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Item Development of a rapid liquid freezer : a thesis presented in partial fulfilment of the requirements for the degree of Doctor of Philosophy in Food Technology at Massey University, Manawatū, New Zealand(Massey University, 2021) Morel, JolinSmall sheep dairy farms often make insufficient volumes of milk for economic daily collection and are limited by transport distances to processors. A method of long-term on-farm storage of milk would enable the industry to grow. Freezing would allow extended milk storage on farms. But existing methods of freezing for on-farm applications have shortcomings around materials handling, labour requirements and product quality. The project reported in this thesis aimed to develop the engineering science behind an economically viable freezing method that would improve on current methods. The first period of this project focused on two freezer designs which were thought to be promising: Rolling Droplet Freezing (RDF) and Falling-Film Flake freezing (FFF). RDF was selected as the initial focus of the research program and consisted of a system where droplets of milk would roll down an angled super-hydrophobic surface against a cold air flow and freeze. RDF was abandoned due to concerns about construction costs and operating reliability. In a condensing atmosphere, droplets rolling on superhydrophobic surfaces occasionally transitioned from a Cassie-Baxter wetting state to a Wenzel wetting state, which caused the droplets to stick. FFF was then developed further. A pilot scale unit was designed and constructed, and preliminary pilot-scale trials that were conducted with pure water and ovine milk reconstituted from powder. The partition coefficient of FFF was measured as 0.946 at an operating temperature of -30°C. At higher operating temperatures the partition coefficient was reduced. Detaching frozen solids by applying a burst of heat to the freezer/ice interface was studied and this method of detachment was successful with pure water, but ineffective for ovine milk. The development of FFF was put on hold with the conception and development of the continuous tubular freezer. Ice formed in a solution can show morphologies ranging from highly dendritic structures with entrapped solutes, which are homogenous on a gross scale, to large crystals of pure ice with solutes rejected and compressed into inter-crystalline spaces. To investigate which sheep milk components influence ice morphology at various freezing rates, whole milk was separated into skim milk and into a casein-free serum phase. A simulated sheep milk ultrafiltrate was also prepared. The morphology of the ice/sample interface was observed in a custom-built microscope stage at freezing front velocities from <0.5 μms⁻¹ to 50 µms⁻¹ with a spatial temperature gradient of 35-38 Kcm⁻¹. The morphology arising from extremely rapid freezing front velocities was investigated by supercooling slides on a temperature-controlled stage and observing the nucleation and recalescence of the samples. The morphology of ice at the interface changed from a planar to columnar and then to dendritic as freezing front velocity increased, with the transitions from one morphology to another occurring at lower speeds in more complicated solutions. A map of freezing front behaviours was developed. The transition between interface morphologies was at different velocities and transition differed based upon the interface velocity. At lower interface velocities a columnar interface grew directly from a planar starting condition. At higher velocities an intermediate dendritic zone formed, which then settled into a columnar interface. The ice formed by rapid freezing from subcooled solutions was highly dendritic, with ice growth rates of approximately 21,000 μms⁻¹, which was close to the diffusion-limited ice growth rate in water of similar degrees of supercooling. The morphology of frozen ovine milk was also studied by Cryogenic Scanning Electron Microscopy (Cryo-SEM): Milk was frozen by three different methods-slow quiescent freezing (SF), rapid directional freezing (DF), and droplet freezing in LN₂. Ice crystals rejected unfrozen solids into the region between crystals in all samples, including those frozen by immersion into liquid nitrogen. There was a distinct difference in morphology between the SF and DF samples, with the bands of unfrozen solids being significantly smaller in DF samples, and the long axes of ice crystals were aligned with the direction of heat flow. SF samples lacked any particular ice growth direction, and ice crystals were orders of magnitude larger. Lactose crystallisation was observed in some SF samples but was not observed in any DF samples. Fat globules were engulfed in ice crystals in DF samples, but rejected in SF samples. To study the effects of frozen storage temperature and time, samples of raw ovine milk were stored frozen at -10°C, -18°C and -28°C to -30°C for up to 8 weeks. Further samples were stored below -20°C for 6 months. After thawing at 20°C, samples were tested for a range of properties and serum samples were collected by separating the fat phase and micellar casein phase by centrifugation. A gel was observed in milk stored at -10°C for 4 weeks and 8 weeks but was not observed in milk stored at lower temperatures. The gel dispersed under heating and homogenisation. There was no change observed in the pH, or serum protein level of thawed samples after frozen storage at any temperature. The whiteness of the milk decreased during frozen storage and the yellowness increased. Both of these changes were reversed on homogenisation. The serum Ca²⁺ levels in milk stored at -10°C and -18°C dropped over the storage period, while no trend was seen in milk stored below -28°C, indicating that the migration of Ca²⁺ may play a role in the formation of gels after frozen storage. Milk that had been stored below -20°C for 6 months had a similar viscosity and appearance to fresh milk. A possible mechanism for the formation of gels at -10°C, but not -18°C or -28°C lies in the altered solute environment, and the physical agglomeration of milk components in the spaces between ice crystals, driving the gelation of closely packed casein micelles, with Ca²⁺ stabilising this network. It is well established in literature that the viscosity of an unfrozen phase increases by several orders of magnitudes as it decreases in temperature and approaches a glassy state. This increased viscosity reduces protein mobility and solute diffusion, which reduces the rate of gel formation. The tendency for frozen milk particles to bind together during frozen storage was evaluated. Frozen pellets of whole ovine milk were stored under weights at -10°C and -18°C and pellets of frozen concentrated milk stored at -18°C and -28°C. Ovine milk pellets bound together at -10°C but not -18°C, while concentrated milk bound together at -18°C, but not -28°C. This can be linked to the volume and viscosity of the unfrozen phase in these samples. Differential scanning calorimetry was used to determine the fraction of freezable water frozen at any temperature. The melting onset temperature was observed, and this was used to determine the solids content maximally freeze concentrated solution (𝑋𝑠(𝑇𝑚)). 𝑋𝑠(𝑇𝑚)=0.875 for whole ovine milk 𝑋𝑠(𝑇𝑚)=0.85 for skim milk, and 𝑋𝑠(𝑇𝑚)=0.81 for ovine milk serum. This was also determined for whole ovine milk by the magnitude of the overall latent heat release during melting, which gave a value for whole milk of 𝑋𝑠(𝑇𝑚)=0.85±0.016. A partial phase diagram for ovine milk was generated from the data collected. The insights generated from observing both the dendritic morphology of high velocity ice fronts and progressive freezing behaviour led to conceptualising a novel tubular freezer, subsequently constructed. It was hypothesised that reducing the volume or area of ice in contact with the freezer wall, due to the inclusion of unfrozen product, could reduce the adhesion strength between a frozen product and the freezer wall. By controlling the outlet temperature, the volume fraction of unfrozen product could be controlled. The adhesion strength could thereby be controlled, and a set of operating conditions could be found that would allow a mostly frozen product to be extruded as a solid from a cooled tube by a high-pressure pump. This was tested on a benchtop scale (up to 5mL/minute, with a freezer internal diameter of 4.2mm and cooled length of 500mm), with ovine milk, fruit juice, fruit pulp, concentrated coffee, bovine cream and concentrated milks. The system successfully froze all samples. The operating pressure was found to increase with increased frozen fraction, and therefore with decreased operating temperature. The ice morphology of milk and juice frozen by this equipment was imaged by cryo-SEM and by optical microscopy. The ice crystals were radially aligned, increasing in size closer to the centre of the frozen product plug, which was expected due to the heat flows and the relationship between freezing front velocity and feature sizing. This positive preliminary result led to the construction of a larger scale prototype unit which consisted of a spiral tube with a length of 5000 mm, and an internal diameter of 10 mm. This was used successfully for a product flowrate of approximately 6 kghr⁻¹.Item Moisture transport processes and control of relative humidity in refrigerated facilities : a thesis presented in partial fulfilment of the requirements for the degree of Master of Engineering at Massey University, Palmerston North, New Zealand(Massey University, 2007) Sujau, MaumoonIncreasingly air relative humidity (RH) is becoming an important design and operational variable for refrigerated facilities. An integrated dynamic model of the main heat and moisture transfer mechanism in a refrigerated facility was developed. Specific features of the model that