GB2405688A

GB2405688A - Refrigerator

Info

Publication number: GB2405688A
Application number: GB0320856A
Authority: GB
Inventors: Ian David Wood
Original assignee: Applied Design and Engineering Ltd
Current assignee: Applied Design and Engineering Ltd
Priority date: 2003-09-05
Filing date: 2003-09-05
Publication date: 2005-03-09
Also published as: WO2005024314A3; WO2005024314A2; GB0320856D0

Abstract

A refrigerator comprises a refrigeration circuit having a compressor means (4) for receiving refrigerant via a suction line (2), a condenser means (12) for receiving refrigerant from the compressor via a hot gas line, an expansion means for receiving refrigerant from the condenser via a liquid line and an evaporator means for receiving refrigerant from the expansion means and sending refrigerant after evaporation to the compressor means via the suction line (2), wherein the circuit includes a branched portion comprising a plurality of parallel branches each having a respective evaporator (3) of the evaporator means. Each evaporator may be associated with its own capillary tube (1). An accumulator (18) may be heated by a heat exchanger (19) through which the high pressure liquid from the condenser (12) passes. The expansion means may be in the form of a thermostatic expansion valve (5, Fig.4) controlled by a superheat sensor (6, Fig.4). A liquid receiver (10, Fig.4) may be located in the circuit upstream of the expansion means. A hot gas feed line (21, Fig.14) positioned immediately downstream from the compressor may pass hot gas to a hot gas solenoid valve (23, Fig.14) on each branch upstream of the respective evaporator for defrosting purposes.

Description

1 2405688 lMPllOVEMENTS IN OR RF,LATING TO REFR1(,ERATION This invention

relates to the art of refrigeration. In pret'erred embodiment.), the invention relates to refrigerators, freezers, combined refrigerator/freezers, cold storage compartments cont'igurable as both refrigerators and freezers. coolers and air conditioners, all t'or domestic or commercial applications including fixed and mobile appliances and installations. For brevity, all such appliances and installations will he rel'erred to herein collectively as refrigerators unless the context demands otherwise.

Whilst the invention primarily relates to appliances and installations for cold storage and cooling, it does not exclude such appliances or installations that have the additional ability to heat to above-ambient temperatures, for example in defrosting.

The invention finds particular benefit in the context of the Applicant's multi-compartment cold storage appliances disclosed in its co-pending patent applications WO 01/20237, WO 02/()731()4, WO 02/()73105 and WO 02/073107. The compartments ol'those appliances are drawers sealed from one another to minimise cross- contamination, waste of energy and icing. Optionally, there is provision to select dil'fcrent temperatures in different compartments to suit clift'ercnt foodstuffs or other contents and to suit different cold storage regimes such as refrigeration marginally above zero Celsius and freezing signil'icantly below zero C'clsius. Indeed, it is possible for a given compartment to be converted readily from rct'rigerator to freezer and back again, thereby to vary the proportion of refrigerator space to freezer space in the appliance as a whole. In this way, the appliance can respond to changing cold-storage needs.

The basic elements of a typical refrigerator arc a compressor, a condenser, a metering or expansion device and an evaporator, connected in that order in a circuit through which refrigerant cycles in use. The compressor compresses gaseous refrigerant that enters the compressor at low pressure via a suction or return line. The high-pressure hot gaseous rcl'rigerant emanating from the compressor via a hot gas line flows through the condenser wl1ere it cools to a liquid also at high pressure, the condenser most commonly rejecting heat to atmosphere. Else cool highpressure liquid refrigerant emanating from the condenser via a liquid line is forced through the metering or expansion device to reduce its pressure so that its boiling point drops to a level suitable for cooling. 'I'he effect of the

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metering or expansion device its to maintain thc necessary pressure difference between thc conclenser and evaporator. The cool k'N;-prcssure liquid refrigerant emanating from the metering or expansion device flows through the evaporator where it evaporates in a low- pressurc environment to draw heat from a storage compartment cooled by the evaporator.

Finally, the low-pressure gaseous rel'rigerant emanating from the evaporator is drawn back into the compressor via the suction line to start the cycle again.

Appendix I hereto is a sheet of symbols used in the diagrams of this specification, accompanied by notes where appropriate. Anions other information, these notes explain that there are many variations of the basic elements described above, such as plate-type evaporators and forced air fan-coil type evaporators.

To put the invention into context, there now follows a brief description of prior art known in the cold storage field. To aid understanding of that prior art, reference will now be made to Figures I to 3 of the drawings which are diagrams of prior art refrigerator circuit arrangements.

Figures I and 2 of the drawings show that capillary tubes 1 (although there is no capillary effect in the true sense) or other fixed orifices are commonly employed as metering or expansion devices in unitary appliances that have a single cold storage compartment, albeit a compartment that may be partitioned by shelves, drawers or the lilac. 'l'he theory behind the operation ol' metering or expansion devices is well known essentially causing the required pressure drop due to the restrictive elect of the narrow aperture and/or the frictional efl'ect of the length of narrow tube which may be coiled as shown for compactness.

