# A Dimensionless Correlation for Heat Transfer

A Dimensionless Correlation for Heat Transfer

A Dimensionless Correlation for Heat Transfer

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CO 2 Project Report-Phase I1. SUMMARYThe project “CO 2 in Supermarket Refrigeration” is collaboration among IUC, KTH/AppliedThermodynamic and Refrigeration division, Ahlsell, Huurre, AGA, WICA and ICA. Theproject is financed by Energimyndigheten.The objective of this project is to develop, test, and evaluate an energy efficient supermarketsystem working with CO 2 as the refrigerant. Based on the experience in designing the system,running and evaluating it, modifications should be applied in order to conclude an efficientoptimized CO 2 system <strong>for</strong> a medium size supermarket in Sweden. Emphasize is on usingenvironmentally friendly refrigerants and the choice was to use natural fluids.A refrigeration system solution <strong>for</strong> a medium size Swedish supermarket has been built in IUClaboratory in Katrineholm. The system is equipped with extensive instrumentations to collectdata and per<strong>for</strong>m online diagnosis. Several variations of the system solution are applied <strong>for</strong>validation and possible modifications.In this report we present the system under investigation and some of the experimental resultsthat have been obtained under the project period. Overall system validation and evaluations ofthe main components are described.2

CO 2 Project Report-Phase I2. FORWARDThis report is a product of the research program Effektivare Kyla which is organized andmanaged by IUC in cooperation with KTH. The program is a shared financing betweenEnergimyndigheten (STEM) and the involved companies.We would like to extend our appreciation <strong>for</strong> the positive spirit of cooperation from all theproject partners. We thank everyone who worked <strong>for</strong> realizing the idea of the project andmaintain its activities.Samer SawalhaArash SoleymaniJörgen RogstamJanuary 20063

CO 2 Project Report-Phase I3. TABLE OF CONTENT1. Summary ........................................................................................................................ 22. Forward .......................................................................................................................... 33. Table of content.............................................................................................................. 44. Introduction .................................................................................................................... 55. Objectives....................................................................................................................... 76. The case study ................................................................................................................ 86.1 System Requirements and Boundaries....................................................................................... 86.2 The System Solution .................................................................................................................. 87. The experimental rig .................................................................................................... 108. Ovearall system analysis .............................................................................................. 128.1 Load Measurements ................................................................................................................. 128.2 Energy Balance Test ................................................................................................................ 12Cascade condenser cooling capacity ........................................................................................ 15Ammonia compressor capacity and efficiencies ...................................................................... 19CO 2 compressor capacity and efficiencies ............................................................................... 248.3 System Efficiency .................................................................................................................... 29Low stage COP......................................................................................................................... 29High stage COP........................................................................................................................ 30Total COP................................................................................................................................. 309. System variations and experimental results ................................................................. 329.1 Pump Circulation Ratio and Gravity Circulation..................................................................... 329.2 Cascade Condenser .................................................................................................................. 379.3 Flashing Gas in Liquid Lines ................................................................................................... 409.4 Freezing Cabinets Control........................................................................................................ 419.5 Safety Tests.............................................................................................................................. 509.6 Flooded Versus Direct Expansion Evaporation ....................................................................... 5010. Theoretical model..................................................................................................... 5110.1 Model Description.................................................................................................................. 5110.2 Overview of Calculations and Assumptions .......................................................................... 5110.3 Design Calculations ............................................................................................................... 5210.4 Some Simulation Results ....................................................................................................... 5311. Conclusions and future work.................................................................................... 5412. References ................................................................................................................ 564

CO 2 Project Report-Phase I4. INTRODUCTIONIn the early stages after the revival of CO 2 as a refrigerant it was applied in low temperatureapplications in supermarkets as a secondary refrigerant in indirect systems. The usage of CO 2in cascade and trans-critical systems <strong>for</strong> this application has been suggested along with theindirect system, but the limitations <strong>for</strong> the application of the cascade and trans-critical systemshave been related to the scarcity of components that can efficiently handle CO 2 . Also therehave been many unanswered questions related to how to handle the highly pressurized systemand how safe it is to deal with CO 2 in this application. In the indirect system it was possible touse conventional components to handle CO 2 ; this is mainly due to the low operating pressurein the indirect loop at low temperature applications, 12 bars at -35°C.The pumping power needed <strong>for</strong> CO 2 in the indirect system is very small compared toconventional brine systems due to the small volume flow rate and pressure drop of CO 2 in thecircuit. The small volume flow rate is a result of the phase changing process on the CO 2 side,which also contributes to having small pressure drop in the pipes and heat exchangers.Gaining experience and confidence in working with CO 2 in indirect systems combined withthe knowledge that is gained through extensive research work on CO 2 in mobile airconditioning and hot water heat pumps brought the attention of the industry <strong>for</strong> the necessityof producing components which are specially designed, or modified, to handle CO 2 . As aresult, cascade and trans-critical systems became reasonably applicable.In supermarket applications the difference between evaporating and condensing temperaturesis large, there<strong>for</strong>e, the cascade or other two-stage systems become favorable and they are welladaptable <strong>for</strong> the two-temperature level requirement <strong>for</strong> chilled and frozen products in thesupermarket. The indirect system requires an additional heat exchanger (primary refrigerantevaporator/CO 2 condenser), which implies that there is an additional temperature differenceacross the heat exchanger and the resulting evaporating temperature will be lower than if adirect expansion had been per<strong>for</strong>med. In cascade or multi-stage CO 2 systems, CO 2 evaporatesdirectly in the evaporators of the display cases, which minimizes the required temperature liftand reduces the energy consumption.In the direct expansion solution, the CO 2 pump is not required; despite the fact that the powerconsumption of the pump in the CO 2 indirect solution is generally very small compared toconventional brine pumps and relative to the total power consumption of the entire system,still the elimination of the pump is advantageous to reduce the installation cost. Practically thepumping power usually is higher than necessary as it is difficult to find CO 2 pumps to matchmedium capacity systems and as the pumps usually are larger than needed a bypass line isintroduced to obtain the required CO 2 flow rate in the evaporators.The low critical point <strong>for</strong> CO 2 of 31ºC implies that it will operate with better theoretical COPbetween low temperature ranges further below the critical point compared to high stageconditions. There<strong>for</strong>e, the application of CO 2 in the low stage of a cascade system yields areasonable theoretical COP compared to other refrigerants, see Figure 1a and Figure 1b. InFigure 1b, the COP values above the critical point are obtained at optimum pressure on thehigh pressure side.5

CO 2 Project Report-Phase I1210COP11109876Evaporating Temperature=-35 °CR134aAmmoniaR404ACO2COP987654Evaporating Temperature=-10 °CAmmoniaR134aR404ACO253423-15 -10 -5 0 5 10 15Condensing temperature (°C)(a)115 20 25 30 35 40 45Condensing temperature (°C)Figure 1 COP of CO2 compared to other main refrigerants in low stage (a) and in high stage (b) operations(b)Several factors contribute to improve the COP of CO 2 . The favorable thermophysicalproperties of CO 2 results in low pressure and temperature drops in the system. From the heattransfer point of view, the low surface tension will make boiling easier and there<strong>for</strong>e willimprove the heat transfer. Also, due to the low pressure drop, (Zhao, Molki et al. 2000), thecomponents of the system will be smaller while the mass flow rate of the refrigerant will becomparable to R404A, R22, R502 and R134a refrigerants, which will result in high mass fluxof CO 2 in the heat exchangers. Another improvement to the COP comes from the improvedvolumetric efficiency of the CO 2 compressor compared to conventional refrigerants; this isdue to the lower pressure ratio across the CO 2 compressor.6

