Technical Papers. 37th Annual Meeting. International Institute of Ammonia Refrigeration. March 22 25, 2015

Transcription

1 Technical Papers 37th Annual Meeting International Institute of Ammonia Refrigeration March 22 25, Industrial Refrigeration Conference & Exhibition San Diego, California

2 ACKNOWLEDGEMENT The success of the 37th Annual Meeting of the International Institute of Ammonia Refrigeration is due to the quality of the technical papers in this volume and the labor of its authors. IIAR expresses its deep appreciation to the authors, reviewers and editors for their contributions to the ammonia refrigeration industry. ABOUT THIS VOLUME IIAR Technical Papers are subjected to rigorous technical peer review. The views expressed in the papers in this volume are those of the authors, not the International Institute of Ammonia Refrigeration. They are not official positions of the Institute and are not officially endorsed. International Institute of Ammonia Refrigeration 1001 North Fairfax Street Suite 503 Alexandria, VA (voice) (fax) Industrial Refrigeration Conference & Exhibition San Diego, California

3 Tech Paper #7 Applying Evaporative Condensers to Subcritical CO 2 Condensing and Transcritical CO 2 Gas Cooling J. Ball P O Box 4424, Ringwood, Vic, 3134, Australia Tel: john@antal-air.com.au K. Visser KAV Consulting Pty Ltd, P O Box 1146, Kangaroo Flat, Vic, 3555, Australia Tel: Fax: Mobile: kavconsult@bigpond.com Abstract A mathematical model of an evaporative condenser that allows for unusual properties of CO 2 has been constructed to show that subcritical CO 2 can be condensed at 30 C (86 F) with ambient air at 35 C (95 F) dry bulb and 24 C (75 F) wet bulb. World-wide climate conditions are examined to show that evaporative condensers can be used to condense subcritical CO 2 at 30 C (86 F) for the entire or most of the year in large areas of many countries such as USA, China, Canada, Australia almost all of Europe including Spain, Italy, Greece, and Turkey. Reduction in water consumption is demonstrated using a dry cooler in conjunction with an evaporative condenser. It is also demonstrated that evaporative CO 2 gas cooling of transcritical fluid results in high electrical energy efficiency signified by high Coefficients Of Performance (COP) at supercritical pressures of 7.5 MPa (1088 psia) and higher, particularly when employing parallel compression for medium temperature and low temperature refrigeration duties in, for example, supermarkets by providing AC cooling as part of the parallel compressor operation at 5 to 10 C (41-50 F) CO 2 evaporating temperatures. IIAR

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5 Applying Evaporative Condensers to Subcritical CO 2 Condensing and Transcritical CO 2 Gas Cooling Introduction CO 2 properties, such as enthalpy, heat capacity, viscosity and thermal conductivity of isobaric CO 2 near its critical temperature, change markedly with temperature. Profiles of CO 2 temperature and heat transfer coefficients in heat exchangers are very different from other refrigerants. One approach necessary in any case for evaporative cooler design, is to use a numerical solution for rating designs. Using our mathematical model, we show that, with an ambient air condition of 35 C (95 F) dry bulb and 24 C (75 F) wet bulb, evaporative condensers can condense subcritical CO 2 at 30 C (86 F) and 7.2 MPa (1044 psia) and can cool supercritical CO 2 fluid to 31 C (88 F) with an ambient wet bulb temperature of 28 C (82 F). We show that water consumption of an evaporative condenser can be significantly reduced by a dry coil that removes a large part of the unusually high proportion of sensible heat from CO 2 condensing at 30 C (86 F). Similar water saving can also be achieved in cooling supercritical CO 2. Other benefits of applying evaporative condensers to subcritical CO 2 are examined. They are lower design pressures compared to transcritical operations, lower energy consumption with attendant lower carbon emissions, and lower owning and operating costs. Hot gas defrosting would also become a standard feature of subcritical CO 2 refrigeration plant operations. An examination of the climatic conditions for the United States in Table 1 Chapter 24 of the 1989 ASHRAE Fundamentals Handbook shows that most of the USA defined approximately by a line running South from South Dakota/Nebraska border through the Colorado/Nebraska to the Mexican border with the northern boundary being a line West to East along the latitude of the South Dakota/Nebraska border, and including the Atlantic states South of New Jersey, has 1% summer design wet bulb temperature conditions ranging from 25 to 27 C (77 to 81 F) with Biloxi, MS and Technical Paper #7 IIAR

6 2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA Alice and Beeville, TX as high as 28 C (82 F). The rest of the USA and Canada have no locations with a 1% wet bulb design above 24 C (75 F) and therefore evaporative condensers will facilitate operating CO 2 systems subcritically all the time unless heat is required, in which case the AC/parallel and or high stage compressors may be selected to operate in trans-critical mode. Furthermore, an examination of average climate conditions using the website and Chapter 24 of the ASHRAE Fundamentals Handbook shows that virtually all of Europe, including Spain, Italy, Greece and Turkey, has a climate where evaporative condensers may be applied for the condensing of CO 2 at subcritical conditions at a condensing temperature of 30 C (86 F) or lower 100% of the time. CO 2 at a saturation temperature of 30 C (86 F) can be condensed in an evaporative condenser, but there are special problems with CO 2. When entering at a discharge temperature of, for example, 77 C (171 F) and condensed to 30 C (86 F), approximately 68% of total rejection is desuperheating sensible heat. Consequently CO 2 condenser design requires a subtle approach, since a) CO 2 flows inside tubes and temperature and phase conditions change along tubes, b) the vapor phase heat transfer coefficients are much lower than condensing phase heat transfer coefficients, and c) properties such as heat capacity, thermal conductivity and viscosity change much more with temperature than other refrigerants. Because the sensible heat fraction is so high with CO 2, a hybrid exchanger can deliver reduced water consumption at a reasonable cost. A dry cooler mounted on top of the evaporative condenser receives moist air at a temperature a few degrees higher than ambient wet bulb. Heat is exchanged from hot CO 2 gas entering at the compressor discharge temperature. Consequently, although air-side heat transfer 4 IIAR 2015 Technical Paper #7

