J. agric. Engng Res. (1987) 36, 275-285

A Dairy Refrigeration Heat Recovery Unit and Its Effects on Refrigeration Operation

G. E. STINSON*?; C. J. STUDMAN*; D. J. WARBURTON*$

Concern about increased energy costs prompted an investigation into refrigeration heat recovery as one conservation alternative for reducing water heating costs on farm dairies. A theoretical energy balance was conducted, from which the potential for recovery of refrigeration condenser heat was estimated to be up to 60% of the water heating energy requirements.

Preliminary tests with heat exchangers led to the development and testing 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 1 of water at 60°C from a 2.25 kW refrigeration system which cooled 2100 1 of milk per day, had a surface area on the refrigerant side of 0.84 rn’, and an overall thermal conductance of 750 W m-* “C- ‘. It was inserted between the compressor and the condenser of the refrigeration plant and tested with two condensing systems (air and water), together with varying conditions of condenser pressure and milk temperatures at inlet and final cooling. In addition, tests on the receiver pressure and suction superheat were performed to determine their effect on the overall system performance.

Increasing the condenser pressure from 6.5 bar to 12 bar increased cooling times. In extreme circumstances the system failed to comply with the New Zealand milk cooling regulations. The average coefficient of performance (C.O.P.) of the refrigerator (with the heat exchanger in the circuit) decreased with increasing pressure, varying from 3.0 to 2.3 over this range of pressures for the water cooled condenser system. Values for the air cooled condenser system were 0.3 to 0.4 lower due to fan power consumption.

1. Introduction

In 1986 the New Zealand dairy industry consisted of 1.97 million cows on 12 516 seasonal process supply farms and 166000 cows on 1368 domestic supply farms. Stringent hygiene requirements exist for all milking equipment on dairy farms and the consequent water heating requirements cost the industry approximately $NZ8 million per year for 500 GJ (1.4 x lo* kWh) of electricity.’ Further forseeable hygiene regulations and inflation will continue to increase this figure. In addition, intensification and larger herd sizes will increase the cost to individual farmers. Prior to the fuel shortages of the 197Os, electricity was considered the most suitable means of heating water, but rapidly increasing prices have increased the interest in alternative sources of energy. Prominent among these have been solar heating and refrigeration heat recovery. Solar heating, although successful, has the disadvantages of (a) requiring significant capital outlay, and (b) being of variable capacity. There are also siting and operating requirements which may not always suit the farm.’

Refrigeration heat recovery extracts some of the heat normally rejected to the atmosphere at the condenser during the milk cooling operation. The heat is transferred to the water via a heat exchanger installed as part of the refrigeration circuit. This system has the advantages, compared with solar water heating, of lower capital cost, a reliable and consistent heat source, no major siting limitations, and it is potentially capable of providing hot water at

* Agricultural Engineering Department, Massey University, Palmerston North, New Zealand t Now at Global Energy Technologies Ltd, Frankton, Hamilton, New Zealand 1 Now at Papakura, Auckland, New Zealand

Received 5 August 1985; accepted in revised form 10 May 1986 275

00?1-8634/87~040275+ I I $03.00/O 0 1987 The British Society for Research in Agricultural Engineering

276 HEAT RECOVERY UNIT-DESIGN AND EFFECTS

temperatures of the order of 60°C every milking day. However, in New Zealand, heat recovery systems are only available, at present, as add-on units for air cooled refrigeration systems, and the potential for improved heat recovery performance using a water cooled condenser as a means of preheating the water prior to its passage through the heat recovery heat exchanger, has not been investigated.

