EVAPORATOR HEAT EXCHANGER FOR PREVENTING ICE BUILD-UP

Abstract

A system includes a compressor for increasing the pressure of a refrigerant; a condenser heat exchanger arranged downstream of the compressor for receiving a high pressure refrigerant output from the compressor and for transferring heat from the high pressure refrigerant to a source of water; an expansion device for reducing the pressure of a refrigerant; and an evaporator heat exchanger for extracting heat from ambient air. The evaporator heat exchanger includes a first tube bank having a first inlet arranged to receive a high pressure refrigerant output from the condenser and a first outlet fluidly coupled to the expansion device; and a second tube bank having a second inlet arranged to receive a low pressure refrigerant output from the expansion device, and a second outlet fluidly coupled to an inlet of the compressor.

Claims

1. An air-to-water heat pump system comprising: a compressor for increasing the pressure of a refrigerant; a condenser heat exchanger arranged downstream of the compressor for receiving a high pressure refrigerant output from the compressor and for transferring heat from the high pressure refrigerant to a source of water; an expansion device for reducing the pressure of a refrigerant; and an evaporator heat exchanger for extracting heat from ambient air, the evaporator heat exchanger comprising: a first tube bank having a first inlet arranged to receive a high pressure refrigerant output from the condenser, and a first outlet fluidly coupled to the expansion device; and a second tube bank having a second inlet arranged to receive a low pressure refrigerant output from the expansion device, and a second outlet fluidly coupled to an inlet of the compressor; wherein the first and second tube banks are arranged in close proximity to one another such that, in use, heat from the high pressure refrigerant passing through the first tube bank is transferred to the second tube bank in order to limit and/or prevent ice build-up on an external surface of the second tube bank.

2. The air-to-water heat pump system as claimed in claim 1, wherein a gap exists between the first and second tube banks, the gap being 1 mm or less in size, preferably 0.5 mm or less in size.

3. The air-to-water heat pump system as claimed in claim 1, wherein the first tube bank is arranged in a heat exchange relationship with ambient air such that, in use, heat is transferred from the high pressure refrigerant to the ambient air, thereby cooling the high pressure refrigerant and warming the ambient air; and/or wherein the second tube bank is arranged in a heat exchange relationship with ambient air such that, in use, heat is transferred from the ambient air to the low pressure refrigerant, thereby warming the low pressure refrigerant and cooling the ambient air.

4. The air-to-water heat pump system as claimed in claim 1, wherein the first and/or second tube bank comprises one or more rows arranged in parallel with one another and fluidly coupled to one another in series.

5. The air-to-water heat pump system as claimed in claim 5, wherein each row of the first and/or second tube bank comprises a plurality of first and/or second tubes connected in series with one another, optionally wherein the plurality of first and/or second tubes are arranged in a serpentine or coil shape.

6. The air-to-water heat pump system as claimed in claim 1, wherein the first and/or second tube banks comprise a plurality of refrigerant circuits.

7. The air-to-water heat pump system as claimed in claim 1, wherein the first tube bank comprises a first plurality of fins and the second tube bank comprises a second plurality of fins, the first plurality of fins and the second plurality of fins being in close proximity to one another.

8. The air-to-water heat pump system as claimed in claim 1, wherein the system is configured such that the freeze limit of the system is less than or equal to 10° C., preferably less than or equal to 7° C., the freeze limit being the minimum ambient temperature at which ice will not form on the heat exchanger.

9. The air-to-water heat pump system as claimed in claim 1, wherein the system is configured such that, in use, the temperature of the external surface of the first tube bank is at least 10° C. higher than the temperature of the external surface of the second tube bank.

10. The air-to-water heat pump system as claimed in claim 1, comprising a fan arranged to create a flow of air over the evaporator heat exchanger, preferably wherein the fan is configured to flow air over the first tube bank before the air flows over the second tube bank.