A capillary tube passes liquid much more readily than vapour due to the increased friction with the vapour; as a result, it is a practical metering device. When a capillary tube is sized to permit the desired flow of refrigerant, the liquid seals its inlet. If the system becomes unbalanced, some vapour (uncondcosed refrigerant) enters the capillary tube. This vapour reduces the mass flow of refrigerant considerably, which increases condenser pressure and causes sub-cooling at the condenser exit and capillary tube inlet. The result is an increase of the mass flow of refrigerant through the capillary tube. If properly sized for the application, the capillary tube compensates automatically for load and system variations and gives acceptable performance over a range of operating conditions. common flow condition is to have suh-coolcd liquid at the entrance to the capillary tube (see Figure 2).

Thus, the capillary tube controls the refrigeration system by allowing either liquid refrigerant in case of high evaporator duty or gaseous refrigerant. in case ol' IONV evaporator duty, to enter the inlet to the tube Most simple domestic refrigerators, freezers and small air conditioning units employ a single capillary tube 1 as shown in Figures l told 2. 'I'hese systems are Charge sensitive' to the amount ol'refrigerant hi the system and operate Within a narrow band of evaporating and condensing temperatures, at small flow rates. Consequently, a system configured for a refrigerator cannot easily be turned into a freezer without components and the refrigerant charge being modified. In other words, capillary systems do not naturally lend themselves to multi-compartment cold storage involving widely-variable temperature (refrigerator-to freezer) under a wide range of load and ambient operating conditions.

The sole diff'ercnce between Figures l and 2 is that Figure 2 has the added refinement of heat transfer from the liquid line at the capillary tube I to the suction line 2 leaving the evaporator 3 en route to the compressor 4. This suction liquid heat exchanger increases the capacity of the refrigeration system by transferring heat *om the liquid in the capillary tube 1 to the suction vapour in the suction line 2 returning to the compressor 4.

Consequently, the enthalpy of the refrigerant entering the evaporator 3 is reduced, increasing its refrigerant effect. Ignoring some small losses, the enthalpy increase in the vapour refrigerant in the suction line 2 is equal to the enthalpy decrease of the liquid refrigermt entering the evaporator 3.

I'he suction liquid heat exchanger shown in Figure 2 has a major efl'ect on the operation of the capillary tube 1. If the heat exchange takes place before vapour starts to form, it sub- cools the liquid refrigerant. This greatly increases the mass flow ol'refrigcrant through the capillary tube 1. If the heat exchange takes place after vapour has started to form, the liquid temperature will have decreased. The temperature dit'fcrence between the liquid refrigerant and the ref'rigera1lt vapour in the suction line 2 will therefore be less and less heat will be transferred.

Generally, the liquid is sub-cooled to prevent bubbles forming when it is expanding (thus hlcreasing flo\N rate) and vapour in the suction line 2 is heated or even superheated. with the et'fect that the overall system capacity is increased.

Figure 3 introduces a thenostatic expansion valve 5 (REV) as a metering device in place of the capillary tube of Figures I and 2. The TEV 5 provides a variable orifice that regulates refrigerant flow based upon the degree of superheat in vapour leaving the evaporator 3. The degree ol' superheat is the extent to which vapour temperature is above the saturation temperature determined by pressure in the evaporator 3, and is detected via a I () phial or bulb sensor 6 associated with the T12V 5 that is located on the suction line return to the compressor to feed hack actuating pressure to the TEV 5 via a pipe 7.

IJnder optimal conditions, a TEV S ensures that liquid refrigerant does not return to the compressor 4, which liquid could otherwise cause hydraulic damage to the compressor 4.

The TEV 5 also allows for greater fluctuations in demand than would a capillary tube 1, provided that those fluctuations are not too rapid. I, arger systems, or systems with more than one evaporator 3, commonly employ a liquid receiver after the condenser to provide a reservoir of refrigerant to meet changing fluctuations in demand. In general, systems employing 'I'F.Vs are not as charge-sensitive (i.e. to the amount of refrigerant in the system) as capillary systems.

I'he present invention resides in the concept of a refrigerator as herein defined, comprising a refrigerant circuit having a compressor means for receiving refrigerant via a suction line, a condenser means l-or receiving refrigerant from the compressor via a hot gas line, an expansion means for receiving refrigerant from the condenser via a liquid line and an evaporator means for receiving refrigerant from the expansion means and sending refrigerant after evaporation to the compressor means via the suction line, wherein the circuit includes a branched portion comprising a plurality of parallel branches each having a respective evaporator of the evaporator means. The evaporators cater for independent cooling of the compartments in the Applicant's multi-compartment cold storage appliances as disclosed in its co-pending patent applications W() 01/2()237, WO 02/073104, WO 02/073 l 05 and WO 02/073107.

