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Ice-Bank Air Conditioner for Fresh Produce Storage

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Abstract and Figures

On site refrigerated storage for fresh produce has become a vital feature, and significant contributor, to the energy use of small scale vegetable farms. Certain high volume crops, like apples and carrots, are stored throughout the winter months in a cold and humid environment. Much of this storage takes place when ambient conditions can be used to save energy. Presented in this report is a model design for a system that uses ambient air to chill and freeze water, which in turn is used to deliver cold and humid air to a fresh produce storage facility.
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An ASABE Meeting Presentation
Paper Number: 11 11686
Ice-Bank Air Conditioner for Fresh Produce Storage
David R. Bohnhoff, Ph.D. P.E., Professor
Justin C. Banach, Undergraduate Student
Adam J. Gardebrecht, Undergraduate Student
Aaron J. Lofy, Undergraduate Student
Michael D. Muehlbauer, Undergraduate Student
Luke P. Syse, Undergraduate Student
Christofer A. Sindunata, Undergradaute Student
Scott A. Sanford, Senior Outreach Specialist
Biological Systems Engineering Department, University of Wisconsin-Madison
460 Henry Mall, Madison, WI 53706
Written for presentation at the
2011 ASABE Annual International Meeting
Sponsored by ASABE
Gault House
Louisville, Kentucky
August 7 – 10, 2011
Abstract. On site refrigerated storage for fresh produce has become a vital feature, and significant
contributor, to the energy use of small scale vegetable farms. Certain high volume crops, like apples
and carrots, are stored throughout the winter months in a cold and humid environment. Much of this
storage takes place when ambient conditions can be used to save energy. Presented in this report is
a model design for a system that uses ambient air to chill and freeze water, which in turn is used to
deliver cold and humid air to a fresh produce storage facility.
Keywords. Air Conditioner, Fresh Produce Storage, Ice Bank Cooler, Heat Exchanger, Cold
Storage, Humidification
2
Introduction
In recent years consumer trends in buying food have changed. More and more people are
buying higher quality, locally grown fruits and vegetables (Biermacher and others, 2007). As a
result, small scale vegetable producers have seen an increased need to extend the marketable
season of their produce into the winter months. Certain crops harvested during the fall months
have the ability to be stored until March without perishing. The preservation of late season
vegetables allows producers to increase revenue by selling at higher prices during off peak
periods. More importantly, winter storage allows producers to stay in contact with their
customers year-round (Pincus, 2010).
As is evident by the table in Appendix A, different crops require different conditions for long term
storage. In almost all cases, produce is stored below 13°C (55°F) and at a relative humidity
between 95% and 100%. Cooler temperatures reduce the rate of respiration, while humid air
prevents dry out. Some crops commonly stored during winter require a storage temperatures
near 0°C (32°F). This would include apples, beets, carrots (figure 1), rutabaga, and turnips.
Providing a high-humidity environment at temperatures near the freezing point of water can be
challenging.
Figure 1. Under proper conditions, this bin of carrots harvested in November can be stored until
the following spring.
Humidification Methods
Methods used to maintain a humid environment can be categorized as assisted evaporation,
natural evaporation, and concealment (Table 1).
Table 1. Humidification Methods
General Type ID Example
Assisted Evaporation A Rotary atomizers, evaporative pads
Natural Evaporation N Water sprayed/dumped on floor
Concealment C Batches of produce encased in plastic bags
3
Assisted evaporation systems are those that utilize specialized mechanical equipment to
humidify air, and include rotary atomizers and evaporative pads. Rotary atomizers (figure 2)
add moisture to the air by forcing water through a screen by means of a spinning wheel, and are
often used in conjunction with refrigeration. While rotary atomizers are capable of humidifying
large volumes, they are usually oversized for the smaller rooms used by some farmers. Often
one atomizer is needed per storage room so the cost is also multiplied by the number of cold
and humid storage rooms a producer uses. Condensation on the walls of the cooler and directly
on the produce is a problem for rotary atomizers.
Figure 2. Rotary atomizer.
Humidifiers with evaporation pads (figure 3) both cool and humidify by moving unsaturated air
over a water-covered mesh with a relatively large surface area. Evaporative pads are often
used to both cool and humidify air being brought into a facility - something they do quite well as
long as the intake air is above the freezing point of the water on the pads.