7 Applying Evaporative Condensers to Subcritical CO 2 Condensing and Transcritical CO 2 Gas Cooling coefficients are low, the temperature difference between CO 2 and the ambient air is good. If a dry cooler reduced CO 2 temperature from 77 C to 33 C (171 F to 91 F) then heat rejection would be evenly split between the dry cooler and the evaporative condenser, thus approximately halving water consumption. Since the revival of CO 2 refrigeration started 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) has been applied almost universally. This has resulted in many CO 2 refrigerating systems needing to run in transcritical mode because the cooling air temperature is close to or exceeds the CO 2 critical temperature of 31 C (88 F) in many cases. More often than not the summer design CO 2 temperatures for air cooled gas coolers are higher than the critical temperature and this results in the compressors needing to operate at a discharge pressure of 9 MPa or higher to ensure a reasonable COP. The summer design COPs of supercritical CO 2 compressors are generally lower than those of air cooled HFC or evaporatively cooled ammonia systems. As a consequence, to date many CO 2 refrigeration applications have been in the form of CO 2 /HFC cascade systems with air cooled HFC condensing. Some larger industrial plants have employed CO 2 /NH 3 cascades with evaporative NH 3 condensing. The obvious solution is to reduce the temperature of the condenser cooling medium to allow a complete subcritical CO 2 refrigeration cycle. This is easily accomplished with an evaporative condenser, where the effective cooling medium temperature is the ambient air wet bulb temperature rather than the ambient air dry bulb temperature in the case of an air-cooled condenser or gas cooler. Disadvantages would be the need for a water supply, consumption, and treatment, and control of a minimum condensing temperature as currently mandated by some compressor suppliers. Technical Paper #7 IIAR

8 2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA Air cooled gas cooling invariably results in gas cooler exit temperatures well above the critical temperatures at summer design conditions with the consequence that systems are designed and operated as transcritical systems at pressures well above the critical point to compensate for the deleterious effect of a high gas cooler exit temperature at pressures close to the critical pressure. This in turn has led to a lot of research into expanders to recover some of the energy by compressing CO 2 vapor. The latest development in this pursuit are ejectors. Both expanders and ejectors are rendered obsolete by evaporative CO 2 condensers and gas coolers. Using discharge air from the evaporative condenser as the cooling supply air to the air cooled CO 2 finned coil virtually eliminates the danger of Legionella bacteria contamination, an unexpected bonus from a hybrid evaporative condenser. However, it must be said that the Legionella danger is grossly exaggerated and an entirely manageable problem. It will be shown that the application of evaporative CO 2 condensing and gas cooling in CO 2 refrigerating systems with parallel compression results in higher COPs than COPs of all conventional HFC, ammonia and hydrocarbon compressors. Furthermore, existing CO 2 /HFC or CO 2 /NH 3 cascade systems may be converted to multifunction two stage transcritical CO 2 refrigerating systems when HFCs need to be phased down by 79% below mid 2013 levels by 2030 according to a G20 agreement. The Montreal Protocol has been requested to accept this onerous and difficult task after their astonishing success with the CFC phaseout by 2010! 6 IIAR 2015 Technical Paper #7

9 Applying Evaporative Condensers to Subcritical CO 2 Condensing and Transcritical CO 2 Gas Cooling Rating Model for a Subcritical CO 2 Hybrid Condenser Rating example Figure 1 is a schematic flow sheet of a hybrid condenser that we describe in this section. Water is recycled over the tube bank, so spray water temperature is the same as the basin water temperature. Humidified air passes through eliminators (not shown) to the dry cooler. Specified parameters are: a) total rejection is 600 kw (2047 MBtu/h), b) air face velocity is 3 m/s (9.8 ft/s) at 35/24 C (95/75 F) dry/wet bulb temperature, c) spray water mass flux is 3.0 kg/m2 s (0.61 lb/ft 2 s), d) bundle dimensions as shown, e) entering CO 2 is superheated vapour at 77 C (171 F) and 7.2 MPa (1044 psia) f) leaving CO 2 is saturated liquid at 30 C (86 F) and 7.2 MPa, g) CO 2 leaves the dry cooler at close to 36 C (97 F). Specifying total rejection and CO 2 temperature leaving the dry cooler (g) means heat transferred in the evaporative condenser section and in the dry cooler section is fixed so the program is used to try different bundle length or circuits or passes until calculated values closely match the specified design values. Technical Paper #7 IIAR

10 2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA Figure 1. Hybrid condenser schematic diagram Mass and energy balances in evaporative coolers Among others, Qureshi (2006) and Heyns (2009) have published five simultaneous nonlinear differential equations describing air water process fluid interactions in evaporative cooling. The equations are listed in Table 1. Note that in Equation (4) the overall heat transfer coefficient is defined using the difference between water temperature on the tubes and CO 2 temperature inside, not ambient air to CO 2 temperature. The air water exchange is calculated using a mass transfer coefficient for water vapor moving through air, Equation (3). 8 IIAR 2015 Technical Paper #7

11 Applying Evaporative Condensers to Subcritical CO 2 Condensing and Transcritical CO 2 Gas Cooling (1) (2) (3) (4) (5) (6) Table 1. Differential equations for heat and mass transfer in an evaporative exchanger We believed that to model heat transfer from CO 2 we needed to use a numerical solution that would allow for the unusually large change in gas properties with temperature as the gas flows through the dry cooler and evaporative condenser. For example, Figure 2 is a plot of isobaric NIST data (Lemmon, 2014) for vapor phase Prandtl numbers over temperatures from saturation at 30 C to 80 C (86 F to 176 F). Even larger variations occur with supercritical CO 2 ; see Figure 9. Technical Paper #7 IIAR