ОБОЗРЕНИЕ ПСИХИАТРИИ И МЕДИЦИНСКОЙ ПСИХОЛОГИИ № 2, 2012Исследованияпредположить, что при использовании эти шкалбудут как ложноположительные, так и ложноотрицательныерезультаты. Следует отметить, чтоданная проблема свойственна всем самоопросникамв целом [65]. Большинство шкал, рассматриваемыхв данном исследовании, имеют хорошиепоказатели конкурентной валидности с другимиметодиками определения депрессии. BDI и EPDSтакже имеют хороший уровень корреляции сошкалами тревоги, что важно, учитывая тот факт,что тревога является одним из основных симптомовпослеродовой депрессии [66]. Выбор скрининговогоинструмента, как мы уже отмечали,является сложной задачей. В связи с этим важнорассмотреть результаты тех исследований, которыепопытались оценить эффективность и полезностьсовместного применения разных шкал. Так,при совместном использовании GHO-12 и EPDSэффективность обеих шкал была выше по сравнениюс применением их по отдельности [44]. Такжеодновременное применение нескольких шкалувеличивало время и трудозатраты исследования,что нежелательно для проведения исследованийв послеродовом периоде. Поэтому всегда предпочтительноиспользование для скрининга минимальногоколичества шкал. Также способствуетпроведению более эффективной диагностикиопределение групп риска по возникновению послеродовойдепрессии.Как показал обзор различных исследований,наиболее изученной и доступной для примененияна разных языках является шкала EPDS. К томуже данная методика имеет достаточно хорошиепоказатели чувствительности и специфичности,что подтверждалось при проведении исследованияу российских женщин. В нашем исследованиичувствительность и специфичность даннойшкалы имели умеренное и хорошее значение.В то же время ПЦПР была достаточно невысокой,что, возможно, связано с малой выраженностьюдепрессивной симптоматики в изучаемойгруппе и требует дальнейшего изучения. Такжедоcтаточно перспективным видится применениеPDSS и BPDS.Другим важным вопросом является выбороптимального времени для проведения исследования.Как упоминалось ранее, согласно критериямDSM-IV послеродовой считается депрессия,развивающаяся в первые четыре неделипосле родов. У большинства женщин (до 85 %)отмечаются явления послеродового блюза. Данноесостояние является преходящим и не считаетсяотклонением от нормы [47]. Проведениедиагностики оправдано начиная со второй неделипосле родов. На более ранних сроках результатыбудут недостоверными [67]. В ряде научныхработ исследователи изучали послеродовойпериод до 18-го месяца после родов [16, 68, 27,52]. Было отмечено, что со временем уменьшаетсяколичество общесоматических симп томовпослеродовой депрессии и, соответственно, наболее поздних сроках может быть использованболее широкий спектр диагностических методик.В целом целесообразным представляется проведениедиагностического исследования начинаясо второй недели после родов и далее на разныхэтапах послеродового периода.Не менее важен вопрос оценки результатов, полученныхв интернациональных межкультуральныхисследованиях. В частности, при проведенииисследований в Австралии и Европе средний баллпо EPDS и BDI был существенно ниже по сравнениюс данными, полученными в Азии и ЮжнойАмерике, где средний балл по изучаемым шкаламбыл максимально высоким [16]. Также могут различатьсяданные, полученные у женщин из разныхкультурных групп в одной стране. Это делаетнеобходимым установление оценочных норм длякаждой страны [17, 49].В целом, сравнивая полезность примененияразличных психометрических шкал для скринингапослеродовой депрессии, можно отметить болеевысокие показатели надежности шкал специальноразработанных для скрининга ПРД (EPDS,PDSS, BPDS) по сравнению со шкалами, традиционноприменяемыми для диагностики депрессивнойсимптоматики. Особенно если речь идето диагностике на ранних сроках послеродовогопериода. Со временем, по мере уменьшения влиянияродов и связанных с этим физиологическихизменений в организме женщины, список диагностическихинструментов расширяется. Показателиэффективности методик как специфичныхдля послеродового периода, так и не специфичныхстановятся вполне сопоставимыми. Каждаяиз перечисленных выше шкал имела как свои достоинства,так и недостатки. «Золотого стандарта»диагностики послеродовой депрессии сегодня несуществует.В настоящее время сохраняется необходимостьпроведения дополнительных исследований,направленных на изучение клинических особенностей,нозологической принадлежности депрессивнойсимптоматики в послеродовом периоде иразработка комплексных подходов к диагностикепослеродовой депрессии.Литература1.2.Milgrom J., Gemmill A.W., Bilszta J.L., HayesB., Barnett B., Brooks J., Ericksen J., Ellwood D.,Buist A., 2008. Antenatal risk factors <strong>for</strong> postnataldepression: a large prospective study. J. Affect. Disord.108 (1–2), p. 147–157.Field T. (1997). The treatment of depressedmothers and their infants. In: Murray E.L.,3.Cooper P.J. (eds). Postpartum depression andchild development. Guil<strong>for</strong>d Press, New York,p. 221–236.Teti D.M., Gelfand D.M. (1991). Behavioral competenceamong mothers of infants in the first year:the mediational role of maternal self-efficacy. ChildDev. 62, p. 918–929.47

CO 2 Project Report-Phase I6. THE CASE STUDY6.1 System Requirements and BoundariesThe system solution under investigation has been chosen to be applicable to a medium sizesupermarket installation in Sweden. The refrigeration system in the supermarkets in Swedenusually operates to satisfy the evaporating temperatures to maintain products at twotemperature levels, around +2ºC <strong>for</strong> cold food and -18ºC <strong>for</strong> frozen products. Despite the lowambient temperature in Sweden the condensing temperature is usually kept constant aroundthe year at a value of about 40 ºC. The cooling capacities of medium size supermarkets aretypically around 50 kW <strong>for</strong> freezing and 150 kW <strong>for</strong> cooling at the medium temperature level.This estimate is based on contacts with major installers of supermarket systems in Sweden.Accordingly, the system has been designed to operate between the temperature boundariesmentioned above and to provide a cooling capacity which is scaled down while trying to keepa load ratio of about 3. The low temperature side has a rated capacity of 7.4 kW and themedium temperature side was designed to have a capacity of 20 kW.6.2 The System SolutionIn the choice of the CO 2 system the aim was to develop an efficient system with good coolingper<strong>for</strong>mance, safe and in accordance with the regulations on the use and release ofrefrigerants. From environmental point of view, the amount of synthetic refrigerants that canbe used to fill the system is limited and the high taxation to prevent leakage makes itexpensive to use. For natural refrigerants such as ammonia and propane there is always thesafety concern when used in applications where people might be exposed to a leakageaccident. CO 2 is considered as a relatively safe refrigerant and classified in group A1according to ASHRAE Handbook-Fundamentals (1997). CO 2 gas that is used in refrigerationis a by-product of the chemical industry and its use in the refrigeration system can beconsidered as a delayed step be<strong>for</strong>e its unavoidable release to the environment. There<strong>for</strong>e,CO 2 becomes an interesting solution as a refrigerant from environmental and safety points ofview especially in supermarket refrigeration systems where large quantities of refrigerant arerequired and direct contact, in case of leakage accident, with large number of people mightoccur.The system that has been chosen is a cascade system with NH 3 at the high stage and CO 2 atthe low stage, at the medium temperature level CO 2 is pumped to provide the required coolingload. Figure 2 is a schematic diagram of the CO 2 circuits in the system.The usage of the cascade system offers the possibility of utilizing two different refrigerantswhere each refrigerant is selected to fit the operating range. Using NH 3 in the high stagemeans that it will be easy to deal with a leakage accident as ammonia can only leak into themachine room which should be equipped with proper safety devices. Using CO 2 in the lowstage results in reasonable operating pressure levels in the CO 2 circuit, 28 bars at -8°C. Thefavorable pressure drop characteristics of CO 2 suits this application where long distributionlines are usually needed. Also this implies that the size of the distribution lines is also smallerthan <strong>for</strong> other refrigerants which reduces the cost of the piping system.8

CO 2 Project Report-Phase IThe evaporators at the medium temperature stage are flooded with CO 2 which is circulatedvia a pump; this is expected to produce better per<strong>for</strong>mance due to the good heat transfercharacteristics of the completely wetted evaporator and there<strong>for</strong>e the evaporator temperaturewill be higher than if a direct expansion concept has been used.Figure 2 Schematic diagram of NH3/CO2 cascade system with CO2 at the medium temperature level9

CO 2 Project Report-Phase I7. THE EXPERIMENTAL RIGThe 20 kW design cooling capacity at the medium temperature level is divided over twodisplay cabinets with 5 kW each and the other 10 kW’s are supplied by an electric heater inwhat is referred to as a load simulator. The installed electric heater at the medium temperaturelevel managed to provide a maximum of 6.6 kW, which makes the maximum total load at themedium temperature level reduces to 16.6 kW. The electric heater load can provide three loadsteps of 2.2, 4.4 and 6.6 kW.On the deep freeze side, the load is divided over two freezers with 2.5 kW each and theelectric heater can provide a maximum load of 3 kW. The electric heater load can be providedon three equal load steps in a similar way to the medium temperature load simulator. Themaximum cooling capacity of the compressor is 7.4 kW. The freezers are equipped withelectronic expansion valves.The compressor is a Copeland scroll type with operating temperatures between -37°C and -8°C and a displacement of 4.1 m 3 /h. The accumulation tank has a capacity to contain 180 L ofCO 2 and is equipped with an electronic level indicator. It can stand a pressure up to 40 barswhich corresponds to an operating temperature of about 6 °C. The system is equipped with asafety release valve that is triggered when the pressure in the system reaches 38 bars. Toavoid the opening of the release valve and the loss of significant charge from the system, ableed valve is installed which opens <strong>for</strong> periodical release of CO 2 at lower pressure than theset value <strong>for</strong> the release valve, 35 bars, so the pressure in the system will be reduced. If thepressure increase in the system is higher than the rate that the bleed valve can handle, then therelease valve will open and release the system’s charge.The CO 2 pump that is used is a hermitic one with capacity higher than the highest circulationrate desired; there<strong>for</strong>e a by-pass is used to reduce the flow rate pumped into the mediumtemperature circuit. About 1.5 meters head over the pump is respected to prevent cavitation.The ammonia unit uses a Bock reciprocating compressor with displacement of 40.5 m 3 /h; itcan run at 50% reduced capacity by unloading half of its cylinders. <strong>Heat</strong> is removed from theammonia evaporator via a thermosyphon loop which required a certain height of the unit. Thecapacity control of both compressors is achieved by a frequency converter.The cascade condenser is a plate type heat exchanger that is specially selected to handle thepressure difference that will exist between CO 2 and ammonia, at -8 °C CO 2 will have about28 bars while ammonia will have a pressure of about 2.7 bars at -12 °C.The medium temperature display cabinets are defrosted using the conventional electric defrostmethod. The deep freeze cabinets will be defrosted using hot gas defrost by passing the hotgas through the cabinet, since the condensing temperature of the CO 2 at the compressordischarge pressure will be low then heating of the evaporators will be achieved via thesensible heat of the hot gas.Figure 3 is a detailed schematic of the test rig with most of the measuring points indicated. Ascan be seen in the schematic several by-pass lines and the high number of valves indicate the10

CO 2 Project Report-Phase Ipossible variations in the system which will be used <strong>for</strong> testing and modifications. The loadsimulators with the electric heaters can be seen in the diagram in the medium and lowtemperature circuits. The electric heater provides heat to a brine loop which exchanges theheat with the refrigerant in a plate heat exchanger.The schematic shows that one of the freezers is electrically defrosted while on the rig bothfreezers are equipped with the hot gas defrost.SafetypipetttpMpLeakage testpointChargingpointptttptttdPtHighe andlow levelcutoutMinimumflow nozeltp6,6 kW elOil draintEl Defrost5 kWtdPtttpMEl Defrost5 kWtdPttppttttdPdPdP3 kW eltt2.5 kWEl Defrost2.5 kWHot gasDefrostFigure 3 Detailed schematic diagram of the NH3/CO2 cascade system test rig11