11. A method of preventing and/or limiting the build-up of ice on an evaporator heat exchanger in situ in an air-to-water heat pump, the method comprising: passing a high pressure refrigerant through a first tube bank of the evaporator heat exchanger in order to cool the high pressure refrigerant though heat exchange between the high pressure refrigerant and ambient air; reducing the pressure of the cooled high pressure refrigerant, so as to provide a low pressure refrigerant; and passing the low pressure refrigerant through a second tube bank of the evaporator heat exchanger in order to warm the low pressure refrigerant through heat exchange between the low pressure refrigerant and ambient air; wherein the first and second tube banks are arranged in close proximity to one another such that heat from the high pressure refrigerant passing through the first tube bank is transferred to the second tube bank in order to heat at least a portion of an external surface of the second tube bank, thereby limiting and/or preventing ice from building up on the external surface of the second tube bank.

12. The method as claimed in claim 11 comprising: increasing the pressure of the refrigerant with a compressor, thereby providing the high pressure refrigerant, before passing the refrigerant to a condenser heat exchanger; and prior to passing the refrigerant to the first tube bank, using the condenser heat exchanger to cool the high pressure refrigerant through heat exchange between the high pressure refrigerant and a water source, thereby warming the water source.

13. The method as claimed in claim 12, drawing air over the first and second tube banks, preferably wherein the air is drawn over the first tube bank before flowing over the second tube bank.

14. The method as claimed in claim 12, wherein the refrigerant is R-32 refrigerant or propane.

15. A method of manufacturing an air-to-water heat pump system, comprising connecting, in series: a compressor for increasing the pressure of a refrigerant, a condenser heat exchanger for receiving a high pressure refrigerant output from the compressor and for transferring heat from the high pressure refrigerant to a source of water, a first tube bank of an evaporator heat exchanger, the evaporator heat exchanger being for extracting heat from ambient air, an expansion device for reducing the pressure of a refrigerant received from the first tube bank of the evaporator heat exchanger, and a second tube bank of the evaporator heat exchanger, wherein the first and second tube banks of the evaporator heat exchanger are arranged in close proximity to one another such that, in use, heat from a high pressure refrigerant passing through the first tube bank will be transferred to the second tube bank in order to limit and/or prevent ice build-up on an external surface of the second tube bank.

Description

BRIEF DESCRIPTION OF THE DRAWINGS

[0057] Certain embodiments of the disclosure will now be described by way of example only and with reference to the accompanying drawings in which:

[0058] FIG. 1 is a schematic diagram of an air-to-water heat pump system;

[0059] FIG. 2 is a schematic diagram of an alternative air-to-water heat pump system;

[0060] FIG. 3 is a schematic diagram of an evaporator heat exchanger of the air-to-water heat pump system of FIG. 2; and

[0061] FIG. 4 is a schematic diagram of a row of an evaporator heat exchanger tube bank.

[0062] FIG. 5 is schematic diagram of a first tube bank and a second tube bank of an evaporator heat exchanger.

DETAILED DESCRIPTION OF THE INVENTION

[0063] As shown in FIG. 1, an air-to-water heat pump system 20 comprises a compressor 22, a condenser heat exchanger 24, an evaporator heat exchanger 26, and an expansion device 28 arranged in a fluid circuit to permit a flow of refrigerant through the system 20. The heat pump system may comprise a fan 30, which may be configured to create a flow of air A over the evaporator heat exchanger 26. Typically, in air-to-water heat pump systems such as the heat pump system 20, the evaporator heat exchanger 26 is located outdoors such that it can extract heat from the ambient air. The heat pump system 20 may further comprise a controller (not shown), which may be configured to control certain aspects of the heat pump system 20. For example, the controller may be configured to control the compressor speed and/or the opening degree of the expansion device 28.

[0064] During operation of the heat pump system 20, the compressor 22, which may be a scroll compressor, compresses a refrigerant to produce a high pressure refrigerant, which may be partly or completely in gaseous form. The high pressure refrigerant exits the compressor 22 and enters the condenser heat exchanger 24, where it exchanges heat with a flow of water 32. As the high pressure refrigerant has a higher temperature than the flow of water 32, heat is transferred from the high pressure refrigerant to the flow of water 32. Thus, the flow of water 32 is heated whilst the high pressure refrigerant is cooled. The heated flow of water 32 may travel to an indoor area, for example to a radiator, and may provide heating to that indoor area. The condenser heat exchanger 24 may be configured to condense at least a portion of the high pressure refrigerant, converting that portion from the gaseous phase to the liquid phase. However, a portion of the refrigerant may remain in the gaseous phase upon exiting the condenser heat exchanger 24.