12 2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA Vapour Prandtl number vs temperature for dtl subcri:cal CO2 at 7.2 MPa (30 C satura:on) O CO2 2 temperature, C C Figure 2. Prandtl number vs temperature for isobaric subcritical CO 2 saturated at 30 C (86 F) We described in an earlier IIR conference paper, Visser (2014), how the evaporative condenser equations were solved. We constructed a program using a fourth order Runge-Kutta numerical solution of differential equations in which pass length was divided into forty equal intervals and properties. Heat transfer rates and pressure drop were calculated in each interval. The evaporative condenser solution is necessarily trial and error because the basin water temperature at air entry has to be guessed and adjusted iteratively until it is the same as the calculated water temperature at the air exit. The solution process is very unstable often final changes in the trial water temperature needed to be just one millionth of a degree. The solution proceeds backwards along a tube pass from CO 2 exit to entry, starting at the air inlet with saturated liquid refrigerant at 30 C (86 F), and proceeds up, as if heating, ending with superheated vapour at a design intercooling temperature. The program has to allow for both two-phase condensation and single phase vapor desuperheating. The dry cooler performance can be calculated when temperature and humidity of air leaving the evaporative condenser are known. The mathematical model is much simpler, essentially 10 IIAR 2015 Technical Paper #7

13 Applying Evaporative Condensers to Subcritical CO 2 Condensing and Transcritical CO 2 Gas Cooling m r C pr ΔT r = UAT(T r T a ) =m a C pa ΔT a To calculate T r and T a along a pass we used another Runge-Kutta routine with the gas in a cooler pass moving through twenty vertical air chimneys ; CO 2 enters at the uppermost tube pass. Our model allows properties to change with temperature and avoids crossflow issues. Typically with aluminium fins the fin efficiency is 55%. Verifying the model There is no analytical solution to the five equations by which the numerical solution can be validated. However two findings give comfort; when the leaving and entering water temperatures are equal, (a) the CO 2 enthalpy change is equal to the moist air enthalpy change, and (b) the heat duty calculated for other refrigerants condensing at 30 C (86 F) is within 9% of duty computed using the simplified Merkel model (Merkel,1926) which is based on constant condensing temperature. Model predictions for subcritical CO 2 With the evaporative condenser/dry cooler conditions shown in Figure 1, Figures 3, 4 & 5 show calculated conditions as CO 2 flows along a tube pass. In Figure 3 CO 2 enters the evaporative condenser at interval zero as vapor at 36 C (97 F) with low heat capacity, CO 2 leaves as liquid at interval 40. The rapid heat capacity change around interval 18 is where CO 2 is close to its critical point; heat capacity is constant in twophase condensation between intervals 19 to 40. Technical Paper #7 IIAR

14 2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA Water temperature, C CO2 heat capacity, kj/kg K Interval number along a tube pass Figure 3. Evaporative condenser: water temperature and CO 2 heat capacity along a tube pass CO2 temperature, C 2 Uo, overall heat transfer coeff., W/m2 K Interval number along a tube pass Figure 4. Evaporative condenser: CO 2 temperature and overall heat transfer coefficient 12 IIAR 2015 Technical Paper #7

15 Applying Evaporative Condensers to Subcritical CO 2 Condensing and Transcritical CO 2 Gas Cooling Figure 4 illustrates how the gas properties affect the heat transfer coefficient and Figure 5 shows computed CO 2 temperature and water temperature along a tube pass. In spite of the change in CO 2 properties and overall coefficient around interval 18, the gas temperature gradient is smooth. Water temperature, C CO2 temperature, C Interval number along a tube pass Figure 5. Evaporative condenser: CO 2 and water temperatures along a tube pass Conditions in the dry cooler are shown in Figure 6. There is considerable but smooth change in both CO 2 heat capacity and inside heat transfer coefficient [gas at 77 C (171 F) enters the dry cooler at interval zero]. The gas condition in the dry cooler is always desuperheating; there is no condensation. Technical Paper #7 IIAR

16 2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA CO2 temperature, C CO2 heat capacity, J/kg K Interval number along a tube pass Figure 6. Dry cooler: CO 2 temperature and heat capacity along a tube pass Water consumption Water consumption with no dry cooler was computed; Figure 7 is a schematic diagram of an evaporative condenser with no dry cooler that has similar design conditions as the hybrid of Figure 1, namely 600 kw rejection, CO 2 entering at 77 C (171 F), leaving at 30 C and same air face velocity, dry bulb temperature and wet bulb temperature; the difference is that CO 2 is transcritical at 7.5 MPa. Water consumption is 1,043 kg/h, compared to 687 kg/h with the dry cooler. 14 IIAR 2015 Technical Paper #7

17 Applying Evaporative Condensers to Subcritical CO 2 Condensing and Transcritical CO 2 Gas Cooling Figure 7. Evaporative condenser (no dry cooler) with 600 kw rejection To show the effect of design value for CO 2 temperature leaving a dry cooler on water consumption, some calculations are shown in Table 2 for subcritical CO 2 leaving the dry cooler at 32, 34, 36 and 38 C (90 F, 93 F, 97 F, and 100 F). Water consumption and evaporative condenser surface area increase by 28% and 22% respectively from 32 to 38 C (90 F to 100 F), but big effect is on dry cooler surface area; it dramatically increases as the CO 2 dry cooler leaving temperature decreases (see figure 8) because of reducing heat transfer temperature difference. Technical Paper #7 IIAR