CO 2 Project Report-Phase I8. OVEARALL SYSTEM ANALYSISThe system under investigation is a scaled down real installation where the main discussion isweather or not this system is a suitable replacement <strong>for</strong> traditional technologies. In order toprovide answers about the current system solution, it is important to per<strong>for</strong>m the overallanalysis of the system where the capacities are properly measured, the energy balance isverified, and the system’s efficiencies are calculated according to the measurements. This isan important step since some of the measurements are based on the components data andplanned experiments depend on the accuracy of measuring cooling loads and capacities ofsome of the main components.8.1 Load MeasurementsThe two compressors are used to determine the mass flow rate of the refrigerant which willthen be used to calculate the cooling capacities in the corresponding circuits. The compressormanufacturer data have been used as guidelines <strong>for</strong> the calculations which are based onknowing the geometry and the efficiencies of the compressors at certain operation conditions.Measuring the rotational speed of the compressor, and the temperatures and pressures aroundit gives all the data needed to calculate the mass flow of the refrigerant. Consequently, it willbe possible to calculate the energy consumption of the compressor and the cooling capacitiesof the evaporators/cabinets.At the return line of the medium temperature level the flow is a two phase one, there<strong>for</strong>e it isnot possible to calculate the load at the medium temperature by measuring the mass flow ofthe refrigerant. By calculating the cooling capacity at the cascade condenser and <strong>for</strong> the lowstage cabinets it will be possible to calculate the total load at the medium temperature level.8.2 Energy Balance TestThe simulators at the medium and low temperature levels provide a fixed known coolingcapacity via the electric heaters which can be used to verify the method of calculating thecooling capacity at the medium and low temperature levels using the compressorsmanufacturers’ data. The medium temperature simulator provides a maximum of 6.6 kW, andthe low temperature simulator provides a maximum of 3 kW. The two simulators can beswitched to 1/3 and 2/3 of its capacity.The system is run with only the simulators on in addition to the CO 2 pump and compressor.The blue line in Figure 4 shows the active lines and components in this test.12

CO 2 Project Report-Phase ILeakage testpointChargingpointpttptttdPtHighe andlow levelcutouttp6,6 kW elOil draintEl Defrost5 kWtdPttttpMEl Defrost5 kWtdPttpptttttdPdPdP3 kW eltt2.5 kWEl Defrost2.5 kWHot gasDefrostFigure 4 Schematic of the active lines and components in the energy balance testThe electric power consumption of the CO 2 pump is measured and it varied around theaverage value of 0.85 kW. The power consumption of the CO 2 compressor was measured byan electric meter instead of using the manufacturers’ data, this is due to the fact that theisentropic and volumetric efficiencies were much less than the provided data. Calculating themass flow using the compressor data resulted in much higher cooling capacities than the 3kW provided by the simulator. The compressor was running at a constant rate and the powerconsumption was measured to be around 1.7 kW.The system is operated around 32ºC <strong>for</strong> condensing ammonia, -26ºC <strong>for</strong> freezers, and amedium temperature of about -9ºC, a plot <strong>for</strong> the temperatures of the boundary conditions ofthe system during the test period is presented in Figure 5.13

CO 2 Project Report-Phase I3530Temperature (C)2520151050-5Evaporating and Condensing TemperatureCond AmmoniadT Cascade CondCond CO2Evap AmmoniaEvap CO2-10-15-20-25-3012:36:00 13:48:00 15:00:00 16:12:00 17:24:00Figure 5 Plot of the temperatures at the system boundaries during the energy balance test periodThe system is first run with only the medium temperature load (simulator and pump); region 1shown in Figure 6, and then the low temperature circuit is switched on where the lowtemperature simulator and the compressor power capacities are added to the load at themedium temperature, region 2. The ammonia compressor run at reduced volume of 50% byunloading two cylinders and this is the volume that have been used along the test except in theregions 2.1 and 2.3 where the compressor was switched to full volume. Running thecompressor at full volume resulted in occasional stop start operation where the load seemed tobe lower than the lowest load to maintain continuous operation of the compressor whileswitching to half of the compressor effective volume region 2.2 showed that the compressorwas running at full speed and was not able to reduce the pressure in the tank to the set value.In order to maintain continuous operation of the compressor the load at the medium stage wasreduced to 2/3 of the total capacity with half of the ammonia compressor volume, region 3.Further reduction of 1/3 on the medium temperature was per<strong>for</strong>med in region 4 after whichthe simulator was switched off. The CO 2 pump at the medium temperature level was keptrunning in region 5 and then switched off in region 6. In region 7 the low stage was switchedoff and the system was running with only the CO 2 pump on.14

CO 2 Project Report-Phase I16,0015,00Energy Balance14,0013,002.1 2.2 2.312,0011,0010,009,0023Q cascade condenserInput capacitydQ(Losses+Comp Ineffec)dQ_averagekW8,007,006,00145,004,003,00562,001,0070,0012:36:00 13:48:00 15:00:00 16:12:00 17:24:00Figure 6 Energy balance with fixed value of volumetric efficiencyRunning the system at different capacities aims at verifying that the cooling capacitycalculated by the ammonia compressor matches different capacities in the system.Cascade condenser cooling capacityThe ammonia mass flow that is passing through the cascade condenser is calculated using theammonia compressor data, the following equation is used to calculate the refrigerant massflow of the compressor:m& = η ⋅ & ⋅ ρ(1)sV sinWhere V & sis the swept volume flow in m3 /s, ρinis the density of the refrigerant (kg/m 3 ) at theinlet of the compressor, ηsis the volumetric efficiency of the compressor. The compressor hasa displacement ( V & sr) of 40.5 m3 /hr at rated speed ( n r) of 1450. Swept volume flow in m3 /s ata given speed can then be calculated using the relation:n 1V &s= V&sr⋅ ⋅(2)nr3600Where n is the compressor speed in RPM (1/min).The volumetric efficiency of the compressor was extracted from BOCK software by runningthe software at different operating conditions, <strong>for</strong> the same operating conditions the mass flowis calculated <strong>for</strong> an ideal compressor, the ratio of the two values is the volumetric efficiency.An average value <strong>for</strong> different operating conditions was found to be about 85%.15

CO 2 Project Report-Phase IMeasuring the compressor speed, the temperature and pressure at the inlet of the compressorprovide the required data to calculate the mass flow. The conditions be<strong>for</strong>e and after thecascade condenser are also measured and there<strong>for</strong>e the cooling capacity is calculated using therelation:Q & = m&⋅ dh(3)Where dh is the enthalpy difference across the heat exchanger.In Figure 6, the cooling capacity of the cascade condenser is plotted and the differencebetween this capacity and the provided load is also plotted as dQ which is also presented withaverage values (dQ average ) over the different operating regions. Excluding region 2 where itwas hard to reach the set point and load peaks have been observed at transition time, thedifference in load which varies between 1 and 1.8 could be explained as heat losses in thesystem and deviation in the volumetric efficiency from the estimated value from BOCK’ssoftware. The volumetric efficiency value that was used in the calculations presented inFigure 6 is assumed to be constant, which is not the case in practice and it will vary with thepressure ratio. In order to correlate the volumetric efficiency to the operating conditions, arelation suggested by Pierre (1982) <strong>for</strong> “good” ammonia reciprocating compressor is used tocalculate the volumetric efficiency as follows:⎛ P ⎞1η = ⋅⎜−⋅⎟s1.02 exp 0.063(4)⎝ P2⎠P1Where is the pressure ratio.P2Using the volumetric efficiency calculated from equation 4 to calculate the mass flow and thecascade condenser capacities yields the results in Figure 7. It can be seen from the figure thatthe difference in cooling capacity is reduced due to the reduced value of the volumetricefficiency and the value varied around 1 kW with smaller deviation around the average valuecompared to the trend in Figure 6.16

CO 2 Project Report-Phase I16,0015,00Energy Balance: calculated volumetric effeciency14,0013,0012,0011,00Q Cascade CondenserInput LoaddQdQ_average10,009,00kW8,007,006,005,004,003,002,001,000,0012:36:00 13:48:00 15:00:00 16:12:00 17:24:00Figure 7 Energy balance with calculated value of volumetric efficiencyIf the heat sink into the system is neglected assuming that the difference in the load is due tolower volumetric efficiency than the calculated one then it will be possible to calculate howmuch the “actual” volumetric efficiency of the compressor should be by using the known loadas the input value to calculate the mass flow of refrigerant from equation 3. Consequently, thevolumetric efficiency can be calculated using equation 1. The plot in Figure 8 shows thevolumetric efficiency calculated using the relation in equation 4 and the “actual” volumetricefficiency. As can be seen in the plot the difference between the calculated and actual valuesis about 10% less <strong>for</strong> the actual efficiency. There<strong>for</strong>e it will be possible to reduce the valuecalculated in equation 4 by 10% in order to adjust the calculated volumetric efficiency to becloser to the actual value. The third plot in Figure 8 is the adjusted value of the volumetricefficiency.17

CO 2 Project Report-Phase I0,9Ammonia compressor volumetric effeciency0,875Calculated0,85ActualAdjusted0,825Effeciency0,80,7750,750,7250,70,6750,6512:36:00 13:04:48 13:33:36 14:02:24 14:31:12 15:00:00 15:28:48 15:57:36Figure 8 Volumetric efficiency of the ammonia compressorUsing the adjusted value <strong>for</strong> the volumetric efficiency brings the average value <strong>for</strong> thedifference in cooling capacities closer to zero as can be seen in the figure below.18