enabled RH to be predicted were: • Multiple air zones to represent variation of temperature and RH with position. • A single zone evaporator model with dehumidification based on a straight line approach to the saturation condition at the surface temperature. • Condensation and evaporation of water from surfaces and structures in the facilities. • Evaporator defrost assuming that a fraction of the defrost heat melts frost and the rest heats the evaporator and refrigerant mass. • Hot gas bypass with liquid refrigerant desuperheating to prevent the compressor operating into vacuum. • Moisture sorption by packaging associated with the product. The model was validated against data collected from a walk-in cool store 3.3m wide by 4.4m long by 3.0m high. The cool-store was cooled by an air cooled direct expansion HFC-134a refrigeration system with electric defrost, a suction line heat exchanger and electronic evaporation pressure regulating (EPR) valve for temperature control. To mimic the different design and operating conditions extra sensible and latent heat loads were provided by the cool store lights, up to 5 kW of electric heaters, and an ultrasonic humidifier. For the validation room trials fan speed, coil size, sensible load, latent loads and temperature set point were varied. Other conditions were held constant as far as possible and the room was operated for at least two defrost cycles. For the coolstore the model computed about 70 ordinary differential equations and more than 160 algebraic equations which were solved using Matlab 6.5, with the ODE45 solver. The measured and predicted store air temperature, RH, refrigerant suction and discharge temperatures and pressures showed good agreement for most of the trials during both pull-down and the mainly steady-state operation between defrosts. Differences in measured and predicted RH and refrigeration system operating conditions were largely explained by uncertainty in model input data, measurements and calibration; and imprecision of the actual refrigeration control system and particularly the hot gas bypass capacity control and the expansion valves. This suggests that the model is a useful tool for the design and optimisation of passive or active RH control strategies for refrigerated stores. Trials were also undertaken to quantify the effect of defrost frequency on the coolstore performance. Defrost efficiency and defrost duration were both proportional to defrost interval and doubled as defrost interval increased from 6 hours to 30 hours. For short defrost intervals; temperature control was poorer due to the frequent pull-downs. For longer defrost interval the room RH was lower and temperature control was poorer due to frost induced decline of evaporator performance The optimal defrost interval for the particular cool store was 8 to 12 hours. Overall energy use did not change significantly due to the use of EPR temperature control and the low latent heat loads used.Item Performance of a transcritical carbon dioxide heat pump for simultaneous refrigeration and water heating : a thesis presented in partial fulfilment of the requirements for the degree of Master of Technology at Massey University(Massey University, 1998) Yarrall, Michael GeorgeMany industrial processes require both refrigeration to less than 0°C and water heating to greater than 60°C. Traditional independent refrigeration and boiler systems have relatively poor energy efficiency, whilst conventional heat pumps can provide both cooling and heating but are limited in terms of the temperature lift that can be achieved. A novel heat pump using CO 2 as the refrigerant in a transcritical cycle has been proposed as a new technology that can overcome these disadvantages. The use of CO 2 as a refrigerant has many advantages. It is environmentally benign, safe, and has good thermodynamic properties, especially compared with fluorocarbons. The transcritical cycle involves evaporation of CO 2 at constant temperature and pressure below the critical point to provide refrigeration, while cooling of the CO 2 occurs at temperatures and pressures above the critical point to provide heating of water. The objective of this project was to design and construct a prototype transcritical CO 2 heat pump to simultaneously provide refrigeration and water heating, and to test its performance over a wide range of operating conditions. The prototype CO 2 heat pump had a nominal cooling capacity of 90 kW at -6°C and nominal water heating capacity of 127 kW from 10°C to 90°C. The prototype was designed to operate with a suction pressure of 30 bar and discharge pressure of 130 bar. The major components were a gas cooler, recuperator, flooded evaporator, low pressure separator/receiver, compressor, expansion valve, connecting piping and a control system. All components were standard high pressure equipment used by the natural gas processing industry. The gas cooler had a reasonably unique design to ensure