IMPROVED EVAPORATIVE CONDENSER

TECHNICAL FIELD

An improved evaporative condenser and evaporative condensation process are disclosed for use in refrigeration and air-conditioning systems. The condenser and process can be employed with both chemical refrigerants (e.g. hydrofluorocarbons) and natural refrigerants (e.g. hydrocarbons (such as propane & isobutane), C0 ₂ , ammonia, etc).

BACKGROUND ART

Existing evaporative condensers are used to reject heat in a variety of refrigeration and air-conditioning systems through the condensing of a refrigerant. More specifically, evaporative condensers comprise one or more wetted (e.g. sprayed) condensing coils for condensing the refrigerant by the passage thereover of an airstream, and into which a portion of the water is evaporated, thereby removing heat from the refrigerant in the condensing coils and causing the refrigerant to condense therein. Evaporative condensers also comprise drift eliminators (or, more simply, eliminators, "drift" being water that would otherwise pass to atmosphere). Drift eliminators remove free water that passes with the airstream as it flows through the condensing coils and water spray, prior to releasing that airstream to atmosphere.

In existing evaporative condensers the plan area of the condensing coils is matched to the plan area of the drift eliminators to ensure constant air flow rate and airstream velocity through the evaporative condenser.

In existing evaporative condensers, heat exchange efficiency is limited by the velocity of air that flows over the condensing coils. The velocity of air is in turn limited by the ability of the drift eliminators to remove free water from the air passing therethrough. In existing evaporative condensers, such removed water is recycled for reuse in wetting the condensing coils. However, any water that passes with the air flowing through the drift eliminators to atmosphere may contain bacteria, such as legionella, hence the requirement to remove as much free water from the airstream as possible.

For example, in many existing evaporative condensers, it is known to specify a maximum air velocity through the drift eliminators as high as 3.5 to 4 m/s to ensure sufficient water removal, however, it is surmised that, with such a high maximum air velocity, there is still a significant risk of bacteria (such as legionella) passing with non- eliminated free water through the drift eliminators. A safer maximum air velocity through the drift eliminators of 3.5 m/s is proposed. However, this will in turn set a limit to the velocity of air that can flow over the condensing coils.

The above references to the background art do not necessarily constitute an admission that the art forms part of the common general knowledge of a person of ordinary skill in the art. The above references are also not intended to limit the application of the condenser and process as disclosed herein.

SUMMARY OF THE DISCLOSURE

Disclosed herein is an evaporative condenser for use in a refrigeration or air- conditioning system. The evaporative condenser as disclosed herein can condense chemical refrigerants (e.g. hydrofluorocarbons, hydrochlorofluorocarbons,

perfluorocarbons, hydro fluoroolefins, etc) and natural refrigerants (e.g. hydrocarbons such as propane & isobutane, C0 ₂ , ammonia, etc).

The evaporative condenser as disclosed herein comprises one or more condensing coils for condensing therewithin the refrigerant of the system. The one or more condensing coils can be arranged in a condensing coil zone of the evaporative condenser. The condensing coil zone may comprise an air plenum having a constant cross-sectional area.

The evaporative condenser as disclosed herein also comprises a mechanism for wetting the one or more condensing coils (e.g. by spraying them with water). The evaporative condenser as disclosed herein further comprises drift eliminators arranged to remove free water from an airstream that has flowed past the one or more condensing coils and wetting mechanism.

In accordance with the present disclosure, the evaporative condenser as disclosed herein comprises a divergent zone that diverges from the condensing coil zone towards the drift eliminators. The configuration of the divergent zone is such that, once the airstream has flowed past the one or more condensing coils, it flows into and through the divergent zone to the drift eliminators. For example, the divergent zone may comprise an air plenum having a progressively increasing cross-sectional area.

The divergent zone is able to cause the airstream leaving the condensing coil zone to decelerate. This means that the velocity of air passing over the condensing coils can be increased, relative to the velocity of air passing through the drift eliminators. This higher velocity can help to reduce fouling of the tube.

Further, it has been surprisingly discovered that a condensing coil bundle with a reduced plan area, relative to the drift eliminators can be employed. A further consequence of this is that less condensing coil is required for the same condenser performance. This means that a lower cost evaporative condenser can be produced, as the condensing coil bundle represents the single-most expensive component of such a condenser.

In addition, an increased flow of refrigerant can be passed through the condensing coil bundle, because the greater air velocity is able to bring about condensation of a relatively greater amount of refrigerant.

Furthermore, this means that, as an alternative to using known hot-dipped galvanized carbon steel condensing tube, a more expensive and/or stronger material (e.g. stainless steel) can be used to form the one or more condensing coils, with the result that longer life, less corrosion and, optionally, thinner wall material for the coil (tube) can be employed. Notwithstanding, and if preferred, DN 8, 10, 15, and 20 (Schedule 40) seamless, hot-dipped seamless galvanized carbon steel tube can still be used to form the one or more condensing coils. In one embodiment, each of the one or more condensing coils may employ stainless steel tube (e.g. 304 or 316 stainless steel of 4.76 - 31.8 mm outside diameter and 0.5 - 1.6 mm thickness). The use of 304 stainless steel can offer better

conductivity, whereas 316 stainless steel can offer better corrosion resistance. Such tube material can perform favourably in comparison to a known condensing coil tube of galvanized mild carbon steel. The use of very small diameter tube can be suitable for certain small-scale applications.

The use of stainless steel tube material (i.e. due to corrosion/chemical resistance, increased refrigerant pressure capacity, etc) can also allow a natural refrigerant, such as a propane and/or isobutane hydrocarbon, C0 ₂ , ammonia, etc, to be employed.

In one embodiment the one or more condensing coils can be arranged as a bundle (e.g. of two or more nested coils) in the condensing coil zone. For example, the condensing coil zone may comprise a section of the condenser of generally constant cross-sectional area (e.g. an air plenum of circular, square, rectangular, etc hollow section).

In one embodiment the divergent part of the zone can be configured to cause the airstream to decelerate in a gradually decreasing manner.

In one embodiment the divergent zone can comprise a hollow frustum (hollow air plenum) through which the airstream flows. Such a hollow frustum may be located on the air exit side of the condensing coil plenum. For example, when the condensing coil plenum is of circular section, the divergent frustum each comprise a conical frustum, or a square-to-circular frustum-like prism; when the condensing coil plenum is of square section, the divergent frustum may comprise a square frustum; etc.

In one embodiment, the drift eliminators may be immediately located at an air leaving side of the divergent zone.