CO 2 Project Report-Phase I16,0015,00Energy Balance: adjusted volumetric effeciency14,0013,0012,0011,0010,00Q Cascade CondenserInput LoaddQdQ_averagekW9,008,007,006,005,004,003,002,001,000,0012:36:00 13:48:00 15:00:00 16:12:00 17:24:00-1,00Figure 9 Energy balance with adjustment to the calculated value of volumetric efficiencyAmmonia compressor capacity and efficienciesThe ammonia compressor is used to estimate the cooling capacity of the cascade compressorand the there<strong>for</strong>e it is important to evaluate its per<strong>for</strong>mance and capacities. Pierre (1982)suggests a relation to estimate the isentropic efficiency ( ηk) <strong>for</strong> the same compressor relatingit to the volumetric efficiency in equation XX.ηsηk⎛ T ⎞⎜1= exp −1.69⋅ + 1. 97⎟⎝ T2⎠(5)The temperature ratio is the absolute temperatures, in Kelvin, of condensation andevaporation corresponding to exist and inlet compressor pressure. Figure 10 shows thecalculated and measured values of the isentropic efficiency. Using BOCK software theaverage isentropic efficiency obtained is about 76% over a range of different operatingconditions, which is close to the calculated one, around 78%. The higher value of themeasured efficiency may have to do with position of the temperature sensor at the dischargeline; it may sense lower temperature value than the actual one.19

CO 2 Project Report-Phase I0,90Ammonia Compressor Isentropic Effeciency0,880,85CalculatedMeasured0,83Effeciency0,800,780,750,730,7012:36:00 13:48:00 15:00:00 16:12:00 17:24:00Figure 10 Measured and calculated isentropic efficiencies of the ammonia compressorAnother method that can be used to calculate the cooling capacity is to measure the electricpower consumption of the compressor and the enthalpy difference across the compressor isdetermined by measuring the pressures and temperatures across it. Certain electric motorefficiency and heat losses to the environment should be assumed in order to calculate the shaftpower according to the following equation:E &Shaft= η ⋅η⋅ E&(6)thermalelelReferring to Climate Check a 7% of thermal losses ( ηthermal= 93%) is usually assumed.Usually the shaft power is considered to be the power that is provided to the refrigerant andthe mass flow can be calculated according to the equation below, the mass flow is then usedto calculate the cooling capacity in equation 3.E & = m&⋅(7)shaftdh compFigure 11 shows typical efficiency values <strong>for</strong> well per<strong>for</strong>ming electric motor related to theshaft power Granryd et el. (2003). The shaft power in the experiment is ranging between 1.8and 3.3 kW, as seen in Figure 12, according to the figure below the electric motor efficiencyis estimated to be around 80% at the rated speed.20

CO 2 Project Report-Phase I1,0η0,90,80,7η elm0,60,50,2 0,5 1 2 5 10 20 50 100 200 kWE &mFigure 11 Typical values of electric motor efficiency versus the motor shaft power Granryd et el. (2003)In the case when the compressor is running at reduced cylinder capacity then there will beadditional losses attached to running two non-productive cylinders, in this analysis theselosses are included in the electric motor efficiency, there<strong>for</strong>e it will be lower than theestimated value.The input electrical power is measured and plotted in Figure 12 along with the shaft powercalculated from the refrigerant side; the adjusted shaft power is the power that is calculatedbased on adjusted volumetric efficiency presented in Figure 8.6,005,00Ammonia Compressor Power ConsumptionShaft powerShaft Power adjustedElectric Power4,00KW3,002,001,000,0012:28:48 13:04:48 13:40:48 14:16:48 14:52:48 15:28:48 16:04:48Figure 12 Calculated ammonia compressor shaft power and measured electric motor efficiency21

CO 2 Project Report-Phase IUsing equation 6 to calculate the efficiency of the electric motor produces the results in theplot of Figure 13. The electric motor efficiency value is dependant on the motor speed whichis also shown in the figure.8075Electric Effeciency vs RPM2 80070652 30060RPM%1 80055501 3004512.1 2.2 2.3 3 44012:28:48 13:04:48 13:40:48 14:16:48 14:52:48 15:28:48 16:04:48800Electirc Effeciency Electirc Effeciency Adjsuted Ammonia RPMFigure 13 Calculated electric motor efficiency <strong>for</strong> the energy balance testIn regions 1, 2.2, 3 and 4 the compressor is running at reduced capacity with two unloadedcylinders which will add some mechanical losses; these losses are included in the electricmotor efficiency presented above. In addition to that in most of the regions the compressorwas running at partial speed which reduces the efficiency of the electric motor. In the end partof region 2.3 the compressor was running at full cylinder capacity. Where the operation wasclose to being stable the efficiency increased due to less mechanical/friction losses, but stillthe compressor was running at partial speed which reduces the electric motor efficiency.In another test the compressor was running at higher cooling capacity with all the cylindersactive. The motor speed was varying due to test conditions and the electric motor efficiency iscalculated in a similar manner to the approach above. The plot in the figure below shows thatthe electric motor efficiency is higher than the values presented in Figure 13. This is mainlydue to the fact that all the moving cylinders are producing work. It can be seen more clearly inthe figure below how the efficiency changes with the speed of the motor. The trend of changeis plotted in Figure 15 where it indicates that the efficiency tend to increase with motor speed.The results in the figure should not be read as quantitative; the aim of the plot is show thetendency of behavior.22

CO 2 Project Report-Phase I100Electric Effeciency vs RPM2 80090802 30070%RPM1 80060501 3004014:16:48 14:52:48 15:28:48 16:04:48 16:40:48 17:16:48 17:52:48Electirc Effeciency Electirc Effeciency Adjsuted Ammonia RPM800Figure 14 Calculated electric motor efficiency <strong>for</strong> ammonia compressor with all cylinders running120110Electric Motor Efficiency vs RPMElectric Effeciency (%)100908070605040800 900 1 000 1 100 1 200 1 300 1 400 1 500RPM (1/min)Figure 15 Calculated electric motor efficiency <strong>for</strong> ammonia compressor at different motor speeds23

CO 2 Project Report-Phase ICO 2 compressor capacity and efficienciesThe same method that is used with ammonia unit to calculate the mass flow of refrigerantfrom the compressor manufacturer data was used <strong>for</strong> the CO 2 compressor; the results showeda large deviation from the expected values, the resulting cooling capacity was much higherthan the provided 3 kW load. This indicates that the actual volumetric efficiency of thecompressor is lower than the 90% value used from the manufacturer data. There<strong>for</strong>e, the CO 2compressor power that is used in the energy balance was measured by the electric energymeter; the average recorded value was 1.62 kW which can be considered as constant since theload at the low stage was constant and equal to the electric heater capacity of 3 kW.Knowing the capacity and the conditions around the freezer simulator the mass flow ofrefrigerant can be calculated and then can be used to calculate the power consumption of theCO 2 compressor. This way of calculating the compressor power resulted in lower value ofabout 15% than the measured electric power consumption. This can be seen in the figurebelow.1,801,60CO2 Compressor Power Consumption1,401,201,00kW0,80CalculatedIsentropic Eff=60%Average Measured0,600,400,200,0013:55:12 14:24:00 14:52:48 15:21:36 15:50:24 16:19:12 16:48:00 17:16:48-0,20Figure 16 Calculated CO 2 compressor shaft power and measured electric motor efficiencyThe reason <strong>for</strong> the deviation may be partly due to the difference in the actual dischargetemperature and the measured value. The point where the temperature is measured is veryclose to the compressor exit but due to the very high discharge temperature and the bigdifference with the room temperature it might be possible that some of the deviation is due to24

CO 2 Project Report-Phase Ithis difference. In order to estimate how much the actual discharge temperature should be, themeasured compressor power is used to calculate the enthalpy at the exit of the compressor andconsequently the temperature can be calculated by using the discharge pressure. Figure 17shows the measured and the expected actual value.200190CO2 compressor discharge temperature180170160T_hg_calcT_hg_measTemperature C150140130120110100908070605014:24:00 14:52:48 15:21:36 15:50:24 16:19:12 16:48:00 17:16:48Figure 17 Calculated and measured CO 2 compressor discharge temperatureIn order to evaluate the isentropic efficiency of the compressor the manufacturer’s data iscompared to the measured ones. Figure 18 shows the curve from the manufacturer data wherethe isentropic efficiency at the running pressure ratio of about 1.7 results in an estimatedisentropic efficiency of 60%, assuming that the curve does not drop sharply be<strong>for</strong>e theoptimum point. Using this value to estimate the power consumption of the compressor in caseof a “good” compressor results in the power consumption presented in Figure 16. It was notpossible to reach the optimum condition <strong>for</strong> the pressure ratio due to the fact that reducing theevaporation pressure resulted in the discharge temperature increasing to very high levels.25

CO 2 Project Report-Phase IIsentropic Efficiency1,000,900,800,700,600,500,401,00 2,00 3,00 4,00 5,00 6,00Pressure RatioFigure 18 Isentropic efficiency of the CO 2 compressor at different pressure ratiosUsing the calculated mass flow from the simulator side and knowing the compressorgeometry and RPM it will be possible to calculate the volumetric efficiency that is presentedin Figure 19 along with the measured isentropic efficiency which is calculated by measuringthe pressures and temperatures around it.26