In some arrangements of the invention, the expansion means comprises a thermostatic expansion valve in each branch situated upstream ot' the evaporator of' that branch, a superheat sensor associated with the thermostatic expansion valve being situated downstream of that evaporator. Preferahly the superheat sensor is also on the branch associated with the associated thermostatic expansion valve.

The invention also contemplates arrangements in which the expansion means comprises a thermostatic expansion valve situated upstream of the branched portion, a superheat sensor associated with the thermostatic expansion valve being situated downstream of the branch portion. '['hat thermostatic expansion valve may be substantially solely responsible for expansion ot' the refrigerant before the refrigerant encounters the evaporator means.

In inexpensive variants of the invention' the expansion means comprises a capillary in each branch upstream of the evaporator of that branch and there may also be means t'or heat exchange between the capillary and the suction line.

I'he refrigerator of the invention preferably comprises a cooling control valve in each branch situated upstream ol' the evaporator ol' that brancll. 'I'hat cooling control valve may be an on/ol'f valve that is cycled in use to control cooling by the evaporator of that branch.

It is also, or alternatively' possible to control evaporator cooling by a cooling control fan means that acts upon the evaporators. The cooling control fan may be varied in speed and/or cycled to control evaporator cooling in use.

I'he refrigerator of the invention may further include an accumulator downstream of the evaporator means to receive refrigerant from the evaporator means and t'rom which the compressor draws refrigerant vapour. 'I'hat accumulator preferably includes means for heat exchange with the liquid line downstream of the condenser. In that case, the superheat sensor associated with the optional thermostatic expansion valve may be situated downstream of the accumulator.

In some arrangements ot' the invention, a liquid receiver may be situated downstream of the condenser means to receive refrigerant from the condenser means in a reservoir from which refrigerant passes to the evaporator means. In that case, the optional thermostatic expansion valve may be situated downstream of the condenser means and the liquid receiver. It is also possible for the suction line to include the liquid receiver, refrigerant being drawls by the compressor means in use from a vapour cavity above liquid in the liquid receiver.

Advantageously, the refrigerator of the invention further comprises head pressure control means associated with the condenser. 'I'hat head pressure control means may comprise one or more Cans acting on the condenser, in which case the or each fan operates cyclically or at variable speed to control head pressure. Alternatively, the head pressure control means may comprise a pressure regulating means in the liquid line. 'I'he pressure regulating means may, for example, be an automatic valve or a means for switching between dif'f'erently- sizcd fixed orifices in the liquid line.

The refrigerator of the invention may also include a circulation pump to impel refrigerant through the evaporator means.

Suction pressure control means may be provided, responsive to suction pressure control logic. For example, in response to the suction pressure control logic, the suction pressure control means advantageously selects an evaporating pressure/temperature appropriate for the evaporator with the lowest set temperature among the evaporators. More generally, the suction pressure control logic suitably takes input from a look-up table recording absolute pressure, evaporator temperature and bar gauge pressure appropriate to refrigeration temperature levels to be achieved by an evaporator.

For defrost purposes, any ol'the embodiments of the invention may further comprise a hot gas feed taken from the hot gas line downstream of' the compressor means to supply hot gas to the evaporator means. In that case, the hot gas feed pref'crably joins the refrigerant circuit at a junction upstream of' the evaporator means, and there may be a hot gas control valve upstream of' the junction. 'he hot gas feed suitably branches to John each of the branches of the branched portion at a respective junction, in which case each branch of the hot gas teed preferably has a respective hot gas control valve upstream of the junction.

I-lowever, it is possible to have a single hot gas control valve situated upstream of where the hot gas feed branches, in which case a further control valve is preferably provided in the refrigerant circuit downstream of the or each junction between the hot gas feed and the refrigerant circuit.

To aid understanding of the invention, reference will no\v be made by way of example to Figures 4 to, 16 of the drawings which are diagrams of refrigerator circuit arrangements according to the invention, and to l'igure 17 which is an example of a look-up table for use by control logic in the invention.

I;igure 4 shows a way of adapting the ahovementioned prior art to suit the Applicant's i preferred multiple-compartmeut variahle-temperature storage system. Like numerals are used t'or like parts. 1-lere, a fourcompartment arrangement is illustrated but any practical number of compartments is possible. Each compartment is cooled by a respective evaporator 3 on respective parallel branches 8 of the circuit. In this circuit, any compartment can be used as a refrigerator or as a freezer by virtue of mass control achieved by cycling a respective solenoid shutol'l' valve 9 serving each evaporator 3. Each branch of the circuit is served by a respective TEV 5 whose superheat sensor 6 is ] 5 downstream of' the evaporator 3 of that branch 8. ()therwise, the circuit is much the same as the basic circuits illustrated in l''igures I to 3, apart prom the routine additions of a high- pressure liquid receiver 1 () as a reservoir in the licluid line I I downstream of the condenser 12 to ensure continuous refligeranl supply, a Alter drier 13 in the liquid line downstream of the liquid receiver 10 to maintain refrigerant quality, and a sight glass 14 in the liquid line downstream of the filter drier 13 to monitor the refrigerant condition. The compressor 4 and condenser 12 are referred to in the singular for brevity but, in all embodiments, they can be plural and/or variable speed to meet variable load and duty requirements in use.