Figure 3. Humidifier (left) that utilizes evaporation pads (right).
Spraying or dumping water on a storage room floor to humidify the storage is herein referred to
as natural evaporation. Although it is a very low cost option, it requires active involvement of
the farmer for routine water applications, and maintaining a wet floor surface is unsafe for those
working in the facility.
4
Concealment of produce in plastic bags (figure 4) is a very effective way to maintain a moisture-
saturated air space around the produce. The one major drawback is that ethylene gas, which
accelerates ripening, is also trapped in the surrounding air space. Also, care must be taken not
to tear or otherwise compromise the bag.
Figure 4. Concealment method of maintaining moisture-saturated air around produce.
Cooling Methods
Methods for winter cooling of storage facilities can be categorized as those that rely on cyclic
refrigeration, those that use ambient air directly, and those that use tempered ambient air (Table
2).
Table 2. Cooling Methods
General Type ID Description
Cyclic Refrigeration R Conventional equipment that relies on compression
and expansion of a refrigerant
Direct Ambient Air D Fans blow cold ambient air directly into storage room
Tempered Ambient Air T Fans draw ambient air through earth tubes and into
storage room
Cyclic refrigeration, specifically vapor-compression refrigeration, is the most commonly used
method for cooling produce. Cyclic refrigeration is convenient, but has a relatively high initial
and operating cost (Lung, 2006). Also refrigeration systems are not good for the environment
as they contribute measurably to the world’s CO2 production (James and James, 2010).
During winter, use of ambient air can be a viable and inexpensive means to cool produce. One
of the problems with using ambient winter air directly, is that it is generally drier than desired
and it requires careful mixing to avoid freezing of (and hence catastrophic damage to) fresh
produce.
When operators desire to maintain a storage temperature closer to 13°C (55°F), it's
advantageous to temper ambient air by drawing it into the storage facility through earth tubes.
In many respects, a system that utilizes the temperature and thermal mass of soil in this manner
is a take-off of root-cellaring techniques used in the past to store vegetables that keep well at a
temperature near the average annual ground temperature.
5
When operators are attempting to maintain storage temperatures closer to 0°C, soil thermal
energy could be used to temper ambient air that is originally below 0°C. This requires a more
elaborate control system, and in reality, it makes more sense to use ambient air directly (i.e.
non-tempered air) and simply modify its flow rate as its temperature changes. It is also
important to note that earth tubes are relatively expensive to install and can harbor bacteria
(EERE, 2011).
A Proposed System
After an assessment of current systems, we focused on developing a low-cost system that
would enable farmers to store fresh produce during late fall and throughout the winter.
The first decision made during this process was to shoot for the most difficult storage conditions.
That is, to achieve storage conditions with saturated air and a temperature near 0°C.
The second decision made was to develop a system that could maintain our desired conditions
throughout late fall and winter without relying on a cyclic refrigeration system. This decision was
made because of the higher initial, operating, and environmental costs of such systems. This
decision, in effect, locked us into the use of ambient air for cooling.
Based on our previous assessment and additional research, we decided the risks of using a
concealed system for humidification and the cost of earth tubes for air tempering to great to
justify their consideration.
At this point we explored options for using a heat exchanger to avoid bringing cold ambient air in
direct contact with stored produce. We dismissed the use of a conventional air-to-air exchanger
because of the difficulty of dealing with icing that occurs when saturated air near 0°C is brought
in contact with a surface below 0°C. We instead settled on an exchange system that utilizes a
pan of water/ice for both energy storage and heat exchange.
Referred to as an ice-bank air conditioner (IBAC), the system we proposed would move ambient
air under a pan of water. This air would only be moved under the pan if it was colder than the
water in the pan and the water in the pan was not yet completely froze. By blowing air from the
produce storage room over the top of this chilled/frozen pan of water, the air in the storage room
is both cooled and humidified/dehumidified. Ice melts as it removes (1) heat of respiration from
the produce, and (2) conductive heat entering through the walls of the storage facility.
This method of air conditioning was chosen because of its temperature buffering capability and
passive method of humidity control.