CO 2 Project Report-Phase I50,0045,00CO2 compressor Isentropic and volumetric effeceincies40,00Isentropic effeceincyVolumetric effeceincy35,00%30,0025,0020,0015,0010,0013:55:12 14:24:00 14:52:48 15:21:36 15:50:24 16:19:12 16:48:00 17:16:48Figure 19 Measured Isentropic and volumetric efficiencies of the CO 2 compressorIt is evident that the volumetric and isentropic efficiencies of the compressor are low and thisis due to the fact that the compressor have been operating with unfavorable conditions duringearly runs of the system. It has been notices that the valve pointed out in Figure 20 wasleaking even when it was firmly closed. When the medium temperature circuit was runningwith the low stage off liquid CO 2 leaked into the oil separator which resulted in badlubrication <strong>for</strong> the compressor. This may have created damage in the compressor which wasrunning with normal sound after the valve was replaced and provided the required coolingcapacity but at a higher cost of energy consumption due to the low volumetric and isentropicefficiencies.27

CO 2 Project Report-Phase ISafetypipetttpMpLeakage testpointChargingpointptttptttdPtHighe andlow levelcutoutMinimumflow nozeltp6,6 kW elOil draintEl Defrost5 kWtdPtttpMEl Defrost5 kWtdPttppttttdPdPdP3 kW eltt2.5 kWEl Defrost2.5 kWHot gasDefrostFigure 20 Schematic diagram of the system shows the leaking line28

CO 2 Project Report-Phase I8.3 System EfficiencyThe overall system efficiency is evaluated using the COP which relates the useful load to thework done to provide it. In this test <strong>for</strong> the energy balance verification the cooling capacitiesare known by running the system on electric heaters in the load simulators as the only load inthe system. The power input to the system has been obtained by measuring the electric powerconsumption of the compressors and pump. This means that all the needed parameters tocalculate the COP are available.Low stage COPThe known supplied load to the low stage via the electric heater and the electric powerconsumption of the CO 2 compressor is measured and the resulting COP is plotted in the figurebelow and denoted as actual value. The calculated COP in the figure is based on the powerconsumption that is calculated using the mass flow obtained from the simulator side.Assuming a good compressor that fulfils the manufacturer’s specifications with the estimated60% isentropic efficiency then the power consumption of the compressor will be much lowerwith a high COP, as can be seen in the figure.8,007,00Low Stage COP6,00Calculated, Eta_v=60%CalculatedActual5,00COP4,003,002,001,000,0014:09:36 14:38:24 15:07:12 15:36:00 16:04:48 16:33:36 17:02:24Figure 21 Calculated and measured low stage COPWith CO 2 compressor that has volumetric and isentropic efficiencies closer to the designvalues the energy consumption of the compressor will be reduced and this will improve thelow and total COP’s.29

CO 2 Project Report-Phase IHigh stage COPIn the case of the ammonia unit per<strong>for</strong>mance evaluation, the capacity of the cascadecondenser is considered as the “useful” load which includes the medium, low stage, CO 2compressor, and pump capacities. The COP of the ammonia presented in the figure below iscalculated by referring to the directly measured electric power consumption which is referredas “actual”. The “calculated” COP is the one that it is obtained by using the mass flow ofrefrigerant to calculate the compressor power; the mass flow is based on the calculatedvolumetric efficiency.1098Ammonia Stage COPUseful cooling loadMeasuredCalculated76COP54321012:28:48 13:04:48 13:40:48 14:16:48 14:52:48 15:28:48 16:04:48Figure 22 Calculated and measured ammonia unit COPThe measured COP’s of the ammonia unit are lower than expected and this is mainly due torunning the system at partial load which is companioned with some mechanical and electricmotor losses. At loads close to full capacity the ammonia compressor will have lower electricpower consumption, consequently, high stage and total system COP’s are expected toimprove.Total COPThe total COP is calculated using two different values <strong>for</strong> the ammonia compressor powerconsumption which are discussed in the ammonia unit evaluation above.30

CO 2 Project Report-Phase I510Total COP4CalculatedActualUseful cooling load983COP72615012:28:48 13:04:48 13:40:48 14:16:48 14:52:48 15:28:48 16:04:484Figure 23 Calculated and measured total COPImproving the efficiencies of the ammonia and CO 2 compressors with favorable operatingconditions on both stages and new CO 2 compressor will reduce the power consumption andreduce the losses at partial load. This will improve the total COP of the system.31

CO 2 Project Report-Phase I9. SYSTEM VARIATIONS AND EXPERIMENTAL RESULTSIn practice, there are many possible variations within the applied CO 2 systems, thesevariations should be investigated in order to modify the system’s design and optimize its costand improve its per<strong>for</strong>mance. In this test rig many variations of the system have been madepossible and the design took into consideration the need to modify and adjust some operatingparameters.9.1 Pump Circulation Ratio and Gravity CirculationAt the medium temperature side the circulation ratio of the refrigerant can be varied so thepressure drop and heat transfer in the display cabinets and distribution lines can beinvestigated, low circulation ratios will result in low pressure drop but may influence the heattransfer in the cabinet. A by-pass line will be used to change the circulation ratio, as using avariable speed pump was impossible due to the fact that it was not possible to find a smallpump that can cover the desired flow range. The pump can also be by-passed so the systemcan be tested <strong>for</strong> gravity circulation operation.The influence of the circulation ratio on the heat transfer and the pressure drop can be firstmeasured and analyzed over the medium temperature simulator where it will be easier toevaluate the effect on heat transfer by measuring the brine temperatures. The obtained resultsfrom the simulator will be used as guidelines to test the proper range of circulation ratios onthe display cabinets. The figure below shows the test circuit.5 kW5 kWFigure 24 Schematic diagram of the circulation rate test circuit32

CO 2 Project Report-Phase IThe available height in the laboratory is limited to 4 m and the ammonia unit is fixed on topof the CO 2 unit due to the thermosyphon loop on the CO 2 side of the cascade condenser. Theammonia unit also requires a certain head <strong>for</strong> its thermosyphon on the cascade condenser side.This means that there will be limited head <strong>for</strong> the CO 2 <strong>for</strong> gravity circulation, about 1.7 m,which might not be sufficient to test the system <strong>for</strong> gravity circulation at full capacity and acompromise <strong>for</strong> smaller capacities might be needed.The medium temperature simulator is used to test <strong>for</strong> different circulation rates. The fixedload provided in the simulator makes it possible to test of a wide range of circulation ratios.Carioles mass flow meter is used to measure the mass flow of refrigerant, the flow rate at lowcirculation ratios of about 2 is very small compared to the measuring range of the mass flowmeter. Due to the continuous generation of bubbles in the CO 2 line it was not possible toper<strong>for</strong>m the zero adjustment reading on the flow meter. For full capacity of the simulator themass flow that is required to result in a circulation rate of 1 is 0.026 kg/s; less than 5% of therange of the mass flow meter where the estimated error is expected to be high.The valve located be<strong>for</strong>e the medium temperature simulator is used to control the massflowing through the simulator. In order to reach the point of circulation 1 the valve wasalmost closed so that superheating is achieved in the simulator, it can be seen in the sightglass that there is no two phase flow in the circuit. Then the valve has been opened on smallsteps until the superheat was eliminated and the sight glass shows no liquid. It was noticedthat slight opening of the valve after this point was reached resulted in liquid drops visual inthe sight glass. The value of the mass flow reading at that point was used as the referencevalue to calculate the circulation ratio. The mass flow showed a reading of about 0.022 kg/s.Circulation rate have been changed between one and slightly over 14, the reason that this highvalue was reached was that the valve was opened gradually to identify at which point the heattransfer starts to change, the valve was fully open at the maximum circulation rate. Figure 25shows the pressure drop across the simulator at different steps of circulation rate. As can beseen from the figure, the waiting time <strong>for</strong> each step is rather short, around 20 minutes,especially <strong>for</strong> the steps with low circulation rates. In order to confirm the accuracy in the datagenerated in this test another test have been run where the run time was longer, at least 1 hour,<strong>for</strong> less steps than the first test, Figure 26.33

CO 2 Project Report-Phase I1,60Pressure drop and CR16151,40141,20dPCR131211dP (bar)1,000,80109CR870,606540,40320,208:48:29 9:24:29 10:00:29 10:36:29 11:12:29 11:48:29 12:24:29 13:00:29 13:36:29 14:12:29 14:48:29 15:24:29 16:00:291Figure 25 Simulator’s pressure drop and circulation ratio <strong>for</strong> short waiting time test1,40Pressure drop151,20131,0011dP (bar)0,800,60CRdP9CR70,4050,2030,00116:55:1216:19:1215:43:1215:07:1214:31:1213:55:1213:19:1212:43:1212:07:1211:31:1210:55:1210:19:1234

CO 2 Project Report-Phase IFigure 26 Simulator’s pressure drop and circulation ratio <strong>for</strong> long waiting time testAs can be seen from the figures the pressure drop across the heat exchanger increases withincreasing the mass flow of refrigerant. Figure 27 relates the pressure drop as a function ofCR, the plots in the diagram shows agreement between short and long run time tests.1,60Pressure drop vs CR1,401,201,00dP (bar)0,80Short Test PeriodLong test Period0,600,400,200,000,00 1,00 2,00 3,00 4,00 5,00 6,00 7,00 8,00 9,00 10,00 11,00 12,00 13,00CRFigure 27 Simulator’s pressure drop at different circulation ratio <strong>for</strong> short and long waiting timeRunning at high circulation rate and having a high pressure drop could be justified if the heattransfer in the evaporator would improve with high circulation rate where an optimumcirculation rate may exist. Measuring the brine temperatures around the simulator it waspossible to evaluate the influence of the circulation rate on the heat transfer in the heatexchanger. Figure 28 shows the brine temperatures around the heat exchanger, theevaporation temperature of CO 2 and the logarithmic mean temperature difference <strong>for</strong> the testwith short waiting time test.35