Theoretically, the system shown in figure 4 should work and, indeed, would work if provided with appropriate TEVs 5. However, eommereially- available TEVs are rated at I too high a duty for the Applicant's purposes. The smallest TEVs are designed to handle a load in excess of 4() ()W whereas the Applieants appliance requires about 1 50W per TEV.

As a result, commercially-available TEVs struggle to control rel'rigerant flow: rather than superheated vapour resuming to the compressor 4 as intended, slugs of liquid may be left in the sueticn line 2 by the '1'EVs 5, which risk compressor damage and l:ailure if they are retuned to the compressor 2. Also. it is diff;eult to maintain a low pressure in the suction line 2 / evaporator 3. This pressure is crucial because it determines how cold the rei'rigerant is when it evaporates, and hence the temperature to which the evaporator 3 can cool the associated compartment. Another drawback is cost: TEVs are expensive, meaning that in a fourcompartment appliance the overall cost of TEVs alone \Nould be in excess of G13I' (about US$320 or ('290) The system shown hi Figure 5 introduces a lovv- pressure accumulator 15 in the suction line 2 between the evaporators 3 and the compressor 4 but is otherwise identical to that shown in Figure 4. 'I'he low-pressure suction line accumulator 15 traps any slugs of liquid emanating from the evaporators 3 that might otherwise be returned to the compressor 4.

I'he accumulator 15 holds any such liquid at the bottom of a pressure Jesse] and allows the compressor 4 only to draw ot'l'vapour above the liquid within the vessel. That vapour is dravv, into the compressor 4 via a 1J-shaped compressor return line 16 that is partially immersed in the liquid within the pressure vessel and has an orifice 17 in the lowest part of the [J to allow a metered flow of oil back to the compressor. 'I'his oil typically emanates from within the compressor casing and is entrained in use as liquid or vapour by the refrigerant passing through the compressor.

Whilst the low-pressure suction line accumulator 15 ol'F'igure 5 ensures that liquid is not returned to the compressor 4, 'l'EVs 5 are still needed for control under most circumstances. This involves substantial expense as outlined above, and the Applicant is aware of Only one commercially-available TEV that is suitable for its needs.

The arrangement shown in Figure 6 is an attempt to reduce the cost problems of the Figure 4 and l igure 5 arrangements by using one 'I'LV 5 to serve all four evaporators 3. To this end, the feedback line 7 of'the TEV 5 bridges the branched part ol'the circuit with the TEV itself being situated upstream of the branch, between the flter/drier 13 and the sight glass 14, and the superheat sensor 6 associated with the TEV 5 being downstream of the branch.

Again, this arrangement should work in theory but the Applicant has encountered significant problems in testing. I 'or example, distribution of liquid refrigerant to the evaporators 3 has been very uneven in that some evaporators 3 receive a lot of refrigerant and others do not, even when all compartments are in equal use for cooling. On partial load, that is with only one evaporator 3 cooling or with fewer than all of the evaporators 3 cooling, the conmercially-available TEVs 5 encounter the same over-capacity problems described earlier, meaning that slugs of liquid can be returned to the compressor 4. A further problem with this system is that downstream of the TEV 5, liquid is distributed to each evaporator 3 at very low temperature requiring aclcltional insulation to prevent condensation and icing on all pipes and inline components. Further inet'ficiencies are caused by the solenoid valves 9 being at low temperature, thereby rejecting heat k' the refit igerant rather than to ambient air. s

The arrangement of l'igure 7 is essentially a hybrid of the arrangements of Figures 5 and 6, in that a low-pressure accumulator 15 is placed in the suction line 2 between the evaporators 3 and the compressor 4, downstream of the superheat sensor 6 associated with the TELV 5. Again, this solves the problem of liquid returning to the compressor 4.

1() However, the distribution ol' liquid refrigerant is still poor, involving oversupply of refrigerant to some evaporators 3 while starving others. Low temperature distribution problems also arise as aforesaid.

I'hus, the arrangements of Figures 6 and 7 demonstrate the disadvantages of a single 'I'EV 5: whilst saving costs a single'I'EV 5 delivers poor performance and control. Conversely, the arTangcrnents of l:'igures 4 and 5 provide better temperature control by using a 'I'EV 5 for each evaporator 3, but at considerable extra cost.