18 2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA CO 2 leaving dry cooler Dry cooler Evaporative condenser C Rejection, kw Surface, m2 Rejection, kw Surface, m2 Water consumption, kg/h Table 2. Water consumption at different CO 2 temperatures leaving the dry cooler EvaporaCve condenser area, m2 Dry cooler area, m Temperature of CO2 leaving the dry cooler, C Figure 8. Dry cooler and evap. condenser surface area at various CO 2 temperatures leaving the dry cooler Table 3 summarizes the calculated performance of the two cases, both rejecting 600 kw from CO 2 condensing at 30 C (86 F) but one having no dry cooler. Water consumption is significantly lower with the dry cooler, as is spray water flow, though fan power is nearly doubled. 16 IIAR 2015 Technical Paper #7

19 Applying Evaporative Condensers to Subcritical CO 2 Condensing and Transcritical CO 2 Gas Cooling Evaporative Evaporative condenser condenser with dry cooler Reduction Water consumption kg/h % Spray water flow kg/s % Air flow m 3 /s % Fan power (70% efficiency) kw % Wet bundle Rejection kw % Surface area m % CO 2 pressure drop kpa % Air pressure drop Pa % Dry bundle Rejection kw Surface area m CO 2 pressure drop kpa 0 45 Air pressure drop Pa Table 3. Performance with and without a dry cooler Rating Model for a Supercritical CO 2 Hybrid Condenser Supercritical CO 2 can be cooled below its critical temperature [approximately 31 C (88 F)] in an evaporative exchanger (refer to figure 7). Properties of supercritical CO 2 change greatly at the critical temperature as illustrated in Figure 9 for gas pressure of 7.5 MPa (1088 psia); our numerical solution by Runge-Kutta routine treats CO 2 as a liquid phase because there is no condensation. In this section we show examples of cooling 7.5 MPa supercritical CO 2 from 77 C to 28 C (171 F to 82 F) with different values of the design temperature leaving the dry cooler [rejection is evenly split between evaporative condenser and dry cooler if the CO 2 leaves the dry cooler at 33.4 C (92 F)]. Technical Paper #7 IIAR

20 2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA 30 Prandtl number of supercri8cal CO2 at 7.5 MPa CO2 temperature, C 2 Figure 9. Prandtl number vs temperature for supercritical CO2 at 7.5 MPa Because 28 C (82 F) is below the critical point, gas properties change considerably in the evaporative condenser; Figure 10 shows the overall coefficient rising and falling over only about 1 C (34 F) change in CO 2 temperature between intervals 10 to 20. W/m2 K Overall heat transfer coefficient Interval along a tube pass CO2 temperature C Figure 10. Evap. condenser: supercritical CO 2 temperature and overall heat transfer coefficient 18 IIAR 2015 Technical Paper #7

21 Applying Evaporative Condensers to Subcritical CO 2 Condensing and Transcritical CO 2 Gas Cooling Table 4 shows calculations for 600 kw rejection in cooling supercritical 7.5 MPa CO 2 from 77 to 28 C (171 F to 82 F). The trends in surface area are similar to Table 2. Evaporative condenser rejection and surface area is practically identical but dry cooler surface area is lower with supercritical CO 2. Surface areas in both tables are not optimised in any way any final design must find a cost balance between number of tube rows and number of tube circuits. For example, increasing tube circuits increases surface area but reduces air-side velocity which reduces air heat transfer coefficient while reducing both air and CO 2 pressure drop. CO 2 leaving dry cooler Dry cooler Evaporative condenser Water C Rejection, kw Surface, m 2 Rejection, kw Surface, m 2 consumption, kg/h Table 4. Supercritical CO 2 : water consumption at different CO 2 temperatures leaving the dry cooler Effect of operating conditions on performance Referring to Figure 11, it is clear that the proportions of sensible heat and latent heat change from 100% sensible at the critical temperature of 31 C (88 F) to 22% sensible at saturation temperature of 16 C (61 F). At 30 C (86 F) saturation, the proportion of sensible heat is 68% and the latent proportion is only 32%. This creates a complex situation when trying to predict the portion of latent heat removed in the condensing section with reducing ambient dry and wet bulb temperatures and reducing air entry temperature into the dry cooler section in the first stage of heat extraction in the air to CO 2 counterflow arrangement. In the interest of reducing water consumption Technical Paper #7 IIAR

22 2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA consistent with desirable low energy consumption this is an area worthy of attention. This is a classic example of the energy-water nexus. 100 Total heat rejec on to condenser, 100% Percentage heat rejected to condenser Calculated percentage propor on of latent heat of condensa on of total heat rejected at CO 2 condenser Calculated percentage propor on of sensible heat of total heat rejected at CO 2 condenser CO 2 Saturated Condensing Temperature, C Figure 11. Heat rejection profile based on a commercial CO 2 compressor at 5 C saturated suction, with 5 K useful suction superheat and 5 C CO 2 liquid temp. Tables 5 and 6 show the performance of a hybrid unit with zero water spray. In Table 5, rejection capacity declines from 619 to 230 kw with increasing ambient temperature from +1 to +20 C (34 F to 68 F) at a constant saturated condensing temperature of 30 C (86 F), compressor discharge temperature of 77 C (171 F) and a dry cooler exit temperature of 32 C (90 F). It is fair to say that the dry condensing capacity of CO 2 hybrid condenser is approximately equal to the difference between entering air temperature and condensing temperature. 20 IIAR 2015 Technical Paper #7

