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PARAMETRIC EVALUATION OF A NH3/CO2 CASCADE SYSTEM FOR SUPERMARKET REFRIGERATION IN LABORATORY ENVIRONMENT

by

Carlos Perales Cabrejas

Master of Science Thesis in Refrigeration Division of Applied Thermodynamics and Refrigeration

2006

Department of Energy Technology Royal Institute of Technology

Stockholm, Sweden

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ABSTRACT A refrigeration NH3/CO2 cascade system has been built in IUC laboratory in Katrineholm (Sweden). The system is equipped with extensive instrumentation in order to collect data and perform online diagnosis under laboratory controlled environment. The system solution under investigation replicates a medium size supermarket in Sweden. The natural and environmental friendly refrigerant CO2 is used in the low stage with several possibilities of system variations and parametric analysis. The CO2 stage is divided into two different levels, one medium-temperature or cooling side at around -8ºC and another low-temperature or freezing side at -36ºC approximately. Overall energy balance has been examined. Cooling and freezing product temperature has been investigated. Several arrangement solutions have been tested for the cascade condenser evaluation in the medium and low temperature sides. Different defrost methods performance has been deep focused.

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ACKNOWLEDGEMENTS The ‘CO2 in Supermarket Refrigeration’ project was initiated as an agreement between Installatörernas Utbildingscentrum (IUC) and KTH/Applied Thermodynamics and Refrigeration Division. The project is managed by IUC and financially supported from the companies Ahlsell, Huurre, AGA, WICA and ICA. This project is also financed by Energimyndigheten (STEM). This thesis work is involved in this project. First of all, I would like to express my sincere gratitude to my supervisor Samer Sawalha since this thesis was only possible due to your help, support and advices, from as technical and as the moral side. Nobody knows all the hours you have dedicated to me and I appreciate it really much. Secondly, I would like to thank the professors from the Applied Thermodynamics Department in the Universidad Politécnica de Valencia (UPV), José Miguel Corberán and José Gonzálvez, and also professor Björn Palm from Applied Thermodynamics and Refrigeration Division of Energy Technology Department in Kungl Tekniska Högskolan (KTH), for giving me the opportunity to take this Erasmus scholarship between both departments and all the help received to fix it. In addition, one part of this gratitude is for my international coordinator in KTH, Rebecca Ljungqvist for her great help in my scholarship procedure. Furthermore, I would like to extend my gratitude to Jörgen Rogstam and Per-Olof Nilsson from Installatörernas Utbildingscentrum (IUC) in Katrineholm, where is placed the NH3/CO2 cascade facility, for all their useful comments and help to manage the installation. And I also would like to thank you Arash Soleimani Karimabad, for being my support, company and friend during the months we spent in the laboratory working together. That is also for you, Laia, for all the chats we had in the department concerning our thesis. On the other hand, I am very grateful to the people from Flemingsberg –Alex, Benny, Emmanuel, Fabian, Felix, Jordi, Menchu and Xavi- and from Lappis –Aina and Marta- for all the marvellous moments we have shared during this Erasmus period here in Sweden. I will never forget. Finally, this thesis is dedicated specially to my parents, Pablo and Lupe, and the rest of my family and friends that, in spite of not having their company here in Stockholm, are always with me wherever I am. Carlos Perales Cabrejas July 2006 Stockholm, Sweden

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NOMENCLATURE COP: Coefficient of performance MC1, MC2: Coolers, medium-temperature or cooling cabinets 1, 2 MTS: Medium-temperature or cooling simulator FC1, FC2: Freezers, low-temperature or freezing cabinets 1, 2 FTS: Low-temperature or freezing simulator Q& : Cooling capacity (kW) m& : Refrigerant mass flow rate (kg/s) Cp : Specific heat (kJ/kgK) dT : Temperature difference (ºC) dh : Enthalpy difference (kJ/kg)

vη : Volumetric efficiency of the compressor (-)

sV& : Swept volume flow (m3/h)

inρ : Density of the refrigerant at the inlet to the compressor (kg/m3)

srV& : Compressor displacement volume n : Compressor speed (rpm)

rn : Compressor rated speed (rpm)

outletP : Discharge pressure (bar)

inletP : Suction pressure (bar)

lossesη : Compressor thermal efficiency (-)

eleccompE ,& : Compressor electrical power (kW)

shaftcompE ,& : Compressor mechanical power (kW)

compdh : Compressor enthalpy difference (kJ/kg)

pumpE& : Pump power (kW) Pr: Product through the cabinets X : Vapour fraction or quality HTC: Heat transfer coefficient HGD: Hot gas defrost EHD: Electric heaters defrost HG: Hot gas

airdT : Air temperature difference (ºC) COP : Coefficient of performance

isη : Isentropic efficiency (-)

SuperheatTΔ : Superheat value between evaporator outlet and suction line t : Time Subscripts W: Water side cond: Condensation Pierre: Based on Pierre’s correlation ce: Complete evaporation

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comp: Compressor shaft: Mechanical shaft work CO2,LT: Carbon dioxide in the low-temperature side evap: evaporator losses: Efficiency losses in the compressor NOP: Refrigeration period or normal operation DOP: Defrosting operation sim: Simulator is: isentropic in: Inlet out: Outlet DC: Defrosting cabinet NDC: Non-defrosting cabinet Non-Defrosting or ND: Simulator and freezer that are not defrosting HG: Hot gas def: Defrost

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DEFINITIONS Overlap Temperature or Temperature Difference Across the Cascade Condenser – Difference of temperature between the working fluids used in both low and high stage in the cascade system. Taken from the Danfoss ‘EKC 414A1 Controller’ Manual [1]: Defrost Stop Temperature – If a defrost sensor has been mounted on the evaporator, the defrost can be stopped at a given temperature. The temperature value is set. If a defrost sensor has not been mounted, the defrost will be stopped based on time. Maximum Defrost Time – If it has been chosen defrost based on temperature, this will constitute a safety period where the defrost will be stopped if no stop based on temperature has taken place by then. If it has selected defrost stop based on time, this setting will be the defrost time. Drip-off Time – Time which is to elapse from the end of a defrost until compressor is to be resumed. It is the time when water is dripping off the evaporator Fan start delay after defrost or fan’s delay – Time which is to elapse from when the compressor is started after a defrost and until the fan may resume operation. It is the time where the water is ‘bound’ to the evaporator Taken from Wikipedia.org: Ozone Depletion Potential (ODP) [2] – The ODP of a chemical compound is the relative amount of degradation to the ozone layer it can cause, with trichlorofluoromethane (R-11) being fixed at an ODP of 1.0. Global Warming Potential (GWP) [3] – The GWP is a measure of how much a given mass of greenhouse gas is estimated to contribute to global warming. It is a relative scale which compares the gas in question to that of the same mass of carbon dioxide (whose GWP is by definition 1). A GWP is calculated over a specific time interval and the value of this must be stated whenever a GWP is quoted or else the value is meaningless.

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TABLE OF CONTENTS ABSTRACT……………………………………………………………………………2 ACKNOWLEDGEMENTS……………………………………………………………4 NOMENCLATURE……………………………………………………………………5 DEFINITIONS………………………………………………………………………….7 TABLE OF CONTENTS……………………………………………………………...8 1. INTRODUCTION………………………………………………………………….10 1.1. Historical use of CO2 as refrigerant……………………………………...10 1.1.1. NH3/CO2 cascade systems……………………………………………...12 1.2. The revival of CO2 as a refrigerant……………………………………….12 1.2.1. Characteristics of CO2………………………………………………….…13 2. BACKGROUND…………………………………………………………………...17 3. OBJECTIVES……………………………………………………………………...19 4. CASE OF STUDY…………………………………………………………………20 4.1. System Requirements and Boundaries………………………………….20 4.2. The system solution…………………………………………………………20 5. THE EXPERIMENTAL FACILITY……………………………………………….23 6. OVERALL SYSTEM ANALYSIS………………………………………………..28 6.1. Volumetric Efficiency for the Ammonia Compressor…………………28 6.2. Load Calculations and Energy Balance………………………………….30 6.3. Overall System Efficiency…………………………………………………..33 6.3.1. Low-Temperature Level COP…………………………………………….33 6.3.2. High-Temperature Level COP…………………………………………....33 6.3.3. Total System COP………………………………………………………....34 7. PRODUCT TEMPERATURE INVESTIGATION……………………………….36 7.1. Freezing cabinets…………………………………………………………….36 7.2. Cooling cabinets…………………………………………………………......39 7.3. Effect of Roof Fans…………………………………………………………..40 8. OPTIMIZATION OF THE CASCADE CONDENSER…………………………47 8.1. Indirect System Arrangements…………………………………………….47 8.1.1. Thermosyphon Loop Arrangement………………………………………50 8.1.2. Forced Condensation Loop Arrangement………………………………50 8.1.3. Comparison between both Indirect System Arrangements……………51 8.2. Cascade Condenser Arrangements………………………………………53 9. DEFROST………………………………………………………………………….60 9.1. Methods of Defrosting Under Investigation………………………….….60

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9.1.1. Hot Gas Defrost Method…………………………………………………..60 9.1.2. Electric Heaters Defrost Method………………………………………....61 9.2. Defrost Tests……………………………………………………………….…62 9.3. Freezing Cabinets……………………………………………………………63 9.3.1. Defrost Sensor Location……………………………………..……………63 9.3.2. Defrosting Intervals…………………………………………..……………69 9.3.3. Effect of Fans during Defrost……………………………………….…….70 9.3.3.1 Electric Heaters Defrost Method……………………………………..……….71 9.3.3.2. Hot Gas Defrost Method………………………………………...…………….74 9.3.3.3. Comparison between HGD and EHD………………………………………..78 9.3.4. Energy Consumption during Defrost……………………….……………78 9.3.4.1. Electric Heaters Defrost Method…………………………………..…………78 9.3.4.2. Hot Gas Defrost Method………………………………………………………79 9.3.4.3. Comparison between Hot Gas and Electric Heaters Defrost in Energy

Consumption…………………………………………………….……………..83 9.3.4.4. Hot Gas Mass Flow during HGD………………………………..……………83 9.4. Cooling Cabinets………………………………………………………..……86 9.4.1. Results and comments………………………………………...………….86 10. CONCLUSIONS…………………………………………………………………89 11. REFERENCES……………………………………………………….………….92 12. BIBLIOGRAPHY……………………………………………………..………….94

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1. INTRODUCTION CO2 has been used as a refrigerant in vapor compression systems for over 130 years. But only in the last decade new interests appeared to exploit its beneficial properties due to restrictions on other refrigerants. The HFCs were once expected to be acceptable refrigerant fluids are now being phased out because of their impact on the environment. Along with trying to find new good chemical compounds, the refrigeration industry had an increasing interest in technologies based on ecologically safe ‘natural’ refrigerants, i.e. fluids like water, air, noble gases, hydrocarbons, NH3 and CO2. Among these, CO2 is the only non-flammable and non-toxic fluid that can also operate in a vapor compression cycle below 0ºC, having the potential to offer environmental and personal safety in a refrigerating system. The HFC refrigerants that were the main cause of the retirement of the CO2 are now paradoxically being partly displaced by this ‘natural’ substance.

1.1. Historical use of CO2 as refrigerant A brief historical overview of using CO2 as refrigerant was described by Donaldson and Nagengast [4], in the following paragraph: CO2 was first proposed as a refrigerant for vapour-compression systems by Alexander Twining, who mentioned it in his 1850 British patent. Thaddeus S.C. Lowe experimented with CO2 for military balloons in the 1860s and recognized the possibilities of using it as a refrigerant. He went on to build refrigerating equipment obtaining British Patent 952 in 1867 and erected an ice machine about 1869 at Jackson, Miss. He also constructed a machine on board a ship for the transport of frozen meat in the Gulf of Mexico. Lowe did not develop his ideas further. Carl Linde also experimented with CO2 when he designed a machine for F. Krupp at Essen, Germany, in 1882. W. Raydt received British Patent 15475 in 1884 for a compression ice-making system using CO2. British Patent 1890 was granted to J. Harrison in 1884 for a device for manufacturing CO2 for refrigerant use. Still, the use of CO2 really did not advance until Franz Windhausen of Germany designed a CO2 compressor, receiving British Patent 2864 in 1886. Windhausen’s patent was purchased by J&E Hall of Great Britain, who improved it, commencing manufacture about 1890. Hall’s CO2 machine saw widespread application on ships, replacing the compressed air machines theretofore used. CO2 machines were used universally on British ships into the 1940s. As it was commented in the previous paragraph, after the first steps of applying CO2 in refrigeration systems, mainly in marine systems but also in air conditioning and stationary refrigeration –[5], [6]–, a peak was reached in CO2 systems production during the 1920s and the early ‘30s, with the advent of fin coils, many small comfort cooling systems were installed in restaurants, hotel, public spaces, nightclubs, hospital operating rooms, etc [6].

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Figure 1 - Percentage use of main primary refrigerants in existing marine cargo

installations [6] With the arrival of the CFC refrigerants by DuPont in the 1930s-1940s (Figs. 1 and 2), these, by then called ‘safety refrigerants’, displaced the NH3, sulphur dioxide and CO2 the ‘old working fluids’ in most refrigeration applications. Kim et alt [6] listed the major factors that contributed to this replacement. In case of NH3 and SO2, they were the well known safety reasons due to toxicity and flammability, but it seems that main reasons of the CO2 displacement was the high-pressure handling problems, capacity and efficiency loss at high temperature (aggravated by the need to use air cooling instead of water). Aggressive marketing of the new CFC products, low-cost tube assembly in competing system and a failure of CO2 system manufacturers to improve and modernize the design of systems and machinery also help to this substitution.

Figure 2 - Timeline or refrigeration development [5] These new products were marketing-announced as the refrigerants able to provide the efficiency and flexibility of NH3 but within the safety and reliability of CO2. With this promotion, the freons quickly took over a large part of the refrigeration market [7]. Although some systems solutions, such as cascade systems, appeared in order to bypass the disadvantages accompanying the high pressure of CO2 were not sufficient to keep the CO2 in the refrigerating field.

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With all these factors, the CFCs almost completely replaced CO2 in marine market and seriously limited NH3’s usage in land market since 1950s.

1.1.1. NH3/CO2 cascade systems As Pearson [5] commented, the Flick Company in 1932 installed a hybrid ‘spit-stage’ system which utilized CO2 in the low-temperature stage. Figure 3 shows the diagram of the Flick’s NH3/CO2 ‘spit-stage’. With this device, it was possible to reduce the NH3 charge in the plant providing the necessary refrigeration to condense the CO2 within moderate temperature and pressure ranges.