Little information is available regarding the merits of alternative heat recovery methods or the changes of efficiency which are brought about by altering conditions of load and environment. Also little is known about the effect which heat recovery equipment has on milk cooling rates and temperature levels. As a result the economic status of heat recovery in farm dairies is not clear. In order to investigate some of these factors a heat exchanger designed to recover waste heat from a refrigeration system was constructed and tested under experimental conditions, based on data for a moderately large herd of 210 cows each producing 10 1 of milk per day. This herd size was chosen because a commercially operated farm of this size was available for field studies, and because it represented a typical unit on which heat recovery systems could be installed. In this paper the equipment is described and the effects on the refrigeration unit’s performance are discussed. In a subsequent paper the effectiveness of the unit for heat recovery is considered.3

2. Theory

2.1. The efSect of changing operating conditions on heat recovery

The rate of heat recovery and the water outlet temperatures from a heat recovery unit depend on the quantity and temperature of the delivery superheat contained in the refrigerant. These can be increased by raising the condenser pressure or by the addition of further heat to the refrigerant after vapourization and before it enters the compressor (suction superheating). However, changing either of these will affect the cycle efficiency and the system performance. Increasing condenser pressure causes (a) an increase in pressure differential across the expansion valve, which reduces the refrigeration effect; (b) an increase in power drawn by the compressor; and (c) an increase in the compression ratio across the compressor, resulting in a decrease in volumetric efficiency and a reduction in the refrigerant flow rate. However, it also improves both the quantity of superheat and, more importantly, the temperature of the superheated vapour. Increasing the degree of suction superheat forces the compression process to operate at a higher level of specific entropy and, as a consequence, produces more delivery superheat at a higher temperature. Thus, it is important to investigate the effects of the heat recovery upon the performance of the refrigeration unit, in order to establish the optimal operating conditions for both functions.

2.2. Refrigeration heat loads in dairy sheds

The total refrigerant load for the milk cooling system consists of the heat to be removed from the milk and any environmental heat load. Currier4 and Vickers5 found that in practice, the latter is small (less than 10% of the total heat load). Allowing for heat losses in the milking plant pipes, milk enters the cooling system at around 33°C. Under current New Zealand regulations it must be cooled to 7°C within three hours of the completion of milking, and maintained at this temperature.lS6 For town supply farms, the figures are 5°C within 3.5 h. Cooling milk from 33°C to 7°C requires that 101 kJ kg-’ be removed. Again under normal New Zealand conditions, approximately 50% of this load is removed by refrigeration, the remainder being removed by a water cooled plate heat exchanger.’ Thus for a 210 cow herd producing 10 1 of milk per cow per day, the energy which must be removed by the refrigeration unit alone is approximately 110 MJ (31 kWh). In the foreseeable future5 the temperature may be reduced to 4°C in which case 122 MJ (34 kWh) must be removed by the refrigeration unit.

G. E. STINSON ET AL. 277

2.3. Energy requiredfor heating water

The volume and temperature of water used for cleaning milking equipment depends on the size of the unit and the cleaning system used. ‘** For a 210 cow unit using the triple hot water system (a system involving a cold rinse, a warm detergent wash, and a hot sanitizing rinse), approximately 300 l/d would be required at 95°C which would use 101 MJ (28 kWh). The energy required is thus of the same order of magnitude as the energy to be extracted from the milk.

3. Experimental design

3.1. Design of test equipment

A laboratory rig was constructed to enable the performance of a heat recovery unit to be examined under controlled conditions (Figs I and 2). The apparatus was based on the size and type of refrigeration equipment which would normally be used on a 210 cow dairy unit,