23 Applying Evaporative Condensers to Subcritical CO 2 Condensing and Transcritical CO 2 Gas Cooling Ambient, C Heat Rejection, kw Operating Parameters Section Total 1 Sat. cond. temp., C 30 Dry Wet Dry & 2 Comp. discharge temp., C 77 Bulb Bulb Dry Wet Wet 3 Dry Cooler exit temp., C Dry Cooler surface area, m Wet cooler surface area, m Superficial air velocity, m/s Total air pressure drop, Pa 205 Total CO 2 pressure drop, kpa at full capacity 11 Table 5. Hybrid CO 2 condenser heat rejection capacity with increasing ambient temperature and zero water spray Condenser Operating Conditions, C Heat Rejection, kw SCT Comp. Dry Section Total Disch. Exit Dry Wet Dry & Wet Operating Parameters 1. Ambient dry bulb + 1 C 2. Ambient wet bulb, 0 C N.B. Items 4 8 in Table 5 are applicable here also. Table 6. Hybrid CO 2 condenser heat rejection capacity at constant ambient air temperature, reducing condensing temperature and zero water spray Table 6 shows the decline in CO 2 condensing capacity of the hybrid CO 2 condenser operating as an air cooled condenser without water sprays with a constant ambient air temperature of 1 C (34 F), the condensing capacity reduces from 619 kw at 30 C (86 F) SCT [Standard Condensing Temperature] to 256 kw at 10 C (50 F) SCT. In both cases no water is consumed and thus the scope for water saving with hybrid CO 2 condensers exists continually with diurnal variations in required Technical Paper #7 IIAR

24 2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA refrigeration capacity and ambient dry and wet bulb temperatures, and seasonal climatic variations. Careful monitoring of operations is necessary to detect the operating conditions where dry operations are possible to minimize cooling water consumption. Effect of ambient wet bulb temperature on CO 2 condenser performance In Table 7 below, the calculated performance of a nominally 470 kw CO 2 evaporative condenser is shown. The following points are worthy of note. At SCT of 30ºC (86 F), rejection increases by 27% with a 33% increase in the air entry wet bulb to SCT approach from 6 to 8 K; a rise of 67% in entry wet bulb to SCT approach from 6 to 10 K gives an increase of 52% in heat rejection capacity. The increase in CO 2 mass flux is proportional to the capacity increase and this would enhance the internal heat transfer in the condensing bundle tubes. This is more than offset by the reducing enthalpy rise per K (Δh/K) as the wet bulb temperature reduces. Total heat rejection Sat. Cond. Temp. Ambient air condition Leaving air condition Mass flux G kg/m 2.s Pressure drop Fan Gross heat (THR) kw (SCT) C DB C WB C DB C RH % CO 2 GCO 2 Air Gair Water Gh 2 o CO 2 kpa Air Pa power kw kw/ m Table 7. Effect of ambient wet bulb temperature on condenser performance flux NB. Tube bundle: 84 circuits, 8 passes, 3,635 mm long, OD x 12 mm WT compressor discharge temperature 77ºC (171 F), m 2 22 IIAR 2015 Technical Paper #7

25 Applying Evaporative Condensers to Subcritical CO 2 Condensing and Transcritical CO 2 Gas Cooling When reducing SCT from 30ºC to 16ºC (86 F to 61 F) at a constant entry wet bulb to SCT approach of 10 K, it is clear that rejection reduces from 712 kw to 407 kw, i.e. a 43% reduction whilst the mass flux reduces by 57%. This is due to the greater effective enthalpy rise available for evaporation in the evaporator with 16ºC (61 F) SCT compared to 30ºC (86 F) SCT. The above is all based on zero liquid subcooling. As the ambient air temperature lowers, the air density increases and as a consequence so do the resulting air mass flux and fan energy consumption. The Δh/K per degree wet bulb reduces as the wet bulb temperature reduces. Effect of ambient wet bulb temperature on the heat rejection and water consumption of both single stage direct and hybrid evaporative condensers Figure 12 shows the performance details of a hybrid CO 2 evaporative condenser specified in the schedule below at ambient wet bulb temperatures declining from 28 C to 24 C (82 F to 75 F). Specification for a hybrid condenser for analysis in Figures 12 and 13 Operating parameters and constants Total Heat Rejection Capacity, kw 648 Saturated condensing temperature, C 30 Ambient dry bulb temperature, C 35 Ambient wet bulb temperature, C Fan power, kw 6.3 Technical Paper #7 IIAR

26 2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA Wet evaporative condenser section 1. Tube bundle length, mm 5, Tube 304 SS OD x wall, mm x mm x No. of circuits No. of passes Total outside tube surface area, m Air flow, m 3 /s Superficial air velocity, m/s Air mass flux, kg/m 2.s Water mass flux, kg/m 2.s Air pressure drop, Pa CO 2 mass flux varies with wet bulb Dry cooler section 1. Tube, 304 SS OD x wall, mm x mm 15.9 x Finned length, mm 5, Finned width 22 row at 50mm pitch 1, Finned depth 4 row at 50mm pitch Fin spacing centres, mm Fin material Al 7. Fin thickness, mm Heat transfer surface area, m Air flow, m 3 /s Air pressure drop, Pa IIAR 2015 Technical Paper #7

27 Applying Evaporative Condensers to Subcritical CO 2 Condensing and Transcritical CO 2 Gas Cooling Dry, wet and total heat rejec on of hybrid and pure evapora ve condenser Total heat rejec on capacity hybrid condenser Total heat rejec on capacity of a pure evapora ve condenser Evap. cond. sec on heat rejec on capacity Dry cooler heat rejec on capacity Entry air temperature into dry cooler Exit air temperature from dry cooler t a CO2 entry temperature into dry cooler O Ambient wet bulb temperature, C Temperaure, C - dry cooler air entry & exit Figure 12. Heat rejection capacity of a hybrid CO 2 condenser and a pure evaporative condenser, dry cooler air entry and exit temperatures with reducing ambient web bulb temperature. Figure 13 shows the performance details of a single stage direct CO 2 evaporative condenser specified in the schedule with ambient wet bulb temperatures declining from 28 C to 24 C (82 F to 75 F). Technical Paper #7 IIAR