Figure 3 - Flick’s ‘split-stage’ NH3/CO2 system [5]

The Frick’s ‘split-stage’ system faced with the problem that most of the operators seems to be willing to accept the hazards associated with running a large NH3 plant, so there was no need to spend the extra expense of installing a cascade heat exchanger and the idea was not further developed.

1.2. The revival of CO2 as a refrigerant After the Montreal and Kyoto Protocols in the late 1990s, the necessity to phase out and eventually eliminate the usage of several groups of halogenated hydrocarbons has renewed the interest in the ‘natural refrigerants’ such as: air, water vapour, NH3, CO2 and hydrocarbons. The first two refrigerants, that is, air and water vapour, have some restrictions on their use in low-temperatures. On the other hand, despite a generally

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excellent safety record, there is a strict limit on allowable charge of hydrocarbons in a plant, which makes them inappropriate for use in large water chillers and industrial systems unless pertinent safety standards could be applied. In many ways, NH3 is ideal for large industrial systems where its soft flammability, pungent smell and low threshold limit value do not present problems. But NH3 is, however, unsuited for domestic, automotive and small commercial refrigeration and heat pump systems. In addition, the evaporating pressure of an NH3 system is below atmospheric pressure when the evaporating temperature is below -35ºC, causing probable air leak into the refrigeration system, leading to short-term inefficiency and the long-term unreliability of the system [8]. These reasons leave CO2 as the only natural refrigerant to be used across these systems [5]. Moreover, CO2 gas that is used in refrigeration is a by-product of the chemical industry and its use in the refrigeration system can be considered as a delayed step before its unavoidable release to the environment [9].

1.2.1. Characteristics of CO2 Table A provides several characteristics and properties of CO2 and compares them with other conventional refrigerants. As can be observed in the table, CO2 (R-744) is non-flammable natural refrigerant with no ozone depletion potential and negligible global warming potential. Its reduced pressure is much higher and its volumetric refrigeration capacity is 3-10 times larger than CFC, HCFC, HFC and HC refrigerants. The critical pressure and temperature of CO2 are 73.8 bar and 31.1ºC, respectively.

Table A - Characteristics of selected refrigerants [6] As can be seen in Figure 4, the pressure and temperature of the triple point of CO2 are -56.6ºC and 5.2 bar, respectively. Furthermore, it can be seen from the figure that the saturation pressure at 0ºC is of around 35 bars.

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Figure 4 - Phase and pressure-enthalpy diagrams of CO2 [6]

Figures 5a and 5b illustrate the saturation pressure and the slope of the saturation temperature of CO2 compared to other refrigerants. The CO2 saturation pressure is much higher than other conventional refrigerants, and its higher steepness produces a smaller decrease in the temperature associated with a given pressure drop in the evaporator.

(a) (b)

Figure 5 - Saturated pressure (a) and its slope (b) for selected refrigerants [6] Furthermore, an important characteristic present in CO2 is its low density ratio (Figure 6a) which is defined as division of the liquid and vapour density. This parameter takes part in an important role in an evaporator because it determines the flow pattern and, therefore, the heat transfer coefficient [6]. The higher vapour density of CO2 results in high volumetric refrigeration capacity. As can be seen in Figure 6b, it increases with the temperature until a maximum value at 22ºC is reached, then it decreases again. By definition, it is zero at the critical point. High volumetric refrigeration capacity results in system designs with smaller volumes.

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(a) (b)

Figure 6 - Density ratio and volumetric refrigeration capacity for some selected refrigerants [6]

Nowadays, hydrofluorocarbons are being used only in restricted amounts and high taxes are enforced. On the contrary, natural refrigerants such as CO2 and hydrocarbons are tax free [10]. Lorentzen [7] exhibits some further advantages of CO2,

- complete compatibility with normal lubricants and common machine construction materials

- easy availability everywhere, independent of any supply monopoly - simple operation and service, no ‘recycling’ required, very low price

On the other hand, CO2 is not exempt from disadvantages. Firstly, using CO2 in refrigeration means that the system will operate with high working pressure. In order to solve this problem, hybrids systems were applied in which CO2 is used in the low stage; this limits the pressure to a level where the requirements for components like compressors, controls and valves can be fulfilled with conventional components [11]. In other case, newer components technology should be investigated. Other problem derived from the high pressure level in CO2 systems is the high off-cycle pressure. In order to avoid this condition, [11] suggest several ways:

• A small, independent refrigeration system can be used to keep the liquid temperature at levels where saturated pressure is less than design pressure

• The volume of the expansion vessel can be of the order to prevent the pressure from exceeding the design pressure

• The system can be design to resist the saturated pressure at the design temperature (approximately 83.56 bar)

• Or, in the simplest way, to release the CO2 to the atmosphere which reduces the system’s pressure

Though CO2 systems work with high pressures, it presents low compression ratio (around 3 or 4 to 1, against approximately 8 to 1 in current HFC

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compressors) which is a major factor contributing to compressor performance [12] Other important characteristic of CO2 which is important to have in mind is that although CO2 is not toxic and non-flammable, on the contrary of NH3, it does not have distinctive smell for detection. Furthermore, CO2 is denser than air and it can displace the oxygen in the space. Both characteristics can be dangerous in case of leakage, especially in reduced spaces. Symptoms associated with the inhalation of air containing CO2 are presented in [11]. Detection and good ventilation systems should be placed in a plant which uses CO2 as refrigerant. If any liquid CO2 leakage happens in the system, it will pass through its triple point (-56.6ºC at 5.2 bar), ‘dry ice’ will appear and the leakage may be sealed by itself. In spite of being a good factor from safety point of view, it represents a risk in case of the need to release CO2 liquid through a relief valve [13].

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2. BACKGROUND1 In the early stages after the revival of CO2 as a refrigerant it was applied in low temperature applications in supermarkets as a secondary refrigerant in indirect systems. The usage of CO2 in cascade and transcritical systems for this application has been suggested along with the indirect system, but the limitations for the application of the cascade and transcritical systems have been related to the scarcity of components that can efficiently handle CO2. Also there have been many unanswered questions related to how to handle the highly pressurized system and how safe it is to deal with CO2 in this application. In the indirect system it was possible to use conventional components to handle CO2; this is mainly due to the low operating pressure in the indirect loop at low temperature applications, 12 bars at -35°C. The pumping power needed for CO2 in the indirect system is very small compared to conventional brine systems due to the small volume flow rate and pressure drop of CO2 in the circuit. The small volume flow rate is a result of the phase changing process on the CO2 side, which also contributes to having small pressure drop in the pipes and heat exchangers. Gaining experience and confidence in working with CO2 in indirect systems combined with the knowledge that is gained through extensive research work on CO2 in mobile air conditioning and hot water heat pumps brought the attention of the industry for the necessity of producing components which are specially designed, or modified, to handle CO2. As a result, cascade and trans-critical systems became reasonably applicable. In supermarket applications the difference between evaporating and condensing temperatures is large, therefore, the cascade or other two-stage systems become favorable and they are well adaptable for the two-temperature level requirement for chilled and frozen products in the supermarket. The indirect system requires an additional heat exchanger (primary refrigerant evaporator/CO2 condenser), which implies that there is an additional temperature difference across the heat exchanger and the resulting evaporating temperature will be lower than if a direct expansion had been performed. In cascade or multi-stage CO2 systems, CO2 evaporates directly in the evaporators of the display cases, which minimizes the required temperature lift and reduces the energy consumption. In the direct expansion solution, the CO2 pump is not required; despite the fact that the power consumption of the pump in the CO2 indirect solution is generally very small compared to conventional 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 the pumping power usually is higher than necessary as it is difficult to find CO2 pumps to match medium capacity systems and as the pumps usually are larger than needed a bypass line is introduced to obtain the required CO2 flow rate in the evaporators. 1 Based on ‘CO2 in Supermarket Refrigeration’. 1st phase report [7]

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The low critical point for CO2 of 31ºC implies that it will operate with better theoretical COP between low temperature ranges further below the critical point compared to high stage conditions. Therefore, the application of CO2 in the low stage of a cascade system yields a reasonable theoretical COP compared to other refrigerants, see Figures 7a and 7b. In Figure 7b, the COP values above the critical point are obtained at optimum pressure on the high pressure side.

Figure 7- COP of CO2 compared to other main refrigerants in low stage (a) and in high

stage (b) operations [9]

Several factors contribute to improve the COP of CO2. The favorable thermophysical properties of CO2 results in low pressure and temperature drops in the system. From the heat transfer point of view, the low surface tension will make boiling easier and therefore will improve the heat transfer. Also, due to the low pressure drop [14], the components of the system will be smaller while the mass flow rate of the refrigerant will be comparable to R404A, R22, R502 and R134a refrigerants, which will result in high mass flux of CO2 in the heat exchangers. Another improvement to the COP comes from the improved volumetric efficiency of the CO2 compressor compared to conventional refrigerants; this is due to the lower pressure ratio across the CO2 compressor.

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3. OBJECTIVES This thesis work is part of a project that aims to develop, test, and evaluate an energy efficient medium size supermarket system working with CO2 as the refrigerant, emphasize is on using environmentally friendly refrigerants, therefore, an NH3/CO2 cascade system is chosen. A laboratory environment allows controlling over the boundary conditions of the system and provides flexibility for modifications. Investigations will focus in first place on overall system evaluation. Secondly, product temperature through both two temperature level cabinets will be examined. Comparison between different cascade condenser modifications will allow optimizing the system for the most efficient solution arrangement. Detailed cabinet defrost analysis will be performed for both medium and low-temperature levels’ cabinets. Performance of electric and hot gas defrost methods in the freezers will be compared. Fans effect, energy spent and time needed for defrost will be the criteria of evaluating the two defrost methods. This thesis is a continuation of the work in the CO2 in Supermarket Refrigeration project [9].

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4. CASE OF STUDY

4.1. System Requirements and Boundaries The system solution under investigation has been chosen to be applicable to a medium size supermarket installation in Sweden. The refrigeration system in the supermarkets in Sweden usually operates to satisfy the evaporating temperatures to maintain products at two temperature levels, around +2ºC for cold food and -18ºC for frozen products. Despite the low ambient temperature in Sweden, the condensing temperature is usually kept constant around the year at a value of about 40 ºC. The cooling capacities of medium size supermarkets are typically around 50 kW for freezing and 150 kW for cooling at the medium-temperature level. This estimation is based on contacts with major installers of supermarket systems in Sweden. Accordingly, the system has been designed to operate between the temperature boundaries mentioned above and to provide a cooling capacity which is scaled down while trying to keep a load ratio of about 3. The low-temperature side has a rated capacity of 7.4 kW and the medium-temperature side was designed to have a capacity of 20 kW.

4.2. The System Solution The absolute ideal refrigerant in every field of application does not exist. All available refrigerant compounds have their own weak points, which must be taken into account in the design and operation of the system. The difficulty of finding one refrigerant which ideally suits both the high and low temperatures and pressure ranges of a system operating with a large temperature lift leads to the concept of a cascade system, that is, two separate refrigerant circuits are connected thermally through a cascade condenser. The cascade condenser heat exchanger is the condenser of the low-temperature stage and the evaporator of the high-temperature stage. Introducing the cascade heat exchange gives a loss in the system’s efficiency due to the necessity of having a temperature difference between both fluids (overlap temperature); however, compressors running with CO2 have greater efficiency and good heat transfer is expected in the heat exchangers. The overall efficiency, therefore, for example, of a NH3/CO2 cascade system is not lower than a traditional NH3 system [11]. In the system under investigation, a cascade system with NH3 at the high-temperature stage and CO2 at the low-temperature stage was chosen, meanwhile at the medium-temperature level; CO2 is pumped to provide the required cooling load. Figure 8 is a schematic diagram of the CO2 circuits in the system.

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Figure 8 - Schematic diagram of NH3/CO2 cascade system with CO2 at the medium and

low temperature level [9] Using NH3 in the high-temperature stage means that it will be easy to deal with a leakage accident as NH3 can only leak into the machine room which should be equipped with proper safety devices. Using CO2 in the low-temperature stage, results in reasonable operating pressure levels in the CO2 circuit (28 bars at -8°C), reduces the risk of water and air penetration (since NH3 pressure will be higher than atmospheric pressure) and minimizes the NH3 charge in the plant. In the chosen system, the NH3 always works above atmosphere pressure and, therefore, this risk does not take place. The high pressure difference between the NH3 and CO2 sides in the cascade condenser is an important practical issue that has to be taken into account. Since the CO2 pressure is higher than that of the NH3, the leakage would occur in the NH3 side. In this case, ammonium carbonate, a solid corrosive substance, would form immediately. To avoid this, good isolation between both fluids must be ensured by proper design of the heat exchanger [15]. The favorable pressure drop characteristics of CO2 suits this application where long distribution lines are usually needed. Also this implies that the size of the distribution lines is also smaller than for other refrigerants, which reduces the cost of the piping system.

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The evaporators at the medium-temperature level are flooded with CO2 which is circulated via a pump; this is expected to produce better performance due to the good heat transfer characteristics of the completely wetted evaporator and, therefore, the evaporator temperature will be higher than if a direct expansion concept has been used.

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5. THE EXPERIMENTAL FACILITY The experimental facility was designed and built for ‘CO2 in Supermarket Refrigeration’ project. This was initiated as an agreement between Installatörernas Utbildingscentrum (IUC) and KTH/Applied Thermodynamics and Refrigeration Division. The 20 kW design cooling capacity at the medium-temperature level is divided over two display cabinets with around 5 kW each (Figure 9) and the rest of the capacity is supplied by an electric heater (Figure 10) in what is referred to as a load simulator. The installed electric heater at the medium-temperature level managed to provide three load steps of 2.2, 4.4 and 6.6 KW. Therefore, that makes the rated total load at the cooling level to around 16.6 kW.

Figure 9 - Pictures of the medium-temperature cabinets (MC). Cabinet MC1 is on the right

and MC2 is on the left.

Figure 10 - Picture of the medium-temperature level simulator (MTS). The arrow indicates the load steps buttons location in the simulator.

On the low-temperature side, the load is divided over two freezers (Figure 11) with around 2.5 kW each and an electric heater (Figure 12) which can provide a maximum load of 3 kW (also three steps of 1 kW each). Moreover, the freezers are equipped with electronic expansion valves.

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Figure 11 - Picture of the freezing cabinets (FC). Cabinet FC1 is on the right and FC2 is on the left.

Figure 12 - Picture of the low-temperature level simulator (FTS). The arrow indicates the load steps buttons location in the simulator.