I i

i’ Kii L iJ a B c G D E F

Fig. I. Schematic diagram of experimental plant. A, Refrigerant receiver; B, solenoid shut-off valve, C, recording,fEow meter; D, liquid-vapour heat exchanger; E, expansion valve; F, evaporator in base of milk vat; G, suction super heater; H, compressor; I, primary heat exchanger (the heat recovery unitJ; J, air cooled secondary heat exchanger (main condenser); K, water cooled secondary heat exchanger (main condenser); L, recording Jlow meter; M, condenser by-pass control valve; N. milk vat (2250 1); 0, milk inlet; P, refrigerant route into water cooled condenser; Q. alternative refrigerant route into air cooled secondary condenser; R, primary heat exchanger by-pass (used for excluding primary in certain tests): S, cold water inlet; T, hot water output from heat recovery unit; U, cold water inlet to heat recovery unit; V, warm water supply (f rom water cooled secondary) to suction superheater (used ,for three runs only); W, jaw switch (directs water either to primary from cold inlet or to suction superheater from water cooled secondary); X, condenser cold water flow control valve [to control water cooled

condenser pressure); Y, condenser pressure control valve (air cooled condenser)

278 HEAT RECOVERY UNIT-DESIGN AND EFFECTS

Fig. 2. Experimental equipment. A, Heat recovery unit (primary heat exchanger); B, air cooled condenser; C, water cooled condenser; D, compressor

and consisted of a 2.25 kW direct expansion system, with Freon 12 refrigerant, cooling a 22501 uninsulated vat which incorporated the evaporator in its base. In the experiment, heated water was used to simulate the milk, but for clarity this will be referred to as “milk” in this paper. Due to specific heat and density differences 2100 1 of milk is equivalent to just over 2000 1 of water, and so a cooling load of 2000 1 of “milk” per day was used, with a 40/60 split between evening and morning milkings. The average flow rate into the vat was 12 l/min during filling, with a filling time of 67 and 100 min respectively.

The apparatus is shown in Figs I and 2. The refrigerant flowed from the receiver reservoir (A) through a solenoid shut-off valve (B) connected to the vat thermostat, then through a flow meter (C), a liquid/vapour heat exchanger and a drier (D), the thermostatically controlled expansion valve (E), and then into the 8 kW evaporator (F). The vapour from the evaporator then passed through the liquid/vapour heat exchanger (and the suction superheater (G) for one run only), before entering the 2.25 kW direct drive sealed compressor unit (H). From the compressor unit, refrigerant passed through the primary heat exchanger (I), one of the secondary heat exchangers (J or K) and the refrigerant flow meter (L) to the receiver.

The condenser system consisted of the primary heat exchanger (I) for heat recovery (in which some condensation could occur) and a secondary heat exchanger for condensation. The secondary heat exchanger could be either air or water cooled. In order to maintain receiver pressure and avoid premature evaporation, a by-pass valve was installed between the compressor outlet and the receiver (M). This valve was set to open if the receiver

G. E. STINSON ET AL. 279

pressure fell below 6 bar. The system was designed to allow flexibility in the routing of the refrigerant and water through the condenser circuits by using shut-off valves to direct the flow. Receiver pressure and condenser pressure for the secondary heat exchangers were controlled by preset control valves manufactured by Danfoss.

Pressures were measured with Bourdon gauges, and temperatures were recorded at the inlet and outlet of all major components using copper constantan thermocouples. The total power consumption for the compressor and the fans was recorded, together with the voltage and current per phase, and the power factor. Turbine flow meters were used to measure the refrigerant flow entering the receiver and the evaporator, and the primary water flow was measured with a low head turbine flow meter. All equipment was calibrated prior to installation in the rig.

3.2. Heat exchanger design

For the milk and hot water volumes selected, theoretical calculations showed that a complete condensing “once-through” system was not a viable alternative.’ This was supported by the findings of Fleming and O’Keefe.’ Therefore, a de-superheater system was selected, where the superheat, rather than the latent heat, of the refrigerant is collected. Based on available information, a tube-in-tube heat exchanger was selected as the basic design for the de-superheater unit. To improve heat transfer, a stainless steel tube with spiral copper finning pressed against the tube was used, with the refrigerant flowing over the fins. Increased turbulence in the water tube was achieved with cores. To establish performance characteristics of the selected design, tests were conducted on a 1 m sample length of finned tube. These tests showed that the head loss on the refrigerant side could be high, ranging from O-6 bar to 2.5 bar over the range of flows tested. The overall thermal conductance (or U value) of the finned tube was 750 W m-’ ‘C-l, which at the flow rates involved, was considerably higher than those quoted by other workers. Based on the results from these tests, a heat recovery unit was designed and constructed. This is shown schematically in Fig. 3. The design criteria adopted are given in Table 1.