[0065] After exiting the condenser heat exchanger 24, the high pressure refrigerant passes into the evaporator heat exchanger 26. Specifically, the high pressure refrigerant passes into a first tube bank 34 of the evaporator heat exchanger 26. Here, the high pressure refrigerant may exchange heat with the ambient air, such that the high pressure refrigerant is cooled by the ambient air, and the ambient air is heated by the high pressure refrigerant. Preferably, the refrigerant is cooled sufficiently such that it is all, or at least predominantly, in liquid form as it exits the first tube bank 34. The high pressure refrigerant exits the first tube bank 34 and enters the expansion device 28, where it is reduced in pressure, resulting in the lowering of the temperature of the refrigerant. In the expansion device 28, the refrigerant becomes a low pressure refrigerant.

[0066] The low pressure refrigerant exits the expansion device 28 and enters a second tube bank 36 of the evaporator heat exchanger 26. Here, the ambient air may exchange heat with the cooled, low pressure refrigerant, such that the low pressure refrigerant is heated and the ambient air is cooled. This may result in at least a portion of the low pressure refrigerant in the second tube bank 36 being evaporated, causing that portion to change phase from a liquid refrigerant to a gaseous refrigerant. The refrigerant may then exit the second tube bank 36 and be passed back to the compressor 22, beginning the cycle again.

[0067] When the ambient air temperature is low, there is a risk of ice forming and building up on the evaporator heat exchanger 26. This may occur when water vapour present in the ambient air encounters the cold outer surface of the evaporator heat exchanger 26 (that has been cooled by the presence of the low pressure refrigerant flowing therethrough), causing the water vapour to condense and, if the air is cooled to below the freezing point of the water (e.g. 0° C.), freeze. As a result, ice may begin to build-up on large portions, if not all, of the outer surface of the evaporator heat exchanger 26. This may prevent the ambient air from directly contacting the evaporator heat exchanger 26, thus limiting heat exchange between the ambient air and the low pressure refrigerant flowing within the second tube bank 36. An excessive build-up of ice on the evaporator heat exchanger 24 may therefore reduce the efficiency of the heat pump system 20, and/or reduce its coefficient of performance (COP). The build-up of ice on the heat exchanger 26 can be a major problem in countries with colder climates, or during the winter months, as outdoor temperatures are frequently sufficiently low that further cooling of the air through interaction with the evaporator heat exchanger 26 causes ice formation on the evaporator heat exchanger 26. In typical systems, ice build-up may begin to occur when the ambient air temperature is lower than 10° C., and the rate of ice build-up will generally increase as the ambient air temperature reduces further.

[0068] To counter this issue, heat pump systems frequently have external heaters arranged in close proximity to the evaporator heat exchanger. These heaters may be turned on at given intervals, or may be turned on in response to an indication of excessive ice build-up. This use of external heaters not only requires additional energy in order to power the heater, but also increases the footprint of the heat pump system. Such bulky systems are generally undesirable. Alternatively, a “defrost” mode may be employed by the system when ice build-up is excessive. A “defrost” mode is typically when the flow of refrigerant through the heat pump system is reversed, causing the evaporator heat exchanger and the condenser heat exchanger to switch functions. That is, the flow of refrigerant through the system is typically reversed so that a relatively warm, high pressure refrigerant is flown through the evaporator heat exchanger and a relatively cool, low pressure refrigerant is flown through the condenser heat exchanger. In this way, heat is emitted from the high pressure refrigerant within the evaporator heat exchanger, thus melting ice that may have built-up on the surface of the heat exchanger. However, a lot of energy is required for such a defrost mode, and the use of such a mode disrupts the normal heating function of the heat pump system.