In one embodiment, the condenser may comprise an air inlet chamber located at an air entry side of the condensing coil zone.

In one embodiment, the mechanism for wetting the one or more condensing coils may comprise one or more spray nozzles. The spray nozzles may be arranged with respect to the divergent zone to spray water onto the one or more condensing coils in a direction that is counter to the airstream flow through the one or more condensing coils. For example, the spray nozzles may be arranged in the divergent zone, and may spray the water generally as a liquid cone into the condensing coil zone.

Alternatively, the mechanism for wetting the one or more condensing coils may comprise water distribution channels, such as those with serrated edges, internal slots, etc.

The condenser can comprise a water collection zone (e.g. located at a base of an air inlet chamber). The collection zone can collect water that has passed through condensing coil zone.

The condenser can further comprise a recycling system for recycling the collected water to the wetting mechanism, to maximize condenser efficiency. In one embodiment the recycling system can comprise a pump for pumping the collected water via pipework to the wetting mechanism. For example, an offtake pipe can extend from the base of the air inlet chamber to the pump, and a delivery pipe can extend from the pump outlet to the wetting mechanism (e.g. to spray nozzle, distribution pipework, etc).

In one embodiment, the recycling system can further comprise, as necessary, a water make-up mechanism for maintaining a predetermined amount of water (e.g. in the water collection zone) for effective operation of the evaporative condenser. Such makeup water can include that eliminated (captured) by the drift eliminators.

In one embodiment, the evaporative condenser can further comprise a heat exchanger (e.g. a separate, laterally located discrete heat exchange unit). The collected water can be passed through the heat exchanger prior to recycling it to the wetting mechanism. In addition, the condensed refrigerant can be passed through the heat exchanger to exchange heat with the recycled collected water. Such a heat exchanger can be used to sub-cool the condensed refrigerant to further improve the operational efficiency of the evaporative condenser.

Also disclosed herein is an evaporative condenser that comprises the collection zone for collecting water that has passed through condensing coil zone, and that comprises the heat exchanger through which the collected water is passed prior to recycling it to the wetting mechanism, and through which the condensed refrigerant is passed to exchange heat with the recycled collected water.

Also disclosed herein is an evaporative condensation process forming part of a refrigeration or air-conditioning cycle.

The process comprises passing refrigerant through one or more condensing coils. The process also comprises wetting the one or more condensing coils with water. The process further comprises passing an airstream over the one or more wetted condensing coils whereby refrigerant is caused to condense within the coils, and whereby a portion of the water is caused to evaporate into the airstream. The process additionally comprises eliminating water that is present in the airstream leaving the one or more condensing coils.

In accordance with the present disclosure, the process is conducted such that the velocity of the airstream leaving the one or more condensing coils is caused to decelerate prior to eliminating the water that is present in the airstream.

As outlined above, this can lead to a reduced plan area (and hence a lesser amount) of the one or more condensing coils relative to the drift eliminators (with the attendant advantages as outlined above).

Also disclosed herein is an evaporative condensation process in which the water that passes through the one or more condensing coils is collected and recycled to wet the one or more condensing coils with water. Further, in such a process, heat can be exchanged between the condensed refrigerant and the collected water, prior to recycling it to wet the one or more condensing coils.

The process as disclosed herein can take place in an evaporative condenser as set forth above.

In the process as disclosed herein, the refrigerant condensed in the one or more condensing coils can comprise a natural refrigerant (e.g. a hydrocarbon such as propane and/or isobutane, C0 ₂ , ammonia, etc) or a chemical refrigerant (e.g. a

hydro fluorocarbon, a hydrochlorofluorocarbon, a perfiuorocarbon, a hydrofluoroolefin, etc). BRIEF DESCRIPTION OF THE DRAWINGS

Notwithstanding any other forms which may fall within the scope of the condenser and process as set forth in the Summary, specific embodiments will now be described, by way of example only, with reference to the accompanying drawings in which:

Figure 1 shows cross-sectional side schematic of an evaporative condenser having a condensing coil zone in which one or more condensing coils are arranged, and a divergent zone extending away from the condensing coil zone;

Figure 2 shows a detail of Figure 1, to illustrate a variant of the evaporative condenser that further comprises a side heat exchanger; and

Figures 3A and 3B respectively show cross-sectional and side schematics of an evaporative condenser having a convergent-divergent zone in which one or more condensing coils are arranged;

Figure 4 shows a cross-sectional side schematic of an evaporative condenser that is similar to Figure 1 , but for different process parameters in accordance with the Examples;

Figure 5 is a graph showing C0 ₂ and water temperature profiles in accordance with the Examples;

Figure 6 is a graph showing C0 ₂ heat capacity profile in accordance with the Examples;

Figure 7 is a graph showing water flow down the tube bundle in accordance with the Examples;

Figure 8 is a graph showing overall heat transfer coefficient & pressure loss in accordance with the Examples;

Figure 9 is a graph showing a heat rejection profile based on a commercially available transcritical C0 ₂ compressor at 5°C sat. suction, with 5 K useful suction superheat and 5°C C0 ₂ liquid temperature, in accordance with the Examples;

Figure 10 is a graph showing performance of a commercially available transcritical C0 ₂ compressor at 50 Hz. 30 kW/27.2 m ³ /h, in accordance with the Examples; and Figure 1 1 is a graph showing the variation in COPs of NH ₃ , R22 R507A, Propane and R134a with Saturated Condensing Temperature, in accordance with the Examples.

DETAILED DESCRIPTION OF SPECIFIC EMBODIMENTS Specific forms of an evaporative condenser, and an evaporative condensation process that form part of a refrigeration or air-conditioning system/cycle, will now be described.

Evaporative condenser embodiments designated 10 and 100 are respectively shown in Figure 1 & 2 and Figures 3 A & 3B. The evaporative condensers embodiments 10 and 100 are able to employ both chemical and natural refrigerants (as set forth above). Figures 4 to 11 relate to embodiments described in the Examples.

In Figures 1 to 3, similar components of the evaporative condensers 10 and 100 are numbered similarly, but with 100 added to the embodiment of Figure 3. It should be further understood that, for the sake of brevity, the following description does not re- describe those similar or like components that re-appear in the embodiment of Figure 3, which therefore should be taken to have been described.

The preferred evaporative condenser 10 of Figures 1 and 2 comprises two or more nested condensing coil bundles 12 that have flowing (for condensing) therewithin the selected refrigerant of the system. The condensing coil bundles 12 are arranged in a condensing coil zone in the form of a rectangular airflow plenum 13.