28 2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA Specification for a CO 2 evaporative condenser for analysis in Figures 12 and 13 Operating parameters and constants 1. Total Heat Rejection Capacity, kw saturated condensing temperature (SCT), C Ambient dry bulb temperature, C Ambient wet bulb temperature, C Fan power, kw 3.6 Wet evaporative condenser section 1. Tube bundle length, mm 5, Tube 304 SS OD x wall, mm x mm x No. of circuits No. of passes Total outside tube surface area, m Air flow, m 3 /s Superficial air velocity, m/s Air mass flux, kg/m 2.s Water mass flux, kg/m 2.s Air pressure drop, Pa CO 2 mass flux varies with capacity and wet bulb 26 IIAR 2015 Technical Paper #7

29 Applying Evaporative Condensers to Subcritical CO 2 Condensing and Transcritical CO 2 Gas Cooling Water consump on, kg/h Evapora ve condenser water consump on, kg/h Hybrid condenser water consump on, kg/h Evapora ve condenser specific water consump on, kg/h.kw Hybrid condenser specific water consump on, kg/h.kw Reduc on in hybrid condenser specific water consump on, kg/h.kw Ambient wet bulb, C Water consump on & water saving, kg/h.kw Figure 13. Variation in total and specific water consumption for both hybrid and evaporative condenser and difference in specific water consumption The two designs have virtually the same capacity at 30 C (86 F) SCT at an ambient wet bulb temperature of 24 C (75 F). Although the air volume of 24.5 m 3 /s in the direct evaporative condenser is 36% greater than the air volume of 18 m 3 /s required for the hybrid unit, the resistance to air flow in the hybrid condenser dry coil is quite high. Consequently, the fan power demand of 3.6 kw of the evaporative condenser is 2.7 kw lower than the 6.3 kw power demand by the hybrid condenser fans. On the other hand, the hybrid CO 2 condenser uses about 0.72 litre less water per kwh heat rejection than the direct evaporative condenser. Table 3 also confirms this. Table 5 shows the dry operation of an hybrid CO 2 condenser at a constant SCT of Technical Paper #7 IIAR

30 2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA 30 C (86 F) with an increasing ambient dry bulb condition from +1 to 20 C (34 F to 68 F). The capacity reduces from 619 kw at 29 K air entry TD to 230 kw at 10 K air entry TD. Table 6 shows the dry operation of the same hybrid condenser as examined in Table 5. In this case, the ambient dry bulb is constant and the SCT is allowed to float from 30 C to 10 C (86 F to 50 F) during which the capacity reduces from 619 kw at 29 K TD to 256 kw at 9 K TD. Given diurnal variations in both heat rejection capacity and ambient conditions as well as seasonal climatic variations, it is clear that in many climates there will be extended periods when a hybrid CO 2 condenser would be able to run without water on the evaporative condensing stage. This would save cooling water but it could well result in excessive energy consumption by maintaining a higher SCT than would be the case with water applied to the evaporative condenser section of the hybrid condenser. Let us take the simple case of the condensers in Figures 12 and 13 operating for 6,000 hours/year at 75% average load conditions. 1. Condenser rejection at design conditions, kw Average load factor, % Average heat rejection to condenser, kw No. of hours/annum 6, Water application time, % Total water saved including 15% bleed, m 3 /annum 1, Total extra fan energy, 6,000 x 3 kw 16, Water cost, $ kl Electrical energy cost $ kwh Water saving benefit $1, Electrical energy cost $3, IIAR 2015 Technical Paper #7

31 Applying Evaporative Condensers to Subcritical CO 2 Condensing and Transcritical CO 2 Gas Cooling In this case the reduction in water costs are about 53% of the extra fan energy cost. We are comparing the water saving benefits and energy consumption of two condenser types. On the surface, this does not make economic sense. However, it highlights the water-energy nexus which is becoming an increasingly important issue. How much energy should one use to save how much water? When considering the much larger benefit of much reduced energy consumption of the CO 2 refrigeration system compared to air cooled HFC and CO 2 /HFC refrigerating systems, the overall economic benefits are clear. Effect of condensing temperature on cycle performance In Figure 14, five COP plots are shown for a commercial semi-hermetic CO 2 compressor with a swept volume of 27.2 m 3 /h at 50 Hz using a 30 kw four pole motor. Superheat is 5 K. COP Saturated condensing temperature, C (1) +10 C SST, 3 K subcooled (2) +5 C SST, 3 K subcooled (3) - 5 C SST,5 C virtual gc out (4) - 5 C SST,10 C virtual gc out (5) - 5 C SST,5 K subcooled Figure 14. COP vs condensing temperature for 27.2 m 3 /h compressor Referring to curve 1, COP ranges from 6.27 at +30 C (86 F) saturated condensing temperature (SCT) to 18.0 at SCT of +16 C (61 F) at a saturated suction temperature Technical Paper #7 IIAR

32 2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA (SST) of +10 C (50 F). A +10 C (50 F) SST would allow an evaporating temperature of 11 C (52 F) with a suction pressure drop corresponding to 1 K boiling point suppression. 11 C (52 F) is a reasonably efficient evaporating temperature for direct cooling of air allowing a relatively large diffusion in air temperature across the cooling coil, thus limiting the volume of air which needs to be circulated, thus reducing fan energy consumption and the resulting parasitic fan heat load. This in turn leads to a reduction in the required energy input into the compressor thereby raising the overall energy efficiency of the system as a whole. Curve 2 shows the COP ranging from 4.45 to at 30 C to 16 C (86 F to 61 F) SCT at an SST of +5 C (41 F). This will allow chilled water production for retrofitting into existing buildings and application to new buildings. In both the above two cases compressors may act also as parallel compressors for refrigeration duties at 5 C (23 F) SST such as maintaining chill storage temperatures at around 0 C (32 F) and high stage duties for two stage CO 2 systems applied to cold storage and blast freezing applications. In such cases, the high stage compressors would operate with virtual CO 2 gas cooler exit temperatures of +5 C (41 F) and +10 C (50 F) which results in COP curves 3 and 4 respectively. Curve 3 ranges from 4.7 to 7.88 at an SST of 5 C (23 F) at SCTs from +30 to +16 C (86 F to 61 F) and a virtual gas cooler exit of +5 C (41 F). Curve 4 shows the COP ranging 4.45 to 7.04 with a virtual gas cooler exit of +10 C (50 F), and SST of 5 C (23 F) and the SCT ranging from +30 to +16 C (86 F to 61 F). This may be improved with a suction heat exchanger in the compressor suction to bring the performance closer to curve 3. In Figure 15 the COPs for transcritical operations at 7.5, 8.0, 9.0 and 10.0 MPa are shown for a commercial transcritical CO 2 compressor working at 10 C saturated suction with a gas cooler exit temperature of 31 C and 28 C (88 F to 82 F) ambient wet bulb temperature when using a hybrid evaporative condenser. These would be the operating characteristics of a transcritical CO 2 compressor for direct DX AC cooling and parallel compression. 30 IIAR 2015 Technical Paper #7