The arrangement shown in Figure 8 replaces the Tl-,Vs 5 of the Figure 4 arrangement with capillary tubes I to provide individual expansion to each evaporator 3 at a fraction of the cost of the same number of individual Tl'Vs 5. As shown, the system in Figure 8 is completely uncontrolled and risks poor refrigeration conditions and the passage of liquid slugs to the compressor 4.

Figure 8a shows a modification of Figure 8 providing for heat transfer from the liquid line at each capillary tube I to the suction line 2 leaving the evaporator 3 en route to the compressor 4. This improves efficiency and adds heat to the ret'rigerant in the suction line 2 returning to the compressor 4.

'T'hc arrangement of'Figure 9 replaces the high-pressure liquid receiver 10 upstream of the evaporators 3 of Figure 8 with a low-pressure suction line accumulator 15 downstream of the evaporators akin k' that shown in Figures 5 and 7 above. In this system, the capillary tubes 1 are sized such that the evaporator 3 provides some superheat to the refrigerant under most operating conditions l'or the evaporator 3.

Figure 9a shows a hybrid of the arrangements of Figure 8a and Figure 9. As in Figure 8a, there is provision for heat transltr prom the liquid line at each capillary tube 1 to the suction line 2. As in Figure 9, there is a low-pressure suction line accumulator 1 S downstream of the evaporators 3.

Moving on now to l igures 10 and 11, these diagrams show preferred embodiments of the invention that provide inexpensive expansion devices t'or a multiple-compartment cold storage appliance in which each of the compartments may be set at temperatures in the lO range between ambient and substantially below zero Celsius. 130th embodiments are similar to' Figure <3, Figure I 1 the more so because, lilac Figure 9, it has solenoid shut-off valves 9 for each evaporator 3 to Provide capacity control, whereby temperature within each compartment is maintained by cyclically opening and closing the appropriate solenoid. In contrast, Figure 10 omits the solenoid shut-ol'f valves 9 in favour of cycling a I'an (not shown) associated with each evaporator 3 to maintain temperature within each compartment.

I'he arrangements of Figures 10 and 11 both employ a modified lowpressure accumulator 18 in the suction line 2 downstream ot'the evaporators 3. 'I'he accumulator 18 is modified by the provision of a liquid/suction heat exchanger 19 in which the high pressure liquid line I I from the condenser 12 passes through the vessel to boil the accumulated refrigerant liquid and to provide superheat to the gaseous refrigerant returned to the compressor 4 via the suction line 2. 'I'his allows each evaporator 3 to be fully flooded with liquid refrigerant (i.e. liquid in and liquid out) giving greater cooling efficiency. Normally, the capillary tube 1 would be sized so that the refrigerant would boil ofl'completely by the time it reaches the downstream end of the evaporator 3, the evaporator 3 often also adding superheat before the vapour leaves the evaporator 3 and enters the suction line 2 leading to the compressor 4. This protects the compressor 4 from slugs of liquid refrigerant. I lowever, this precaution is not necessary in the systems of lF'igures 10 and 1 I because any excess liquid flows harmlessly to the suction accumulator 18, where it is boiled ol'f by heat exchange from the liquid line 11. Thus, the evaporators 3 can he run fully flooded and so at optimum efficiency.

A capillary tube 1 has further practical advantages in addition to its low cost in comparison with a TEV 5. For example, the small-bore capillary tube 1 can be run to the evaporator 3 by hand without the need for skilled pipe fabrication. Also, very low temperatures are first encountered only at the downstream end portion of the capillary tube I adjacent to the evaporator 3, representing perhaps 1()% to 20% of the capillary length. which makes insulation requirements less onerous. A TEV 5 and its associated phial/bulb superheat sensor 6 are much more dif'i'icuit to insulate.

further addition to the arrangement of Figures] () and I 1 is head pressure control 20 associated with the condenser 12. This is because the condensing temperature/pressure normally varies with ambient temperature due to the At over the heat rejection exchanger.

Specifically, low ambient temperature implies low head pressure and high ambient temperature implies high head pressure. 'I'he capillary tube 1 requires a more constant AP so that it regulates the same flow of'refrigerant to the evaporator 3 under all ambient I 5 conditions. Consequently, head pressure control 20 is pret'erahly employed to stabilise the AP over the capillary tube 1. 'I'his may be via multiple fans cycling in combination on the condenser 12 or with variable-speed fans on the condenser 12. Alternatively, a pressure regulating valve may be employed in the liquid line 11 or means may be provided for switching between a series of differently-sized fixed orifices in the liquid line 11.

Figures 12 and 13 show arrangements akin to those of Figures 10 and 11, in that a low- pressure accumulator 18 in the suction line 2 downstream of the evaporators 3 includes a liquid/suction heat exchanger 19 in which heat prom the high pressure liquid line 11 boils accumulated refrigerant liquid and superheats the gaseous refrigerant returned to the compressor 4 via the suction line 2. It is expected that in the arrangements of Figures 12 and 13, head pressure control 20 would not be required as the automatically operated TEV should compensate for variations in temperature.