The evaporative condenser 10 also comprises a mechanism in the form of spray nozzles 14 formed in a distributor tube 15 for wetting the condensing coil bundles 12 by spraying them with cones 16 of water (e.g. at a rate of 3 kg/m as shown). Alternatively, water distribution channels, such as those having serrated edges or internal slots, can be employed.

The spray nozzles 14 are arranged to spray water onto the condensing coil bundles 12 in a direction that is counter to the airstream flow therethrough as shown.

The evaporative condenser 10 also comprises a fan arranged in a fan housing at an upper end of the condenser. Such an arrangement is actually shown in the embodiment of Figure 3 as a fan 1 18 arranged in a fan housing 120 located at an uppermost end of the condenser (see Figure 3A). The same or similar arrangement can be employed in the embodiment of Figures 1 & 2. In this regard, the fan causes air to be drawn via an air inlet 21 into an air inlet chamber 22 that is arranged towards the lower end of the condenser 10.

In the embodiment of Figures 1 & 2 the airstream A enters at a volumetric flow rate of e.g. 8.1 m /s, first passing through mesh filters and then into the air inlet chamber 22, before it is caused by the fan to flow up to and through the condensing coil bundles 12. The air pressure differential can be maintained by the fan at e.g. 160Pa.

In the embodiment of Figure 3 the airstream A enters with a velocity of e.g. 3 m/s and a wet bulb temperature of e.g. 23°C, first passing through optional mesh filters 124 depending on air contamination and air inlet slots 126 and then into the air inlet chamber 122, before it is caused by the fan 1 18 to flow up to and through the condensing coil bundles 112.

The evaporative condenser 10 further comprises drift eliminators 30 which are arranged within the condenser adjacent to an upper end thereof. The drift eliminators 30 remove free water from the airstream once it has flowed past the condensing coil bundles 12 and spray nozzles 14.

In the embodiment of Figures 1 & 2, the evaporative condenser 10 comprises the rectangular airflow plenum 13 immediately followed by a divergent airflow zone in the form of a frustum-shaped plenum 40. The rectangular airflow plenum 13 can be of square, rectangular, etc hollow section (e.g. of bent and welded plastic or metal sheet/plate). The divergent plenum 40 can also be of hollow section (e.g. of bent and welded plastic or metal sheet/plate), but formed so as define the frustum. When, for example, the plenum 13 is of square section, the divergent plenum 40 comprises a square or rectangular frustum.

However, in the embodiment of Figure 3, the evaporative condenser 100 employs both a convergent airflow zone 135 and a divergent airflow zone 140 located on either side of an intermediate rectangular airflow plenum 1 13 that contains the condensing coil bundles 112. The plenum 113 has a constant cross-sectional area and interconnects the convergent airflow zone 135 and the divergent airflow zone 140. The intermediate airflow plenum 1 13 can again be of square, rectangular, etc hollow section (e.g. sheet/plate). The convergent airflow zone 135 and divergent airflow zone 140 can again be of hollow section (e.g. sheet/plate), but each formed so as define the frustum. When, for example, the intermediate zone 1 13 is of square section, the convergent and divergent frustums may each comprise a square or rectangular frustum.

In the embodiment of Figures 1 & 2, the fan is operated such that the airstream A is already at a higher velocity at the condensing coil bundles 12 relative to the drift eliminators 30. Having flowed past the condensing coil bundles 12, the airstream A flows into the divergent airflow plenum 40, passing through the water cones 16.

Because of the progressively increasing cross-section of the divergent airflow plenum 40, the airflow is able to decelerate to an acceptable velocity before it reaches and passes through the drift eliminators 30. The evaporative condenser 10, and in particular, the divergent airflow plenum 40, is configured such that this velocity is at a level whereby an environmentally acceptable minimum amount of free water in the airstream can be eliminated therefrom. In this regard, the airflow rate at the drift eliminators 30 can decelerate to approximately 3.5 m/s.

It will be seen that the drift eliminators 30 are arranged immediately at the air exit of the divergent airflow plenum 40, whereby the airflow is not permitted to decelerate more than is necessary.

Thus, the embodiment of Figures 1 & 2 does not employ a convergent airflow zone. Rather, the airflow velocity from the air inlet chamber 22, and through the condensing coil bundles 12, is approximately 5 m/s, until the airflow reaches the divergent airflow plenum 40, whereupon the airflow gradually decelerates to approximately 3.5 m s at the drift eliminators 30.

However, in the embodiment of Figure 3, the condensing coil bundles 1 12 are arranged in the intermediate airflow zone 1 13. The configuration of these zones is such that the airstream A flows through and is accelerated in the convergent airflow zone 135 to the condensing coil bundles 1 12 located in the intermediate airflow plenum 113 (e.g. to approximately 5 m/s). Having flowed past the condensing coil bundles 112, the airstream A flows into the divergent airflow zone 140, passing through the water cone 16, and decelerating before reaching the drift eliminators 130. Again, the drift eliminators 130 are arranged immediately at the air exit of the divergent airflow plenum 140.

The convergent airflow zone 135 is configured to cause the airstream A to accelerate, such as in a gradually increasing manner. Conversely, the divergent airflow zone 140 is configured to cause the airstream to decelerate, such as in a gradually decreasing manner. This means that the velocity of air passing through the intermediate airflow zone 1 13 and over the condensing coil bundles 1 12 is increased, relative to the velocity of air that passes into the air inlet chamber 122 as well as through the drift eliminators 130. For example, in the configuration depicted, the airflow rate in the intermediate airflow zone 1 13 is approximately double at ~ 5 m/s, (i.e. about 45% higher than) the 3.5 m/s air velocity through the drift eliminators.

In either embodiment, and as a result of this increased airflow rate passing over the condensing coil bundles 12, 112, it has surprisingly been discovered that a condensing coil bundle with a reduced plan area, relative to the drift eliminators 30, 130 can be employed. As a further consequence of this increased airflow rate, it has surprisingly been discovered that less condensing coil is required for the same condenser heat rejection performance.