Based on Table 1 the expected rate of heat transfer required was 1.3 kW, and the required length of spiral tube was 3.0 m. This was divided into two parallel pathways to reduce total head loss. To keep the heat exchanger compact the final arrangement was four 0.75 m tubes arranged as two parallel units 0.75 m long in series (Fig. 3). Each tube of the exchanger was constructed from 9.5 mm stainless steel copper finned tube (A) with 158 fins per metre, into which a 4.7 mm OD stainless steel core (B) was inserted. The finned tube was placed inside a 31.75 mm OD stainless steel outer tube (C). The four tubes were enclosed in a 180 mm diameter stainless steel cylinder (D), and the cavity was filled with polyurethane foam (E).

Table 1

Design data

Refrigerant flow rate Water inlet temperature Water outlet temperature Water flow rate Overall thermal conductance (U) Maximum head loss in primary Area of refrigerant side of heat exchanger Area of water side of heat exchanger

2.0lmin-’ (l.Olmin-’ per leg)* 30°C 60°C 0.625 I min-’ (0.31 I min-’ per leg)* 750Wm-Z”C-Lt 150 kPa (1.5 bar) 0.297 m* per metre of tube 0.022 m2 per metre of tube

* To avoid excessive head loss the refrigerant flow was divided into two parallel pathways. To maintain counter flow in the heat exchanger, the water path was also split

t Based on water side area

280 HEAT RECOVER

A 8 C

.Y UNIT-DESIGN AND EFFECTS

Fig. 3. Primary heat exchanger. A, 9.5 mm diameter stainless steel tube with copper fins (158 jins per metre); B, 4.7 mm diameter stainless steel core; C, 31.75 mm diameter stainless steel cylinder;

D, 180 mm diameter 18 s.w.g. stainless steel sheet; E, cavity Jilled with polyurethane foam

3.3. Experimental procedure

The apparatus was operated at four condenser pressures in the range between 6.5 and 12 bar. Milk inlet temperatures of 23°C and 18°C were used, and the milk was cooled to either 7°C or 4°C. In addition, a further experiment to assess the effect of superheating the suction vapour using the warmed water discharged from the secondary water cooled heat exchanger was conducted. The primary heat exchanger was excluded for three runs to assess its effect on system performance. The following procedure was adopted for each experimental run: initially the operating conditions (milk loading rate, milk inlet temperature and primary water flow rate) were selected by adjusting the appropriate controls. The refrigeration system was started at the same time as the milk flow began. System loading was 800 1 followed by a further 1200 1 after the first load had been cooled (simulating evening and morning milkings). The experiments were conducted in a large laboratory at normal’temperature (17°C f 3°C) to simulate an average environmental heat load. The interval between loads was ignored in the experiment since the evaporator only needed to operate infrequently and for very short periods in order to maintain the milk temperature. A final adjustment of the condensing pressure and primary water flow rate was made within 20 min of initial start-up. Readings at all data collection points were made at 30 min intervals until the final temperature was reached, at which point only totals of power consumption, time and refrigerant volume were recorded. The environmental heat load was calculated for each run.

Some difficulty was experienced in controlling primary water flow rates for the water cooled condenser system and a mathematical model was used in the calculation of the water outlet temperatures and heat recovery rates to make slight adjustments to correct the data to

G. E. STINSON ET AL. 281

conditions of a constant water flow of 0.625 1 min.- ‘. The model, which has been described by Stinson,’ altered temperatures by typically 3 to 5”C, and heat recovery by not more than 0.2 kW. The results obtained by this method allowed a comparison to be made with the results from the air cooled system. The model also enabled the results to be applied to other sizes of heat exchanger.”