33 Applying Evaporative Condensers to Subcritical CO 2 Condensing and Transcritical CO 2 Gas Cooling 7.5 MPa CO2 pressure 8.0 MPa CO2 pressure 9.0 MPa CO2 pressure 10.0 MPa CO2 pressure O2 O2 O2 O2 COP CO2 exit temperature, C O2 Figure 15. Compressor COP vs CO 2 exit temperature at 10 C SST and 5 K superheat In Figure 16 the COPs for transcritical operations at 7.5, 8.0, 9.0 and 10.0 MPa are shown for a commercial transcritical CO 2 compressor working at 5 C (41 F) saturated suction with a gas cooler exit temperature of 31 C (88 F) and a 28 C (82 F) ambient wet bulb temperature when using a hybrid evaporative condenser. This would be the operating characteristics of a transcritical CO 2 compressor for chilled water cooling for AC and parallel compression. Technical Paper #7 IIAR

34 2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA 7.5 MPa CO2 pressure 8.0 MPa CO2 pressure 9.0 MPa CO2 pressure 10.0 MPa CO2 pressure O2 O2 O2 COP CO2 exit temperature, C O2 Figure 16. Compressor COP vs CO 2 exit temperature at 5 C SST and 5 K superheat Should heat recovery be required, these AC/parallel compressors would be controlled with regulated discharge pressure and suction superheat (SSH) to achieve the desirable discharge temperature for a required hot water temperature. Relative energy efficiency of ammonia, R22, R507A, propane, and R134a The COPs for these refrigerants in a Mycom 8 WB 1,000 rpm compressor operating at -8 C (18 F) SST, 5 K superheat and 5 K subcooling are shown in Figure 17. It confirms that ammonia is the best refrigerant. The big surprise is the low COP of R134a. At SCT s of 16 C and 35 C (61 F and 95 F) the R134a, COPs are respectively 42 and 31% lower than those of ammonia. Furthermore the COP of an R134a compressor at +16 C (61 F) SCT is at 3.84, about the same as the COP of an ammonia compressor at +35 C (95 F) SCT at identical suction conditions. 32 IIAR 2015 Technical Paper #7

35 Applying Evaporative Condensers to Subcritical CO 2 Condensing and Transcritical CO 2 Gas Cooling COP Ammonia R22 R507a Propane R134a Saturated condensing temperature, C Figure 17. COP vs condensing temperature (Mycom compressor, 637 m 3 /h swept volume) This confirms beyond any doubt that R134a has both high direct and indirect Global Warming Potential (GWP). The performance of R507A is 11 to 16% less efficient than R22 at 25 to 35 C (77 F to 95 F) SCT. HFC R507A has no ozone depletion potential like HCFC R22, but the 100 years GWP of R507A is 3,895. This is more than double the 100 year GWP of 1,810 of R22. In Figure 18 we have added curves (3), (4) and (5) from Figure 14 to Figure 17 to get a direct comparison between the CO 2 COPs at +5 C and +10 F (41 F and 50 F) virtual gas cooler exit temperatures and 5 K liquid subcooling. It is clear that at 30 C (86 F) saturated condensing temperature, the COPs with parallel compression at +5 to +10 C (41 F and 50 F) SST CO 2 are 4.48 and 4.73, respectively, which is significantly more efficient than all conventional refrigerants, including ammonia. But the CO 2 condensing at 30 C (86 F) would occur at an ambient wet bulb temperature of 24 C (75 F) at which condition the expected ammonia condensing temperature would be at least 35 C (95 F) with a COP of 4.0. This is considerably lower than the CO 2 COPs of 4.48 to Technical Paper #7 IIAR

36 2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA 8 7 Ammonia R717 HCFC R22 HFC R507A Propane R290 HFC R134a Saturated Suc on Temp.: 8 C Suc on Superheat: 5 K Liquid Subcooling: 3 K 5 C SST, 5K SSH, +5 C virtual gc out (3) from Fig C SST, 5K SSH +10 C virtual gc out (4) from Fig C SST, 5K SSH, 5K Liq. Subc. (5) from Fig COP Saturated condensing temperature, C Figure 18. Comparison of CO 2 COPs at +5 and +10 C (41 F and 50 F) virtual gas cooler exit temperatures and 5K liquid subcooling compared to COPs in Figure 15. In the following Figure 19, 23 maps of the USA and Canada, Latin America, Europe, Asia and Australia are shown with approximate boundaries for 1% incidence Ambient Wet Bulb Design Temperatures (AWBDTs) called Zones. In Zone 1, the 1% AWBDT is 24 C (75 F) In Zone 2, the 1% AWBDT is 25 C (77 F) In Zone 3, the 1% AWBDT is 26 C (79 F) In Zone 4, the 1% AWBDT is 27 C (81 F) In Zone 5, the 1% AWBDT is 28 C (82 F) or higher. 34 IIAR 2015 Technical Paper #7