The result is that a lower cost evaporative condenser can be produced, as the condensing coil bundle represents the single-most expensive component of the condenser. Alternatively, instead of using known thick-wall, hot-dipped galvanized carbon steel condensing tube for the coil bundle 12, 1 12 a more expensive and/or stronger material, such as stainless steel tube, can be used to form the coil bundle 12, 112. In such case, the result is longer coil life, less corrosion and, if desired, thinner wall material for the tube in the coil bundle. In this regard, the coil bundle 12, 1 12 can comprise stainless steel tube, such as 304 or 316 stainless steel of 4.76 - 31.8 mm outside diameter and 0.5 - 1.6 mm thickness. Such tube is observed to perform well in comparison to known condensing coil tube of galvanized mild carbon steel. The corrosion and chemical resistance, as well as increased refrigerant pressure capacity, that is provided by such stainless steel tube materials also allows a natural refrigerant, such as a propane and/or isobutane hydrocarbon, C0 ₂ , ammonia, etc, to be employed in the evaporative condenser 10, 100.

Another consequence of the increased airflow rate over the condensing coils is that an increased flow of refrigerant can be passed through the condensing coil bundle 12, 1 12 because the greater air velocity is able to bring about condensation of a relatively greater amount of refrigerant.

The condenser 10 also comprises a water collection zone in the form of a basin 50 located at a base of (i.e. adjacent to) the air inlet chamber 22. The basin 50 collects excess spray water that has passed through or from condensing coils.

To maximize condenser efficiency, the condenser 10 additionally comprises a recycling system for recycling the collected water to the distributor tube 15 for feeding to the spray nozzles 14. In this regard, the recycling system comprises a pump 52 for pumping the collected water via pipework to the distributor tube 15. The pump 52 draws water out of the basin 50 via an offtake pipe 54. A delivery pipe section 56 then extends from the pump outlet to connect with the distributor tube 15.

The recycling system also comprises water make-up 58 (e.g. at 383 kg/h) for maintaining a predetermined amount of water in the basin 50 for effective operation of the evaporative condenser. Such make-up water can include a supply of water that has been eliminated (captured) by the drift eliminators 30.

In a variation of the evaporative condenser shown in the detail of Figure 2, the condenser 10 can further comprise a side heat exchanger unit 60. The water in the basin 50 can be pumped via pump 52 and into and through the heat exchange unit 60, prior to being recycled to the distributor tube 15 via the delivery pipe section 56. Such a unit can also be fitted to the embodiment of Figure 3.

In this variation, the condensed refrigerant in the condenser tubes can also be passed via refrigerant delivery pipe 62 to and through the heat exchange unit 60 to exchange heat with the recycled water from the basin 50. In the heat exchange unit 60 the relatively cool basin water can sub-cool the condensed refrigerant, for example, from 30°C to around 26.5°C. This can further improve the operational efficiency of the refrigerating system. The refrigerant (e.g. C0 ₂ ) leaving the heat exchange unit 60 as the stream 64 can be at a sub-cooled temperature (e.g. around 26.5°C).

Examples

Non-limiting examples of the present condenser and process will now be provided in order to illustrate the theoretical basis of the condenser and process, and to better understand the condenser and process in operation.

Example 1 - Process Design Model

A design model for the application to subcritical C0 ₂ condensing of evaporative condensers, such as those depicted in Figures 1 to 3, was developed. More specifically, the benefits of applying evaporative condensing techniques for the condensation of subcritical C0 ₂ were examined. Such benefits included lower design pressures compared to trans-critical operations, lower energy consumption, and lower running and operating costs. It was noted that hot gas defrosting could also become a standard feature of subcritical C0 ₂ refrigeration plant operations.

Firstly, however, it was noted that ammonia can be condensed at 30°C in an evaporative condenser with an entering air wet bulb temperature of 24°C. In the developed design model it was shown that an evaporative condenser for subcritical C0 ₂ condensing at 30°C (i.e. 1.1 K below the critical point) was able to be designed for a wet bulb of 24°C.

Secondly, it was noted that average climate conditions in much of Europe, including the warmer climates in Spain, Italy, Greece and Turkey, were suitable for evaporative condensers to condense subcritical C0 ₂ at 30°C. Canada, large parts of the USA and China, and most of Australia below the tropic of Capricorn were also noted to have climates suitable for the application of evaporative condensers to subcritical C0 ₂ condensing. The thermo dynamic and transport properties of subcritical C0 ₂ at 30°C were noted to change significantly with temperature. Thus, the effect these changes have on C0 ₂ temperature profile, heat transfer and pressure loss for a particular design was also shown.

For example, an examination of average climate conditions revealed that much of Europe, including Spain, Italy, Greece and Turkey, has a climate where evaporative condensers may be applied for the condensing of C0 ₂ at subcritical conditions at a condensing temperature of 30°C or lower 100% of the time in many locations. For example, the only location in Europe where the 5% design Wet Bulb temperature incidence exceeded 24° C was Adana in Turkey (where the 1 and 2.5% wet bulb incidence design levels are at 26° C). At Thessaloniki in Greece the 1% wet bulb design incidence is at 25° C, but the 2.5% and 5% wet bulb design incidence levels are at 24° C. The next highest 1% wet bulb design incidence level of 24° C occurred at Gibraltar, Barcelona, Valencia, Milan, Istanbul and Izmir.

Finally, it was concluded that the use of evaporative condensers for C0 ₂ in temperate and many subtropical climates could make C0 ₂ refrigeration as ubiquitous as any chemical refrigerant and would compete successfully with ammonia when it needed to be used in an indirect application (such as the heating and cooling of office buildings and hospitals for example).

When C0 ₂ refrigeration was revived about 20 years ago, air cooled gas cooling (some with adiabatic assistance by spraying water onto the air inlet face of the finned coil gas cooler) was applied almost universally. It was noted that this resulted in virtually all C0 ₂ refrigeration systems needing to run in trans-critical mode because the air cooling temperature is close to, or exceeds, the C0 ₂ critical temperature of 31.1°C.

More often than not the summer design C0 ₂ exit temperatures from an air cooled gas cooler were higher than the critical temperature, and this resulted in the compressors needing to operate at a pressure of 90 bar or higher to ensure a reasonable COP. The summer design COPs of trans-critical C0 ₂ compressors were generally lower than those of air cooled HFC or evaporatively cooled ammonia systems.

It was therefore proposed to reduce the temperature of the condenser cooling medium to a level which would allow a complete subcritical C0 ₂ refrigeration cycle.

This was accomplished with an evaporative condenser, where the ambient air Wet Bulb (WB) temperature was the effective cooling medium temperature, rather than the ambient air Dry Bulb (DB) temperature in the case of an air cooled condenser or gas cooler. Issues noted included the need for a water supply, water consumption and water treatment, and control of a minimum condensing temperature as currently mandated by some compressor suppliers. Another issue was the control strategies to handle inadvertent trans-critical conditions. Recommendations were made to address these issues.