Water is the most used phase change material due to its availability and chemical stability
(Silvetti, 2007). This system works optimally on days in which the night time temperatures drop
below freezing and day time temperatures are above freezing—or when storage coolers are
located in heated buildings that need continuous removal of heat independent of ambient
temperature. The proposed system eliminates many concerns of existing systems by allowing
water to freely diffuse into the circulating air. This will reduce the amount of water condensing in
the cooler and make it a better environment in which to work.
Project Goal
Design and fabricate an ice-bank air-conditioning (IBAC) system for wintertime storage of fresh
produce that utilizes ambient air for ice formation.
6
Approach
The approach we took toward development of an IBAC was to establish basic specifications for
an experimental system. Included in these specifications was a minimum pan size. Next we
designed and fabricated the system. We then ran a set of experiments to determine the heat
transfer coefficients for energy movement to and from a water filled pan.
IBAC Design & Fabrication
Design Specifications
Specifications established for the experimental (a.k.a. model) system included:
Material and labor costs not to exceed $1,000 dollars.
Transportation of the unit should require no more than a pallet jack.
Pan must be made accessible for rapid loading, unloading and cleaning.
Fans must be located and mounted for quick attachment and detachment.
System must be able to provide storage air properties outlined in Appendix A.
Pan should hold approximately 0.17 m3 (50 gallons) of water.
The system should not present any health and/or safety hazards to those working with, or in
the vicinity of the unit.
In practice, such a system must be decay and corrosion resistant. Due to the experimental
nature of our prototype, we did not include decay and corrosion resistance as design
specifications.
The pan was sized so that when containing solid ice a 0°C it could condition a space containing
3630 kg (8000 lbm) of carrots for about three days before all the ice melted. This mass of
carrots equates to 8 pallet boxes. To contain eight crates stacked two high, a room with
dimensions of 3 m x 3 m x 3 m (10 ft x 10 ft x 10 ft) was assumed. For the pan sizing
calculations (Appendix B) the respiration rate of carrots at 0°C was used and a temperature
gradient of 11°C (20°F) across all cooler surfaces was assumed. Thermal resistance (R-value)
was fixed at 4.40 °K·m2/W (25 h·°F·ft2/Btu) for all walls, the ceiling and floor.
A 0.17 m3 volume of ice will melt in fewer than 3 days if the cooler is maintained at a higher
temperature (which increases respiration rate of carrots), the temperature differential across the
cooler enclosure is greater than 11C, individuals enter and leave the cooler, and the cooler
enclosure leaks an appreciable amount of air. Even with these factors taken into account, we
doubt that all the ice would melt in less than a day.
7
Pan Design
The pan, which is at the heart of the IBAC was the first item designed and fabricated. Our initial
pan design featured a corrugated bottom for increased surface area. However, due to
manufacturing limitations, the corrugated bottom was scrapped in favor of a flat one. Fins were
added to the underside of the pan to increase its effective surface area and to create greater
turbulence for an increased heat transfer rate (figure 5).
Figure 5. Three dimensional view showing fins added to the bottom of a flat pan.
Single pass, double pass, and four pass configurations for air movement below the pan were
considered. We decided that a four pass configuration would give the cold, outside air a longer
path to travel and provide the best heat transfer (figure 6).
Warmer Air Out
Cold Air In
Fins
Baffle
Warmer Air Out
Cold Air In
Fins
Baffle
Figure 6. Four pass configuration.
In order to minimize both cost and resistance to heat transfer, we selected 16 gauge steel which
was the thinnest that we, as novice welders, felt comfortable welding. Three sheets of metal
were cut and bent to spec, which formed the pan base and two end caps (figure 7). Welds were
made on both the inside and outside edges of joints to help ensure water-tight seams. This
would likely not be done in commercial production. Three baffles were welded to the base of
the pan followed by 42 fins (figure 8). Detailed drawings of all pan part drawings are shown in
Appendix C. A summary of pan properties is given in Table 4.
Corrosion would obviously need to be addressed for commercial production. Options include
galvanizing the welded assembly or fabricating the pan from aluminum or stainless steel.
8
Figure 7. Completed pan assembly.Figure 8. Fins welded to the bottom of the
pan. A steel baffle appears to the left.