37 Applying Evaporative Condensers to Subcritical CO 2 Condensing and Transcritical CO 2 Gas Cooling In 98% of the world s locations the AWBDT is 28 C (82 F) or less. In only about 2% of the world s locations (the Persian Gulf, Da Nang in Vietnam and Broome in North West Australia) the AWBDT is 30 C (86 F). But less than 1% of the world s population lives in these hot, humid climates. Figure 19. USA and Canada climate zones with approximate percentage incidence of subcritical CO 2 condensing annually Technical Paper #7 IIAR

38 2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA The big surprise is the USA, which has a warm to hot, humid climate in virtually all of the area East of the Rocky Mountains. Please note all of Canada is in Zone 1, [i.e. all CO 2 refrigerating systems may operate in subcritical mode with AWBDTs not exceeding 24 C (75 F) anywhere]. Figure 20. Latin America climate zones with approximate percentage incidence of subcritical CO 2 condensing annually 36 IIAR 2015 Technical Paper #7

39 Applying Evaporative Condensers to Subcritical CO 2 Condensing and Transcritical CO 2 Gas Cooling In Zone 1 it is possible to operate subcritically 100% of the time at an SCT of 30 C (86 F) at 24 C (75 F) ambient wet bulb. As can be seen from Figure 12 and 16, this enables highly efficient CO 2 refrigerating systems to be operated particularly where parallel compression is employed as part of a high evaporating temperature refrigerated cooling duty like AC. It is also clear from Figures 13 and 14 that transcritical refrigeration operations can also be highly efficient provided that the gas cooler exit temperature is 31 C (88 F) which is readily achievable in an evaporatively cooled direct water contact gas cooler with an AWBDT of 28 C (82 F). This must not be confused with an adiabatically assisted gas cooler, where the cooling air for air cooled gas cooling is pre-cooled before entry into a finned tube coil. Except for two locations in South East Europe, all of Europe, including the Mediterranean, is Zone 1 territory. This is a big surprise! About half the Australian territory is Zone 1. This includes about 90% of the Australian urban areas. Figure 21. Virtually all of Continental Europe can enjoy subcritical CO 2 condensing 100% of the time Technical Paper #7 IIAR

40 2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA Figure 22. Asian climate zones with approximate percentage incidence of subcritical CO 2 condensing annually Figure 23. Australia climate zones with approximate percentage incidence of subcritical CO 2 condensing annually 38 IIAR 2015 Technical Paper #7

41 Applying Evaporative Condensers to Subcritical CO 2 Condensing and Transcritical CO 2 Gas Cooling Discussion The mathematical model and computer program used for this paper was carried out by Mr. John Ball who was commissioned by his co-author Visser, after Visser came to the conclusion that the required mathematical skills were beyond his ability. It has been an extremely difficult exercise which took place over a period of nearly fourteen months starting early October, 2013 and ending November, The results are quite stunning! Overall heat transfer factor, Uo Referring to Figure 4, Uo (overall heat transfer coefficient factor) is considerably higher than the Uo Visser considered in the case of ammonia condensing where Uo ranges from about 450 to 550 W/m 2 K (79 to 97 Btu/ft 2 h) at superficial air velocities of 2.6 to 3.05 m/s. A superficial air velocity of 3 m/s 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 4 for the CO 2 evaporative condenser (in Figure 7) is about 1,050 W/m2 K (185 Btu/ft 2 h). This is more than double the average value for ammonia at virtually the same superficial air velocity entering the condensing tube bundle. This is all the more remarkable, when considering that at 30 C (86 F) condensing, 68% of the heat to be removed is sensible superheat and only 32% is actually latent heat of condensation at 30 C (86 F). The high overall heat transfer factor is attributed to the high CO 2 mass flux of kg/m 2 s in the 48 off 66.7 metre length circuits causing a calculated pressure drop of 15 kpa, the maximum value acceptable to facilitate CO 2 condensers to operate in parallel without requiring too great a drop leg to avoid liquid hold up in operating condensers should one condenser not be operating. Technical Paper #7 IIAR

42 2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA In evaporative condensers, the high ΔP/ΔT ratio of CO 2 allows high mass fluxes in the condenser circuits giving high rates of heat transfer, allowing fewer longer circuits. This also makes for more economical manufacturing of the tube bundle. Ammonia mass fluxes in evaporative condensers range from about 25 to 40 kg/m 2 s and are frequently lower than 25. The pressure drop is a concern with ammonia condensers, as excessive pressure drop in an ammonia evaporative condenser lifts the discharge pressure slightly, and thus the saturated condensing temperature resulting in slightly increased energy consumption. Consequences of minimum airflow Again referring to Figure 7, the calculated leaving air dry bulb temperature is 30.2 C (86 F) at 100% RH and hence the leaving wet bulb temperature is also 30.2 C (86 F). This is close to the CO 2 leaving temperature of 30 C (86 F) but heat transfer rates are reasonable because the top tubes are at a temperature of 77 C (171 F) and the high proportion of sensible superheat lifts the mean temperature difference. This does not happen in ammonia evaporative condensers where minimum leaving temperatures differences between ammonia SCT and leaving wet bulb are rarely less than 3 K and never less than 2.5 K at design conditions in Visser s experience. Lower airflow also results in minimum fan energy consumption. Advantages and disadvantages of CO 2 evaporative condensers and hybrid evaporative condensers and gas coolers Advantages 1. Lower discharge pressure mostly below the critical point. 2. Lower CO 2 exit temperature from the gas cooler with subcooling to a minimum of 3 K approach between the CO 2 saturated condensing temperature and the ambient wet bulb temperature. 40 IIAR 2015 Technical Paper #7