The compressor is a Copeland scroll type (Figure 13a) with operating temperatures between -37°C and -8°C and a displacement of 4.1 m3/h. The maximum cooling capacity of the compressor is 7.4 kW. The accumulation tank (Figure 13b) has a capacity to contain 180 L of CO2 and is equipped with an electronic level indicator. It can stand a pressure up to 40 bars which corresponds to an operating temperature of about 6 °C. The system is equipped with a safety release valve which is triggered when the pressure in the system reaches 38 bars. To avoid the opening of the release valve and the loss of significant charge from the system, a bleed valve is installed which opens for periodical release of CO2 at lower pressure than the set value for the release valve, 35 bars, so the pressure in the system will be reduced. If the pressure increase in the system is higher than the rate that the bleed valve can handle, then the release valve will open and release the system charge.

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(a) (b)

Figure 13 - Picture of the (a) CO2 Copeland Scroll compressor and the (b) CO2 accumulation tank used in the installation

The CO2 pump (Figure 14a) is a hermitic one with capacity higher than the highest circulation rate desired; therefore, a by-pass is used to reduce the flow rate pumped into the medium temperature circuit. About 1.5 meters head over the pump is respected to prevent cavitation (Figure 14b).

(a) (b) Figure 14 - Pictures of the (a) CO2 pump and a detail for the (b) safety head of 1.5 m

added. The NH3 unit uses a Bock reciprocating compressor (Figure 15a) with displacement of 40.5 m3/h; it can run at 50% reduced capacity by unloading half of its cylinders. Heat is removed from the NH3 evaporator via a thermosyphon loop which required a certain height of the unit. The capacity control of both compressors is achieved by a frequency converter.

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The cascade condenser (Figure 15b) is a plate type heat exchanger that is specially selected to handle the pressure difference that will exist between CO2 and NH3, at -8 °C CO2 will have about 28 bars while NH3 will have a pressure of about 2.7 bars at -12 °C.

(a) (b)

Figure 15 - Pictures of the (a) NH3 Bock Reciprocating compressor and the (b) cascade condenser heat exchanger installed in the facility The medium-temperature display cabinets have the conventional electric defrost method installed. The freezing cabinets, in their part, can be defrosted either by using hot gas defrost –where the evaporators are heated by passing the hot gas through the evaporator tubes-, or by electric heaters. Figure 16 is a detailed schematic diagram of the facility with most of the measuring points indicated. As can be seen in the schematic diagram, several by-pass lines and the high number of valves indicate possible variations in the system which will be used for testing, modifications and parametric analysis. The load simulators with the electric heaters can be seen in the diagram in the medium and low- temperature circuits. The electric heaters provide heat to a brine loop which exchanges the heat with the refrigerant in a plate heat exchanger. Although the schematic plot shows that one of the freezers is electrically defrosted while the other is based on hot gas defrost, both freezers have the option to run either electric or hot gas defrost methods.

27

Figure 16 - Detailed schematic diagram of the NH3/CO2 cascade system facility [9]

28

6. OVERALL SYSTEM ANALYSIS The system under investigation is a scaled down real installation where the main discussion is weather or not this system is a suitable replacement for traditional technologies. In order to provide answers about the current system solution, it is important to perform the overall analysis of the system where the capacities are properly measured and the energy balance is verified. The mass flow rate of the NH3 is used to calculate the cooling capacity and the COP of the system. Using all the sensors and devices installed in the facility it is possible to measure the mass flow rate of the NH3 and by knowing the geometry of the compressor the volumetric efficiency can be calculated.

6.1. Volumetric Efficiency for the NH3 Compressor A mass flow meter was installed in the water side and two thermocouples in the water/NH3 heat exchanger to know the inlet and outlet temperatures of the water. The water mass flow meter does not give an online measurement to the data capturing system and is not permanently installed on the rig. This requires calculating the NH3 mass flow from the compressor side using the volumetric efficiency, compressor speed and density in the suction line. By running energy balance on the NH3 condenser the mass flow rate of NH3 can be obtained using the following formula:

condNHWWNHW dhdTCpQ ,3NH3W3/ m m ⋅=⋅⋅= &&& (1) Using the NH3 mass flow that is passing through the cascade condenser, the actual volumetric efficiency is calculated using the refrigerant mass flow of the compressor:

insNH Vm ρη ⋅⋅= && v3 ins

NH

Vm

ρη

⋅=&& 3

v (2)

The compressor has a displacement srV of 40.5 m3/h in full capacity mode and 20.25 m3/hr for reduced capacity (half capacity) mode both at rated speed rn of 1450 rpm. Swept volume flow in m3/s at a given speed can then be calculated using the relation:

36001

⋅⋅=r

srs nnVV& (3)

In order to calculate the volumetric efficiency at different pressure ratios a correlation, equation 4, was suggested by Pierre [16] for “good” NH3 reciprocating compressor:

29

)063.0exp(02.1,inlet

outletPierrev P

P⋅−⋅=η (4)

Comparing the results obtained from both actual (‘nu_actual’ line in Figure 17) and Pierre’s (’nu_Pierre’ line in Figure 17) methods, it can be seen that the actual value of the volumetric efficiency is 7% higher than the Pierre’s. As can be seen in the figure, the actual value of the volumetric efficiency presents a fluctuation since it is based on the NH3 mass flow rate that is passing through the NH3 condenser according to equations 1 and 2. This, on the other hand, is calculated by the NH3 high pressure and as can be verified in Figure 18 (‘P_NH3_L’ line) the high pressure of the NH3 shows that fluctuation.

0

0,1

0,2

0,3

0,4

0,5

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1,2

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12:2

5:37

12:3

3:10

12:4

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12:4

8:13

12:5

5:43

13:0

3:14

13:1

0:48

13:1

8:18

13:2

5:50

13:3

3:23

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1:01

13:4

8:32

13:5

6:10

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3:40

14:1

1:15

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14:2

6:17

14:3

3:51

14:4

1:21

14:4

8:53

14:5

6:24

15:0

3:57

15:1

1:29

15:1

8:59

15:2

6:33

15:3

4:06

15:4

1:36

15:4

9:09

15:5

6:41

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4:11

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1:45

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9:18

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6:48

16:3

4:22

16:4

1:55

16:4

9:29

16:5

6:59

17:0

4:35

17:1

2:05

Time (hh:mm:ss)

Volu

met

ric e

ffici

ency

(-)

nu_actual

nu_Pierre

nu_corrected

Figure 17 - Plot of different methods of calculating the volumetric efficiency of the NH3

compressor

30

0

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4:06

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4:11

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16:2

6:48

16:3

4:22

16:4

1:55

16:4

9:29

16:5

6:59

17:0

4:35

17:1

2:05

Time (hh:mm:ss)

Pres

sure

(bar

)

P_NH3_L

P_NH3_H

P_CO2_L

P_CO2_H

Figure 18 - Pressures for the NH3 and CO2 during the test

In order to have online calculating of the cooling capacity based on the compressor data, Pierre’s correlation have been adjusted to account for the 7% difference in volumetric efficiency (‘nu_corrected’ line in Figure 18). Using the average of the actual volumetric efficiency, the following formula is produced:

)063.0exp(094.1,inlet

outletcorrectedv P

P⋅−⋅=η (5)

6.2. Load Calculations and Energy Balance The NH3 compressor is used to determine the mass flow rate of the refrigerant which will then be used to calculate the cooling capacity of the system. The compressor manufacturer data have been used as guidelines for the calculations which are based on knowing the geometry and the efficiencies of the compressors at certain operating conditions. Measuring the rotational speed of the compressor, and the temperatures and pressures around it gives all the data needed to calculate the mass flow of the refrigerant. Consequently, it will be possible to calculate the energy consumption of the compressor and the cooling capacities of the evaporators/cabinets. On the CO2 side, the energy balance around the compressor is also used to obtain the total mass flow in the low-temperature side. The power consumption of the CO2 compressor was measured by an electric meter. In this test, the compressor was running at a constant rate and the power consumption was measured to be around 2.02 kW during the test.

31

At the return line of the medium-temperature level the flow is a two phase one, therefore it is not possible to calculate the load at the medium temperature by measuring the mass flow of the refrigerant. By calculating the cooling capacity at the cascade condenser and for the low-stage cabinets it will be possible to calculate the total load at the medium temperature level. The simulators at the medium- and low-temperature levels provide a fixed known cooling capacity via the electric heaters which can be used to verify the method of calculating the cooling capacity at the medium- and low-temperature levels using the compressors manufacturers’ data. The medium-temperature simulator provides a maximum of 6.6 kW, and the low-temperature simulator provides a maximum of 3 kW. The two simulators can also be switched to 1/3 and 2/3 of its capacity. The system is run with the cabinets and the simulators on. The electric power consumption of the CO2 pump is measured and it varied around the average value of 0.85 kW. The system runs at around 34.4ºC for condensing NH3, -36ºC for freezers, and a medium temperature of about -9ºC. A plot for the temperatures of the boundary conditions of the system during the test period is presented in Figure 19.

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-10

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4:13

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6:14

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2:18

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8:18

13:2

4:20

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13:3

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4:32

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0:40

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2:45

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8:45

14:2

4:47

14:3

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4:54

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0:57

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6:57

15:1

2:59

15:1

8:59

15:2

5:03

15:3

1:03

15:3

7:06

15:4

3:09

15:4

9:09

15:5

5:09

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1:11

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7:15

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3:15

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9:18

16:2

5:18

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1:22

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7:22

16:4

3:29

16:4

9:29

16:5

5:29

17:0

1:35

17:0

7:35

17:1

3:38

Time (hh:mm:ss)

Tem

pera

ture

(ºC

)

T_NH3_cond T_NH3_evap

T_CO2_cond T_CO2_evap

Figure 19- Boundary temperatures during energy balance test

In the beginning of the test, the system is run with the whole capacity of the medium-temperature simulator (6.6 kW) and the load needed in the low temperature side to run the CO2 compressor constant (region 1 in Figure 20). In this region, the NH3 compressor was running at full capacity mode since the

32

load added was too high to run at reduced capacity. Then, the load in medium-temperature side was reduced by switching off 1/3 of its simulator’s capacity (region 2 in Figure 20). It was decided to put in reduced capacity mode for the NH3 compressor since the compressor did not maintain continuous operation as can be verified in Figure 20. Another button was removed in region 3, only running with 2.2 kW in the medium-temperature simulator. After that, the whole capacity from the medium-temperature side was removed except of the CO2 pump as can be seen in region 4 in the figure.

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9:03

12:5

5:13

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1:24

13:0

7:34

13:1

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13:2

6:10

13:3

2:23

13:3

8:41

13:4

4:52

13:5

1:02

13:5

7:20

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5:55

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2:07

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6:53

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5:27

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1:39

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15:2

4:03

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0:13

15:3

6:26

15:4

2:36

15:4

8:49

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4:59

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1:11

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7:25

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3:35

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9:48

16:2

5:58

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16:3

8:22

16:4

4:39

16:5

0:49

16:5

6:59

17:0

3:15

17:0

9:25

Time (hh:mm:ss)

Cap

acity

(kW

)

Q_dot_3/3MTS+LTS Q_dot_2/3MTS+LTS

Q_dot_1/3MTS+LTS Q_dot_LTS

Input Load_CO2s Q_dot_CC_av

1

23

4

Figure 20 - Load input and adjusted cooling capacity through the cascade condenser

In order to calculate the low-temperature cooling capacity the CO2 compressor power consumption can be used to obtain the CO2 mass flow though the compressor. That is,

compLTCOelcomplossesshaftcomp dhmEE ⋅=⋅= ,2,, &&& η (6) Where the elcompCOE ,2

& is the electrical power consumption measured and compCOdh 2 the enthalpy difference between inlet and outlet of the CO2 compressor. Furthermore, it was assumed 7% losses in the compressor power due to heat losses to the environment. On the other hand, in the region 4, the cooling capacity at the low stage can be calculated as follows:

evapLTCOLT dhmQ ⋅= ,2&& (7)

33

Therefore, adding the capacity from the CO2 compressor and pump using equation 7 yields the total load given in the region 4 to the system.

pumpCOelcompCOlossesLTpumpcompCOLT EEQQ 2,2&2&&&& +⋅+=+ η (8)

By adding the known capacity from the medium-temperature simulator to the total load from the low-temperature stage, it will be possible to know the entire load added in the whole the test. As is shown in Figure 20, the adjusted formula for the volumetric efficiency correlation fits quite well with the supplied capacity during the whole test. The region which has less accuracy –with the highest deviation from the input load- is the second one; this is due to the fact that in this period the NH3 compressor was running at full speed and it could not remove the load added, gradually increasing the tank pressure.

6.3. Overall System Efficiency The overall system efficiency is evaluated using the COP which compares the useful refrigerating effect to the work supplied to provide it. The total COP and the COP of each temperature level is calculated in order to compare the efficiency of each stage during the test.

6.3.1. Low-Temperature Level COP In this case, both freezers and simulator were running to maintain constant speed in the CO2 compressor. The freezing side COP can be calculated by the expression:

compCO

LTLT E

QCOP

,2&

&= (9)

The cooling capacity from the low-temperature side was obtained using the CO2 mass flow rate and the conditions before and after the evaporator as was expressed in equation 7.

6.3.2. High-Temperature Level COP The cooling load of the cascade condenser is considered as the ‘useful’ load which includes the refrigerating effects from the medium- and low-temperature level and the capacities from the CO2 compressor and pump. Therefore, the COP of the NH3 side can be obtained by the following equation:

34

compNH

CCNH E

QCOP

,33 &

&= (10)

where the power consumption of the NH3 compressor, shaft power, was calculated using the NH3 mass flow rate and the inlet and outlet conditions of the compressor.

6.3.3. Total System COP The COP of the overall system is evaluated using the useful refrigerating capacities from both medium- and low-temperature levels divided by the sum of the CO2 compressor and pump power consumption plus the NH3 compressor’s as follows:

compNHcompCOpumpCO

LTMTCO EEE

QQCOP

,3,2,22 &&&

&&

+++

= (11)

Figure 21 illustrates the COP of the high and low-temperature levels and of the total system.

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4,5

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12:1

8:06

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5:37

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3:10

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5:43

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3:14

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8:18

13:2

5:50

13:3

3:23

13:4

1:41

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9:12

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6:50

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4:20

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9:25

14:2

6:57

14:3

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14:4

9:33

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16:5

7:39

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5:15

17:1

2:48

Time (hh:mm:ss)

Coe

ffici

ent o

f Per

form

ance

(-)

COP_LT

COP_NH3

COP_total

Figure 21 - Coefficients of performance for the system

As Figure 21 shows, the COP of the CO2’s low-temperature level (‘COP_LT’ line) is almost constant during the whole test, being approximately 3.4. This is due to the fact that the freezing load was almost constant in order to maintain the CO2 compressor at stable speed as was commented before.