Exchanger box insulation appears to the
right and below the fins
Table 4. Pan Physical Properties
Property Value
Volume 48.5 gal.
Bottom area of pan 0.91 m2 (9.8 ft2)
Number of fins/baffles 42/3
Total surface area of fin and baffles 1.10 m2 (11.8 ft2)
Heat Exchanger Box Design
To house the pan and isolate it from environment conditions, we fabricated an insulated wooden
box. Overall dimensions (figure 9) where dictated by the size of the pan. Extruded polystyrene
with a 50 mm (2 in.) thickness and a thermal resistance of 1.76 °K·m2/W (10 h°F ft2/Btu) was
used for thermal isolation. Ducts were 203 mm (8 in.) in diameter. Various views of the
exchanger box are shown in figures 10 through 13.
Figure 9 Cross-section of heat exchanger box. Pink color represents extruded polystyrene.
9
Ambient Air Intake
Ambient Air Outlet
Conditioned Air Outlet
Conditioned Air Return
Ambient Air Intake
Ambient Air Outlet
Conditioned Air Outlet
Conditioned Air Return
Figure 10. Closed box. Hose exiting at lower
left rear of box is the pan drain tube. Figure 11. Open box showing all four
inlets/outlets.
 
Figure 12. Pan housed in box. Note spray
foam insulation used for additional sealing
between upper and lower plenum spaces
Figure 13. Top corner of heat exchanger box.
Note use of self-adhesive foam tape around
top edge of insulation to obtain a more
complete seal
Storage Cooler Design
A 1.2 m x 1.2 m x 2.4 m (4 ft x 4 ft x 8 ft) model storage cooler (figures 14 and 15) was built to
help create a dynamic system for testing. One-inch (25 mm) thick extruded polystyrene
insulation (R = 0.88 °K·m2/W = 5 h·°F·ft2/Btu) was selected as the wall material to provide a
reasonable heat loss rate for evaluation of the IBAC.
Air Movement
A small in-line duct fan (figure 16a) was initially purchased to move ambient air, but was found
incapable of moving sufficient air through the plenum space beneath the pan. It was
subsequently replaced by a model 1TDT2 Dayton PSC blower (figure 16b) with a rated capacity
under a zero static pressure of 15.6 m3/min (550 ft3/min). The manufacture's fan curve for this
blower is shown in figure 17.
10
Figure 14. Cross-section of storage cooler.Figure 15. Completed storage cooler.
(a)
(b)
Figure 16. (a) In-line
duct fan, and (b)
model 1TDT2 Dayton
PSC blower
Figure 17. Fan performance curve for model 1TDT2 Dayton
PSC blower
11
Fabrication
The bulk of the fabrication work, testing and analyses were performed by the undergraduate
student authors of this paper. Original design and construction of the IBAC was completed as
part of a senior design class. Figure 18 shows students engaged in fabrication and testing.
Appendix D contains a summary of fabrication costs.
Figure 18. Undergraduate students at work.
Clockwise from upper left: Luke checks door
on storage cooler; Aaron, Justin, and Mike
putting foam insulation around a duct inlet;
Chris collecting summer data; Aaron setting
up temperature sensors; Justin looking for
electrical shrink wrap.
12
Testing
Original tests were performed during the winter (December, 2010 throughout February, 2011)
weather permitting and during June, 2011, to determine the performance of the IBAC.
Specifically, tests were conducted to determine:
Heat transfer coefficient for energy transfer from ambient air to the water in the pan.
Heat transfer coefficient for energy transfer from between cooler storage air and stored
ice/water.
Experiments took place in an indoor laboratory that was heated and ventilated. For winter tests
(figure 19a), the air intake duct drew air from outside the laboratory and the outlet duct
exhausted it back outdoors. For summer tests (figure 19b), room air was drawn into the intake
and exhausted back to the room.
(a) (b)
Figure 19. Experimental set-up during (a) winter and (b) summer.
Temperature and humidity data were collected and stored with Hobo data logging equipment
and analyzed with Boxcar software. The location of temperature and humidity sensors was
varied depending on the purpose of the test.