A Rating Model for a CO? Evaporative Condenser

Rating example

Figure 4 shows a schematic flow sheet of an evaporative condenser that will now be further described. In the flow sheet of Figure 4 water was recycled over the tube bank, so that spray water temperature was the same as the basin water temperature.

The specified parameters were: (a) air velocity and wet and dry bulb

temperatures, (b) spray water flow rate, (c) bundle dimensions, and (d) leaving C0 ₂ comprising saturated liquid at 30°C and 7.2 MPa.

Mass and energy balances in evaporative coolers

Qureshi (2006) and Heyns (2009) published five simultaneous non-linear differential equations describing air-water-process fluid interactions in evaporative cooling.

h _d (W _int - W)

da

1) dw 1 dm _w

da m _a da

2) dh _r U ₀

da

4) dT _w 1

m _a * dh _a - Cp _w T _w dm _w — m _r Cp _r dT _r )

da m _w Cp _< w 5)

The equations were solved by writing a program using a fourth order Runge- Kutta routine written in Microsoft's VBA behind a Microsoft Excel spreadsheet with pass length divided into forty intervals. The solution was trial and error because the basin water temperature at air entry was guessed and adjusted iteratively until it was the same as the calculated water temperature at the air exit.

The solution proceeded "backwards" along a tube pass from C0 ₂ exit to entry, starting at the air inlet with saturated liquid refrigerant at 30°C, and proceeded up, as if heating, ending with superheated vapour at a calculated discharge temperature. The program allowed for both two-phase condensation and single phase vapour de- superheating.

Verifying the model.

There was no analytical solution to the five equations by which the numerical solution can be validated. However, two findings were noted: when the leaving and entering water temperatures were equal; (a) the C0 ₂ enthalpy change was equal to the moist air enthalpy change, and (b) the heat duty calculated for ammonia condensing at 30°C was within 9% of duty computed using the simplified Merkel model

(Merkel,1926) which is based on constant condensing temperature.

Model predictions

Figures 5 and 6 show C0 ₂ and water temperature profiles. The shape of the C0 ₂ temperature profile (Figure 5) was a surprise - it was much flatter than expected. Over about 37% of exchanger surface, the C0 ₂ vapour temperature only reduced from 32 to 30°C at interval number 29. This was a consequence of the very high heat capacity just above 30°C (Figure 6). The model predicted that 67% of exchanger surface would be needed for sensible cooling. Enthalpy data for C0 ₂ at 30°C, close to the critical point, showed that 68% of the rejection was sensible cooling, unlike ammonia where it is only 10%.

Water temperature profile was skewed to the left, compared to profiles where sensible cooling was relatively small, reflecting the larger proportion of sensible cooling for C0 ₂ near the critical point.

Water evaporation

Figure 7 shows the water flow down the tube bundle. Another surprise was that water did not evaporate to air at the top of the bundle. Here, water temperature was low though rising, and water contacted air with a humidity ratio higher than the humidity ratio at the air-water interface, so some condensation occurred and water flow increased.

Effect of property changes

Figure 8 illustrates the effects of changes in heat capacity, density, viscosity and thermal conductivity with temperature on overall heat transfer coefficient and pressure loss over each solution interval. Interval zero was where the hot discharge gas enters.

There was considerable rise in overall heat transfer coefficient as the C0 ₂ temperature approached 32°C, with a corresponding decrease in pressure loss per metre as the vapour phase transitioned to two phases.

In the model 0.845 was used for the Lewis number in equation (3). With a

Lewis number of 1.00, the surface area required for the same heat duty was reduced by only 1.4%.

Remarks

The case modelled was extreme in the respect that C0 ₂ condensing at 30°C was very close to its critical point. It was noted that at lower condensing temperatures the proportion of sensible cooling would reduce and the variation of properties with temperature would be much reduced. Reference was made to Figure 9.

It was further noted that heat rejection predicted by the Merkel simplified model was about 22% lower than the differential model with C0 ₂ condensing at 30°C, which was not unexpected given the significant proportion of sensible cooling. CO2 Compressor Subcritical Energy Performance

Effect of condensing temperature on cycle performance

In Figure 10, five COP plots were produced for a commercially available, semi hermetic, trans-critical C0 ₂ compressor with a swept volume of 27.2 m ³ /h at 50 Hz and 30 kW 4 pole motor.

Referring to curve 1 in Figure 10, the COP ranged from 6.27 at +30°C Saturated Condensing Temperature (SCT) to 18.0 at an SCT of + 16°C at a Saturated Suction Temperature (SST) of +10°C. A +10°C SST would allow an Evaporating Temperature (ET) of +11°C with a suction pressure drop corresponding to 1 K boiling point suppression. 1 1°C was noted as a reasonably efficient evaporating temperature for direct cooling of Air Conditioning (AC) air, allowing a relatively large diffusion in air temperature across the cooling coil, thus limiting the volume of air which would need to be circulated, and thereby reducing fan energy consumption and the resulting parasitic heat load. This in turn would lead to a reduction in the required energy input into the compressor thereby lifting the overall energy efficiency of the system as a whole.

Curve 2 showed the COP ranging from 4.45 to 11.67 at 30°C to 16°C SCT at an SST of +5°C. This would allow chilled water production for AC for retrofitting into existing buildings and application to new buildings.

In both the above two cases the AC compressors could act also as parallel compressors for refrigeration duties at -5°C SST, such as maintaining chill storage temperatures at around 0°C and high stage duties for two stage C0 ₂ systems applied to cold storage and blast freezing applications.

In such cases the high stage compressors would operate with virtual C0 ₂ gas cooler exit temperatures of +5°C and +10°C, which resulted in COP curves 3 and 4 respectively. COP curve 3 ranged from 4.7 to 7.88 at an SST of -5°C at SCTs ranging from +30 to +16°C and a virtual gas cooler exit of +5°C. COP curve 4 showed the COP ranging 4.45 to 7.04 with a virtual gas cooler exit of +10°C, and SST of ^~ 5°C and the SCT ranging from +30 to +16°C. It was noted that this could be improved with a Suction Heat Exchanger (SHEX) in the compressor suction to bring the performance closer to curve 3. Effect of ambient Wet Bulb temperature on condenser performance

Figure 4 shows the general details on C0 ₂ , air and water. The Effect of ambient Wet Bulb temperature on condenser performance is shown in the results set forth in the following table:

Relative energy efficiency of ammonia, R22, R507A, propane, and R134a

The COPs for these refrigerants at identical operating conditions were shown in Figure 1 1. The results confirmed that ammonia was the best refrigerant of these. A surprise was the low COP of R134a. At SCTs of 16 and 35°C the R134a COPs were respectively 42 and 31% lower than those of ammonia. Furthermore the COP of an R134a compressor at +16°C SCT was, at 3.84, about the same as the COP of an ammonia compressor at +35°C SCT at identical suction conditions. This confirmed that R134a had both high direct and indirect Global Warming Potential (GWP). The performance of R507A was 1 1 to 16% less efficient than R22 at 25 to 35°C SCT. HFC R507A had no Ozone Depletion Potential like HCFC R22, but the 100 years GWP of R507A was 3,895, more than double the 100 year GWP of 1,810 of R22.