35

As can be observed in the figure, the COP of the high stage (‘CO2_NH3’ line) increases when the load from the medium-temperature simulator decreases, having a value between 3.7 and 3.9. It has been noticed that the medium-temperature level has an important influence in the overall system COP (‘COP_total’ in the plot). The total system COP is around 2.8 when the simulator’s entire load is added and approximately 1.4 when the simulator is turned off. Therefore, although the high stage performance worsens when medium-temperature loads are supplied; however, their positive effect on the overall system performance is higher.

36

7. PRODUCT TEMPERATURE INVESTIGATION One of the main objectives of this thesis work is to know the optimum cabinet operating conditions to obtain good product temperature between 0 and 4ºC in the cooling cabinets and under –18ºC in the freezing cabinets.

7.1. Freezing Cabinets Figure 22 illustrates product temperature measurements in both freezing cabinets. For the measurement, thermocouples were inserted between two product dummies. The selected measuring point positions were chosen based on the assumption that these points will have the warmest product temperature due to the fact that they are close to the warm inlet air side, and therefore will have relatively fast response to the changes in air temperature. Additionally, they are at the top of the products package where it will be exposed to radiation from the surroundings.

(a) (b)

Figure 22 - Product temperature measurement distribution through the freezers – Freezer 2 (a) and Freezer 1 (b)-

Figures 23 and 24 show product temperature measuring points in both freezers. Positions of air inlet and outlet to the evaporator are indicated.

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Time (hh:mm:ss)

Tem

pera

ture

(ºC

)

Air,in FC1

Air,out FC1

Pr1,FC1

Pr2,FC1

Pr3,FC1

Pr4,FC1

Figure 23 - Air and product temperatures in freezer 1 (FC1)

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0:42

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0:55

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1:07

:26

Time (hh:mm:ss)

Tem

pera

ture

(ºC

)

Air,in FC2

Air,out FC2

Pr1,FC2

Pr2,FC2

Pr3,FC2

Figure 24 - Air and product temperatures in freezer 2 (FC2)

As can be extracted from the Figure 22, there is different product temperature distribution. The coldest product line (blue line in Figure 22) is the closest to the cold air entering the cabinet after passing across the evaporator. The warmest product line (pink line in Figure 22) is the furthest to the air outlet and is near the air inlet to the freezer’s evaporator. This is due to the fact that the air in that region has taken heat from the products and mixes with the room’s warm air. The temperature difference between the coldest and hottest product line was around 5ºC.

38

Furthermore, on the warm products line up to 2ºC temperature difference was observed in the first freezer (FC1). This maximum temperature difference is insignificant. As could be extracted from the results of FC1, the point in the middle of the cabinet is colder than the other points at the same line (Pr1, in Figure 23). On the other hand, the environment conditions in each side of the cabinet have important influence in their product temperature. That is, the products placed close to the wall (Pr2 in Figures 22b and 23) and near the humidifier (Pr3) sides have higher temperature than in the middle. In the second freezer (FC2) the product temperature distribution on the warmest side was different. In this case, the coldest product place was near the wall (Pr2, Figure 22a), meanwhile the central product points (Pr1) was the warmest. In this case, 2ºC temperature difference on average was observed. The reason of that has to do with the air movement in the room and around the cabinets. Depending also on how long is the distance between the wall and each measuring point. However, the difference is considered to be insignificant especially at steady state conditions. The controlling parameter of the cabinet was the air inlet in order to maintain the product temperature below -18ºC. Figure 25 illustrates the temperature profiles across the low-temperature evaporator and on the air side. As can be seen in the plot, air enters to the evaporator frontal inlet at approximately -23ºC, then, by passing through the evaporator tubes and fins cools down to around -34.5ºC at the evaporator outlet at least 8ºC is obtained for the superheat and around 2ºC for the approach temperature.

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Freezing Evaporator Length

Tem

pera

ture

(ºC

)

Warmestproduct

temperature

Air temperature

Refrigerant temperature

Approachtemperature

(2ºC)

Superheat(At least 8ºC)

Figure 25 - Different temperature parameters through the freezers

39

7.2. Cooling Cabinets On the medium-temperature level, only one cabinet (MC1) had product temperature measurements installed. Figure 26 shows the temperature measurement points in the cabinet, the points are indicated by the red arrows.

Figure 26 - Product temperature measurement distribution through the cooling cabinet In this case, the aim is to keep the chilled products in the range 0 to +4ºC. The same criteria as in the case of the freezers, described above, are used in distributing the temperature measurement point in the medium-temperature cabinets. Figure 27 shows air, product and evaporating temperatures in the cooling cabinets. The controlling parameter of the cooling cabinets was the air inlet temperature in the same way as in the freezers. In order to reach the desired product temperature the setting temperature value for the entering air was changed until the optimal performance was achieved. In this particular case, the sensitive part was to have the coldest point which is at the rear end of the bottom of the cabinet higher than 0 so the product will not freeze.

40

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12:0

9:52

12:1

5:02

12:2

0:15

12:2

5:25

12:3

0:39

12:3

5:49

12:4

1:02

12:4

6:12

12:5

1:22

12:5

6:36

13:0

1:46

13:0

6:59

13:1

2:09

13:1

7:23

Time (hh:mm:ss)

Tem

pera

ture

(ºC

)Air,in Air,out T_Pr1 T_Pr2 T_Pr3 T_Pr4 Tevap

Figure 27 – Air, product and evaporating temperatures in MC1

The way the test was done was as follows:

1. A certain temperature value was set in the display of the cabinet’s controller for the entering air

2. After the steady-state conditions had been reached, the product temperature range was observed. Then, the setting value was increased or decreased depending on whether any product temperature was equal or below 0ºC or above +4ºC, respectively.

3. The test was finished when every measured product temperature was within the desired temperature range

Looking at the results from the figure, all products measuring points were within the expected temperature range. Around 1.5ºC insignificant difference between the maximum and minimum temperature products in the measured points has been observed.

7.3. Effect of the Roof Fans When open display cabinets are used there is a difference in temperature between the cold environment of the cabinet and the warm exterior. Then, there is a natural exchange of air between the two regions (infiltration). Figure 28 illustrates the effect of infiltration between a refrigerated room and a warm exterior.

41

Figure 28 – Infiltration due to the exchange of air between the cold refrigerated room and

the warm environment [17] As can be seen in the figure, the air pressure of the both cold and warm spaces equalize at the neutral elevation. Above and below this elevation, the air pressures differ due to the static force of the column of air as function of the density. Then, this air pressure difference will cause the cold air to flow out from the refrigerated room at the bottom of the opening and the warm air to flow into the cold room at the top of the opening [17]. The effect of this infiltration produces a loss of energy due to the loss of cold air and the access of the warm air from the surroundings. In order to avoid the infiltration fans were placed in the roof of the cooling cabinets. These fans create an air layer which protects the cold product environment. Figure 29 shows a schematic diagram of an opened vertical display cabinet [18] when an air curtain is used.

42

Figure 29 – Plot of the cooling cabinet with air curtain installed [19]

As is shown in the figure, the warm air enters to the cabinet in the return air grille where the air inlet measurement is placed. Then it passes through the fans that send the entering air to the evaporator to be cooled down. Once the cold air has passed across the evaporator (where the air outlet measurement is located), it goes up by the back-panel which is perforated and allows diffusing the cooled air through the product shelves. The remaining air arrives to the ceiling of the cabinet being discharged downwards to the air inlet grille in order to close the air cycle. When the roof fans are on there is an air jet protection (air curtain) which prevents infiltration. Therefore, the effect of the air curtain on the performance of the cooling cabinets was tested. To compare this, the roof fans from the medium-temperature cabinets were disconnected for a sufficient period to get steady-state conditions. Figures 30 and 31 show air inlet and outlet temperatures in the cabinet respectively for both roof fans switched-on and off cases. The defrost period was, in both cases, 12 hours. The air temperatures in the cabinet have an important influence on the performance. That is, when the air inlet temperature reaches a certain maximum value the solenoid valve opens allowing the refrigerant entering to the evaporator then cooling down the air. When the air inlet reaches the minimum setting value, the valve closes stopping the cooling.

43

In case of having high air temperatures that means that the valve has to open more frequently in order to cool the air to achieve the settings. As can be seen in both figures, although the air frequency is slightly shorter in first half period of the test for the roof fans-on case, later it becomes longer; while, the air cycle frequency in roof fans-off case remains almost constant during the test period, being on average approximately the same value in both cases. Although there is not curtain circulating the air across the cooling cabinet in the switched-off case, observing Figures 31 and 32, there is almost no difference in the air temperature in the cabinet. Figures 32 and 33 illustrate the product and room temperatures when the roof fans were on and off, respectively. Table B summarizes the maximum and minimum product temperature values comparing both cases. As can be seen in those figures, although the air temperatures in the cabinet have almost the same value in both cases, the product temperatures when the roof fans were disconnected are higher than when the air curtain was connected in about 0.7ºC. This may indicate that there is an infiltration of warm exterior in the upper part of the opening as was commented above which increases the product temperatures. Looking at the room temperature in Figure 33 (‘Room_Temp_OFF’ line), when the tests were run, the roof fans off had a room temperature between 17 and 19ºC. In the case when the air curtain was connected (‘Room_Temp_ON’ line in Figure 32), the room temperature was between around 19 and 21ºC. Therefore, in spite having even lower room temperature when the air curtain was disconnected, the roof fans off method had higher product temperatures

44

0

1

2

3

4

5

6

7

8

9

10

14:0

7:59

14:2

6:20

14:4

4:34

15:0

2:55

15:2

1:12

15:3

9:33

15:5

7:47

16:1

6:08

16:3

4:28

16:5

2:44

17:1

1:04

17:2

9:19

17:4

7:39

18:0

5:59

18:2

4:15

18:4

2:35

19:0

0:50

19:1

9:10

19:3

7:26

19:5

5:47

20:1

4:10

20:3

2:25

20:5

0:46

21:0

9:01

21:2

7:23

21:4

5:38

22:0

4:00

22:2

2:21

22:4

0:37

22:5

8:59

23:1

7:14

23:3

5:36

23:5

3:58

0:12

:14

0:30

:37

0:48

:52

1:07

:14

1:25

:30

1:43

:53

2:02

:15

Time (hh:mm:ss)

Air

Inle

t Tem

pera

ture

(ºC

)

Air,in_OFF

Air,in_ON

Figure 30 - Air inlet in the medium-temperature cabinets depending on the switched-on

and off methods

-8

-6

-4

-2

0

2

4

6

8

10

14:0

7:59

14:2

6:20

14:4

4:34

15:0

2:55

15:2

1:12

15:3

9:33

15:5

7:47

16:1

6:08

16:3

4:28

16:5

2:44

17:1

1:04

17:2

9:19

17:4

7:39

18:0

5:59

18:2

4:15

18:4

2:35

19:0

0:50

19:1

9:10

19:3

7:26

19:5

5:47

20:1

4:10

20:3

2:25

20:5

0:46

21:0

9:01

21:2

7:23

21:4

5:38

22:0

4:00

22:2

2:21

22:4

0:37

22:5

8:59

23:1

7:14

23:3

5:36

23:5

3:58

0:12

:14

0:30

:37

0:48

:52

1:07

:14

1:25

:30

1:43

:53

2:02

:15

Time (hh:mm:ss)

Air

Out

let T

empe

ratu

re (º

C)

Air,out_OFF

Air,out_ON

Figure 31 - Air outlet in the medium-temperature cabinets depending on switch-on and

off methods

45

0

1

2

3

4

5

6

7

8

9

10

14:0

4:58

14:2

3:42

14:4

2:29

15:0

1:17

15:2

0:03

15:3

8:50

15:5

7:38

16:1

6:21

16:3

5:11

16:5

3:55

17:1

2:43

17:3

1:32

17:5

0:15

18:0

9:05

18:2

7:55

18:4

6:40

19:0

5:30

19:2

4:15

19:4

3:06

20:0

1:56

20:2

0:42

20:3

9:31

20:5

8:20

21:1

7:06

21:3

5:56

21:5

4:41

22:1

3:32

22:3

2:22

22:5

1:07

23:0

9:58

23:2

8:48

23:4

7:33

0:06

:24

0:25

:10

0:44

:01

1:02

:52

1:21

:38

1:40

:30

1:59

:22

Time (hh:mm:ss)

Prod

uct T

empe

ratu

re (º

C)

15

15,5

16

16,5

17

17,5

18

18,5

19

19,5

20

20,5

21

21,5

22

22,5

23

Roo

m T

empe

ratu

re (º

C)

T_Pr1_ON T_Pr2_ON

T_Pr3_ON T_Pr4_ONRoom_Temp_ON

Figure 32 – Product and room temperatures when the roof fans were connected

0

1

2

3

4

5

6

7

8

9

10

14:0

7:59

14:2

7:00

14:4

5:54

15:0

4:55

15:2

3:57

15:4

2:53

16:0

1:53

16:2

0:48

16:3

9:48

16:5

8:48

17:1

7:44

17:3

6:45

17:5

5:45

18:1

4:39

18:3

3:40

18:5

2:40

19:1

1:36

19:3

0:36

19:4

9:31

20:0

8:32

20:2

7:35

20:4

6:30

21:0

5:31

21:2

4:33

21:4

3:28

22:0

2:30

22:2

1:31

22:4

0:27

22:5

9:29

23:1

8:24

23:3

7:26

23:5

6:28

0:15

:24

0:34

:27

0:53

:28

1:12

:24

1:31

:26

1:50

:28

2:09

:25

Time (hh:mm:ss)

Prod

uct T

empe

ratu

re (º

C)

13

13,5

14

14,5

15

15,5

16

16,5

17

17,5

18

18,5

19

19,5

20

20,5

21

Roo

m T

empe

ratu

re (º

C)

T_Pr1_OFF T_Pr2_OFFT_Pr3_OFF T_Pr4_OFFRoom_Temp_OFF

Figure 33 – Product and room temperatures when the roof fans were disconnected

Method

Parameter

Roof fans ON

Roof fans OFF

Max. product temp. (ºC)

3.6 4.4

46

Min. product temp. (ºC) 1.9 2.6 Product glide2 (ºC) 1.7 1.8

Table B - Comparison between product temperatures depending on roof-fans method

used As a recommendation, the usage of the air curtain will have a positive influence on the product temperature. The cooling capacity of the cabinet should be improved; this has not been measured in the test but the increase in the product temperature is an indication for it.