Three different methods were used to estimate airflow. A vane type anemometer was first used
to determine air velocity in the air outlet duct. The duct’s circular cross-section was divided into
five equal areas using concentric rings. Four measurements, each 90 degrees apart, were
taken in each ring, and averaged to obtain a velocity for that ring. Average velocities for each
ring were multiplied by ring area to get the volumetric airflow for each ring. These values were
then summed to obtain total volumetric airflow. For summer tests, a pitot-static tube was used
in a manner identical to the vane anemometer to provide a second estimate of volumetric
airflow. A third estimate of airflow was obtained by measuring the static pressure drop across
the blower and inputting the resulting pressure drop into the fan performance equation (figure
17).
13
Heat Transfer Between Pan and Ambient Air
Both winter and summer tests were conducted to determine the rate at which heat could be
transferred between ambient air and the water in the pan. This was accomplished by
measuring the change in temperature of liquid water in the pan with the ducts on the top of the
conditioning box blocked with insulation as shown in figure 19b. Tap water was placed in the
pan for winter tests, and the drop in pan temperature monitored as cold, outdoor air was blown
under the pan. Ice water was placed in the pan for summer tests, and the increase in pan water
temperature monitored as warm, room air was circulated under the pan. The components of
heat transfer for both winter and summer tests are shown in Figure 20.
Tfloor
Tupper
qin
(main at Tin
and Win)
qlw
qt
qf
qout
(maout at Tout
andWout)
Troom
qp
Winter Test
Tlower
quw
qws
Mw
Tw&p
Qw&p
Tfloor
Tupper
qin
(main at Tin
and Win)
qlw
qt
qf
qout
(maout at Tout
andWout)
Troom
qp
qp
Winter Test
Tlower
Tlower
quw
qws
Mw
Tw&p
Qw&p
Mw
Tw&p
Qw&p
Tfloor
Tupper
qin
(main at Tin
and Win)
qlw
qt
qf
qout
(maout at Tout
andWout)
Troom
qp
Summer Test
Tlower
quw
qws
Mw
Tw&p
Qw&p
Tfloor
Tupper
qin
(main at Tin
and Win)
qlw
qt
qf
qout
(maout at Tout
andWout)
Troom
qp
qp
Summer Test
Tlower
Tlower
quw
qws
Mw
Tw&p
Qw&p
Mw
Tw&p
Qw&p
Figure 20. Heat flow designations for winter and summer tests used to determine the effective
heat transfer coefficient for the underside of the pan.
Heat flows shown in figure 20 are defined and calculated as follows;
qt = conductive heat flow through the top of the box, W (Btu/h)
= At (TroomTupper)/Rt
quw = conductive heat flow through the upper walls of the box, W (Btu/h)
= Auw (TroomTupper)/Ruw
qlw = conductive heat flow through the lower walls of the box, W (Btu/h)
= Alw (TroomTlower)/Rlw
qf = conductive heat flow through the floor of the box, W (Btu/h)
= Af (TfloorTlower)/Rf
qws = net heat flow between air above the pan and water in the pan, W (Btu/h)
= hws Aws (TupperTw&p)
qp = net heat flow between the air under the pan and the pan, W (Btu/h)
= hpan Apan (Tlower - Tw&p ) [will produce negative value for winter tests]
qin = enthalpy of moist air entering inlet air duct, W (Btu/h)
= mair hin
qout = enthalpy of most air exiting outlet air duct, W (Btu/h)
= mair hout
Qw&p = change in heat content of the pan and water in the pan, W (Btu/h)
= (Mw cw + Mp cs )
Tw&p /
t [will produce negative value for winter tests]
14
where:
Troom = temperature of room air, °C (°F)
Toutside = temperature of outdoor air, °C (°F)
Tupper = temperature of air above the water surface, °C (°F)
Tw&p = water and pan temperature, °C (°F)
Tlower = average temperature of air in the lower box (i.e., under the pan), °C (°F)
Tfloor = temperature of air between the box and the floor, °C (°F)
Tin = temperature of air entering inlet duct, °C (°F)