CO 2 Project Report-Phase I531LMTDT Brine InT Brine OutT CO2Temperature (C)-1-3-5-7-9-117:12:00 8:24:00 9:36:00 10:48:00 12:00:00 13:12:00 14:24:00 15:36:00 16:48:00Figure 28 Temperatures around simulatorAs can be seen from the figure, within the tested range the circulation ratio has almost noinfluence on the heat transfer in the heat exchanger; LMTD had almost constant value ofabout 3.6ºC along the test. Plot of the LMTD versus the circulation ratio <strong>for</strong> the two tests ispresented in Figure 29 where the influence of the circulation ratio can be seen in a clearerway, the influence is very small where the LMTD caries within a range of 0.2ºC which isinsignificant.36

CO 2 Project Report-Phase I4LMTD vs CR3,93,8LMTD C3,73,6Short Test PeriodLong Test Period3,53,43,30,00 1,00 2,00 3,00 4,00 5,00 6,00 7,00 8,00 9,00 10,00 11,00 12,00 13,00CRFigure 29 Simulator’s LMTD at different CR <strong>for</strong> long and short waiting time testsAs can be seem from the results above increasing the mass flow of refrigerant had aninsignificant improvement on heat transfer. Increasing the circulation ratio resulted in anincrease in the pressure drop across the heat exchanger without improvement on the heattransfer; this indicates that the circulation ratio should be chosen as low as possible to ensurecomplete evaporation at the highest load expected. There was no optimum operatingcirculation ratio found similar to the case of conventional brines in secondary systems, thisgives more flexibility in choosing the operating conditions <strong>for</strong> CO 2 secondary system,especially that its pressure drop is much lower than it is <strong>for</strong> brines.Similar test <strong>for</strong> low circulation rates will be conducted on the medium temperature leveldisplay cabinets. The pressure drop will be evaluated and it will be easier to compare it totheoretical calculation since the geometry in the cabinets’ tubes is much easier to apply in thetheoretical model. The heat transfer is not expected to change in a measurable way in thecabinets and this will be observed from any change in air temperatures around the cabinet’sevaporator.9.2 Cascade CondenserOn the cascade condenser there are three main arrangements to be tested concerning the wayCO 2 condenses. The first arrangement is the one in Figure 30, where the hot gas return fromthe low stage is passed directly through the cascade condenser after mixing with the saturatedvapor from the CO 2 tank. A line has been introduced, shown in the figure as red dashed line,where some of the pumped liquid in the medium temperature level can be flashed into the hot37

CO 2 Project Report-Phase Igas line be<strong>for</strong>e entering the cascade condenser so it will be de-superheated and that may resultin more favorable temperature profile in the heat exchanger.Figure 30 Schematic of a cascade condenser arrangementThe other arrangement is where CO 2 is condensing in a thermosyphon loop, Figure 31a,where the two phase CO 2 return line ends in the accumulation tank and the hot gas returnfrom the low stage passes through the liquid of the tank so the hot gas will be de-superheatedby boiling off some of the tank’s liquid.The third arrangement, Figure 31b, is where the return line from the medium temperaturelevel is passing directly through the cascade condenser after mixing with the hot dischargefrom the low stage.(a)(b)Figure 31 Schematic diagrams of two cascade condensation arrangementsThe system has been all the time running on the thermosyphon arrangement described inFigure 31a. Due to the resulted high discharge temperature of CO 2 the two other arrangementshave not been tested yet. It will be possible to per<strong>for</strong>m the tests with better CO 2 compressorthat has higher isentropic efficiency. The temperature at the compressor inlet should bereduced and this can be achieved by running the freezers at the minimum superheat setting38

CO 2 Project Report-Phase Iwhere the system will have stable operation, also the external superheat which is due to heatsink from the environment should be reduced, in average about 7°C was observed.Figure 32 shows the temperature values across the cascade condenser; the difference betweenthe “hot” CO 2 and the “cold” ammonia is slightly higher than 2°C. In this test the load wasonly from the low temperature side and the medium temperature was off.54Cascade Condenser Temperatures321Temperature (C)0-1-2-3-4-5-6Temperature differenceCO2Ammonia-7-8-9-10-11-12Time span 09:10:34-15:23:26-1375,00 175,00 275,00 375,00 475,00 575,00 675,00Figure 32 Temperatures across the cascade condenserIn the applications where CO 2 is used in indirect circuit the two arrangements that have beenused are shown in the schematics below. The arrangement A in the figure below has beenused in the early stages of applying CO 2 in indirect solutions <strong>for</strong> freezing temperatureapplications and then has been abandoned in favor of the natural circulationsolution/thermosyphon.Figure 33 Basic schematic of two possible arrangements <strong>for</strong> the CO 2 indirect circuit39

CO 2 Project Report-Phase IThe arrangement in Figure 33b has been tested and the results are similar to the ones plottedin Figure 32. The influence on heat transfer in the cascade condenser by using the otherindirect system arrangement will be tested by applying the system arrangement in Figure 34.The per<strong>for</strong>mance of the cascade condenser will be tested <strong>for</strong> different loads on botharrangements which will change the mass flow of refrigerants on both sides of the cascadecondenser and the effect on heat transfer will be analyzed in a similar way in the Figure 32.The stability of operation which has been one of the reasons <strong>for</strong> favoring the thermosyphonarrangement will be investigated.Figure 34 Basic schematic of <strong>for</strong>ced condensation arrangement on the cascade condenserThe temperature drop that is measured and observed while operating the cascade condenserindicates a good heat transfer that yields a low temperature difference across the cascadecondenser. This is mainly due to the favorable conditions <strong>for</strong> exchanging heat on the sides ofthe heat exchanger. CO 2 enters the cascade condenser as saturated vapor and ammonia boilsoff along the heat exchanger from saturated conditions at the inlet. Further evaluation atdifferent loads will be per<strong>for</strong>med and other arrangements available in the system will betested.9.3 Flashing Gas in Liquid LinesLong supply lines to the freezing cabinets are usually the case in this application which meansthat there will be a pressure drop and some of the refrigerant might flash be<strong>for</strong>e the expansionvalve resulting in unstable operation. An internal heat exchanger is usually used to tackle thisproblem by adding some sub-cooling be<strong>for</strong>e the distribution lines, Figure 35. Anotherpossibility is to add some head to the liquid and this is achieved by connecting the liquidsupply to the expansion valves after the medium temperature circulation pump, Figure 36, inthis case the additional pumping power accompanying this solution will be very small sincethe CO 2 pump is larger than needed. Both solutions are planned to be tested.40

CO 2 Project Report-Phase IFigure 35 Basic schematic of internal heat exchanger solutionFigure 36 Basic schematic of the solution where head is added to the liquid be<strong>for</strong>e expansion valveSo far the system has been running without pump heat or internal heat exchanger. Thepressure drop and heat leaking in the liquid supply line proved to be small and the availablehead in the tank was enough to overcome it. Both arrangements in the figure above have beentested and they work properly, when running with internal heat exchanger the discharge gastemperature was noticed to be very high. This has to do with the low isentropic efficiency ofthe CO 2 compressor.9.4 Freezing Cabinets ControlThe two freezers in the system are identical and equipped with electronic expansion valves.The freezers were delivered with a control system that measures the evaporation temperatureon the surface of the evaporator tube. In this case the thermocouple is placed at the first bendafter the expansion valve. Figure 37 is a schematic diagram of the freezer evaporator wherethe positions of the measuring points are indicated.41

CO 2 Project Report-Phase It_sht2Figure 37 Basic schematic of the freezer’s evaporator with the measuring points indicatedThe freezers showed occasional unstable behavior, especially at start up, while running thesystem with this control method, this was more likely to happen at low superheat set point.While running the system at the lowest set point of 3°C the air temperature in the freezersstarted to increase up to a certain value and then decrease again and the behavior is repeatedin a cyclic manner. This is plotted in Figure 38 where it can be seen that the air temperaturedifference is decreasing which indicates loss in cooling capacity.2010Superheat=3°Temperature (C)0-10-20-30-4013:48:00 14:16:48 14:45:36 15:14:24 15:43:12 16:12:00Superheat Air in Air out Freezer surface-controller FreezeFigure 38 Temperatures around the freezer cabinet with temperature based controller. Superheat set value is 3ºC.42

CO 2 Project Report-Phase IThe controller starts controlling the system <strong>for</strong> the highest superheat set value, usually setbetween 10-15°C, and then the valve opening is adjusted targeting the lowest superheat setvalue, where the system is usually operating at in a stable condition. When the system reachesthe lowest superheat set point this behavior is triggered. This can be seen in the superheatvalue in the figure above where the controller is trying to control the system to operate at thesuperheat value of 3°C. During the unstable operation there was not enough refrigerantflowing in the evaporator which led to the refrigerant being superheated at the point where theevaporation temperature should be measured. The evaporating temperature measured by thepressure transducer is plotted in the diagram which indicates the difference between themeasurement of the real evaporating temperature from the pressure reading and the one that ismeasured on the surface of the heat exchanger tube.The superheat set point of 3°C is very low compared to practice but the same behavior wasrepeated with higher superheat set values but with fewer oscillations after the start up period.A sample of the tests is shown in Figure 39 where the superheat set value of 7°C was used.The unstable behavior occurred only once at start up of the test.10Temperature (C)0-10-20Superheat=7T SH SurfaceT Air InT Air outT Freeze SurfaceT Freeze Sat Press-30-4011:24:00 11:52:48 12:21:36 12:50:24 13:19:12 13:48:00 14:16:48 14:45:36 15:14:24Figure 39 Temperatures around the freezer cabinet with temperature based controller. Superheat set value is 7ºC.Another method of measuring the evaporation temperature is by measuring the evaporationpressure and this will be the input signal to the expansion valve controller. A schematicdiagram of the pressure based controller method is presented in Figure 40. In this case the43