[0069] As can be seen from FIG. 1, the first tube bank 34 of the evaporator heat exchanger 26, which is on the high pressure side of the heat pump system 20, and the second tube bank 36 of the evaporator heat exchanger 26, which is on the low pressure side of the heat pump system 20, are in close proximity to one another. This allows heat to be exchanged between the two tube banks 34, 36. During normal operation, the first tube bank 34 will contain a high pressure refrigerant that is at a higher temperature than the low pressure refrigerant present in the second tube bank 36. The temperature of the high pressure refrigerant may be such that the temperature of the outer surface of the first tube bank 34 may be, for example, 50° C. or greater. The outer surface of the second tube bank 34 will typically be much cooler due to the presence of the relatively cold, low pressure refrigerant therein. As such, heat is transferred from the first tube bank 34 to the second tube bank 36. This transfer of heat may occur through conduction (due to the close proximity of the tube banks 34, 36). Additionally or alternatively, heat may be transferred between the two tube banks 34, 36 via convection or radiation. The first tube bank 34 therefore heats an outer surface of the second tube bank 34, for example such that the temperature of the outer surface of the second tube bank 36 is greater than 0° C. As a result, ice build-up on the outer surface of the second tube bank 34 is limited and/or prevented.

[0070] As will be appreciated, this arrangement offers various advantages over prior heat pump systems. Evaporator heat exchangers typically have multiple tube banks. By connecting one of these tube banks to the high pressure side of the heat pump system, as in the present invention, the heat from the high pressure refrigerant can be effectively utilised to prevent and/or limit ice build-up on the evaporator heat exchanger without the need to introduce an external source of heat. The footprint of the heat pump system also remains the same. That is, the build-up of ice can be limited and/or prevented without the need to increase the footprint of the evaporator heat exchanger. Additionally, it may no longer be necessary to employ a “defrost” mode as excessive ice build-up can be avoided, or at least the frequency at which a “defrost” mode needs to be used can be significantly reduced, as the rate of ice accumulation can be decreased. As a result, the efficiency of the heat pump system can be increased by up to 20% as compared with prior art systems.

[0071] As mentioned above, prior art heat pump systems typically begin to encounter ice build-up when the ambient air temperature is approximately 10° C. or lower. The minimum ambient temperature at which ice will not form on an evaporator heat exchanger is known as the “freeze limit”, and for typical systems this is around 10° C. In the heat pump system 20 shown in FIG. 1, the freeze limit may be reduced significantly as compared to prior art systems. For example, the freeze limit of the heat pump system 20 may be lower than 10° C., and preferably lower than 7° C. The heat pump system 20 may therefore not accumulate ice unless the temperature is lower than 7° C. Further, the rate of ice build-up on the evaporator heat exchanger 26 at a given temperature may also be greatly reduced, even at ambient air temperatures below the freeze limit, as compared to prior art systems, meaning that excessive ice build-up is less likely, and the need for a “defrost” mode is eliminated or reduced.

[0072] The heat pump system 20 may be a reversible heat pump system. In other words, it may be possible to reverse the flow of refrigerant in the heat pump system 20 depending on whether a user wants a cooling mode or heating mode. In FIG. 1, the heat pump system 20 is shown in a heating mode, with the flow of refrigerant through the system being shown by the arrows in the refrigerant flow path. In this mode, the refrigerant in the evaporator heat exchanger 26 extracts heat from ambient air, which is then expelled and provided to the water source in the condenser heat exchanger 24, thereby heating the water. If the flow were reversed, the evaporator heat exchanger 26 would function as a condenser heat exchanger and vice versa, thus causing the system 20 to extract heat from the water (e.g. from inside a building) and expel heat to the ambient air outdoors. Since, in this cooling mode, the evaporator heat exchanger 26 is heated by the refrigerant flowing therethrough, the cooling mode may be utilised as a “defrost” mode when necessary in order to melt ice that has built up on the evaporator heat exchanger 26 during operation of the system in the heating mode. A valve, such as a four-way valve, may be utilised in order to reverse the flow, though this is not shown in FIG. 1.