4. Results

4.1. Operation of refrigeration system without heat recovery

The performance of the refrigeration system varied considerably in response to changing operating conditions. It was found that increasing the condenser pressure from 6.5 bar to 12 bar reduced the average C.O.P., and increased cooling times to the point where the system failed to meet the New Zealand milk cooling regulations. Average values of C.O.P. were between 3.2 and 2.4 for both air and water condenser systems. The compressor power consumption varied between 2.2 and 2.7 kW, increasing linearly with condenser pressure over the range tested. The air cooled system required an additional 0.36 kW to drive the fans, which reduced the C.O.P. by yp to 0.4. A receiver pressure below 6.0 bar resulted in incorrect operation of the expansion valve. Thus for optimal utilization of refrigeration energy without heat recovery, a water cooled system operating at the lowest possible condenser pressure consistent with the correct operation of the expansion valve should be selected.

4.2. Results with heat recovery unit installed

4.2.1. Vat temperature

In Figs I and 5 curves of vat temperature against time are given for different condenser pressure settings. The left-hand curve represents the first milk load (the evening milking) and

1200 I load

0 60 I20 180 0 60 I20 180 240 3 Time since start, mln

10

Fig. 4. Effect of condenser pressure on vat temperature. Curves are for a water cooled condenser system. with a milk inlet temperature of 23°C and cooling to 4°C. The curves to the left are for the initial 800 I load and those on the right arise when the 1200 1 load is added. 4, 12 bar condenser pressure: 0, IO bar

condenser pressure,’ A. 6.5 bar condenser pressure

282 HEAT RECOVERY UNIT-DESIGN AND EFFECTS

800 I load 1200 I load

0 60 I20 I80 0 60 120 180 240 3 Time snx start, mln

)O

Fig. 5. Effkct of milk cooling temperature d@erential on vat temperature. Curves are jOr a condenser pressure of 7.5 bar. v, water cooled condenser, 23%4°C; ??. water cooled condenser, 23-7°C; A. water cooled condenser, I&4°C; +, water cooled condenser, I&7°C; 7, air cooled condenser, 23-4°C; A, air

cooled condenser, I&4°C

the right-hand curve is the result of adding the second (morning) milking. Fig. 4 shows a typical set of curves for various condenser pressures for fixed milk cooling differential (i.e. cooling the milk from 23°C to 4”C), while Fig. 5 shows the results for different milk inlet and outlet temperatures. All curves are of the same general shape: the first milk load was cooled rapidly to 4°C while the total load cooled more slowly, in some cases approaching the regulation limit of 180 min after the end of milking. The highest temperature reached by the mixture was around 10°C which occurred as the last milk was added. In the curves shown the milk was added at 23°C and cooled to 4°C (Fig. 4). If the milk was only cooled to 7°C in the first run, the peak temperature in the second run was approximately 1°C higher. The overall effect of condenser pressure or condenser system on vat temperature was negligible. Reducing milk inlet temperatures by 5°C to 18°C had the most effect on vat temperature, with temperatures being reduced by 3 to 4°C on average (Fig. 5). The mix temperature for the 1200 1 loading was reduced by 2.5”C. While there is no legislation restricting mix temperatures in milk vats, it is generally recommended that they do not rise above 10°C for milk quality reasons.4

4.2.2. Cooling times

It was found that cooling times increased under conditions of increasing condenser pressure, milk inlet temperature, and environmental load, and decreasing final milk temperature. In particular, a combination of high condenser pressure (12 bar), high milk inlet temperature (23”C), significant environmental load (8500 kJ) and low milk final temperature (4”C), would result in the second fill failing to meet the 180 min regulation cooling time by 17 min. Under the same environmental load the cooling time at 10 bar condenser pressure would just meet the regulations. If the environmental load were reduced to negligible proportions by insulating the vat, then in all cases the cooling time would fall within the regulation period. The most significant factor governing cooling times was the