CO 2 Project Report-Phase Iactual evaporation temperature is measured instead of the surface temperature which is about2°C higher than the real value; this can be seen in the plots in the two figures above.Figure 40 Basic schematic of the freezer’s evaporator with the measuring points indicatedThe controller was changed on one display cabinet while the other was kept on the surfacetemperature measurement method. The freezers were run at the same conditions and at thesame set point. The influence of the superheat value is tested, a value of 8°C was used and thetwo cabinets were running with good stability as can be seen in Figure 41 which shows the airand brine temperatures <strong>for</strong> the cabinets and the simulator. The simulator used the pressurebased control all the time.Temperature (C)20,0010,000,00-10,00Freezers and Simulator TemperatureSuperheat=8-10°CAir in Press ControlAir out Press ControlAir in Temp ControlAir out Temp ControlBrine in Press ControlBrine out Press Control-20,00-30,00Time span 09:10:34-11:12:430,00 100,00 200,00 300,00 400,00 500,00 600,00 700,0044

CO 2 Project Report-Phase IFigure 41 Temperatures around the freezers and simulator. Superheat set value is 8ºC.Figure 42 and Figure 43 shows evaporating temperatures, superheat values and the superheattemperature measured at the exit of the evaporator <strong>for</strong> the freezers and the simulator. Thevariation in the superheat temperature in the freezer with surface temperature control, inFigure 42, may be due to the fact that the surface temperature controller senses lowersuperheat than the pressure one and starts regulate at earlier time than the pressure basedcontroller.As can be seen in Figure 42, the measured superheat temperature at the exit of the evaporatorin case of the simulator is fluctuating at start up which may be due to the small size of the heatexchanger where the response to the expansion valve opening or closing is faster and moreprominent than the case of the long freezer heat exchanger.40,0030,0020,00Freezers Controlers ParamtersSuperheat=8-10°CSuper<strong>Heat</strong> Press ControlSuper<strong>Heat</strong> Temp ControlEvap exit Press ControlEvap exit Temp ControlFreezing Temp ControlFrezer Press ControlTemperature (C)10,000,00-10,00-20,00-30,00-40,00Time span 09:10:34-11:12:430,00 100,00 200,00 300,00 400,00 500,00 600,00 700,00Figure 42 Temperature values input to the controller in the freezers. Superheat set value is 8ºC.45

CO 2 Project Report-Phase I40,0030,0020,00Simulator Controlers ParamtersSuperheat=8-10°CSuper <strong>Heat</strong>Evap exitFrezeringTemperature (C)10,000,00-10,00-20,00-30,00Time span 09:10:34-11:12:43-40,000,00 100,00 200,00 300,00 400,00 500,00 600,00 700,00Figure 43 Temperature values input to the controller in the simulator. Superheat set value is 8ºC.When the superheat temperature is reduced to 5°C in an attempt to trigger the unstablebehavior both freezers started to show cyclic increase and decrease in the air temperature,Figure 44 and Figure 45 shows the air and superheat temperatures on both cabinets. In case ofsurface temperature control the oscillations had higher peaks and the air exit temperature wasclose to 0°C, while in the other case the value did not go over -12°C. And the superheat in thesecond case is oscillating between 2 and 10°C while in the first case the superheat value goesbelow zero and then increase up to 20°C. The superheat high set value <strong>for</strong> the controller is10°C in both cases.46

CO 2 Project Report-Phase I20,0015,00Superheat=5-10°CFreezers Temperature and Superheat10,005,00Temperature (C)0,00-5,00-10,00SuperheatAir inAir out-15,00-20,00-25,00-30,0011:45:36 12:14:24 12:43:12 13:12:00 13:40:48 14:09:36 14:38:24 15:07:12Figure 44 Temperatures around the freezer with pressure based controller. Superheat set value is 5ºC.20,0015,00Superheat=5-10°CFreezers Temperature and Superheat10,005,00Temperature (C)0,00-5,00-10,00SupeheatAir inAir out-15,00-20,00-25,00-30,0011:45:36 12:14:24 12:43:12 13:12:00 13:40:48 14:09:36 14:38:24 15:07:12Figure 45 Temperatures around the freezer with temperature based controller. Superheat set value is 5ºC.47

CO 2 Project Report-Phase IThe parameters that the controllers read in both cases are presented in Figure 46 and Figure47, as can be seen in Figure 46 the evaporating temperature read is increasing due torefrigerant superheat and the controller reads the wrong evaporation temperature and fails tocontrol properly.20,0015,00Superheat=5-10°CFreezers Controllers Parameters10,005,00Temperature (C)0,00-5,00-10,00Super <strong>Heat</strong>FreezerEvap Exit-15,00-20,00-25,00-30,0011:38:24 12:07:12 12:36:00 13:04:48 13:33:36 14:02:24 14:31:12 15:00:00Figure 46 Temperature values input to the pressure based controller in the freezer. Superheat set value is 5ºC.2015Superheat=5-10°CFreezers Controllers Parameters105Temperature (C)0-5-10-15Freezer Press ControlSuper <strong>Heat</strong>Freezer Temp ControlEvap Exit-20-25-3011:38:24 12:07:12 12:36:00 13:04:48 13:33:36 14:02:24 14:31:12 15:00:0048

CO 2 Project Report-Phase IFigure 47 Temperature values input to the temperature based controller in the freezer. Superheat set value is 5ºC.In the case of pressure based controller it measures a higher superheat value than the actualone, which can be seen in the figures above as a difference between surface measuredevaporation temperature and the pressure related one. This is true assuming that the sametemperature difference exists at the superheat measuring point. This gives the pressure basedcontroller higher margin of stability, on the other hand the temperature based controller maybe closer to measuring the real superheat value but in the cases discussed above at someconditions it measures wrong evaporation temperature. Also the response to changes orfluctuations in evaporation temperature would be faster in case of the pressure basedcontroller.In the case of the simulator the controller managed to regulate the superheat value adequatelyat the 5°C set value, this can be seen in Figure 48. This may be due to the fast response at theevaporator exit to any changes to the degree of opening of the expansion valve whichconsequently readjusts itself accordingly.30,0025,0020,0015,0010,00Freezer Simulator Brine Temperature and SuperheatSuperheat=5-10°CSuperheatBrine inBrine outTemperature (C)5,000,00-5,00-10,00-15,00-20,00-25,00-30,0011:45:36 12:14:24 12:43:12 13:12:00 13:40:48 14:09:36 14:38:24 15:07:12Figure 48 Temperatures around the simulator with superheat set value of 5ºC.The measurements and experiences in running the display cabinets with the two controllersindicate a more favorable conditions in case of the pressure based controller. The controllersuccessfully kept the superheat value within the set range, it had lower stability at lowsuperheat set values but still not as unstable as the temperature based controller. Measuringthe evaporation temperature on the surface of the tube of the evaporator proved to be thesource of instability in the controller; the point measured superheated vapor when the massflow was small in the heat exchanger.49

CO 2 Project Report-Phase I9.5 Safety TestsThe main two safety concerns when dealing with CO 2 in such an application is the problem ofthe rising pressure at standstill and CO 2 concentration level during and after a leakageaccident. Several valves at different points of the system where CO 2 exists as gas, liquid, andtwo-phase are installed. The valves will be opened in order to simulate leakage of a pipebreaking. The rate of leakage will be observed together with the rate of increase inconcentration at different levels in the room. The concentration will be measured at differentzones in the space to try to identify the severity of an accident and the best place to install theCO 2 indicators <strong>for</strong> the alarm system. The <strong>for</strong>mation of dry ice at the leakage point may blockthe leakage or reduce its rate.The pressure in the system will be allowed to rise to levels where the bleed valves will beopened and then the time <strong>for</strong> the system to reach this state will be observed and theconcentration rise in the room will be recorded. Also, the frequency of this operation and theamount of lost charge will be reported.9.6 Flooded Versus Direct Expansion EvaporationThe per<strong>for</strong>mance analysis of the medium temperature flooded evaporators will be evaluatedagainst the direct expansion freezers, accordingly, it will be possible to decide about how aflooded freezer may per<strong>for</strong>m and estimate the improvements that may be gained in eachsolution. In the preliminary design it is made possible to transfer the evaporation in thefreezers from direct expansion to flooded type by adding a low pressure tank, a pump and bypassingthe expansion valves.50

CO 2 Project Report-Phase I10. THEORETICAL MODELAlong with the experimental rig, a mathematical model has been developed to simulate theexperimental system. At the design stage of the system, the theoretical model facilitateddesign calculations concerning sizing of components and it will be evaluated against theexperimental results when the system will be running. If the theoretical model agrees with theexperimental results, this will give confidence in evaluating the per<strong>for</strong>mance of the system byvarying certain parameters and will help in evaluating design modifications.10.1 Model DescriptionThe model is written in Engineering Equation Solver (EES) and uses the manufacturers’ datato calculate the per<strong>for</strong>mance of each component. The boundary conditions and the coolingcapacities of the system are input variables to the model which are inserted in the per<strong>for</strong>manceequations and efficiency curves to calculate energy consumption, capacities, and efficienciesof different components and the system as a whole.10.2 Overview of Calculations and AssumptionsTwo phase flow pressure drop calculations use Friedel’s correlation which is based on thetwo-phase multiplier. The Clapeyron equation is used to convert the pressure drop to anequivalent change in saturation temperature. Single-phase flow pressure drop is calculatedusing the Gnielinsky correlation <strong>for</strong> the friction factor in turbulent flow. More details aboutthe pressure and temperature drop calculations <strong>for</strong> CO 2 can be found in (Sawalha and Palm2003).The heat exchangers’ per<strong>for</strong>mances have been simulated by calculating the effectivenessbased on the supplied manufacturer’s per<strong>for</strong>mance data <strong>for</strong> the heat transfer coefficient andthe heat transfer area (Arias 2005). These properties are assumed constant and the evaporatingand condensation temperature along the heat exchangers are also assumed constant. In thecascade condenser and in the thermosyphon mode at the CO 2 side the logarithmic meantemperature difference method has been calculated since the temperature is constant at bothsides of the heat exchanger.The compressors are presently assumed to have constant isentropic and volumetricefficiencies but in a later stage the efficiencies will be introduced in the model based on theper<strong>for</strong>mance curves <strong>for</strong> the compressors.In the display cabinets the flow is assumed to be equally distributed in the cabinets’ loops.The pressure drop in the medium temperature display cabinets is calculated assuming singlephase flow (liquid) in the first half of the loop’s pipe and two phase flow with the exit qualityin the other half, consequently the total pressure drop is the sum of the two pressure drops. Inthe deep freeze cabinets, where dry expansion takes place, the pressure drop is calculatedassuming two phase flow in the first half of the pipe with the inlet quality to the evaporatorand in the second half it is assumed to be single phase flow with the exit conditions ofsuperheated vapor, the total pressure drop is the sum of the two values. These assumptionsare made to simplify the calculations and are only used <strong>for</strong> the estimation of the pressuredrops.51