Figure 13 is identical to Figure 11 save for the provision of a single TEV 5 upstream of the 3() evaporators 3 whose feedback line 7 extends from a superheat sensor 6 on the suction line 2 downstream of' the low-pressure suction line accumulator 18. Thus, superheat is measured by the 'I'EV 5 on the suction line 2 to the compressor 4 having picked up heat t'rom the liquid/suction heat exchanger 18. lL'igure 13 therefore introduces the concept of two-stage expansion where some of the expansion is carried out by the fixed capillary tubes 1, but the '1'L2V S provides automatic (even if crude) control of refrigerant expansion.

This arrangement its proposed to improve liquid distribution and overcome variations in ambient temperature, negating head control.

The arrangement of Figure 12 has the same REV arrangement as shown in Figure 13 but in this instance, the capillary tubes] are omitted entirely so that the TEV 5 is fully responsible for the pressure drop frown highpressure liquid leaving the condenser 12 to low pressure liquid entering each evaporator 3. 'I'hus, the arrangement of Figure 12 replaces the capillary tubes 1 with a common REV 5. The system still allows for the evaporators 3 to be flooded with refrigerant, even where only one evaporator 3 is being used for cooling: the oversized TEV 5 can still provide crude control without liquid returning to the compressor 4 and compensate for variations in ambient temperature.

The arrangements of 1 igures 12 and 13 are described to demonstrateexamples of alternatives that could be developed from the basic concepts shown in Figures 10 and 11.

Turning now to Figures 14 and 15, these detail hot-gas defrost systems for a multiple- compartment cold storage appliance or installation.

I'he arrangement in Figure 14 is as for Figure 13 apart from omission of the TEV 5 and provision of a hot gas feed 21 *tom the hot gas line 22 immediately downstream of the compressor 4. The hot gas feed 21 branches via a hot gas solenoid valve 23 on each branch to supply hot gas to a junction 24 with the main circuit immediately upstream of the evaporators 3.

I'he hot gas solenoid valve 23 to each evaporator 3 in Figure 14 is normally closed. On defrost, that valve 23 is opened and the corresponding liquid refrigerant solenoid valve 9 is closed to introduce hot gas to the associated evaporator 3 and melt accumulated ice. Each evaporator 3 can be defrosted separately by virtue of independently controllable valves 23 on each branch. ITowevcr, it is not essential to have an individual hot gas solenoid valve 23 on each branch ot' the hot gas feed 21. As Figure 15 shows, it is possible to have a common hot gas solenoid valve 23 upstream of the branch in the hot gas l'eed 21. In this case' each branch of the main circuit has a liquid refrigerant solenoid valve 9 downstream of its junction 24 with the associated branch of the hot gas feed 21. When the common hot gas solenoid valve 23 in the hot gas feed 21 is open, the liquid refrigerant solenoid valve(s) 9 associated with the evaporator(s) 3 to be defrosted must be open and the other liquid refrigerant solenoid valves 9 must be closed.

It will be apparent to those skilled in the art that the defrost systems of Figures]4 and 15 may be applied to various arrangements described herein and derivatives thereof, in addition to the above-described systems that are applied to arrangements akin to that of Figure 13. 1()

Rel'erring now to l'igure l 6, in this arrangement a low-pressure receiver 25 receives cold liquid refrigerant l'rom the liquid line I I downstream ot' the condenser 12 and outputs that liquid ret'rigerant to the evaporators 3 via a filter drier 13, a circulation pump 26 and a sight glass 14. Vapour refrigerant Lowing from the evaporators 3 is returned to the receiver 25 ] 5 above the liquid level therein and is drawn ol'f from above the liquid level into the suction line 2 that feeds the compressor 4. Although not shown in Figure 16, it would of course he possible to employ a U-shaped compressor return line 16 as first described above in relation to Figure 5, in which the lowest part of the U immersed in the liquid within the receiver has an orifice 17 to allow a metered flow back to the compressor 4 of oil entrained 2() by the rel'rigerant.

Whilst still a vapour, the ref'rigcrant being drawn ol'f from the receiver 25 through the suction line 2 has been exposed to the cold liquid refrigerant beneath and so will be somewhat colder than the vapour entering the receiver from the evaporators 3. By the same token, the liquid refrigerant leaving the receiver 25 en route to the evaporators 3 will be somewhat warmer than the liquid refrigerant flowing in to the receiver 25 from the condenser 12 via the liquid line 11. This is by virtue of heat exchange with the warmer vapour within the receiver 25 above the body of liquid refrigerant.