[0073] Typically, reversible systems require a receiver to be located in the refrigerant flowpath between the condenser heat exchanger and the evaporator heat exchanger. This can be required because the volume of condenser heat exchangers is typically much smaller than the volume of evaporator heat exchangers, which can result in excess liquid refrigerant within the system when the system is switched from the heating mode to the cooling mode, and this is stored in the receiver before (in a defrost mode) it is passed to the condenser heat exchanger. However, the heat pump system 20 may not require a separate receiver, as the first tube bank 34 may function as a receiver to retain excess liquid refrigerant during the cooling mode. As such, the need for a receiver may be eliminated, further simplifying the system and reducing its footprint.

[0074] In FIG. 1, the fan 30 is arranged to cause an air flow A to pass firstly over the second tube bank 36, and then over the first tube bank 34. This may be achieved by positioning the fan 30 adjacent to the second tube bank 36. However, the fan 30 can alternatively be arranged to create a flow of air A that passes firstly over the first tube bank 34 and then over the second tube bank 36. This may be achieved by positioning the fan adjacent to the first tube bank 34, as shown in FIG. 2. The flow of air A may therefore be heated through interaction with the relatively warm, high pressure refrigerant in the first tube bank 34 before being passed over the second tube bank 36, where it may provide heat to the outer surface of the second tube bank 36 and also provide additional heating for the low pressure refrigerant within the second tube bank 36. By first passing the air over the first tube bank in this way, the temperature of the air that is passed over the second tube bank 36 can be raised above the ambient air temperature, and may be raised above the freeze limit of the system. As a result, subsequent cooling of the air through heat exchange with the low pressure refrigerant in the second tube bank 36 may not be sufficient to cause ice to form on the second tube bank 36. Even in instances where ice does still form on the second tube bank 36, the rate of ice formation may be reduced compared to if air at the temperature of the ambient air (i.e. air that had not been warmed) had been passed over the second tube bank 36. The warmed air may also raise the temperature of the second tube bank 36 as it passes over the second tube bank 36, thus helping to limit and/or prevent ice build-up on the second tube bank 36 and also helping to heat the low pressure refrigerant.

[0075] In FIG. 1, the second tube bank 36 is shown as comprising three rows, 36a, 36b, 36c, as will be described in more detail below with reference to FIG. 3. This system 20 is arranged such that, during the heating mode, refrigerant is first passed through row 36a, which is positioned furthest from the first tube bank 34, and is then passed through row 36b and then row 36c in series before exiting the second tube bank 36. It will be appreciated that in this arrangement the refrigerant flows towards the first tube bank 34 as it flows through the rows 36a-c of the second tube bank 36. Hence, the refrigerant flows in generally the same direction as the air flow A over the evaporator heat exchanger 26. This may be termed “co-current” flow. Typically, co-current flow is employed in reversible heat pump systems in order to avoid significant performance loss when the system is operated in a cooling mode. Hence, co-current flow may be used where it is desired to provide a good level of performance when the system is operated in the heating mode as well as when the system is operated in the cooling mode. Co-current flow may be employed more regularly in warmer climates, such as Mediterranean climates, where the ambient air temperature is relatively warm.

[0076] An alternative arrangement is shown in FIG. 2, in which the second tube bank 36 is arranged such that during a heating mode refrigerant is passed through the rows 36a-c in a direction generally opposite to the direction of the air flow A. This may be termed “counter-current” flow. Counter-current flow typically leads to optimised heating of the refrigerant in a heating mode by the air that is passed over the evaporator heat exchanger, and therefore optimised performance of the system 20 in a heating mode. This may be particularly beneficial in cooler climates where the ambient air temperature is low, such as in Nordic climates. Counter-current flow typically provides for a greater degree of heating compared to co-current flow.

[0077] Whilst the system shown in FIG. 1 is arranged for co-current flow, it will be appreciated that it could alternatively be arranged for counter-current flow. Similarly, the system 20 of FIG. 2 may be arranged for co-current flow, rather than counter-current flow.