‘3. E. STINSON ET AL. 283

temperature differential through which the milk was cooled. For example, when the milk inlet temperature was reduced by 5°C the total cooling times were reduced by 100 min on average (about 25%, as would be expected). Since the final milk temperature is set by regulation, effective precooling of the milk with a water cooled plate heat exchanger (or equivalent device) is required to ensure that the cooling requirements are met, given the present sizing of refrigeration plants for vat cooling. Operating the system with either an air cooled or water cooled condenser had no effect on cooling times. Isolating the primary heat exchanger from the circuit, or increasing the pressure in the receiver tank (normally at 6 bar for all experiments) to 7 bar, also had no effect of any significance on cooling times.

4.2.3. Rqfrigerant,flow

Over the entire set of experiments the refrigerant flow was in the range of 1.7 to 2.5 1 min- ‘. The only variable affecting the flow rate significantly was milk temperature in the vat. Over each run the flow rates dropped through the measured range as the temperatures dropped. Reducing the milk inlet temperature by 5°C reduced flow rates by approximately 10%. The effect brought about by changing the condenser pressure was minor, with a slight decrease in refrigerant flow with increasing condenser pressure for both condenser systems. There were no measurable differences in flow rate between the two condenser systems, and no noticeable effects due to increasing receiver pressure or the inclusion of the primary heat exchanger. The mass of refrigerant flow by-passing the condenser system through the by-pass valve (M, Fig. 1) was found to be less than 5% of the total flow rate and was therefore of little practical importance to the potential level of heat recovery.

4.2.4. Refrigeration efkct

The refrigeration effect (or evaporator enthalpy change) is a measure of the amount of heat removed from the vat per kilogram of refrigerant flowing through the system. It varied between 140 and 150 kJ kg-’ with highest values occurring at high milk temperatures in the vat. Condenser pressure change had only a small effect on refrigeration heat transfer with a variation of 4 kJ kg-’ at the maximum vat temperature recorded. The effect of changing the condenser system was also small, with the air cooled system giving slightly higher values than the water cooled system. This was because the condenser control system increased the degree of subcooling in the liquid refrigerant backed up in the condenser by the control valve. Reducing the milk inlet temperature reduced the average refrigeration effect slightly but changing the milk final temperature, or increasing the receiver pressure, had no marked effects.

4.2.5. Head loss in the primary heat exchanger

The head loss in the primary heat exchanger depended on the refrigerant volumetric flow rate through the primary. Although the mass flow rate changed little with condenser pressure, a larger volume of refrigerant gas had to pass through the primary at lower pressure, resulting in a greater head loss. The loss varied between 0.4 and 1.1 bar during the 12 bar runs (being greatest at the start when refrigerant flow was highest), and between 1.0 and 1.5 bar for the 6.5 bar runs.

4.2.6. Cot@cient of performance (C.O.P.)

The C.O.P. fell as the condenser pressure increased. It also varied during each run as the milk cooled and the refrigerant flow changed. Highest values were 3.3 at the start of the 6.5 bar runs, falling to 2.8 towards the end (an average C.O.P. of 3.0), while the lowest C.O.P. occurred at 12 bar, varying from 2.5 to 2.2 during the run. This represents a reduction of about 0.1, compared with operation without the heat recovery unit. Excluding the power consumption required by the fans, there were no significant differences between

284 HEAT RECOVERY UNIT-DESIGN AND EFFECTS

condenser systems. The inclusion of fan power requirements reduced the C.O.P. for the air cooled system by between 0.3 and 0.4. The changes in C.O.P. in response to differing condenser pressures were the result of the combined changes in refrigerant flow, refrigeration effect and compressor power consumption.