Overall heat transfer factor, Uo Referring again to Figure 8, Uo was very considerably higher than the familiar Uo in the case of ammonia condensing, where Uo ranges from about 450 to 550 w/m ² .K at superficial air velocities of 2.6 to 3.05m/s. A 3m/s superficial air velocity was chosen as a maximum in the models to ensure that the drift eliminators would be able to catch most of the free water suspended in the upward air draft.

The average Uo in Figure 8 for the C0 ₂ evaporative condenser in Figure 4 was about 1,050 w/irT.K. This was noted to be more than double the average value for ammonia at virtually the same superficial air velocity entering the condensing tube bundles. This was remarkable, when it was considered that at 30°C condensing 68% of the heat to be removed is sensible superheat and only 32% is actually latent heat of condensation at 30°C as shown in Figure 9. The high overall heat transfer factor was attributed to by the high C0 ₂ mass flux of 338.7 kg/m ² .s in the 76 off 55 metre equivalent length circuits causing a calculated pressure drop of 15 kPa. This is the maximum value acceptable to facilitate C0 ₂ condensers operating in parallel without requiring too great a drop leg to avoid liquid hold up in operating condensers should one condenser not be operating.

As with evaporators, the high ΔΡ/ΔΤ ratio of C0 ₂ allowed high mass fluxes in the condenser circuits giving high rates of heat transfer, allowing fewer longer circuits, which also made for more economical manufacture of the tube bundle.

It was noted that ammonia mass fluxes in evaporative condensers range from about 25 to 40 kg/m2.s and are frequently lower than 25. The pressure drop was a concern with ammonia condensers, as excessive pressure drop in an ammonia evaporative condenser lifts the discharge pressure, and thus the Saturated Condensing Temperature (SCT), resulting in increased energy consumption.

Consequences of minimum airflow

Again referring to Figure 4, the calculated leaving air Dry Bulb Temperature is 29.3°C at 100% RH and hence the leaving Wet Bulb Temperature was also 29.3°C. This was only 0.7°K lower than the SCT of 30°c. This was possible because the top tubes were at a temperature of 77°C and the high proportion of sensible superheat ensured that there was a high leaving approach TD available of 47.7°K. It was noted that this was not possible in ammonia evaporative condensers, where minimum leaving temperatures approaches between the ammonia SCT and the leaving Wet Bulb are rarely less than 3K and not less than 2.5K at design conditions. Little airflow also resulted in minimum fan energy consumption.

Conclusion

Subject to satisfactory performance testing of a full scale prototype C0 ₂ evaporative condenser, it was concluded that the application of the evaporative condensers in the higher latitude subtropics with maximum design Wet Bulb (WB) temperatures of 24 to 25°C showed a great deal of promise. The C0 ₂ evaporative showed even more promise in areas with more temperate and cool to cold climates where ambient WB temperatures are lower.

According to the above conclusion, suitable areas for the application of evaporative condensers to the condensing of subcritical C0 ₂ compressor discharge gases was thus feasible in virtually all of Europe (including the Mediterranean countries), the USA except for the Southern States bordering the Gulf of Mexico and the Atlantic Ocean, and many of the Mid West States as far North as Minnesota.

Experimentation also showed that evaporative gas cooling with ambient Wet Bulb temperatures of 28 to 29°C and ambient air WB to C0 ₂ exit temperature approaches of 3 K were entirely feasible. This was ascribed to the fact that, in trans- critical mode, there was only sensible heat transfer at a larger LMTD without the condensing phase (Figure 9) and relatively high trans-critical fluid densities coupled with high heat capacities similar to that shown in Figure 6.

The application of evaporative cooling to both the condensing of C0 ₂ at subcritical and gas cooling at trans-critical C0 ₂ resulted in efficient refrigeration at high COPs which were comparable to, and in many cases higher than, the COPs achieved with conventional refrigerants operating below their critical points. This opened the way for the worldwide application of C0 ₂ refrigeration. This is particularly true in applications where C0 ₂ is used for Air Conditioning duties at +5 and +10°C

compressor Saturated Suction Temperature for chilled water, and DX or pumped C0 ₂ AC applications respectively. It was further noted that the AC compressors may also act as parallel compressors for any remaining refrigeration duties in a facility such as a supermarket where both chilling and freezing duties are required at high to very high COPs as shown in Figure 10.

Indeed, when comparing Figure 10 and 1 1 it was clear that, in the subcritical condensing phase, C0 ₂ outperformed conventional chemical refrigerants such as R22, R507A and R134a, as also found by Pearson (2010). Additionally, C0 ₂ rivalled or outperformed ammonia and propane under most operating conditions and particularly so where parallel compression was involved.

At a high Wet Bulb temperature of e.g. 28°C conventional evaporative condensers would be able to operate at 40°C SCT resulting in COPs of 3.37, 3.34, 2.71, 2.96 and 2.38 for NH ₃ , R22, R507A, propane and R134a respectively as shown in Figure 9. Thus, to develop highly efficient C0 ₂ refrigeration larger compressors are needed, for example, modified versions of CNG fuel compressors.