[0078] The arrangement of the evaporator heat exchanger 26 may provide a certain degree of subcooling to at least a portion of the high pressure refrigerant within the first tube bank 34. For instance, the temperature of the refrigerant may be reduced by up to 20° C. as it is passed through the first tube bank 34. This may be as a result of being cooled through heat exchange with the ambient air and/or the second tube bank 36 (and the refrigerant therein). This subcooling may ensure that a majority, if not all, of the high pressure refrigerant is in liquid form before entering the expansion device 28 and, subsequently, the second tube bank 36 of the evaporator heat exchanger 26. As such, there may be no, or very little, gaseous refrigerant entering the expansion device 28 and/or the second tube bank 36. This may improve the efficiency of the heat pump system 20, as a greater amount of refrigerant may be evaporated in the second tube bank 36 and thus a greater amount of heat may be extracted from the ambient air.

[0079] The arrangement of the evaporator heat exchanger 26 may provide a certain degree of superheat to at least a portion of the low pressure refrigerant within the second tube bank 36. For instance, the temperature of the refrigerant may be increased by up to 20° C. as it is passed through the second tube bank 36. This may be as a result of being warmed through heat exchange with the ambient air and/or the first tube bank 34 (and the refrigerant therein). This may ensure that the entirety, or at least a large proportion, of the low pressure refrigerant is evaporated and is therefore in gaseous form as it leaves the second tube bank 36. This may ensure that there is no, or very little, liquid refrigerant remaining in the refrigerant that is passed to the compressor 22. This may be important to ensure that the compressor 22 continues to operate correctly, as the presence of too much liquid refrigerant within the compressor can create faults and/or could damage the compressor 22. Any drops in efficiency that may result from improper function of the compressor may therefore be avoided.

[0080] Hence, the evaporator heat exchanger 26 of the heat pump system 20 may be considered to be a combined evaporator, subcooler, and superheater.

[0081] FIG. 3 shows a schematic diagram of the evaporator heat exchanger 26 of FIG. 2. The first tube bank 34 is shown on the right of the diagram, and the second tube bank 36 is shown on the left of the diagram. The flow of air A is shown flowing from right to left such that it passes over the first tube bank 34 before it passes over the second tube bank 36, similar to the arrangement shown in FIG. 2. The evaporator heat exchanger 26 may be a round tube plate fin (RTFP) heat exchanger, and optionally a microchannel heat exchanger. The evaporator heat exchanger 26 may therefore comprise a plurality of fins 46, 48, such as plate fins. The plate fins may lie in a plane that is perpendicular to the direction of refrigerant flow through the heat exchanger 26, as shown in FIGS. 3 and 4. The first tube bank 34 may comprise a first plurality of fins 46 and the second tube bank may comprise a second plurality of fins 48. The evaporator heat exchanger 26 may comprise copper tubes and/or aluminium fins.

[0082] High pressure refrigerant output from the condenser heat exchanger 24 enters the first tube bank 34 through a first inlet 38. The high pressure refrigerant then travels through the first tube bank 34 and exchanges heat with the second tube bank 36. The high pressure refrigerant may also exchange heat with the air flow A. In the first tube bank 34, the high pressure refrigerant is cooled as it expels heat. After passing through the length of the first tube bank 34, the high pressure refrigerant exits through a first outlet 40 and is passed to the expansion device 28 (not shown in FIG. 3). Although shown with only one row in FIG. 3, it will be appreciated that the first tube bank 34 may comprise multiple rows, similar to the second tube bank 36 shown in FIG. 3.

[0083] Low pressure refrigerant output from the expansion device 28 enters the second tube bank 36 through a second inlet 42. The second tube bank 36 is shown in FIG. 3 as comprising three rows 36a, 36b, 36c; however, it will be appreciated that the second tube bank 36 may comprise more or fewer rows than this. The low pressure refrigerant enters the first row 36a and then passes sequentially from the first row 36a to the second row 36b and finally to the third row 36c. In FIG. 3, the final row 36c is shown adjacent to the first tube bank 34; however, the first row 36a may instead be adjacent to the first tube bank 34, with the final row 36c being furthest from the first tube bank 34. The low pressure refrigerant then exits the second tube bank 36 via second outlet 44, and is passed to the compressor 22. The rows 36a-c of the second tube bank 36 are heated by the first tube bank 34. As a result, as it travels through the rows 36a-c of the second tube bank 36, the low pressure refrigerant is heated, and at least a portion may be evaporated. The low pressure refrigerant may also be heated via heat exchange with the ambient air, such as air flow A passing over the second tube bank 36. The rows 36a-c may be arranged in parallel with one another and may be arranged in parallel with the rows of the first tube bank 34. As such, a large surface area of each tube bank 34, 36 may be in close contact with one another.