2 Temperature difference between maximum and minimum product temperatures in the cabinet

47

8. OPTIMIZATION OF THE CASCADE CONDENSER The usage of the cascade condenser heat exchanger produces a penalty in comparison to a two-stage system using one refrigerant. In a cascade system there must be an overlap of temperatures between the condensing temperature of the low temperature stage and the evaporating temperature of the high temperature stage. This will result in lower evaporating temperature in the cascade system and higher power consumption at the high stage. Condensation of the CO2 can take place in different arrangements. There is a need to evaluate and compare the performance of each arrangement and condition in order to optimise the system.

8.1. Indirect System Arrangements Figure 34 shows the two different arrangements tested for the medium-temperature loop solution, arrangement in figure (a) is called ‘thermosyphon’ (Th) and (b) is ‘forced condensation’ (Fc).

(a) (b)

Figure 34 - Schematic diagrams for the cascade condenser arrangements at medium loop: (a) Thermosyphon, (b) Forced condensation [9]

The thermosyphon arrangement operates by connecting the return line from the medium-temperature evaporator to the accumulation vessel and, then, a loop delivers the saturated vapour from the vessel to the cascade condenser so that it condenses and returns back to the vessel. The forced condensation arrangement passes directly the two-phase return refrigerant from the cooling evaporator through the cascade condenser and then it accumulates in the vessel. In the thermosyphon arrangement the refrigerant entering the cascade condenser will be saturated vapour which indicates good heat transfer conditions. In case of the forced condensation arrangement, the quality of CO2 at the entrance of the cascade condenser will be lower than 1 and the heat transfer is expected to be worse than the thermosyphon case. However, the refrigerant mass flow will be higher which may improve the heat transfer over

48

the thermosyphon case. Therefore, it is necessary to investigate both arrangements in order to conclude which solution yields better performance from heat transfer and stability points of view. The test consists of estimating the vapour quality at the exit of the medium-temperature cabinet’s evaporators. Depending on the load, the mass flow of refrigerant was calculated in order to obtain the desired circulation ratio. That is, 4 kW cooling capacity of each medium-temperature cabinet was assumed. Then, the known medium temperature simulator load was provided in steps in order to increase the cooling capacity. Using the complete evaporation enthalpy difference ( cedh ) at the cooling evaporator conditions (-8ºC and 28 bar), this yields,

kgkJbarChdh fgce /42,253)28,º8( =−= (12) Hereby, the mass flow for complete evaporation process was obtained as follows:

ce

simulatorcabinetsce dh

QQm&&

&+

= (13)

Therefore, depending on the desired vapor fraction to have at the evaporator outlet (cascade condenser inlet), the following formula was used to adjust the mass flow of refrigerant in the medium-temperature loop:

Xmm ce

actual&

& = (14)

Furthermore, equation 15 expresses the relation between the vapour fraction and the circulation ratio:

XCR 1

= (15)

During the tests both medium-temperature loop arrangements were studied were investigated for two vapor quality values, X=0.145 (low) and X=0.33 (high), and with different simulator loads. Figure 35 illustrates the pressure in both CO2 and NH3 sides inside the cascade condenser throughout both Th and FC loops tests. As can be seen in the plot, the pressure that was changing during the test was the NH3’s because the NH3 unit controls the pressure in the CO2 tank.

49

26

26,5

27

27,5

28

28,5

29

12:0

1:31

12:2

1:24

12:4

1:21

13:0

1:13

13:2

1:05

13:4

0:57

14:0

0:53

14:2

0:46

14:4

0:37

15:0

0:28

15:2

0:27

15:4

0:24

16:0

0:15

16:2

1:08

16:4

1:05

17:0

0:58

17:2

0:50

17:4

0:48

18:0

0:40

12:2

7:31

12:4

7:23

13:0

7:18

13:2

7:14

13:4

7:08

14:0

7:03

14:2

6:56

14:4

6:55

15:0

6:49

15:2

6:49

15:4

6:42

16:0

6:36

16:2

6:37

16:4

6:32

17:0

6:29

17:2

6:27

17:4

6:23

18:0

6:23

18:2

6:18

Time (hh:mm)

CO

2 Pr

essu

re (b

ar)

2

2,2

2,4

2,6

2,8

3

3,2

3,4

3,6

3,8

4

NH

3 Pr

essu

re (b

ar)

P_CO2_Th P_CO2_FC

P_NH3_Th P_NH3_FC

Figure 35 - Pressure for both CO2 and NH3 sides in the cascade condenser for the

thermosyphon and forced condensation arrangements Figure 36 presents the CO2 mass flow circulated in the medium-temperature level loop for both thermosyphon and forced condensation arrangement tests. Each estimated mass flow circulation was obtained by opening and closing valves until the desired valued was reached. The room temperature remained almost constant.

0

0,025

0,05

0,075

0,1

0,125

0,15

0,175

0,2

0,225

0,25

0,275

0,3

0,325

0,35

0,375

0,4

0,425

0,45

12:0

1:31

12:2

1:34

12:4

1:41

13:0

1:43

13:2

1:45

13:4

1:47

14:0

1:53

14:2

1:56

14:4

1:57

15:0

1:59

15:2

2:07

15:4

2:14

16:0

2:46

16:2

3:23

16:4

3:25

17:0

3:28

17:2

3:30

17:4

3:39

18:0

3:48

12:3

0:41

12:5

0:47

13:1

0:48

13:3

0:54

13:5

0:59

14:1

1:03

14:3

1:06

14:5

1:15

15:1

1:21

15:3

1:29

15:5

1:32

16:1

1:36

16:3

1:47

16:5

1:56

17:1

1:59

17:3

2:09

17:5

2:13

18:1

2:23

18:3

2:30

Time (hh:mm:ss)

MC

s C

O2

Mas

s Fl

ow (k

g/s)

19

19,5

20

20,5

21

21,5

22

22,5

23

23,5

24

24,5

25

25,5

26

26,5

27

Roo

m T

empe

ratu

re (º

C)

ForcedCondensation_MCsMassFlow

Thermosyphon_MCsMassFlow

ForcedCondensation_Room Temp

Thermosyphon_Room Temp

High vapor quality in both arrangements

Low vapor quality in both arrangements

Higher simulator's load for low X

Higher simulator's load for high X

Figure 36 - Refrigerant mass flow rate in the medium-temperature level and room

temperature during both thermosyphon and forced condensation arrangements tests

50

8.1.1. Thermosyphon Loop Arrangement The test for the thermosyphon arrangement was repeated in order to verify that there is no difference between low and high quality cases in the refrigerant through the cooling evaporator, as can be seen in Figure 37 with reasonable accuracy. That is logical since the indirect loop is only delivering saturated vapour towards the cascade condenser from the accumulating tank whichever is the vapor fraction at the outlet of the medium-temperature evaporator. Therefore, in spite of having different vapor quality at the return line of the evaporator, actually there is quality equals to 1 at the cascade condenser entrance in any case.

0

0,5

1

1,5

2

2,5

3

3,5

4

13,5 14 14,5 15 15,5 16 16,5 17 17,5 18 18,5 19 19,5 20 20,5 21 21,5 22

Cooling Capacity (kW)

Tem

pera

ture

Diff

eren

ce A

cros

s th

e C

asca

de C

onde

nser

(ºC

)

Thermosyphon_Low_X

Thermosyphon_High_X

Figure 37 - Temperature difference across the cascade condenser for different cooling

capacities for thermosyphon arrangement As can be seen in Figures 36 and 37, the mass flow increases by about 30% with the high capacity which should improve the HTC. This improvement seems to be very small since the temperature difference increase with higher capacity. Furthermore, it has been noted that this arrangement was quite stable within the whole test.

8.1.2. Forced Condensation Loop Arrangement On the contrary of the thermosyphon loop, at the forced condensation arrangement, it was found important differences between the low and high qualities as Figure 38 illustrates. The biggest differences are at lower capacities.

51

0

0,5

1

1,5

2

2,5

3

3,5

4

4,5

5

5,5

15,5 16 16,5 17 17,5 18 18,5 19 19,5 20 20,5 21 21,5 22 22,5 23 23,5

Cooling Capacity (kW)

Tem

pera

ture

Diff

eren

ce A

cros

s th

e C

asca

de C

onde

nser

(ºC

)

ForcedCondensation_Low_X

ForcedCondensation_High_X

Figure 38 - Temperature difference across the cascade condenser for different cooling

capacities for forced condensation arrangement Looking at Figures 36 and 38, the high quality case increases the HTC with high capacity; the improvement may be due to the increase of mass flow rate. Regarding to the low vapor fraction, in this case happens the same as was commented before for the thermosyphon loop. Moreover, it has noticed that when this arrangement is used, the system presents certain instability which implies longer time to get steady-state conditions. This instability was higher with high vapor qualities.

8.1.3. Comparison between both Indirect System Arrangements Figure 39 illustrates a comparison between both Th and Fc arrangements at two different qualities and cooling capacities. The figure shows that both the thermosyphon and forced condensation arrangements present approximately the same heat transfer through the cascade condenser at low vapor fraction values as was pointed out above. On the other hand, when higher vapour qualities are reached at the exit of the evaporator, the forced condensation arrangement’s heat transfer is worse than the thermosyphon’s.. This can be explained by looking at Figure 40. Figure 40 shows local HTC value based on Shaw’s correlation [20] at different circulation ratios for both arrangements. In the figure, the entering conditions to the cascade condenser (or evaporator’s outlet conditions) in both forced condensation’s cases are indicated with an ellipse.

52

0

0,5

1

1,5

2

2,5

3

3,5

4

4,5

5

14 14,5 15 15,5 16 16,5 17 17,5 18 18,5 19 19,5 20 20,5 21 21,5 22 22,5

Cooling Capacity (kW)

Tem

pera

ture

Diff

eren

ce a

cros

s th

e C

asca

de C

onde

nser

(ºC

)

Thermosyphon

ForcedCondensation_Low_X

ForcedCondensation_High_X

CR=1.8X=0.55

CR=1.9X=0.52 CR=2.1

X=0.47

CR=4.4X=0.22

CR=4.4X=0.22

CR=4X=0.24

CR=1 X=1

Figure 39 - Comparison between both direct and indirect medium-temperature loop

arrangements at different cooling capacities

Figure 40 – Local HTC based on Shaw’s correlation [20] with different CR for the

arrangements under investigation One important aspect to have into account about Figure 40 is that the HTC values from Shaw’s correlation are overpredicted twice as high as the experimental values observed by Jang and Hrnjak [21]. This means that the

53

actual lines would be closer to the vapor quality axis. Furthermore, experimentally has been observed that mass flow does not improve the HTC as much as the correlation predicts. This indicates that the actual mass flux lines for each CR are closer to each other. Shaw’s correlation also fail to capture the effect of heat flux since Jang and Hrnjak’s results showed strong influence on heat transfer. Anyhow, Shaw’s correlation can be used as tend line for the current evaluation. Considering the comments from above, the HTC of the case of Fc with low CR (high quality) might be higher with high capacities than the other Fc’s case due to the steepness of the slope of each curve. This would explain the high improvement of the HTC and the decrease in the overlap temperature. Therefore, similar heat transfer performance has been noticed between Fc with high CR and thermosyphon. Moreover, high capacity values the difference among all the investigated arrangement seems to decrease. But, from the stability point of view, thermosyphon has been observed more stable. Then, thermosyphon loop should be applied.

8.2. Cascade Condenser Arrangements Further arrangements were consequently performed to evaluate and compare variations in the low-temperature level. In the first arrangement (Figure 41a), called thermosyphon, the hot gas discharge from the CO2 compressor passes though the liquid accumulated in the tank, allowing the hot gas to be de-superheated by boiling off some of this liquid. Meanwhile, the two-phase returning from the medium-temperature evaporators goes directly to the tank. The saturated vapor in the tank circulates in a thermosyphon loop where the condensate returns back to the accumulation tank. In the second arrangement (Figure 41b), the return line from the cooling evaporators is mixed with the hot discharge gas from the freezing side just before entering directly to the cascade condenser. And, finally, in the third arrangement, called forced condensation, (Figure 41c), the medium-temperature side is connected in the same way as in the thermosyphon arrangement. But, in this case, the hot discharge gas from the low-temperature level passes directly to the cascade condenser after mixing with the saturated vapor from the accumulation tank.

54

(c)

Figure 41 - Different diagrams for the cascade condenser arrangements at low-temperature level: (a) Thermosyphon (b) Second arrangement and (c) Forced

condensation [9] The second arrangement will result in quality at the inlet of the cascade condenser but, in this case, the mass flow will be higher and the heat transfer would be better as it was shown in some cases in the indirect system arrangements. Since this arrangement is quite similar to the forced condensation arrangement in the indirect system loop (Figure 34b), which was previously evaluated, priority was given to the forced condensation arrangement. The thermosyphon arrangement is the base arrangement that the system is run with. Test for both thermosyphon and forced condensation arrangements are done by maintaining the low-temperature level at a specific capacity which keeps the CO2 compressor running at constant speed. Then load is being added from the medium-temperature simulator and only the first medium-temperature cabinet was running all the time during the test. Figure 42 shows the pressure boundaries of each refrigerant (NH3 and CO2) and the room temperature during both tests. As can be seen in the plot, the pressures were almost constant during the whole test. Condensing pressures were around 13 bars and 27.5 bars for NH3 and CO2 respectively. Evaporating

55

pressures were 11.4 bar and 2.8 bar for CO2 and NH3 respectively. There was only around 2ºC temperature difference in the room due to running activities in the laboratory.

0

2

4

6

8

10

12

14

16

18

20

22

24

26

28

30

9:57

:10

10:0

6:05

10:1

4:57

10:2

3:47

10:3

2:42

10:4

1:33

10:5

0:28

10:5

9:18

11:0

8:09

11:1

7:00

11:2

6:00

11:3

4:52

11:4

3:42

11:5

2:36

12:0

1:30

12:1

0:20

12:1

9:55

12:2

9:14

12:3

8:10

12:4

7:03

12:5

5:53

13:0

4:48

13:1

3:42

13:2

2:32

13:3

1:32

13:4

0:22

13:4

9:14

13:5

8:09

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7:01

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sure

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m T

empe

ratu

re (º

C)

P_NH3_H P_NH3_L

P_CO2_H P_CO2_L

Room Temp

Forced Condensation Loop Thermosyphon Loop

Figure 42 - Evaporating and condensing pressures from both NH3 and CO2 side and

room temperature during the test Figure 43 illustrates the CO2 and NH3 compressors’ speeds. As can be observed in the figure, the CO2 compressor maintained stable speed since the load from the low-temperature level was constant during the whole test. However, the compressor speed of the NH3 (pink line in Figure C) had different speeds based on its frequency control according to the load added during the test.