5. Conclusions

Equipment has been constructed to study the effect of a heat recovery unit on the performance of a farm dairy vat refrigeration system. The equipment simulated milk loads and conditions on a 210 cow dairy unit producing 10 1 of milk per cow per day using 2000 1 of water as an equivalent cooling load. The performance of the refrigeration system, as expressed by the C.O.P., varied considerably with changing operating conditions. Increasing the condenser pressure from 6.5 bar to 12 bar reduced the C.O.P. by about 23% due to the increase in compressor power consumption. The C.O.P. for the water cooled system was 10 to 18% higher, depending on the condenser pressure, than the air cooled system due to the power required to operate the fans. Changing the milk inlet or final temperature had no major effect on the C.O.P. Inclusion of the heat recovery unit reduced the C.O.P. values by approximately 0.1 (3 to 4%).

The refrigerant flow rate depended only on the milk temperature in the vat. The mass of refrigerant by-passing the condenser system through the by-pass valve was less than 5% of the total flow rate, while the effect of condenser pressure on refrigerant effect was very small. Head loss in the primary heat exchanger varied from a maximum of 1.1 bar at the start of cooling at 12 bar condenser pressure, to a maximum of 1.5 bar at 6.5 bar condenser pressure.

The ability of the refrigeration system to meet the requirements of the cooling regulations was independent of condenser system but was dependent upon condenser pressure, environmental heat load, and milk inlet and final temperature. When operating with a significant environmental heat load of 8500 kJ, with condenser pressures above 10 bar, and at milk inlet temperatures of 23°C the system failed to cool the final 1200 1 of milk to 4°C within 180 min of the completion of loading. However, elimination of the environmental heat load by insulating the vat would have allowed the regulations to be met at all pressures.

Acknowledgements

The authors would like to thank the Massey University Agricultural Research Foundation for financial support. We would also like to thank the Manawatu Co-op Dairy Company, Fridge Heat Ltd, and McAlpine-Prestcold Refrigeration (NZ) Ltd for the use of refrigeration equipment, and Messrs J. Bowen (Fridge Heat Ltd), J. Jenkins (Protech Engineering) and S. E. Compton and R. E. Belgrave (Agricultural Engineering Department, Massey University) for their advice and assistance in the construction of the experimental equipment.

References

’ Stinson, G. E. Heat recovery refrigeration in New Zealand dairy sheds. M. Agri. Sci. Thesis, Agricultural Engineering Department, Massey University, 1982

* Studman, C. J. A once-through solar water heater for farm dairies. Journal of Agricultural Engineering Research 1979, 24: 149-156

3 Stinson, G. E.; Studman, C. J.; Warburton, D. J. The performance and economics of a dairy refrigeration heat recovery unit. Journal of Agricultural Engineering Research 1987, 36: 287-300

4 Currier, J. R. Refrigeration energy in raw milk. Paper presented at the joint meeting of Commissions C2, Dl, D2, D3 and El, at the International Institute of Refrigeration, Australia, 1976

5 Vickers, V. T. New Zealand Dairy Research Institute, Personal communication, 1980 5 Milk Industry Regulations. New Zealand Dairy Industry Regulations, Government Printer, 1977

G. E. STINSON ET AL. 285

’ Carter, R. S.; Fisher, R. K. The evaluation of a heat exchanger used to recover refrigeration waste heat. Paper presented at the joint meeting of Commissions, C2, Dl, D2 and D3 of the International Institute of Refrigeration, New Zealand, 1982

a New Zealand Gazette. Requirement for heating plant and hot water in farm dairies. Notice No. 603 Ag 50202 B, 1973

g Fleming, M. G.; O’Keefe, J. Energy can be saved on the dairy farm. Journal of Farm and Food Research, Nov.-Dec., 1977

lo Studman, C. J. Refrigeration heat recovery and utilization in New Zealand dairy sheds. Research Publication Series No. 8, Massey University Agricultural Research Foundation, Massey University, 1983