[0084] As will be appreciated, the temperature of the row 36a will be increased through the exchange of heat from the first tube bank 34 the least amount by virtue of it being furthest from the first tube bank 34. However, the outer surface temperature of the first row 36a may still be increased to higher than 0° C. due to the heating caused by the presence of the relatively high temperature first tube bank 34 and the air flow A. In this way, it may be ensured that the build-up of ice is prevented and/or limited over the entire surface of the second tube bank 36.

[0085] The system 20 may be configured such that the row of the first tube bank 34 that is closest to the second tube bank 36 has a surface temperature of at least 50° C. This has been found to provide adequate heating to the second tube bank 36 to prevent and/or limit ice build-up.

[0086] In order to provide for heat exchange between the first and second tube banks 34, 36, they should be in close proximity to each other. As can be seen from FIG. 3, a gap may exist between the first tube bank 34 and the second tube bank 36. This gap may have a size D. Though shown as a gap with constant size in FIG. 3, the gap may vary in size along the length of the tube banks. For example, the gap may have a size of 0 mm along at least a portion of the length of the tube banks, such that the tube banks are in direct contact for at least a portion of their length. Regardless, the gap may have a size D that is no larger than 1 mm (or preferably no larger than 0.5 mm) at any point along the tube banks. As shown in FIG. 3, the gap D may be measured between the extremities of the first plurality of fins 46 and the second plurality of fins 48. This closeness in proximity may aid heat transfer between the two tube banks 34, 36. In some cases, the gap may be sized such that water droplets, e.g. condensed from the ambient air, may bridge the gap between the first and second tube banks 34, 36 to enable heat to be transferred between the first and second tube banks via the water droplet(s).

[0087] FIG. 4 shows an example of a single row of a first and/or second tube bank 34, 36. FIG. 3 may be considered a top view of one of the rows 36a-c shown in FIG. 4.

[0088] As can be seen, the row may comprise one or more tubes 50 arranged in parallel. Two or more of these tubes 50 may be connected to one another in series to form a refrigerant circuit 52. The tubes 50 may be connected to form, for example, a serpentine shaped or a coil shaped refrigerant circuit 52. A single row of the tube bank 34, 36 may comprise one or more refrigerant circuits 52. In the example shown in FIG. 4, the tube bank 34, 36 comprises three pairs of tubes 50, i.e. there are six tubes 50 in total. Each pair of tubes 50 is connected in series to form a serpentine-shaped refrigerant circuit 52. Hence, the row includes three serpentine-shaped circuits 52. Whilst FIG. 4 shows the specific example of six tubes 50 and three serpentine-shaped refrigerant circuits 52, it will be appreciated that any number of tubes 50 and refrigerant circuits 52 is possible, and any shape of refrigerant circuit 52 is possible. A plurality of fins 46, 48 may extend between the tubes 50 in order to aid heat transfer.

[0089] The row shown in FIG. 4 is arranged for use in a tube bank 34, 36 that includes only a single row, and has a refrigerant distributor 54 for receiving refrigerant fed into tube bank 34, 36 via an inlet 38, 42. The refrigerant distributor 54 acts to divide up the flow of refrigerant into different refrigerant streams and separate the refrigerant between each of the refrigerant circuits 52. After having travelled through the tubes 50 of a circuit 52, the refrigerant may then be re-combined in a refrigerant header 56 before it exits through the outlet 40, 44.