CO 2 Project Report-Phase I10.3 Design CalculationsPressure and temperature drop calculations have been used to size the distribution lines of thesystem. Figure 49a and Figure 49b show results <strong>for</strong> the medium circuit calculation with acirculation ratio of 3 and the length of each line being 20 m. The figures show plots of thepressure and temperature drops <strong>for</strong> the liquid supply and the two-phase return lines usingdifferent pipe diameters. It is possible to change the circulation ratio and obtain such plots <strong>for</strong>each value; the results <strong>for</strong> the value of 3 have been selected because it fits almost in themiddle of the range of the values <strong>for</strong> the circulation ratio to be tested. The plotted totalpressure drop includes the losses in the fittings along the line, which were calculated using theequivalent pressure drop coefficient method (Stoecker 1998).301,42004x10 025L=20 m1,2dP [kPa]2015105dP totdP pipedT tot10,80,60,40,2dT [°C]dP [kPa]15010050L=20 mdP totdP PipedT tot3x10 02x10 010 0dT [°C]0010 12 14 16 18 20 22 24 26d [millimeter]00x10 010 12 14 16 18 20 22 24 26d[millimeter](a)(b)Figure 49 Pressure and temperature drops <strong>for</strong> different tube sizes in medium temperature liquid supply (a) and return(b) linesSimilar calculations have been made <strong>for</strong> the low stage and a sample of the results <strong>for</strong> thesingle phase vapor in the suction line is presented in Figure 50.401,630L=20 m1,41,2dP tot1dP [kPa]20dP PipedT tot0,80,6dT [°C]100,40,20010 12 14 16 18 20 22 24 26d [millimeter]Figure 50 Pressure and temperature drops in the suction line <strong>for</strong> different tube diameters52

CO 2 Project Report-Phase IAs can be seen from the plots, the pressure and temperature drops <strong>for</strong> CO 2 are very small,even with small pipe sizes. The sizes of the major lines in the rig are 5/8” (15.8 mm) <strong>for</strong> theliquid supply line of the medium temperature stage, the return line has 1 1/8” (28.6 mm), theliquid line of the low stage is 3/8” (9.5 mm) and the suction line has a size of 5/8”. Based onthese sizes it has been estimated that gravity circulation will be possible to achieve with theavailable head of 1.7 m.At standstill the system will be absorbing heat from the warm surroundings. The main volumeof the system is assumed to be the tank since at standstill the charge will be accumulating inthe tank. The heat flow from the 25ºC room into the -8ºC tank with different insulationthicknesses covering the tank is plotted in Figure 51. Assuming 50% of the tank filled withliquid and using 50mm of insulation the time <strong>for</strong> the system to reach the pressure of 35 bars,which is the limit that will trigger the bleed valves, is calculated to be about 12 hours.At standstill conditions the CO 2 in the tank is cooled down by using cold brine circuitavailable in the laboratory. By using heat exchanger with a capacity of 1 kW the time requiredto remove the heat added from the ambient to the tank since standstill started is estimated tobe about 50 min. In the calculations, the temperature of the tank is assumed to be constantduring the heat transfer into the system, there<strong>for</strong>e in real conditions the time of pressure risewill be longer than calculated.450400350300Q tot [W]2502001501005000 10 20 30 40 50 60 70 80 90 100w insulation [millimeter]Figure 51 <strong>Heat</strong> sink into the accumulation tank with different insulation thickness covering the tank10.4 Some Simulation ResultsAt this stage of the development of the model with the available manufacturing data someper<strong>for</strong>mance calculations can be obtained to estimate the overall per<strong>for</strong>mance of the system.The COP of the low stage is about 4, <strong>for</strong> the ammonia unit the COP is 2.6 and the total COPis 2.1. The total COP is calculated as the ratio of the total cooling capacity and the total powerconsumption. With a circulation ratio of 2, the pump power consumption is calculated to be12 W with 50% of assumed pump efficiency.53

CO 2 Project Report-Phase I11. CONCLUSIONS AND FUTURE WORKThe chosen NH 3 /CO 2 system <strong>for</strong> supermarket refrigeration is being built in a laboratoryenvironment which can simulate the conditions in a real installation. Different variations inthe system have been implemented in order to compare different options and find an optimumsystem design <strong>for</strong> the usage of CO 2 in the low and medium temperature stages of asupermarket refrigeration system. Several tests have been run <strong>for</strong> overall evaluation of thesystem and <strong>for</strong> the main components. A mathematical model has been developed to simulatethe per<strong>for</strong>mance of the system and facilitate detailed system design calculations. Theexperimental and computational models will be adjusted while running the system in order tobring the two models closer to each other to be more comparable. The closer the theoreticalmodel will get to simulate the real case the more confidence it will give to use it <strong>for</strong>adjustments in the system <strong>for</strong> the optimum operation.The energy balance in the system has been verified and the efficiencies of the compressorshave been measured and calculated. The obtained efficiencies of the compressors will be usedto calculate cooling capacities and shaft power consumption of the ammonia compressor. Dueto unfavorable operating environment the CO 2 compressor had to run in its volumetric andisentropic efficiencies have been measured to be much less than expected. System’s highstage, low stage and total COP’s are measured and calculated. The measured COP’s of thesystem are lower than expected and this is mainly due to running the system at partial loadwhich is companioned with some mechanical and electric motor losses; at loads close to fullcapacity ammonia unit and total system COP is expected to improve. With CO 2 compressorthat has the design efficiencies low and total COP’s are expected to improve.The circulation ratio in the simulator has been changed and its influence on heat transfer andpressure drop has been evaluated. Increasing the mass flow of refrigerant to have circulationratios higher than one had insignificant improvement on heat transfer. The pressure dropacross the heat exchanger increased and no optimum operating point have been found.There<strong>for</strong>e, the circulation ratio should be chosen as low as possible to ensure completeevaporation at the highest load expected. Similar test <strong>for</strong> low circulation rates will beconducted on the medium temperature level display cabinets, mainly to evaluate the pressuredrop in its tubes.The temperature drop that is measured and observed while operating the cascade condenser inthe thermosyphon mode indicates good heat transfer conditions. This is mainly due to thefavorable conditions <strong>for</strong> exchanging heat on the sides of the heat exchanger; CO 2 enters thecascade condenser as saturated vapor and ammonia boils off along the heat exchanger fromsaturated conditions at the inlet. Further evaluation at different loads will be per<strong>for</strong>med andother arrangements available in the system will be tested.The measurements and experiences in running the display cabinets with the two controllersindicated more favorable conditions in case of the pressure based controller. The controllersuccessfully kept the superheat value within the set range, it had lower stability at lowsuperheat set values but still not as unstable as the temperature based controller.54

CO 2 Project Report-Phase ISafety analysis of the system with different scenarios will be per<strong>for</strong>med both theoretically andpractically and will provide more detailed in<strong>for</strong>mation about the behavior of CO 2 and allowdevelopment of methods <strong>for</strong> safe handling.Based on the results of the tests per<strong>for</strong>med up to this stage, weak and strong points of thesystem have been identified where the system will be modified <strong>for</strong> the best of its per<strong>for</strong>manceat the selected basic arrangement. The analysis to follow will be focusing on testing thepossible arrangements available in the test rig and to evaluate parametric analysis of the mainsystem’s components.55

CO 2 Project Report-Phase I12. REFERENCESASHRAE 1997. ASHRAE Handbook, Fundamentals.Sawalha, S. and Palm, B. 2003, Energy Consumption Evaluation of Indirect Systems WithCO 2 as Secondary Refrigerant in Supermarket Refrigeration, Proceedings of the 21st IIRInternational Congress of Refrigeration, Washington, D.C., USAPierre, B.: ”Kylteknik, allmän kurs”, Inst. Mekanisk värmeteori och kylteknik, KTH,Stockholm, 1982 (in Swedish)Granryd, E., Ekroth, I., Ludqvist, P., Melinder, Å., Palm, B., and Rohlin, P. 2003,Refrigeration Engineering, Department of Energy Technology, KTH, Stockholm, Sweden.Arias, J. 2005. Energy Usage in Supermarkets Modelling and Field Measurements, Doctoralthesis, Department of Energy Technology, Royal Institute of Technology. Stockholm, Sweden.Stoecker, W. F. 1998. Industrial Refrigeration Handbook, McGraw-Hill.Zhao, Y., M. Molki, et al. 2000. Flow Boiling of CO 2 in Micro channels. ASHRAETransactions 106(1): 437-445.56