close to pure counter-current heat exchange between the cooling CO 2 gas and the water being heated, both of which had relatively low flowrates. The compressor used was an open crankcase, reciprocating type with special gas seals on the piston rod to prevent CO 2 leakage. Refrigeration capacity (suction pressure) was controlled by varying the compressor speed. Water heating capacity was controlled by both using the expansion valve to control the CO 2 discharge pressure and varying the water flowrate through the gas cooler. The main problem encountered during commissioning of the prototype was CO 2 leakage through the compressor piston rod seals. Alternative sealing systems were tried, but the leakage remained an on-going problem that prevented prolonged operation of the prototype, such as would be necessary in industrial applications. Performance of the prototype was determined by energy balances based on measurements of CO 2 and water flowrate and temperature when it operated at steady-state. The energy balances generally agreed to within 6%. Trials were performed with suction pressures from 29.6 to 35.5 bar, discharge pressures from 80 to 130 bar, with hot water outlet temperatures from 65°C to 90°C, and evaporator water inlet temperatures from 11°C to 21°C. When heating water to 90°C and providing refrigeration at 1°C (35.5 bar suction pressure), the maximum overall Coefficient of Performance (COP) achieved was 5.4 at a discharge pressure of 114 bar. Below this optimum discharge pressure, the COP declined due to gas cooler heat transfer limitations (lower compressor discharge temperature led to lower temperature difference in the gas cooler and high CO 2 outlet temperature). Above the optimum, the decline in thermodynamic and compressor efficiency as pressure ratio increased caused the COP to decrease. The maximum heating and cooling capacities were about 13% less than the design values. This was attributed to the lower than expected volumetric efficiency of the compressor. The performance of the heat exchangers were generally close to the design values when allowances for lower than design water flowrates were taken into account. As expected, when suction pressure was reduced to 29.6 bar (-6°C), there was up to a 10% decrease in optimum COP as well as reduced heating and cooling capacity. When heating water to 65°C rather than 90°C, the optimum COP was about 20% higher. When suction pressure or hot water outlet temperature was decreased, the optimum discharge pressure became slightly lower due to the gas cooler heat transfer being less of a limitation on overall system performance. Addition of oil to the CO 2 did not reduce the CO 2 leakage sufficiently to allow long-term operation without recharging, and had minimum impact on the performance of the gas cooler, recuperator and compressor. However, oil fouling caused a significant drop in heat transfer performance of the evaporator. The measured prototype performance agreed well with process simulations of the equipment and with results for similar laboratory scale equipment reported in the literature. Therefore, simulations could be used to optimise component and system design with a reasonable level of confidence. It was shown that the biggest increase in COP could be achieved by improving compressor isentropic efficiency rather than increased heat exchanger size. Overall, the concept of the transcritical CO 2 heat pump for simultaneous refrigeration and water heating was proven and the required energy efficiency was sufficiently high that the heat pump is likely to be economically competitive with traditional heating and cooling systems. Further work should concentrate on improving compressor design to eliminate CO 2 leakage and to improve both isentropic and volumetric efficiency.Item The effect of favourable and unfavourable frost on air cooling coil performance : a thesis presented in partial fulfilment of the requirements for the degree of Master of Technology, at Massey University(Massey University, 1994) O'Hagan, Anthony NoelThe most common type of air cooling coil used in the refrigeration industry is the finned tube heat exchanger. The performance of such coils can be greatly hindered by frost formation, which will occur when the coil surface temperature is both below the dewpoint of the air passing over it, and below 0°C. Frost reduces performance, both through the increased thermal resistance of the frost layer, and by reduction of the air flow through the coil. Whilst frosting on coils is influential on performance, there is comparatively little information available on the performance of finned tube heat exchangers under frosting conditions. Smith (1989) has proposed an "unfavourable" frost formation theory. The theory states that unfavourable frost formation occurs when the line representing the temperature and humidity of the air passing through the coil, crosses the saturation line of the psychrometric chart. This criteria