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sor S

peed

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Lower cooling capacity Lower cooling capacity

Figure 43 - NH3 and CO2 compressors’ speeds during the test

Figure 44 presents the temperatures of two important points in the cascade condenser in the forced condensation arrangements. One measures the temperature of the compressor discharge after the junction which separates each loop arrangement (point 1 in Figure 44). The other shows the temperature at the inlet of the cascade condenser in the CO2 side (point 2 in Figure 44). The latter indicates the temperature after mixing the refrigerant in both medium-temperature loop and hot gas discharge in the direct arrangement.

57

Figure 44 - Schematic diagram where measurements before the cascade condenser are

indicated Figure 45 shows the measurements before the cascade condenser obtained from the tests. These measurements are the temperature of the indicated points in Figure 44. As can be observed in the ’Point 1 Temperature’ line in the figure with Fc loop running, when higher load is supplied from the medium-temperature simulator lower temperature is reached at the inlet of the cascade condenser in the CO2 side. This is logical since when higher simulator’s load is added, higher mass flow is provided to the medium-temperature level. The higher saturated vapor mass flow is delivered from the CO2 tank to the cascade condenser which is mixed with the hot discharge gas. The results from the plot indicate a 28.7ºC at the inlet of the cascade condenser when no simulator’s load is added. 22.2ºC temperature was reached when one simulator’s step was supplied. The next step provided from the simulator resulted in about 17ºC The hot discharge gas was stable during the whole test for this arrangement at a value around 100ºC as can be seen in Figure 45. Once the arrangement was changed to the thermosyphon’s, ’Point 1 Temperature’ line indicates the saturating temperature of the vapor delivered from the accumulation tank.

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pera

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)

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Forced Condensation Loop Thermosyphon Loop

28.7ºC

22.2ºC 17ºC -7.3ºC

Lower cooling capacity

Lower cooling capacity

Figure 45 - Temperatures before the cascade condenser from the low-temperature side

The CO2 mass flow rate through the cascade condenser can be seen in Figure 46.

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CO

2 M

ass

Flow

Rat

e (k

g/s)

m_dot_CC_CO2_Fc

m_dot_CC_CO2_Th

Lower cooling capacity Lower cooling capacity

Thermosyphon LoopForced Condensation Loop

Figure 46 - CO2 mass flow rate through the cascade condenser during both

arrangements Looking at the figure it can be observed that the CO2 mass flow rate in the thermosyphon arrangement was between 14-6.5% higher than the forced condensation arrangement’s during the whole test.

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Figure 47 illustrates the temperature difference across the cascade condenser for different cooling capacities. The points in the plot were taken using the average in the stable conditions.

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14 14,5 15 15,5 16 16,5 17 17,5 18 18,5 19 19,5 20 20,5 21 21,5 22 22,5 23Cooling Capacity (kW)

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ce A

cros

s th

e C

asca

de C

onde

nser

(ºC

)

Forced Condensation

Thermosyphon

Figure 47- Comparison between different arrangements for the cascade condenser at

low-temperature side The results extracted from the test in the thermosyphon arrangement confirm earlier results as can be verified comparing the ’thermosyphon’ line in Figure 47 with Figure 39. As can be seen in Figure 47, there is a temperature difference between both arrangements of 0.2ºC all along the range. This temperature difference is considered to be insignificant and negligible from the overall system perspective. Moreover, it has been noticed that both arrangements presented good stability within the whole range of study. Consequently, both arrangements have similar heat transfer performance and no optimum arrangement has been found.

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9. DEFROST There is a necessity for periodically defrosting air-cooled evaporators that operate at sub-zero temperatures which will cause frost to collect on the evaporator’s surface, obstructing the air flow through the evaporator and degrading the heat transfer. The period the evaporator should be defrosted depends on the nature of the installation and the method of defrosting. In general, the length of the defrost period is determined by the degree of frost accumulation on the evaporator and by the rate at which heat can be applied to melt off the frost. Generally, the degree of frost accumulation will depend on the humidity in the air, the season of the year and the frequency of defrosting. As a general rule, the more frequently the evaporator is defrosted, the smaller is the frost accumulation and the shorter is the defrost period required [22]; but in that case, higher energy is needed and higher increase in product temperature results. There must be an optimum value which is one of the points of this test.

9.1. Methods of Defrosting Under Investigation In the experimental rig, there are two different ways of defrosting installed in the freezing cabinets: hot gas and electric heaters defrost; while the cooling cabinets are equipped with conventional electric defrost method.

9.1.1. Hot Gas Defrost Method Figures 48a and 48b is a schematic diagram illustrating the active lines and components in the low-temperature level loop in normal and hot gas defrost operations, respectively.

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(a) (b)

Figure 48 - Schematic diagram of the active lines and components at the low-temperature level in (a) normal operation and (b) hot gas defrost operation

Hot gas defrost (HGD) utilizes the hot gas directly discharged from the compressor as a source of heat to defrost the evaporator. The HGD consists of a bypass equipped with a solenoid valve installed between the compressor discharge and the evaporator (point 2 in Figure 48). Just before the defrost period starts, the refrigeration is stopped closing the valve from the refrigerant inlet line of the cabinet (point 1 in Figure 48). When the solenoid valve is opened, a fraction of the hot gas from the compressor discharge bypasses the condenser and enters the evaporator at a point just beyond the refrigerant control. This method uses the sensible heat of the hot gas and no condensation occurs.

9.1.2. Electric Heaters Defrost Method Electric heaters defrost (EHD) uses electric resistance heaters along the evaporator to melt the frost accumulated on its surface. The EHD cycle is initiated by closing a solenoid valve in the liquid line causing the evaporator to be evacuated. At the same time, the heating elements in the evaporator are energized. After the evaporator is defrosted, the heaters are de-energized and the system put back in operation by opening the liquid line solenoid.

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Both methods are automatically operated and a timer is used to shut down the freezing unit for a fixed period of time at regular intervals. Both the number and the length of the defrost periods can be adjusted to suit the individual installation. Since it is undesirable to keep the system out of service for any longer than is necessary, the length of the defrost period should be carefully adjusted so that the cabinets are properly defrosted and the system is back in service as soon as possible. There are two ways to stop the defrost process: by using the maximum defrost time or by a temperature signal placed in the coldest part of the evaporator which indicates that defrost is completed.

9.2. Defrost Tests To know the optimum defrost procedures, different periods and frequency of defrost have been tested. The settings to stop defrosting have been investigated and the position of the defrost sensor has been examined. In the beginning, different defrost periods were tested and an optimum preliminary interval was obtained for each defrost method, but the humidity in the laboratory was lower than the real one in supermarket operation approximately 18%. Humidity was removed from the room by frost forming on the evaporator’s cold surfaces. Therefore, in order to have conditions similar to a real supermarket case, a humidifier (Figure 49) was placed in the laboratory in order to maintain a good level of humidity of about 30%3.

Figure 49 - Pictures of the humidifier and the pipe extension installed in the laboratory

3 Based in personal contacts in refrigerating companies such as WICA for a humidity value in real supermarket case

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An extension pipe, pointed out in Figure 49 to diffuse the humidity in the laboratory was used to obtain a more homogeneous distribution in the room environment.

9.3. Freezing Cabinets The criteria for optimization of the defrost interval and method were to evaluate the success of removing the accumulated frost and the increase of product temperature after defrost.

9.3.1. Defrost Sensor Location One important aspect related to defrost is the location of the defrost sensor. This sensor stops the defrost operation when a temperature value is reached and, therefore, its location inside the cabinet has crucial importance. The defrost sensor is an indicator that the accumulated frost has been completely removed and the evaporator is being heated up. In the beginning 6ºC defrost stop temperature was set according to provider’s data; but it was changed to 12ºC in order to be totally sure that the entire evaporator was clean of frost. Figure 50 shows an example of how the defrost sensor stops defrosting in the cabinet when the 12ºC setting value is reached and the cabinet, after 10 minutes of drip off, is put back to refrigeration.

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Time (hh:mm:ss)

Tem

pera

ture

(ºC

)

Inlet Evaporator Temperature

Outlet Evaporator Temperature

Defrost Sensor Temperature

Defroststarts

Defrostfinishes

Drip-offtime

Refrigerationstarts

Figure 50 - Example of how the defrost sensor acts in stopping defrosting at 12ºC

As can be seen in the plot, when the defrost sensor temperature reaches 0ºC it remains at that temperature around 10 minutes (green line in the figure), indicating that the melting process of the frost is taking place. After that, the

64

steepness of the defrost sensor temperature changes radically and reaches quite fast the setting value of the defrost stop temperature. Since the way of defrosting is different for the methods investigated, different optimal places for the defrost sensor has been found. In the beginning, the defrost sensor was placed in the upper side of the fins of the refrigerant outlet according to the provider suggestions, as indicated in Figure 51. This preliminary place for the defrost sensor was due to the fact that the cabinets were designed for defrosting by electric heaters. The heaters are placed in the frontal area of the evaporator and it was assumed that this area would be heated up first meanwhile the rear area would be the colder. Then, the sensor was kept in a relatively cold place.

Figure 51 - First place of the defrost sensor. The red-blue arrow indicates the flow of the air across the evaporator

After early tests, it was found that the frontal side of the evaporator was the area where the frost accumulation accumulates most. Figure 52 shows how under normal operation the frost is accumulating on the frontal area of the evaporator while the rear area remains clean after defrosting.

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Figure 52 - Picture of the evaporator after defrosting when the defrost sensor was placed on the rear area

Therefore, the defrost sensor was moved from the original place to the frontal side as can be seen in Figure 53.

Figure 53 - Second place of the defrost sensor and final for EHD Figures 54a and 54b illustrate before and after electric heaters defrosting when the defrost sensor is placed on the frontal area of the evaporator, respectively. As can be observed in the plots, the frost has been melted letting clean the frontal area of the evaporator. Only some frozen droplets remain on the evaporator’s surface after refrigeration starts. In this case, it was the optimum place for electrical heaters defrost method.

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(a) (b)

Figure 54 - Pictures of (a) before and (b) after defrosting when the defrost sensor is placed on the frontal side of the evaporator

For the HGD method when the sensor is located in the upper side of the frontal area was not sufficient to remove the frost from the evaporator. Figures 55a, 55b and 56 illustrate with detail the blockage of the frontal side of the evaporator and the temperature profile of the defrost sensor in hot gas defrosting, concluding that this area is the most important part to take care of in defrosting. When the frontal side is blocked, then, there is less air flowing though the evaporator and there will be loss in cooling capacity. If the frost continues accumulating throughout the evaporator, considerable frost thickness is reached. Therefore, it takes more time to remove the frost or even not to melt it at all. Once the frontal side is totally blocked, the humid air accumulates more and more frost as can be observed in Figure 57.

(a) (b)

Figure 55 - Pictures of the evaporator in HGD method when the defrost sensor is placed in the upper side of the frontal area

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Evaporator Outlet Temperature

Defrost Sensor Temperature

Defroststarts

Defrostfinishes

RefrigerationRefrigeration

Figure 56 - Evaporator and defrost sensor temperatures for hot gas defrosting with 6ºC stop temperature when the defrost sensor is placed in the upper side of the frontal area

Figure 57 - Frontal area of the freezing evaporator totally blocked of frost after several cycles of hot gas defrost

Since there was an accumulation of frost on the cabinet’s floor, the defrost sensor was placed in lower place as Figure 58 illustrates in order to be sure that this frost melts during defrost. But it was not enough and the frontal area of the evaporator was totally blocked again after several defrost cycles. So the defrost sensor was lowered even further closer to the cabinet’s floor, as Figure 59 indicates.

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Figure 58 - Detail of the third place for the hot gas defrost sensor and the beginning of the process of the frost accumulation from the drain pan

As can be seen in Figure 58, the melted frost is accumulated in the drain pan and starts the process of partly blocking the lower side of the evaporator’s frontal area. The reason of why the liquid frost is not drained is due to two possible reasons: the used drip-off time is not longer enough or the slope of the drain pan is not high enough. To increase the drip-off long time enough would imply more delay to put the system back to refrigeration mode, increasing the product temperature and it may not assure complete removal of the liquid. To increase the steepness of the drain pan would be a redesign problem which would imply space limitations.

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Figure 59 - View of the freezing cabinet’s evaporator after hot gas defrosting and the optimum place for hot gas defrost sensor

As a conclusion, the evaporator’s area where the frost accumulates the most has been identified. So the defrost sensor measures the temperatures of the coldest area where frost melts last, which should be a proper indication to terminate defrosting. Furthermore, due to the different nature of defrosting it has found different sensor location for the defrost methods under investigation. Figures 53 and 59 show the optimum places for the defrost sensor in the frontal area for both EHD and HGD, respectively.

9.3.2. Defrost Intervals Once the defrost sensor was fixed in its optimal place for defrost, the optimum defrost interval was tested. As was pointed out in the introduction of the defrost section, the defrost period of an evaporator should have an optimum length which achieves the aim of removing completely the accumulated frost in the evaporator’s surface with the lowest energy consumption required. The criteria used to obtain the optimum defrost period for defrosting was to set a defrost interval and after several cycles of defrost to observe if the setting value was enough to clean the evaporator from frost. The cabinets were opened before and after defrost in order to see the state of the evaporator and some photos were taken. If the defrost period could not remove the accumulated frost from the evaporator, a shorter defrost interval was set. The test started with a setting value of 12 hours but it could not fulfil the aim of clean the evaporator of frost. The following tests were made in the way explained in the previous paragraph. Then, reducing the defrost period it was

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found a good setting value of 8 hours for both hot gas and electric heaters defrost methods. Figure 60 illustrates the insufficiency of the defrost interval value since as can be seen in the figure the frontal side of the evaporator is completely covered with frost after several cycles of defrost.

Figure 60 - Frost accumulation in the evaporator due to the insufficiency of the defrost period

Figures 61a and 61b show the conditions before and after defrost takes place for the optimum interval respectively.

(a) (b)

Figure 61 - Photos of the freezing evaporator (a) before and (b) after defrost for the optimal defrost interval

9.3.3. Effect of Fans during Defrost As Dossat [22] suggests in the explanation of the EHD, it should have switched-off the fans during defrost operation. But, the current freezers were designed to

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have connected the fans then. Therefore, a test was suggested in order to evaluate the effect of the fans when EHD is taking place on the performance of the freezing cabinets. Moreover, the study went further investigating also their influence of the HGD. In order to obtain the optimum behaviour of each defrost method and to know their effect, both EHD and HGD methods were tested with and without fans running during defrost operation.