= Toutside for winter tests
= Troom for summer tests
Tout = temperature of air leaving outlet duct, °C (°F)
Mw = mass of water in pan, kg (lbm)
Ms = mass of pan without baffles and fin, kg (lbm)
= 22 kg (48 lbm)
At = inside surface area of box lid
= 1.11 m2 = 12.0 ft2
Auw = inside surface area of upper walls of box
= 1.61 m2 = 17.3 ft2
Alw = inside surface area of lower walls of box
= 1.45 m2 = 15.6 ft2
Af = inside surface area of box floor
= 1.11 m2 = 12.0 ft2
Aws = effective area of the water surface
Apan = effective area of the underside of the pan
Rt = thermal resistance of box lid
= 2.214 °C·m2/W = 12.58 h·°F·ft2/Btu
Ruw = thermal resistance of upper walls of box
= 2.158 °C·m2/W = 12.26 h·°F·ft2/Btu)
Rlw = thermal resistance of lower walls of box
= 2.054 °C·m2/W = 11.67 h·°F·ft2/Btu)
Rf = thermal resistance of box floor
= 2.028 °C·m2/W = 11.52 h·°F·ft2/Btu)
hws = effective heat transfer coefficient for water surface, W/(m2°C) or Btu/(h ft2°F)
hpan = effective heat transfer coefficient for the underside of the pan, W/(m2°C)
or Btu/(h ft2°F)
mair = mass flowrate of moist air, kg of dry air/h (lbm of dry air/h)
hin = enthalpy of moist air entering the inlet air duct, kJ/kg of dry air (Btu/lbm of dry air)
= 1.006 Tin + Win(2501 +1.805 Tin) for hin in kJ/kg and Tin in °C
= 0.240 Tin + Win(1061 +0.444 Tin) for hin in Btu/lbm and Tin in °F
hout = enthalpy of moist air exiting the outlet air duct, kJ/kg of dry air (Btu/lbm of dry air)
= 1.006 Tout + Wout(2501 +1.805 Tout) for hout in kJ/kg and Tout in °C
= 0.240 Tout + Wout(1061 +0.444 Tout) for hout in Btu/lbm and Tout in °F
Win = humidity ratio of air entering the inlet air duct, kg of water/kg of dry air
= humidity ratio of room air for summer tests
= humidity ratio of outdoor air for winter tests
Wout = humidity ratio of air exiting the outlet air duct, kg of water/kg of dry air
cw = specific heat of water = 4.19 kJ/(kg°C) = 1.00 Btu/(lbm °F)
cs = specific heat of steel = 0.51 kJ/(kg°C) = 0.122 Btu/(lbm °F)
Tw&p /
t = Change in pan and water temperature over a time period
t, °C/h (°F/h). This
value is positive when the pan and water temperature increases
15
Three energy balances that can be written for the system shown in Figure 20 include one for the
airspace above the pan, one for the pan with its water, and one for the airspace below the pan.
These balances are expressed as equations 1, 2 and 3 respectively.
qt +quw - qws = 0 [1]
qws + qp =
Qw&p [2]
qf + qlw + qin - qout - qp = 0 [3]
Although the
Qw&p term accounts for the energy required to change the temperature of the
water and pan, equations 1 and 3 ignore the energy required to increase the temperature of the
box itself. This is justifiable because the energy required to change the temperature of the box
is a fraction of a percent of the total energy being transferred. This in turn can be attributed to
the fact that the wood on the outside of the box barely changes in temperature during the
duration of a test, and the foam insulation lining the inside of the box has a very low heat
capacity because of its minimal mass.
Adding equations 1 and 2 and rearranging yields:
qp = qt + quw

Qw&p [4]
or
hpan = (qt + quw

Qw&p) / [(Tlower - Tw&pr) Apan ] [5]
Equation 5 was used to calculate the coefficient of heat transfer hpan as a function of water
temperature for four tests - a winter test and three summer tests. For these calculations the
effective area of the pan, Apan , was set equal to 0.91 m2 (9.8 ft2) which is the bottom area of the
pan (Table 4), and thus does not include the area of the sides, baffles or fins. The results of
these calculations (figure 21) show that hpan is not constant but increases with the temperature
of the water and pan.