[0090] Either or both of the first tube bank 34 or the second tube bank 36 may comprise one or more rows. If a tube bank 34, 36 includes more than one row, the refrigerant may pass through all rows before it is re-combined. In this way, a refrigerant circuit 52 may span across more than one and/or all of the rows in a tube bank 34, 36. It will therefore be appreciated that not each row in the tube bank 34, 36 will require a distributor 54 and/or a header 56. Rather, only the first row (i.e. having the inlet 38, 42) may include a distributor 54 and only the last row (i.e. having the outlet 40, 44) may include a header 56. The use of one or more refrigerant circuits 52 may provide redundancy to the system. The use of the one or more refrigerant circuits 52 may also lead to lower pressure drops in the refrigerant, and increase the thermal efficiency of the tube bank 34, 36. In this way, the heat exchange performance of the evaporator heat exchanger 26 may be optimised.

[0091] FIG. 5 shows a cross schematic view of an evaporator heat exchanger 26 comprising a first tube bank 34 with a single row 34a and a second tube bank 36 with a plurality of, in this case three, rows 36a-c. In this Figure, the layout of the refrigerant circuits 52 can be seen.

[0092] In FIG. 5, the first tube bank 34 includes a first refrigerant distributor 54a into which high pressure refrigerant may enter via the first inlet 38. The first distributor 54a acts to divide the flow of refrigerant and pass it to a plurality of high pressure refrigerant circuits 52a, which each comprise a plurality of, in this case two, high pressure refrigerant tubes 50a fluidly connected in series. After passing through the high pressure refrigerant circuits 52a, the refrigerant from each of the high pressure refrigerant circuits 52a is passed to a first refrigerant header 56a, where the refrigerant is recombined into a single flow before exiting the first tube bank through the first outlet 40. In FIG. 5, the first tube bank 34 comprises a single row 34a comprising twelve refrigerant circuits 52a that each comprise two high pressure refrigerant tubes 50a. However, it is envisioned that the first tube bank 34 may comprise any number of rows 58a with any number of refrigerant circuits 52a.

[0093] The second tube bank 36 includes a second refrigerant distributor 54b in which low pressure refrigerant may enter via the second inlet 42. The second distributor 54b acts to divide the flow of refrigerant and pass it to a plurality of low pressure circuits 52b. As shown in FIG. 5, each of these low pressure refrigerant circuits 52b may span across multiple rows 36a-c of the second tube bank 36. Each row 36a-c comprises multiple low pressure refrigerant tubes 50b. Each low pressure refrigerant circuit 52b may comprise a plurality of these low pressure refrigerant tubes 50b connected in series. FIG. 5 shows a second tube bank 36 that comprises twelve refrigerant circuits 52b that extend across three rows 36a-c. In the illustrated example, each circuit 52b includes two low pressure tubes 50b in the first row 36a, two low pressure tubes 50b in the second row 36b, and two low pressure tubes 50b in the third row 36c. However, it will be appreciated that the second tube bank 36 may comprise any suitable number of refrigerant circuits 52b, comprising any number of tubes 50b and extending over any suitable number of rows 36a-c. After being passed through the low pressure refrigerant circuits 50b, the refrigerant is passed to a second refrigerant header 56b. The recombined low pressure refrigerant may then exit the second tube bank 36 through the second outlet 40.

[0094] The numbers of tubes 50, circuits 52, and rows 60 may be chosen based on a number of factors, including (but not limited to) performance target, heat exchanger footprint, refrigerant properties, and fin density.

[0095] As the two refrigerant flows pass through the first tube bank 34 and the second tube bank 36 respectively, they exchange heat with one another, thus heating the exterior surface of second tube bank 36 such that ice build-up on the exterior surface of the second tube bank 36 is prevented and/or reduced.

[0096] The evaporator heat exchanger 24 of the present invention prevents and/or eliminates ice build-up by utilising heat from the high pressure refrigerant in the first tube bank 34 to maintain and/or increase a temperature of the second tube bank 36. This evaporator heat exchanger 24 can be achieved by simply modifying an existing heat exchanger such that one or more of its rows may be utilised as a first tube bank (for a high pressure refrigerant) and one or more of its rows may be utilised as a second tube bank (for a low pressure refrigerant). This offers advantages in terms of space, complexity, and cost, as well as leading to improvements in efficiency for the heat pump system 20. A simple and efficient heat pump system 20 is therefore provided by the use of such an evaporator heat exchanger 24.