9.3.3.1. Electric Heaters Defrost Method Figure 62 illustrates the electric heater distribution in the freezing evaporator. The first heater is placed at the inlet air channel in order to warm the air entering the cabinet mainly warming the frontal area of the evaporator. The second and third heaters warm the evaporator’s body and the air flowing through it for proper distribution of the heat. Therefore, the electric heater system was designed to have the air circulating through the cabinet.

Figure 62 - Electric heater distribution in the freezers Figures 63 and 64 illustrate the temperature profiles in the freezers during both electric heaters defrost with and without fans operation. Both figures show the air, evaporator and defrost sensor temperatures. These temperature measurements indicate how the cabinet performs during defrost. Furthermore, Table C adds the information of the increase of the product temperature and summarizes the temperature parameters from the plots. Observing both figures and table, it can be observed that to have the fans running during defrost slightly increases the defrost length and increases the product temperature by about 1.5-2ºC. This is due to the fact that when the fans are off the uneven heat distribution makes the area around the heaters heats up

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faster, especially that it is situated close to the electric heaters. When the fans are on the heat is removed from the area around the sensor and more even distribution is achieved. Furthermore, with the fans on, the circulated air is heated up across the electrical heaters then cooled down when it passes through the refrigerated space, heating the products. This can be confirmed by looking at the air inlet and outlet temperature difference in Table C.

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Time (hh:mm:ss)

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pera

ture

(ºC

)

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Air,out_ON EHD

Tevap,in_ON EHD

Tevap,out_ON EHD

Defsensor_ON EHD

Defroststarts

Defrostfinishes

Drip-offtime

Refrigeration

Figure 63 - Electric Heaters defrost with fans running during defrosting (12ºC stop

temperature, 10 min. drip-off time) In figure 63 where the fans are on, it can be clearly seen that all the temperatures, except the exit air temperature, are following the same trend which indicates a good distribution of heat along the evaporator. This is due to the fact that the heaters are placed in the frontal area of the evaporator and there is an additional heater in the air supply channel which heats up the entering air as was commented above. It can be seen in the plot that the defrost sensor has a constant value of 0ºC for a period of time which is due to the melting frost around the sensor; this indicates that the sensor is in a good position.

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Air,out_OFF EHD

Tevap,in_OFF EHD

Tevap,out_OFF EHD

Defsensor_OFF EHD

Defroststarts

Defrostfinishes

Dripp-offtime

Refrigeration

Fans delay

Figure 64 - Electric Heaters defrost without fans running during defrosting (12ºC stop

temperature, 10 min. drip-off time, 5 min. fan delay) In figure 64, it can be seen that when the fans are off the temperatures in the frontal side of the evaporator are much higher than on the rear part, about 25°C difference when defrost finishes. This indicates a bad temperature distribution and with time it may have frost accumulating in the rear part of the evaporator, this can be seen by looking at the temperature of the evaporator exit (Tevap_out). The defrost sensor temperature reaches higher value than the set point, when the heaters are switched off the temperature of the sensor keeps rising. With the fans off , the heater located in the inlet air channel does not have the effect that it is installed for; it does not properly contribute to heating the evaporator.

EHD Fans method

Parameter

ON

OFF

Increase in prod. temp. (ºC) 3.99 2.31 Defrost time (min) 34.2 31.9 Max. air inlet T (ºC) 12.06 13.74

Max. air outlet T (ºC) 22.98 7.5

Table C - Different values for the cabinet parameters during EHD fan method It has been noticed that after the fans turned-off method, there was some frost remaining in the rear area of the evaporator mainly due to the fact that the heater distribution in the cabinet was not designed to have the fans disconnected during defrost. Otherwise, as can be seen in the figures, the fans switched-off method would be a good way to melt the accumulated frost since it separates the evaporator from the refrigerated space.

74

Therefore, to have the fans running during defrost contributes to achieve a good homogeneous warming of the evaporator but it has the penalty of increase the product temperature more than needed. On the contrary, with fans on the increase of product temperature is smaller and the energy consumption of the fans is saved but uneven heat distribution in the evaporator is reached. Then, a good solution would be to have the fans off during defrost but heaters distribution redesign is needed. Furthermore, a still better solution would be to delay the fans start after defrost is finished and the system is put back to refrigeration; by this way, it is possible to cool the warm air which surrounds the evaporator and to avoid rejecting it into the refrigerated space.

9.3.3.2. Hot Gas Defrost Method Figures 66 and 67 show the air, evaporator, defrost sensor and evaporator’s hot gas inlet temperatures profiles in the freezers in hot gas defrosting when fans are on and off, respectively. Moreover, Table D adds the information of the increase of the product temperature and summarizes the temperature parameters from the plots. It has been noticed that the hot gas mass flow (valve opening) was also a sensitive parameter for defrost and its opening was critical for the operation. Low mass flow resulted in bad defrost which stopped by time not temperature and large opening will cause the evaporating temperature to increase which affects the other cabinet’s cooling. Then a compromise solution was taken. Figure 65 shows a schematic diagram indicating the temperature measurements in the evaporator in the HGD. The hot gas line is located just after the expansion valve as was commented above when the hot gas defrost method was introduced. One thermocouple (point 1 in the figure) measures the hot gas temperature when entering the evaporator. The second measurement (point 2) indicates the temperature of the first bend in the evaporator’s tube as Figure 65 indicates. Finally, there is another thermocouple (point 3 in Figure 65) which measures the evaporator’s outlet temperature.

75

Figure 65 - Temperature measurements in the freezing evaporator

When the hot gas valve (HGV in Figure 65) is opened the hot gas is entering the evaporator’s inlet tube at around 54ºC (‘Tdef_evap_in’ line in Figures 66 and 67). The hot gas starts to heat the evaporator’s first tube and then cools down quickly. The heat is transferring from the hot gas to the tube’s surface of the evaporator by convection. Then it gradually warms up the fins by conduction. Moreover, there is heat conduction along the tube which helps in the longitudinal heating of the tube. Observing the difference among the hot gas inlet temperature and the evaporator’s inlet and outlet can have an idea of how the hot gas is heating the evaporator and, then, the melt of the frost is achieved. This process is quite slow as can be seen in the plots and table, especially compared with the defrost process in EHD method (Figures 63 and 64, and Table C). As a result from the figures and table, the defrost length and increase in product temperature are higher having connected the fans during defrost. It does not reach the defrost stop temperature and cuts by maximum time.

76

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Time (hh:mm:ss)

Tem

pera

ture

(ºC

)

-50

-40

-30

-20

-10

0

10

20

30

40

50

60

70

HG

Eva

pora

tor i

nlet

Tem

pera

ture

(ºC

)

Air,in_OFF HGD

Air,out_OFF HGD

Tevap,in_OFF HGD

Tevap,out_OFF HGD

Defsensor_OFF HGD

Tdefr_evap,in_OFF HGD

Defroststarts

Defrostfinishes

Drip-offtime

Fans delay

Refrigeration

Figure 66 - Hot Gas defrost without fans running during defrosting (12ºC stop

temperature, 10 min. drip-off time, 5 min. fans start delay) Looking at figure 66 can be seen that with the fans off the difference between the air and the coldest area of the evaporator –which the defrost sensor is located- is approximately 10ºC when defrost finishes. Moreover, even the air temperatures are below 0ºC during defrosting period. These indicate a good separation between evaporator and air. As can be observed in the figure, around 3ºC temperature difference along defrost is found between the defrost sensor and evaporator tube’s surface. This illustrates how is the heat transfer by conduction from the evaporator’s tubes until the extreme of the fins where the defrost sensor is placed.

77

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Time (hh:mm:ss)

Tem

pera

ture

(ºC

)

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0

10

20

30

40

50

60

70

HG

Eva

pora

tor I

nlet

Tem

pera

ture

(ºC

)

Air,in_ON HGD

Air,out_ON HGD

Tevap,in_ON HGD

Tevap,out_ON HGD

Defsensor_ON HGD

Tdefr_evap,in_ON HGD

Defroststarts

Defrostfinishes

Drip-offtime

Refrigeration

Figure 67 - Hot Gas defrost with fans running during defrosting (12ºC stop temperature,

10 min. drip-off time, 5 min. fans start delay) The temperature profile for the hot gas defrosting with fans on, similar to Figure 66, is shown in Figure 67. As a result from the plot and table, the defrost length and increase in product temperature are higher having connected the fans during defrost. It does not reach the defrost stop temperature and cuts by maximum time. As can be observed in the plot, the temperature difference between the air and evaporator in the freezer is smaller when the fans are off during defrost, being around 3ºC in this case. This indicates that the air instead of helping to melt the frost is being opposed to the hot gas effect. The hot gas acts from inside the evaporator’s tubes and its heat reach the tips of the fins by conduction, at the same time heating the air around the evaporator and between the fins. Therefore, the hot gas melts the frost off the evaporator’s surface. With the fans on, the air passes through the products heating them up and then passes across the evaporator cooling it down which delays the defrost process.

HGD Fans method

Parameter

ON

OFF

Increase in prod. temp. (ºC) 14.12 8.45 Defrost time (min) 180 (max.) 94.5

Max. evaporator outlet T (ºC) 12.01 14.95 Max. air inlet T (ºC) 8.86 3.22

Max. air outlet T (ºC) 9.78 0.62

Table D - Fan effects in the parameters around freezing cabinets.

78

Comparing figures and table above, it is concluded that there are really significant differences in time and performance between the HGD’s cases with fans on and off. Having the fans off during HGD gives good defrost results and abolishes the need to separating the evaporator from the refrigerated space not to heat up the products.

9.3.3.3. Comparison between HGD and ELD Since the main point of the earlier tests was to optimise both systems of EH and HG defrosts depending on the fans’ effect, this section compares the two defrost methods optimised based on the above analysis. This is, to compare EHD running fans (EHwF) with HGD without fans (HGwoF) during defrost. Defrost sensors are in appropriate locations in both cases. Comparing Figures 63 and 66, and Tables C and D can extract that the EHwF is more efficient method than the HGwoF because EHwF can remove the frost from the evaporator in shorter defrost time. The HGwoF defrost time is almost three times the EHwF’s. Furthermore, in spite of separating the evaporator from the refrigerated space during defrost in HGwoF case, its increase in product temperature is higher (around twice) than EHwF’s due to the longer defrost time. Therefore, it has been noticed that the optimum defrosting method for this installation, based on increase in product temperatures and defrost length, is EHD with the fans running during defrost. Possible improvement would be to isolate the evaporator from the products’ space during defrost. This would allow circulating the hot air across the evaporator increasing the efficiency of defrost and it would avoid to reject the warm air to the refrigerated space. Then, after defrost is finished the refrigeration would start cooling the warm air until it is cold enough to be circulated again across the products.

9.3.4. Energy Consumption During Defrost In the previous section the discussion was about comparing the performance of both defrost methods, however, it is important to complete this analysis by evaluating the energy consumption during defrost in both cases.

9.3.4.1. Electric Heaters Defrost Method The electric energy consumed by the heaters is calculated by measuring the electric current and voltage during defrost period. The energy consumption of the fans, which are required to run this method, is not included in this analysis. The electrical intensity and voltage per line in the heaters are around 2.42 and 226 amperes and volts, respectively. Therefore, an electrical power of around 0.9473 kW is supplied during electrical defrost by the heaters. Since the defrost time is 34.2 minutes, then, it yields an energy consumption of 0.54 kWh each defrost period.

79

9.3.4.2. Hot Gas Defrost Method In order to calculate the energy added to the system during the hot gas defrost, there are two different ways. In both methods, the low-temperature total mass flow is obtained from the conditions before and after the CO2 compressor, as the following formula shows:

compcompCOeleccomplosses dhmE ⋅=⋅ ,2, &&η (16) Where the compressor heat losses ( lossesη ) are assumed as 7%. The energy balance in the freezers and the freezing simulator during normal operation (NOP) yields:

NOPFCNOPFCNOPsimNOPTotal QQQQ ,2,1,,&&&& ++= (17)

where NOPTotalQ ,

& is the cooling capacity from the low-temperature side during the refrigerating period. Assuming that both freezers have the same cooling capacity and the capacity from the simulator was fixed to 2 kW,

NOPfreezerNOPsimNOPTotal QQQ ,,, 2 &&& ⋅+≅ (18) Clearing the cooling capacity from the freezer from the above equation,

2,,

,NOPsimNOPTotal

NOPfreezer

QQQ

&&& −

= (19)

Furthermore, the cooling capacity from the freezer cabinet can be calculated using the air side as follows:

NOPairairairNOPfreezer dTCpmQ ,, ⋅⋅= && (20) It is possible to obtain the air mass flow from the fans through the freezing evaporators:

NOPairair

NOPfreezerair dTCp

Qm

,

,

⋅=

&& (21)

which can be considered as constant. In addiction, the isentropic efficiency of the CO2 compressor is calculated during normal operation case:

80

inout

inis

comp

isis hh

hhdhdh

−−

==η (22)

Method 1 In the first method, the low-temperature total mass flow from the conditions of the CO2 compressor is utilized to calculate the total cooling capacity from the low-temperature side during defrosting operation (DOP):

DOPcompDOPLTCODOPeleccomplosses dhmE ,,,2,, ⋅=⋅ &&η (23)

DOPNDCDOPDCDOPsimDOPTotal QQQQ ,,,,&&&& ++= (24)

Then, the capacity from the defrosting cabinet ( DOPDCQ ,

& ) is obtained from the total cooling capacity reducing it by the cooling capacity from the simulator and from the non-defrosting cabinet ( DOPNDCQ ,

& ):

DOPDefrostingNonDOPTotalDOPNDCDOPsimDOPTotalDOPDC QQQQQQ ,,,,,, )( −−=+−= &&&&&& (25) where the latter is the cooling capacity from the freezing simulator and the non-defrosting cabinet. On the other hand, the COP during defrost period is the total capacity from the low-temperature side divided by the CO2 compressor power supply,

DOPcomp

DOPTotalDOP E

QCOP

,

,

&

&= (26)

This can be used to calculate the part of the compressor power consumption from the non-defrosting side as follows:

DOP

DOPDefrostingNonDOPDefrostingNon COP

QE ,

,−

− =&

& (27)

In order to calculate the defrost power consumption, it has to reduce the total compressor power by the non-defrosting part,

DOPDefrostingNonDOPcompHGD EEE ,, −−= &&& (28) Using this method, the CO2 compressor power required to add for defrosting is around 1.32 kW and since the defrost time was 94.5 minutes, 2.08 kWh is obtained for the energy consumption in the hot gas defrost.