5 1015202530
Summer 2 (Tdp = 23 C)
Summer 1 (Tdp = 18 C)
Winter
Water and Pan Temperature, °C
0
20
40
60
80
100
120
Coefficient of Heat Transfer, hpan, W/(m2 °C)
Summer 3 (Tdp = 21 C)
140
5 1015202530
Summer 2 (Tdp = 23 C)
Summer 1 (Tdp = 18 C)
Winter
Water and Pan Temperature, °C
0
20
40
60
80
100
120
Coefficient of Heat Transfer, hpan, W/(m2 °C)
Summer 3 (Tdp = 21 C)
140
Figure 21. Coefficient of heat transfer between the pan and air in the lower box.
16
During summer tests, the underside of the pan functions as a dehumidifier as warm, moist air is
brought into contact with a pan initially filled with ice water (figure 21). Once all the ice has
melted and the temperature of the pan and water begin to increase, the amount of moisture
condensing on the underside of the pan slows. As shown in figure 20, this resulted in decrease
in the coefficient of heat transfer. Once the temperature of the pan and water equals the dew
point temperature of the incoming air, condensation stops, and the coefficient of heat transfer
was found to increase. Note that dew point temperatures listed in figure 20 were calculated
from wet and dry bulb temperatures of the room air taken with a sling psychrometer before and
after each summer test.
Figure 21. Moisture that condensed on the underside of the pan during a
summer test is shown here as it drips onto the floor beneath the box.
For summer tests 2 and 3 a stirring device (figure 22) was added to the IBAC. This was only
done to obtain a more uniform temperature distribution within the pan. The smoother curves in
figure 20 for summer tests 2 and 3 are a direct result of this agitation. It is also quite likely that
the slightly higher coefficient of heat transfer associated with these two tests may be due to this
agitation. Obviously such agitation could not be used during a winter test when ice is being
created during the test.
Figure 22. A fiberglass rod on the end of a plastic shaft was used to stir water
during the last two summer tests. The stirring rod is shown here above a mixture of
ice and liquid water at the start of a test.
17
Power consumed by the stirring device was measured under load and unloaded. The difference
was in the 2 to 3 watt range, an indication that the energy imparted to the water by the stirring
device was negligible.
In actual use, air would only be allowed to flow under the pan when the pan is warmer than the
incoming air – a situation mirrored by the winter tests. When the incoming air is cooler than the
pan, moisture does not condense out of the incoming air onto the fins and baffles. When there
is no condensation, the moisture content of the air (i.e., the humidity ratio of the air) does not
change as it moves under the pan, meaning that Win and Wout as previously defined are equal.
This simplifies calculation of the amount of energy added (or removed) from the air as it
circulates under the pan.
During a summer test where moisture condenses out of incoming air onto surfaces under the
pan, Win and Wout will not be equal. If the rate at which moisture is condensing on surfaces
under the pan exceeds that rate at which it is being evaporated from surfaces under the pan,
then Win will exceed Wout. If the rate at which moisture is condensing on surfaces under the pan
is less than the rate at which it is being evaporated from the same surfaces, then Win will be less
Wout. Plotted in figure 23 are temperature and humidity ratio values associated with summer
test 3. The humidity ratio curves show how Win initially exceeded Wout as condensation
occurred under the pan. As the water in the pan approached the dew point temperature of the
incoming air, condensation slowed, while evaporation of water that had pooled on the floor of
the IBAC as well as the bottom of baffles and fins began to increase. After about 40 minutes
into the test, Wout surpasses Win..
0
5
10
15
20
25
30
35
01234
Tw&p 0
0.005
0.010
0.015
0.020
0.025
0.030
0.035
Tupper
Tout
Tin = Troom
Tfloor
Wout
Wint
Temperature T, C
Humidity Ratio W
Elapsed Time, h
0
5
10
15
20
25
30
35
01234
Tw&p 0
0.005
0.010
0.015
0.020
0.025
0.030
0.035
0
0.005
0.010
0.015
0.020
0.025
0.030
0.035
Tupper
Tout
Tin = Troom
Tfloor
Wout
Wint
Temperature T, C
Humidity Ratio W
Elapsed Time, h
Figure 23. Temperatures and humidity ratios recorded during summer test 3.
When moisture is condensing and/or evaporating under the pan, the energy balance equation
for the airspace below the pan (equation 3) should be used with caution as it does not take into
account energy flows associated with the condensation and evaporation of the water. This is
shown in figure 24, which contains plots of the qp (i.e., the magnitude of heat transferred from
the airspace to the pan) for summer test 3 that were calculated using both equations 3 and 4.