A TE:V 5 is situated in the liquid line I 1 between the condenser 12 and the low-pressure receiver 25 and its superheat sensor 6 is situated on the suction line 2 upstream of the compressor 4. Thus, the 'I'EV 5 operates in a simple circuit with the compressor 4, condenser] 2 and receiver 25. The low-pressure receiver 25 acts as a reservoir to damp out sudden fluctuations in refrigerant floss, thus enabling more stable operation of the REV 5.

The circulation pump 26 simply impels cold liquid refrigerant through any evaporator 3 whose associate<:! solenoid valve 9 is open.

In the Figure 16 arrangement, problems may occur in balancing the flow of liquid refrigerant between branches and evaporators 3, especially if the aggregate flow rate is not sut'ficicnt to over-i:ted all four evaporator coils with reL'rigcrant. Problems may also occur with the compressor suction temperatures being so cold. In this respect, the low-pressure receiver 25 doubles as an accumulator for any liquid refrigerant that may return from the evaporators 3, and so will prevent liquid entering the suction line 2 to the compressor 4, provided that the receiver 95 is large enough. However, the suction line temperatures will still be very cold, risking condensation and frosting on the suction line 2 unless good insulation is used.

Another challenge with the arrangement of Figure 16 is control of the 'I'EV 5 at low or partial duties to maintain a low enough suction pressure for low refrigerant temperature in the evaporators 3.

In general, requirements such as much thicker insulation on most pipe work, and solenoid 9() valves 9 and a pump 26 suitable for use at low temperatures, would probably make the arrangement of Figure 16 expensive to produce. IIowever, it has been identified as a possible way of coping with very variable heat loads whilst maintaining accurate temperature control.

Mention has already been made of variations that are possible within the inventive concept.

Many other variations are possible. look example, using the hot gas defrost system set out in Figures 14 and 15 above, it would be possible to use one or more of the refrigerator compartments to rapid-thaw product. As each compartment of the Applicant's system as outlined in its copending patent applications WO 01/20237, WO 02/073104, WO 02/073105 and WO 02/073107 has its own insulation and temperature control, it is possible to use the compartments for processes other than chilling or freezing. Thus, while one compartment is freezing product, t'or example, another compartment could be used for the rapid thawing of product.

During a defrost cycle, it is envisaged that only one hot gas solenoid valve need be opened at any one time. I<or the duration of' the defrost cycle, all of the liquid rei'rigerant solenoid valNes can remain closed to prevent a rise in suction pressure, resulting from the hot gas solenoid valve opening, from adversely ai'i'ecting the other drawer temperatures.

Alternatively, in order to use hot gas continuously through one or more of the evaporators whilst one or more of' the other evaporators are being used for cooling, the hot gas could be supplied at a lower pressure.

I'o ensure that any product being rapidly thawed is only thawed rather than cooked, hot gas may only need to be pulsed periodically into the appropriate evaporator. Meanwhile, evaporator air circulation lans can run continuously in all compartments. It would therefore he possible to close all ot' the liquid rei-'rigerant solenoid valves for the pulse duration of hot gas solenoid valve opening without greatly affecting the product storage temperatures experienced in the chilled and i-'rozen compartments.

Solenoid valve switching could be reduced by fitting non-return valves to the downstream exit of each evaporator before the branches converge into the common suction line.

Another possibility is a blast chill facility, in which case alternative evaporator air circulation fans would be required in order to obtain the increased air flow-rate required.

Either multi-speed or supplementary fans could be used to provide the increased air flovv-- rate as and when it was required.

I'he low-pressure accumulator system suggested for certain arrangements above allows the evaporator coils to run not only fully-flooded, but actually to be overfed with liquid under normal operation without fear of liquid returning to the compressor. This over-capacity should accommodate a blast chill l'acility by maintaining the evaporating temperature despite the large heat load made possible by higher air flow-rates over the evaporator coil.

Turning i'inally to Figure 17 of the drawings, this table shows how suction pressure control logic selects an evaporating pressure/temperature appropriate for the compartment with the lowest set point. 'I'hat is, the control logic will maintain a higher suction pressure/temperature if all drawers are set above zero than if they are all set below zero.

On its left sicle, the table in Figure 17 shows the relationship between absolute pressure, temperature and bar gauge. On its right side, the table shows typical temperature/pressure settings that arc provided in a look-up table in the control logic.

Claims

1. A refrigerator as hereinbeforc defined. the refrigerator comprising a refrigerant circuit having a compressor means for receiving refrigerant via a suction line, a condenser means for receiving refrigerant front the compressor via a hot gas line an expansion means for receiving refrigerant from the condenser via a liquid line and an evaporator means for receiving refrigerant Prom the expansion means and sencling refrigerant after evaporation to the compressor means via the suction line, wherein the circuit includes a branched portion comprising a plurality of' parallel branches each having a respective evaporator of the 1() evaporator means.

2. 'I'he refrigerator of Claim 1, wherein the expansion means comprises a thermostatic expansion valve in each branch situated upstream of the evaporator of that branch, a superheat sensor associated with the thermostatic expansion valve being situated downstream of that evaporator.