81

Method 2 On the other hand, in the second method, the non-defrosting mass flow during defrost (DOP) is extracted from the air side,

DOPNDevapDOPNDDOPairairairDOPfreezer dhmdTCpmQ ,,,,, ⋅=⋅⋅= &&& (29) Using the conditions before and after the freezing simulator, can also be calculated the simulator mass flow, because the cooling capacity from the simulator is known.

DOPsimDOPsimDOPsim dhmQ ,,, ⋅= && (30) Then, the non-defrosting mass flow is the sum of both parameters,

DOPsimDOPNDDOPDefrostingNon mmm ,,, &&& +=− (31) Hereby, the hot gas flow through the defrosting cabinet can be obtained as the subtraction of the non-defrosting mass flow from the total one:

DOPDefrostingNonDOPTotalHG mmm ,, −−= &&& (32) From then, it is estimated the compressor conditions working only with the non-defrosting side in normal operation or refrigeration period without the influence of the defrosting cabinet. For this, it is assumed an external superheat value ( SHextT ,Δ ) of around 7ºC and a suction temperature using the following formula:

SHextDOPDefrostingNon

outsimDOPsimoutNDCDOPNDCestimatedinC T

mTmTm

T ,,

,,,,,, Δ+

⋅+⋅=

−&

&& (33)

The enthalpy difference between the inlet and the isentropic outlet ( estimatedisdh , ) is calculated by this temperature ( estimatedinCT ,, ) and the inlet and outlet conditions of the compressor. So, using the isentropic efficiency obtained from the normal operation period (equation 22), it is possible to extract the enthalpy difference in the compressor.

is

estimatedisestimatedcomp

dhdh

η,

, = (34)

Thus, the non-defrosting side from the compressor power consumption follows,

estimatedcompDOPDefrostingNonDOPDefrostingNon dhmE ,,, ⋅= −− && (35)

82

This can be used to get the defrosting power consumption by subtracting it from the total measured power consumption during the defrost period, in the same way as in equation 22. Utilizing this second method to calculate the energy supplied for the hot gas yields similar energy consumption as for the first method, which can be verified in Figure 68.

0

0,2

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0,8

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1,2

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1,6

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9

Time (hh:mm:ss)

HG

D P

ower

Sup

ply

Req

uire

d (k

W)

E_dot_Hot_Gas_Defr_FC1_M1

E_dot_Hot_Gas_Defr_FC1_M2

Figure 68 – Comparison between the methods used to calculate de CO2 power

consumption required for HGD Figure 69 shows the defrost capacity of the hot gas. Around 4.6 kW is observed in the figure.

83

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2,5

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Time (hh:mm:ss)

Def

rost

Cap

acity

(kW

)

Figure 69 – Defrost capacity of the hot gas

9.3.4.3. Comparison between Hot Gas and Electric Heaters in Energy Consumption

Comparing the results obtained from the calculations of the energy consumption during defrost, the energy supplied in the HGD method is greater than that for the EHD by:

%91.7310007.2

54.007.2=×

−

Otherwise, EHD consumes around four times less energy than HGD during defrost. Therefore, it can be concluded that the EHD method is better in this facility than the HGD method from the energy consumed for defrosting point of view.

9.3.4.4. Hot Gas Mass Flow during HGD An important parameter in order to complete the energy analysis and to know the performance of the hot gas defrost method in the freezing cabinets is the hot gas mass flow provided from the CO2 compressor to the defrosting cabinet. This parameter can be used to calculate how much percentage of the total mass flow through the compressor is required to clean of frost one freezing cabinet when it is placed in a real size installation (50 kW). The way to calculate the hot gas mass flow was introduced above in the second method to calculate the energy consumption. This results in approximately

84

0.017 kg/s on average during steady-state conditions, as can be seen in Figure 70.

0

0,0025

0,005

0,0075

0,01

0,0125

0,015

0,0175

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0,0225

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0,0275

0,037:

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Time (hh:mm:ss)

Hot

Gas

Def

rost

Mas

s Fl

ow (k

g/s)

Figure 70 – Hot gas mass flow through defrosting cabinet during HGD

Figure 71 shows the total mass flow rate through the CO2 compressor during HGD. As can be seen in the figure, it gives around 0.0295 kg/s on average. This means that the mass flow rate for HGD is approximately 57.63% from the total supplied.

0

0,005

0,01

0,015

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Time (hh:mm:ss)

CO

2 To

tal M

ass

Flow

Rat

e (k

g/s)

Figure 71 – Total mass flow rate through the CO2 compressor during HGD

85

Furthermore, other important parameter to know is the hot gas defrost mass flow required to complete defrosting in the same time as electric heaters defrost method; that is in about 30 minutes. The process to calculate that is as follows: The energy required to melt the accumulated frost throughout the evaporator with the current hot gas mass flow is,

currentcurrentdefHGD tQE ⋅= ,& (36)

Which should be the same energy spent but, in this case, in only 30 minutes ( requiredt ),

requiredrequireddefHGD tQE ⋅= ,& (37)

Then, equalling both formulas and knowing that,

currentHGcurrentHGcurrentdef dhmQ ,,, ⋅= && (38) This can be able to reach the following equation,

requiredrequiredHGrequiredHGcurrentcurrentHGcurrentHG tdhmtdhm ⋅⋅=⋅⋅ ,,,, && (39) Now, assuming that the enthalpy difference is not changing, this gives,

)/(05355.0017.015.330

5.94,,, skgm

ttmm currentHGrequired

currentcurrentHGrequiredHG =⋅≅⋅=⋅= &&& (40)

where, in the last step, the current hot gas mass flow average was taken. Now, how much mass flow rate of HG required for a normal medium size installation with 50 kW in the low-temperature level is evaluated. Figure 72 shows the total mass flow required for refrigeration operation of both freezers and two steps of the freezing simulator (about 6.4 kW). As can be observed in the plot, this results 0,021 kg/s on average.

86

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0,0025

0,005

0,0075

0,01

0,0125

0,015

0,0175

0,02

0,0225

0,025

0,0275

0,03

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:09

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:56

Time (hh:mm:ss)

CO

2 M

ass

Flow

Rat

e (k

g/s)

Figure 72 – Total mass flow required for normal operation when two simulator’s steps

and both freezers are supplied in the low-temperature level Therefore, the required mass flow in normal operation for a medium size low-temperature level yields,

skgkWkWmm kWkW /164.0

4.650

4.650 ==& (41)

Then, the required HG mass flow to melt the frost in 30 minutes (equation 40) represents the 32.65% of the total during refrigeration operation.

9.4. Cooling Cabinets In the same way as with the freezing cabinets, the optimum defrost interval was investigated for the medium temperature cabinets. In this case, although both cabinets are equipped with conventional electric defrost, instead, a simple method of defrosting is used. This consists of stopping refrigeration at the medium-temperature cabinet until the evaporator warms up enough to melt down the frost, after which refrigeration resumes. This process can also be referred to as drip off instead of defrost.

9.4.1. Results and comments After the humidifier was placed in the laboratory, it was found that both cabinets had different defrost needs. For example, the cabinet number 1 (MC1) did not need any defrost period to avoid the frost accumulation. But, on the contrary, the other cabinet (MC2) needed at least 8-hour interval to remove the frost owing to maybe its position near to the humidifier.

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Figure 73 confirms the effect of frost blockage in the second medium-temperature level cabinet (MC2) when no-defrost period is used.

Figure 73 - Detail of MC2’s evaporator totally covered with frost

Figures 74a and 74b shows the evaporators of the medium-temperature cabinets, MC1 and MC2, when no- and 8-hour defrost period is used in MC1 and MC2, respectively.

(a) (b)

Figure 74 - Cooling evaporators in their optimal defrost periods: a) MC1, no-defrost period; b) MC2, 8-hour defrost

Figure 75 represents the air, evaporating, product and room temperatures when 8-hour defrost was set in the controller of the cooling cabinets.

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-9

-7

-5

-3

-1

1

3

5

7

9

11

13

12:0

0:07

12:0

6:32

12:1

2:52

12:1

9:15

12:2

5:35

12:3

1:59

12:3

8:19

12:4

4:42

12:5

1:02

12:5

7:26

13:0

3:49

13:1

0:09

13:1

6:33

13:2

2:53

13:2

9:16

13:3

5:36

13:4

2:01

13:4

8:21

13:5

4:44

14:0

1:04

14:0

7:28

14:1

3:52

14:2

0:12

14:2

6:36

14:3

2:56

14:3

9:19

14:4

5:39

14:5

2:04

14:5

8:24

15:0

4:47

15:1

1:13

15:1

7:33

15:2

3:56

15:3

0:16

15:3

6:40

15:4

3:00

15:4

9:23

15:5

5:43

Time (hh:mm:ss)

Tem

pera

ture

(ºC

)

-10

-8

-6

-4

-2

0

2

4

6

8

10

12

14

16

18

20

22

24

Roo

m T

empe

ratu

re (º

C)

Air,in Air,out T_Pr1 T_Pr2 T_Pr3 T_Pr4 Tevap Room_Temp

Figure 75 - Temperature measurements in the medium-temperature cabinet when 8-hour defrost is used

The room temperature (‘Room_Temp’ line in the figure) was between 19.7 and 21.5ºC during the whole test as Figure C shows. The evaporating temperature (‘Tevap’ line) has a fluctuation according to the tank pressure, being around -6.3ºC. When defrost stars (approximately at 13:23) the valve closes stopping the refrigeration. This allows removing the accumulated frost in the evaporator’s surface by the entering warm air. This can be seen in the plot looking at both air’s inlet and outlet temperatures. Defrost terminates when the defrost sensor reaches 6ºC starting the refrigeration again. The way how cooling is controlled in these cabinets has to do with the fact that they do not need to be defrosted when they are running under normal conditions. When the set point for the controller is reached then the valve closes and the fans keeps working providing warm air to the evaporator which melts any frost that was formed while the valve was open. If the cooling load is high and the valve does not close as often as it does under normal operation (at start up for example or when new products are added as in a real supermarket) then the valve will be opened for longer period of time and frost will be accumulating on the cold evaporator.

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10. CONCLUSIONS The NH3/CO2 cascade system which replicates a real supermarket refrigeration installation has been tested in a laboratory environment. Tests have been run for overall evaluation of the system. Good correlation for the volumetric efficiency of the NH3 compressor has been found in order to have an accurate calculation of the NH3 mass flow rate. This is used to calculate the cooling capacities and the COP of the system. The calculated COP of the low stage gives a value around 3.4 meanwhile the high stage COP is approximately 3.8. The temperatures during the test were 34.4ºC for condensing NH3, -36ºC for evaporating CO2 and about -9ºC for the intermediate temperature. Product temperatures in both medium and low-temperature level have been investigated. Tests have been made in order to maintain a product temperature between 0 and 4ºC in the cooling cabinets and below –18ºC in the freezers. The experiments showed success to keep the product within the desired range of temperatures in both medium- and low-temperature levels. In the freezers case, different product temperature distribution has been observed. On the medium-temperature side, good product temperature distribution has been noted. Air curtain effect on the performance of the cooling cabinets has been evaluated. It was observed that the air curtain has an influence on product temperature but not inlet air temperature. The circulation ratio in the medium-temperature simulator and cooling cabinets has been changed in different indirect system arrangements for the cascade condenser in order to know its influence on heat transfer. In thermosyphon arrangement, the two-phase return line from the evaporator is connected to the CO2 tank, from which saturated vapor is sent to the cascade condenser in order to condense, after that the condensate returns back to the tank. In forced condensation arrangement, the return line is directly delivered to the cascade condenser, after which the condensate goes back to the tank. Circulation ratio highly influences the heat transfer performance in forced condensation arrangement. The differences in performance of the different indirect systems investigated are reduced when high capacity is added. Furthermore, it has been noticed that the thermosyphon arrangement has better stability. In the forced condensation arrangement high instability was observed when low circulation ratios are used. No significant difference in heat exchanging performances has been observed between different arrangements for the cascade condenser with the low temperature circuit connected. In thermosyphon arrangement the hot discharge gas from the compressor is delivered to the CO2 accumulation tank boiling some liquid refrigerant and then by using a loop the CO2 enters to the cascade condenser as saturated vapor and the NH3 boils off along the heat exchanger from saturated conditions at the inlet. In case of forced condensation

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arrangement the hot discharge gas mixes with the saturated vapor from the CO2 tank at the inlet of the cascade condenser. Hot gas and electric heaters defrost methods have been investigated. A procedure has been applied to find the optimum defrost performance. The defrost sensor location in the freezers has been fixed in the optimum place for each investigated defrost method. It has been noticed that the frontal area of the evaporator is the place where frost accumulates the most. Upper and lower sides of the frontal evaporator’s area have been found as the optimum locations for the electric heaters and hot gas defrost methods, respectively. It is also found that the steepness of the drain pan has high influence in the blockage of the frontal area of the evaporator. Higher slope should be used to improve the drainage of the melted frost and to avoid blocking the evaporator. A systematic procedure has been followed to obtain the optimum defrost interval in the current freezers. A certain value for the defrost period was set in the controller. After several defrost cycles the evaporator’s surface was observed. If the chosen defrost interval is found to be insufficient to melt the frost then the interval is decreased. 8-hour defrost has been found as optimum interval for both hot gas and electric heaters defrost methods. The effect of fans during freezers defrost operation has been evaluated. Tests have been run connecting and disconnecting the fans in both investigated defrost methods. It has been noticed that to connect the fans during EHD reaches good temperature distribution along the evaporator. The uneven temperature distribution observed when switching off the fans was mainly due to the fact that the heaters distribution was designed so air will carry the heat to be distributed along the evaporator. Then, if EHD with fans off method is used, redesign that takes into account redistributing the heaters in the evaporator should be considered. Higher increase in product temperature during fans on case has been observed since the warm rejected from the defrost operation heats up the products; meanwhile in case of having fans off, it results in good separation between the evaporator and the refrigerated space. Concerning the HGD method, important differences have been observed. To have connected the fans during defrost is not good solution for this method since it has been noticed that due to the nature of the HGD method the cold external air supplied by the fans has a negative effect by cooling down the evaporator. Comparing both optimum performance for each defrost method, it has been noticed that HGD takes around three times longer time than EHD. HGD has higher increase of product temperature (about twice EHD’s) mainly due to its longer defrost period. It has been also noticed that from the energy consumption point of view EHD consumes around four times less energy than HGD. Therefore, EHD with fans connected has been found to be the optimum defrost method for the current system and conditions.

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Evaluation of defrosting of the cooling cabinets has been performed. The results from the tests showed that both medium-temperature cabinets have different defrost needs due to different operating and surrounding conditions. 8-hour defrost/drip-off interval was found to be good setting for the cooling cabinets.

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