Nomenclature

In the Examples:

a outside surface area m ²

m _a air flow rate kg dry air s ^"1 m _w water flow rate kg s- ¹

m _r C0 ₂ flow rate kg s ^"1

Imasw saturated air enthalpy at air-water interface J kg ^"1 dry air a air enthalpy J kg ^"1 dry air h _d mass transfer coefficient kg

i _v water vapour enthalpy J kg ^"1

T water temperature °C

Le Lewis number -

T _r C0 ₂ temperature °C

Cpw heat capacity of liquid water J kg^ K ^"1

r ^pa heat capacity of moist air J kg-' K ^"1

Uo overall heat transfer coefficient W m ^"2 K ^"1 W air humidity ratio kg water kg ^" dry

air

air humidity ratio at air-water interface kg water kg ^"1 dry

air

water-tube heat transfer coefficient W m ^"2 K ^"1

hi C0 ₂ heat transfer coefficient W m ^"2 K ^"1

dj tube inside diameter m

d ₀ tube outside diameter m

ff fouling factor K m ² W ^"1

Model parameters

1. NIST (201 1) data were used for thermodynamic and transport properties of saturated and superheated C0 ₂ ;

2. h _w in equation (6) was calculated from Mizushima and Miyasita ( 1967), equation (A.8) in Qureshi and Zubair (2006);

3. h _d in equation (3) was calculated from Mizushima and Miyasita (1967), equation (A.13) in Qureshi and Zubair (2006);

4. For two phase C0 ₂ flow, hi in equation (6) was calculated from Shah's (2009), Qureshi and Zubair (2006) equations (A.6) and (A.7); pressure loss was calculated from Muller-Steinhagen and Heck correlation (ASHRAE, 2005);

5. For single phase C0 ₂ vapour flow, ¾ in equation (6) was calculated from the Dittus-Boelter correlation Nu = 0.023Re ^0'8 Pr ^0'3 ; pressure loss was calculated from a friction factor = 0.079Re ^"0'25 ;

6. Air pressure drop across the tube bank was calculated from Mills ( 1999), section 4.5.1 , p. 316.

The following References were used in formulating the model:

1. ASHRAE, 2005, 2005 Fundamentals, page 4.12- 13

2. Heyns J, Kroger D, 2009, Performance characteristics of an air-cooled steam condenser incorporating a hybrid (dry/wet) dephlegmator, Appendix A,

PIER Report, CEC-500-2013-065-APA 3. Merkel, F., 1926, Verdunstungskuling, VDI-Zeitschrift, Vol. 70, pp. 123

- 128

4. Mills A.F., 1999, Basic Heat & Mass Transfer, 2nd ed., A.F., Prentice

Hall.

5. Mizushima, T., R. Ito and H. Miyasita, 1967, Experimental study of an evaporative cooler, International Chemical Engineering, Vol. 7, pp. 727-732

6. NIST 201 1, http://webbook.nist.gov/chemistry/fluid/ Thermophysical Properties of Fluid Systems

7. Qureshi B, Zubair S, 2006, A comprehensive design and rating study of evaporative coolers and condensers. Part I Performance evaluation, Int. J.

Refrigeration, 29: 645-658.

8. Shah M, 2009, An improved and extended general correlation for heat transfer during condensation in plain tubes, HVAC&R Research, 15 (5)

9. Pearson, S. Forbes, 2010, Use of carbon dioxide for air conditioning and general refrigeration, IIR-IOR 1 ^st Cold Chain Conference, Cambridge, UK.

Example 2 - Design Model Outputs

The following data points were produced by the design model to illustrate condenser capacity variation with the superficial air velocity:

Evaporative Condenser Model 1

Required rejection

Calculated rejection Estimate for Condenser circuits and pressure dro

Rejection 485 kW

Refrigerant ; CO2, TC = 30 =C j | _R744 Try R744 mass velocity 370 kg/m2 s

R744 discharge temperature Te 80 °C R744 mass flow 2.492 kg/s

R744 design condensing temp. Tc 30 °C Condenser circuits 47

mm =

16 ft 7.2

Length of tube bundle L 5,060 in Condenser pressure drop

Ambient air dry bulb Ta 35.0 °C

Ambient air wet bulb Twbjn 24.0 °C

Refrigerant C02, Tc = 30°

Water mass velocity at air inlet Gw 3.00 kg/m2 s Calculated discharge temperature 80.7 ° m2

Water-side fouling allowance ff 0.088 K/kW R744 mass flux 243 k

Atmospheric pressure 101.3 kPa abs Condenser R744 pressure drop 6 k Needs higher water flow to

uniformly wet tubes (Ref 2, 3.3)

Plain or finned tube Plain Condenser exit air dry bulb 28.9 °

Plain tube OD D 15.88 mm Condenser exit air RH 100.0%

Tube ID 13.48 mm Condenser air pressure drop 61 P

Total outside surface 145.4 m2 Fan power ( at 70% efficiency) 1 .8 k

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Substitute Sheet (Rule 26) RO/AU Evaporative Condenser odel 2

Required rejection 631 kW

Calculated rejection 632 Estimate for Condenser circuits and pressure dro

kW Rejection 631 kW

Refrigerant R744 Try R744 mass velocity 370 kg/m2 s

R744 discharge temperature 80 °C R744 mass flow 3.242 kg/s R744 design condensing temp. 30 °C Condenser circuits 61

mm =

16 ft 7.2

nser pressure drop 10 kPa

Ambient air dry bulb 35.0 °C

Ambient air wet bulb °C

Refrigerant C02, Tc = 30°

Water mass velocity at air inlet 3.00 kg/m2 s Calculated discharge temperature 80.9 ° m2

Water-side fouling allowance 0.088 K/kW R744 mass flux 316 k Atmospheric pressure 101.3 kPa abs Condenser R744 pressure drop 9 k Needs higher water flow to

uniformly wet tubes (Ref 2, 3.3)

Plain or finned tube Plain Condenser exit air dry bulb 28.8 " Plain tube OD 15.88 mm Condenser exit air RH 100.1 % Tube ID 13.48 mm Condenser air pressure drop 102 P

Total outside surface 145.4 m2 Fan power ( at 70% efficiency) 4.0 k

Air mass velocity Ga 4.61 kg/m2 s

Ambient air relative humidity 41 % Basin water temperature 25.76 °

Water spray into basin m.w in 20.65 kg/s Rise above ambient wet bulb 1.8 K

Water makeup required 1221 k

Tube material 2 310 stainless

Thermal conductivity 16 W/m K Above Ga of 3.7, water is partially held up on tubes - Ref.