3. 'I'he refrigerator of Claim 2, wherein the superheat sensor its on the branch of the associated thermostatic expansion valve.

4. The rehigerator of' any preceding Claim, wherein the expansion means comprises a thermostatic expansion valve situated upstream of the branched portion' a superheat sensor associated with the thermostatic expansion valve being situated downstream of the branch portion.

5. The ret'rigerator of'Claim 4, wherein said thermostatic expansion valve is substantially solely responsible Nor expansion ol' the refrigerant before the refrigerant encounters the evaporator means.

6. The refrigerator of any preceding Claim wherein the expansion means comprises a capillary in each branch upstream of the evaporator of that branch.

7. The refrigerator of Claim 6, further comprising means for heat exchange between the capillary and the suction line. ls

8. The refrigerator of any preceding Claim, comprising a cooling control valve in each branch situated upstream of the evaporator of that branch.

9. 'I'he refrigerator of Claim 8, wherein the cooling control valve is an on/off valve that is cycled in use to control cooling bV the evaporator of that hrancll.

10. The refrigerator of any preceding ('laim, comprising cooling control fan means acting upon the evaporators.

1 1. The refrigerator of Claim 10, wherein the cooling control fan is varied in speed and/or cycled to control evaporator cooling in use.

12. The refrigerator of any preceding Claim, comprising an accumulator downstream of the evaporator means to receive retigerant from the evaporator means and loom which the compressor draws refrigerant vapour.

13. The refrigerator of Claim 12, wherein the accumulator includes means for heat exchange with the liquid line downstream oi'the condenser.

2()

14. 'l'he rehigerator of Claim 13 when appendant to Claim 4, wherein the superheat sensor associated with the thermostatic expansion valve is situated downstream of the accumulator.

15. The refrigerator of any preceding Claim, comprising a liquid receiver downstream of the condenser means to receive refrigerant from the condenser means in a reservoir from which refrigerant passes to the evaporator means.

16. The refrigerator of Claim 15 when appendant to Claim 4' wherein the thermostatic expansion valve is situated downstream of the condenser means and upstream of the liquid receiver.

17. 'l'he refrigerator of Claim 15 or Claim 16, wherein the suction line includes the liquid receiver, rei'rigerant being drawn by the compressor means in use from a vapour cavity above liquid in the liquid receiver.

18. The refrigerator of any preceding Claim, further comprising head pressure control means associated with the condenser.

19. 'I'he refrigerator of Claim 18, wherein the head,orcssure control means comprises oTle or snore facts acting on the condense'.

2(). 'I'he refrigerator ot' Claim 19 wherein the or each fan operates cyclically or at variable speed to control head pressure.

21. The refrigerator of' Claim 18, wherein the head pressure control means comprises a pressure regulating means in the liquid line.

22. 'I'he refrigerator ol' Claim 21, wherein the pressure regulating means is a valve or a means for switching between differently-sized fixed orifices in the liquid line.

23. The ret'rigerator of any preceding Claim, further comprising a circulation pump to impel refrigerant through the evaporator means.

24. 'I'he refrigerator of any preceding Claim, lilrther comprising suction pressure control means responsive to suction pressure control logic.

25. 'I'he refrigerator of Claim 24 wherein in response to the suction pressure control logic, the suction pressure control means selects an evaporating pressure/temperature appropriate for the evaporator with the lowest set temperature among the evaporators.

26. The refrigerator of Claim 24 or Claim 25, wherein the suction pressure control logic takes input from a lool<-up table recording absolute pressure, evaporator temperature and bar gauge pressure appropriate to rel'rigeration temperature levels to be achieved by an evaporator.

27. The refrigerator of any preceding Claim, further comprising a hot gas liked taken from the hot gas title downstream of the compressor means to supply hot gas to the evaporator mealls.

28. The refrigerator of Claim 37, wherein the hot gas feed joins the refrigerant circuit at a junction upstream of the evaporator means.

29. 'I'he refrigerator of Claim 28, wherein the hot gas feed has a hot gas control valve upstream of the j unction.

30. The refrigerator of Cluing 2X or Claim 29, wherein the hot gas feed branches to join each of the branches of the branched portion at a respective junction. to

31. The refrigerator of Claim 30 when appendant to Claim 29, wherein each branch of the hot gas feed has a respective hot gas control valve upstream of the junction.

32. The rel'rigerator of Claim 30 when appendant to Claim 29, wherein a single hot gas control valve is upstream of where the hot gas toed branches.

33. The refrigerator of Claim 32. wherein a farther control valve is in the refrigerant circuit downstream of the or each junction between the hot gas feed and the refrigerant circuit.

34. A refrigerator as hereinbefore defined, including a refrigerant circuit substantially as hereinbeforc described with reference to or as illustrated in any of Figures 4 to 16 of' the accompanying drawings.