan assembly sufficiently large to provide a useful airflow rate for use in a domestic setting.

The second heat exchanger 49 is located at approximately the same height as the inlet 25 in the housing 17. Consequently, the apertures in the housing 17 may be said to surround the second heat exchanger 49. As a result of this arrangement, a relatively straight, radial path may be taken by the airflow when moving from the inlet 25 to the airflow generator 20. By providing a relatively straight, less contorted path, pressure losses may be reduced and thus a higher flow rate may be achieved for the airflow.

The tray 77 of the condensation collector 77,79 is located beneath the second heat exchanger 49 and above the first heat exchanger 46 (i.e. between the first 46 and second 49 heat exchangers). Locating the tray 77 between the two heat exchangers makes better use of the available space within the fan assembly 10 and thereby may provide a more compact arrangement.

With the fan assembly 10 described above, the thermal store 43 performs two important functions. Firstly, the thermal store 43 stores heat transferred from the first heat exchanger 46, thereby obviating the requirement to expel the heat immediately into the surrounding environment. Secondly, the thermal store 43 absorbs noise generated by the compressor 47. Acoustic emissions from the fan assembly 10 may therefore be reduced without the requirement for separate noise-absorbing materials, such as acoustic foams. In the example described above, the thermal store 43 subtends a central angle 61 of 340°. However, the thermal store 43 may subtend a smaller angle, or indeed a larger angle. For example, the thermal store 380 may subtend an angle of, say, at least 180°. As a result, noise generated by the compressor 47 may be absorbed around at least one half of the fan assembly 10. The fan assembly 10 may then be sited adjacent a wall within a room and oriented such that the portion of the compressor 47 covered by the thermal store 43 is directed towards the centre of the room, whereas the uncovered portion may be directed towards the wall. Significant acoustic improvements may therefore be achieved without necessarily requiring the thermal store 43 to surround wholly the compressor 47.

In the above example the first and second plates 73,75 cover the top and bottom of the chamber 67. However, conceivably the one or both of the plates 73,75 may be omitted. Moreover, the thermal store 43 may be arranged to cover at least part of the top and/or bottom of the chamber 67. This may then further reduce the emission of noise from the fan assembly 10.

In the above example, the fan assembly is used to cool a room. The phase change material is then warmed and melts in cooling mode, and cools and solidifies in regeneration mode. In an alternative example, the fan assembly may be used to heat the room. The phase change material is then cooled and solidifies in a heating mode, and warmed and melts in regeneration mode. In this alternative example, the phase change material may have a melting point below the ambient temperature of the room, e.g. a melting point of 0 °C, such that the phase change material transitions from a liquid to a solid state in heating mode, and vice versa in regeneration mode. Again, in so doing, advantage may be taken of the latent heat capacity of the phase change material to store and subsequently release relatively large amounts of heat for a given temperature range.

In the example described above, the user must first remove one of the filter assemblies in order to access and remove the tray. Figure 9 illustrates an alternative example in which the tray may be removed without having to first remove a filter assembly. The condensation collector of Figure 9 comprises a tray 110 which comprises a further seal 112 which extends around an outside of the tray 110 and is for sealing the tray 110 against the main body 15. The tray 110 is located within a slot 114 which extends beneath the filter assemblies 18,19 and second heat exchanger 49 and has an opening accessible from the outside of the main body 15. As a result, the tray 110 is removable from the main body 15 independently of the filter assemblies 18,19. Therefore, by providing an independently removable tray 110, the user is not required to unnecessarily remove the filter assemblies 18,19 to remove the tray 110. This may improve the ease of use of the fan assembly 10. Additionally, the condensation collector does not comprise a bottle. Instead, the bottle is omitted and the tray 110 of Figure 9 is deeper than the tray 77 of Figure 2. As a result, the tray 110 is capable of capturing a larger amount of condensate. Omitting the bottle may provide a simpler arrangement for the condensation collector.

Although described above in connection with a fan assembly, the refrigeration system and/or the particular packaging of the components may be used in other products, such as a ducted portable air conditioning device or a dehumidifier.

In the above examples, the thermal store surrounds the compressor of the refrigeration system of the product. In other examples, the thermal store may surround an alternative component of the product. For example, where the product comprises an airflow generator or other turbomachine, the thermal store may surround the turbomachine. Accordingly, in a more general sense, the product may be said to comprise a component and a thermal store, and the thermal store may be said to surround at least part of the component to absorb noise generated by the component.

In the above examples, the airflow generator 20 is operable to generate an airflow through the fan assembly 10 having a maximum flow rate of 25 L/s. In other examples, the airflow generated by the air generator 20 may have a different maximum flow rate. However, by having an airflow of at least 20 L/s, the fan assembly 10 may provide effective cooling of a user and/or heating of a relatively large volume, such as a room. The above examples are to be understood as illustrative examples of the invention. Furthermore, equivalents and modifications not described above may also be employed without departing from the scope of the invention, which is defined in the accompanying claims.