is more likely to occur under conditions of high relative humidity, low sensible heat ratio (SHR), and/or high refrigerant-to-air temperature difference (TD). Under unfavourable conditions it is suggested that the frost will be of particularly low density, which would cause coil performance to decline to a much greater extent for the same total frost accumulation, than under "favourable" frosting conditions. The objectives of this study were to measure the change in performance of a cooling coil under frosting conditions, and to assess the validity of the unfavourable frost formation theory. A calorimeter style coil test facility was used, that allowed coil performance to be measured as frost accumulated in a manner consistent with coil operation in industrial practice (i.e. declining air flowrate and a wide range of SHR's). The data collected supported the concept of unfavourable frost formation with a more rapid decline in performance for operation with low SHR, than that at high SHR, for the same total frost accumulation. Some recovery of coil performance was observed when operation at low SHR (with rapid performance deterioration) was followed by a period of high SHR operation. Equations were developed that allowed the theoretical conditions for the formation from favourable to unfavourable frosting to be quantified. The measured change in the rate of coil performance deterioration with frost buildup was dependent on air and coil conditions, in a manner consistent with these equations. The transition between favourable and unfavourable frost formation appeared to be related to the lowest temperature on the coil surface rather than the mean surface temperature. Satisfactory predictions of frost formation types were obtained by using the refrigerant evaporation temperature as an approximation to the lowest coil surface temperature.Item The application of cogeneration systems to the cooling of food and buildings in East Timor : a thesis presented in partial fulfilment of the requirements for the degree of Master of Technology at Massey University(Massey University, 2001) Saldanha, Estanislau de SousaCogeneration is generation of both heat and power simultaneously using a single primary energy input. Cogeneration recovers "waste heat "from a conventional power generation plant to produce useful energy, leading to the increased overall efficiency of fuel input. This also achieves cost savings, and reduces greenhouse gas emissions where fossil fuels are used. The objectives of this study are to assess the technical and economic viability of a cogeneration system for the cooling of food and buildings in East Timor. The findings of this research provide a basis for recommending action and further research to East Timor's decision-makers on energy issues. Technical assessments in this study focus on cooling, electricity demand, and fuel supply as the basis for choosing the type and size of a cogeneration system. The financial viability of the cogeneration system is assessed using net present value (NPV) and sensitivity analysis. The NPV of the cogeneration system is compared with the NPV of conventional energy supply for cooling and electricity. There is low demand for cooling for comfort and food preservation in East Timor, due to low levels of industrial and commercial investment, and the vast majority of people still living in poverty. Although cooling demand is low overall, numerous government and commercial buildings have installed cooling systems. In this study, six buildings (2 office buildings, a bank, a hotel, a university and a mini-market) were selected based on their relatively high cooling demand and their geographic proximity to one another. The cooling demand of these six buildings was modeled based on a room-by- room approach. The results showed that their overall hourly cooling demand averages 600 kilowatt-cooling, while peak load was 707 kilowatt-cooling. This cooling demand was primarily driven by ambient temperature, number of people present and lighting load. Power demand in East Timor is low. The total operable power supply capacity for the entire country is 22 megawatts, of which more than half is located in Dili. Electricity demand is predominantly driven by residential consumption, rather than commercial and industrial consumption. Although there is low electricity demand, East Timor faces an immediate electricity deficit of 24 megawatts, which is higher than the existing operable capacity. In the six selected buildings, the overall peak and average electricity loads were 489 kW and 422 kW respectively. This load was mainly driven by air conditioning, computers, and lighting applications during working hours. Electricity generation relies on diesel, which is imported from Indonesia. Diesel will remain the main source to generate electricity due to a lack of feasible alternatives. East Timor is rich in natural gas both offshore and onshore. However, until now there has been