Horizontal tube pitch: PH to OD ratio 2.25

Tube horizontal pitch = 2.25 x 15.875

mm 35.72 mm

Tube to tube below: PL to OD ratio 3.90

Tube vertical pitch to tube below =

3.90 x 15.875 mm 61 .91 mm

?3 GO

s= Row-to-row pitch angle a 60.0 degrees o ^°

Evaporative Condenser Model 3

Required rejection 756 kW

Calculated rejection 756 Estimate for Condenser circuits and pressure drop

kW Rejection 756 kW

Refrigerant _; C02, TC -_3O=C R744 Try R744 mass velocity 370 kg/m2 s

R744 discharge temperature Te 80 °C R744 mass flow 3.884 kg/s R744 design condensing temp. Tc 30 °C Condenser circuits 74

mm =

nser pressure drop 10 kPa

Tube selection 5/8 ^' tube

Entering air superficial velocity Vo 5.00 m/s

Ambient air dry bulb Ta 35.0 °C

Ambient air wet bulb Twb in 24.0

Refrigerant C02, Tc = 30°C

Water mass velocity at air inlet Gw 3.00 kg/m2 s Calculated discharge temperature 80.3 °C m2

Water-side fouling allowance ff 0.088 K/kW R744 mass flux 378 kg Atmospheric pressure 101.3 kPa abs Condenser R744 pressure drop 13 kP Needs higher water flow to

uniformly wet tubes (Ref 2, 3.3)

Plain or finned tube Plain Condenser exit air dry bulb 28.5 °C

Plain tube OD D 15.88 mm Condenser exit air RH 100.2% Tube ID 13.48 mm Condenser air pressure drop 152 Pa

Total outside surface 145.4 m2 Fan power ( at 70% efficiency) 7.5 k

Air mass velocity Ga 5.76 kg/m2 s

Ambient air relative humidity 41 % Basin water temperature 25.39 °C

Water spray into basin m.w in 20.65 kg/s Rise above ambient wet bulb 1.4 K

Water makeup required 1487 kg

Tube material

Thermal conductivity 16 W/m K Above Ga of 3.7, water is partially held up on tubes - Ref. 2, 3.3

Horizontal tube pitch: PH to OD ratio 2.25

Tube horizontal pitch = 2.25 x 15.875

mm mm

Tube to tube below: PL to OD ratio 3.90

Tube vertical pitch to tube below =

GO 3.90 x 15.875 mm mm

Row-to-row pitch angle a 60.0 degrees

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Evaporative Condenser Model 4

Required rejection 485 kW

Calculated rejection 485 Estimate for Condenser circuits and pressure dro

kW Rejection 485 kW

Refrigerant C02, Tc = 30 ^: C R744 Try R744 mass velocity 370 kg/m2 s

R744 discharge temperature Te 80 °C R744 mass flow 2.492 kg/s R744 design condensing temp. Tc 30 °C Condenser circuits 47

mm =

16 ft 5.3

Length of tube bundle 5,012 in 0.0% OK Condenser pressure drop 10 kPa Number of circuits 56

Width allowance for header clearance mm

Width of tube enclosure 1075 mm I ^►

Number of tube passes 8

33 0.009

Tube selection 5/8 ^' tube

Entering air superficial velocity Vo 4.00 m/s

Ambient air dry bulb Ta 35.0 °C

Ambient air wet bulb 24.0 °C

Refrigerant C02, Tc = 30°C

Water mass velocity at air inlet 3.00 kg/m2 s Calculated discharge temperature 80.6 ° m2

Water-side fouling allowance 0.088 K/kW R744 mass flux 312 kq Atmospheric pressure 101.3 kPa abs Condenser R744 pressure drop 9 k Needs higher water flow to

uniformly wet tubes (Ref 2, 3.3)

Plain or finned tube Plain Condenser exit air dry bulb 28.8 °

Plain tube OD D 15.88 mm Condenser exit air RH 99.6% Tube ID 13.48 mm Condenser air pressure drop 102 P

Total outside surface 112.0 m2 Fan power ( at 70% efficiency) 3.2 k

Air mass velocity Ga 4.61 kg/m2 s

Ambient air relative humidity 41 % Basin water temperature 25.98

Water spray into basin m.w in 16.16 kg/s Rise above ambient wet bulb 2.0

Water makeup required 943

Tube material 2 | 316 stainless

Thermal conductivity 16 W/m K Above Ga of 3.7, water is partially held up on tubes - Ref. 2, 3.3

Horizontal tube pitch: PH to OD ratio 2.25

Tube horizontal pitch = 2.25 x 15.875

mm 35.72 mm

Tube to tube below: PL to OD ratio 3.90

Tube vertical pitch to tube below =

3.90 x 15.875 mm 61 .91 mm

Row-to-row pitch angle a 60.0 degrees

Evaporative Condenser Model 5

Required rejection 485 kW

Calculated rejection 485 Estimate for Condenser circuits and pressure dro

kW Rejection 485 kW

Refrigerant ' ^{C02 Jc = 30'} ~ ^c R744 Try R744 mass velocity 370 kg/m2 s

R744 discharge temperature Te 80 °C R744 mass flow 2.492 kg/s R744 design condensing temp. Tc 30 °C Condenser circuits 47

mm =

16 ft

Length of tube bundle L 5,165 1 1 .3 in 0.0% OK Condenser pressure drop 11 kPa ^ Number of circuits Nc 46

2? Width allowance for header clearance 25 mm

∞ Width of tube enclosure W 896 mm

ON £ Number of tube passes Np 8

! 5?S ^' tube

^ Tube selection 5

d ~

Entering air superficial velocity Vo 5.00 m/s

Ambient air dry bulb Ta 35.0 °C

Ambient air wet bulb Twb in 24.0

Refrigerant C02, Tc = 30

Water mass velocity at air inlet Gw 3.00 kg/m2 s Calculated discharge temperature 80.7

m2

Water-side fouling allowance ff 0.088 K/kW R744 mass flux 380 k Atmospheric pressure 101.3 kPa abs Condenser R744 pressure drop 13 k Needs higher water flow to

uniformly wet tubes (Ref 2, 3.3)

Plain or finned tube Plain Condenser exit air dry bulb 28.6 °

Plain tube OD D 15.88 mm Condenser exit air RH 98.6% Tube ID 13.48 mm Condenser air pressure drop 152

Total outside surface 94.8 m2 Fan power ( at 70% efficiency) 5.0 k

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Substitute Sheet (Rule 26) RO/AU Whilst a number of condenser and process embodiments and models have been described, it should be appreciated that the condenser and process may be embodied in many other forms. For example, the plenum 13 could be of circular section, whereby the divergent plenum 40 comprises a conical frustum, or a square to circular frustum-like prism. However, such a configuration is less favoured, as it does not promote free drainage of water within the condenser.

In the claims which follow, and in the preceding description, except where the context requires otherwise due to express language or necessary implication, the word "comprise" and variations such as "comprises" or "comprising" are used in an inclusive sense, i.e. to specify the presence of the stated features but not to preclude the presence or addition of further features in various embodiments of the condenser and process as disclosed herein.