no plan to provide natural gas distribution pipelines to East Timor. Based on the cooling and electricity demand and fuel availability, diesel was chosen to drive the cogeneration systems. The size of the cogeneration system was selected so as to fulfill both the electricity demand in the six selected buildings and be able to export surplus to the local grid. There are two reasons for employing a larger engine capacity. Firstly, a small engine will not be able to generate sufficient heat to drive an absorption cooling system with a capacity of 600 kilowatt. Secondly, export electricity will increase revenues generated from the cogeneration plant. Financially, the net present value (NPV) of both the cogeneration system and the conventional energy supply system were lower than zero, which means that neither system can be viable financially. The cogeneration system's NPV was lower than that for the conventional energy supply system, due to its higher capital and operating costs. High operating costs were due to fuel costs, with low revenues being due to heavy subsidies on electricity. If fuel and electricity subsidies were removed, a cogeneration system could become a more attractive option compared to a conventional system. However, removing the electricity subsidy would result in the large majority of people being unable to afford electricity.Item Heat recovery refrigeration in New Zealand dairy sheds : a thesis presented in partial fulfilment of the requirements for the degree of Master of Agricultural Science in Agricultural Engineering at Massey University(Massey University, 1982) Stinson, Grant ErrolIncreased energy costs initiated an investigation into refrigeration heat recovery as one conservation alternative available for reducing water heating costs on farm dairies. A theoretical energy balance was conducted, from which the potential of recovering refrigeration condenser heat was estimated at up to 60% of the water heating energy requirements. Preliminary tests with heat exchangers lead to the use of a tube-in-tube, counter flow, heat exchanger with fins on the refrigerant side, and cores on the water side, to improve the heat transfer characteristics. The exchanger, designed to provide 300 litres of 60°C water from a 2.25 kw refrigeration system cooling 2000 litres of milk per day, had an area of 0.84 m2, and an overall thermal conductance of 100 W.m-2.°C-1. This heat exchanger was inserted between the compressor and condenser of the refrigeration plant and tested with two condenser systems (air and water), four condenser pressures (6.5 bar, 7.5 bar, 10 bar and 12 bar), two milk inlet temperatures(23°C and l8°C), and two milk final temperatures (4°C and 7°C). In addition, tests on receiver pressure and suction superheat were performed to determine overall system performance. Increasing condenser pressure increased cooling times from 2 hours 32 minutes to 3 hours 17 minutes, after the completion of the 1200 litre morning milking (thus failing to comply with the 3 hour cooling regulation at high condenser pressures.) Also, C.O.P. decreased from 3.05 to 2.35 for the water cooled condenser system (2.70 to 2.00 for the air cooled condenser system due to fan power consumption). Gross heat recovery rose from 4.2 kWh.day-1 .m-3 to 8.l kWh.day-1 .m-3 for the water cooled system, giving water outlet temperatures of 45°C to 64°C as condenser pressure rose. The corresponding ranges for air cooled condensers were 3.8 kWh.day-1 .m-3, to 6.6 kWh .day-1 .m-3, and 38°C to 55°C. Changing milk inlet and final temperatures gave a proportional change in cooling times and total heat recovery, but had no effect on C.O.P. or heat recovery rates. Suction superheating increased total heat recovery by 15%, and water outlet temperatures by 9%. Increases in gross heat recovery with increasing condenser pressure were partially offset by additional compressor power, and yielded nett heat recoveries of 4.0 kWh.day-1 .m-3 to 6.0 kWh.day-1 .m-3 for water cooled, and 3.6 kWh. day-1 .m-3 to 4.3 kWh. day-1 .m-3 for air cooled, condenser systems. The maximum gross and nett heat recoveries (at 12 bar condenser pressure) were applied to the energy requirements of a monitored 220 cow town supply dairy. This analysis showed that the gross heat recovery was 51% of the water heating requirements, but the nett heat recovery dropped to 17% of the total heating and refrigeration demand. Based on current electricity and equipment prices, it is estimated that the payback period for this level of recovery would be 16-17 years. Changing the electricity pricing structure, to reflect up to a 1:3 differential in favour of water heating power costs, results in the 6.5 bar condenser pressure giving optimum results, but the nett returns are significantly lower than those reported. The potential for improved savings is greater from larger capacity systems as the capital investment is not proportionally increased with an increase in scale.