Equation 4 provides the correct measure of the qp. Equation 3 does not account for the amount
18
of heat energy that is associated with the mass of water that is condensed and then later
evaporated from beneath the pan. Data and calculations for summer test 3 are compiled in
Appendix E.
500
1000
1500
2000
2500
01234
Elapsed Test Time, h
Heat Transfer, W
qp(equation 4)
qp(equation 3)
500
1000
1500
2000
2500
01234
Elapsed Test Time, h
Heat Transfer, W
qp(equation 4)
qp(equation 3)
500
1000
1500
2000
2500
01234
Elapsed Test Time, h
Heat Transfer, W
qp(equation 4)
qp(equation 3)
Figure 24. Calculation of the heat transfer rate qp for summer test 3 using two different energy
balance equations. The rate calculated using equation 3 is incorrect because it does not adjust
for condensation and subsequent evaporation of water from beneath the pan.
The coefficient of heat transfer between the pan and the air in the lower box, hpan, is of primary
interest at temperatures near 0 C; that is, at temperatures at which ice will be made. For this
reason. data from five winter tests were analyzed at temperatures in the 0 to 10C range . This
data, which is summarized in Table 5, shows the effective coefficient for heat transfer from the
air below the pan to the pan is approximately 35 W/(m2°C) or 6 Btu/(h ft2°F).
One observation made during winter tests was that ice first formed on the bottom of the pan at
locations where the pan was welded to baffles and fins. Ice next covered the entire base and
sides of the pan, and then built upward.
19
Table 5. Estimates of hpan from Winter Test Data
Qp&w Tpan - Tlower hpan x Apan, hpan *
Test
Number W Btu/h °C °F W/°C Btu/(h °F) W/(m2 C) Btu/(h ft2°F)
162 554 4.91 8.84 37.0 70.2 40.7 7.16
144 491 4.37 7.87 34.8 66.0 38.3 6.73
140 478 4.13 7.44 28.5 54.0 31.3 5.51
1
127 432 6.00 10.80 21.2 40.2 23.3 4.10
193 659 6.86 12.34 25.3 48.0 27.8 4.90
187 638 7.66 13.79 27.2 51.6 29.9 5.27
186 634 6.56 11.80 28.2 53.4 31.0 5.45
2
181 617 5.98 10.77 30.4 57.6 33.4 5.88
180 614 4.72 8.50 32.0 60.6 35.1 6.18
141 482 5.61 10.1 29.1 55.2 32.0 5.63
138 471 4.67 8.41 28.5 54.0 31.3 5.51
3
117 399 4.10 7.38 28.5 54.0 31.3 5.51
343 1169 9.47 17.04 35.1 66.6 38.6 6.80
334 1139 9.16 16.49 35.1 66.6 38.6 6.80
4 326 1112 9.74 17.53 35.5 67.2 39.0 6.86
253 863 6.97 12.55 35.5 67.2 39.0 6.86
242 827 7.41 13.33 35.5 67.2 39.0 6.86
241 821 7.11 12.79 32.6 61.8 35.8 6.31
233 794 6.81 12.26 33.2 63.0 36.5 6.43
5
219 749 6.47 11.65 33.9 64.2 37.2 6.55
Mean 204 697 6.44 11.60 31.4 59.4 34.5 6.06
Std. Dev. 69 235 1.67 3.06 4.1 7.8 4.6 0.80
* Apan set equal to 0.91 m2 (9.8 ft2)
Heat and Mass Transfer Between Water and Storage Room
To assess the IBAC’s ability to condition air, the model storage cooler was attached to the IBAC
with insulated ducts as shown in figure 19a. An in-line duct fan (figure 16a) was used to move
conditioned air, and the model 1TDT2 Dayton PSC blower (figure 16b) was used to move
ambient air. Temperature and relative humidity values for air in the upper IBAC and in the
model storage cooler that were recorded during one of these tests have been plotted in figure
25. Six sets of points taken off this graph are compiled in Table 6. The three points for each
condition were plotted and then connected with a line on the psychrometric chart in figure 26.
The two resulting lines both lie just slightly above the 32°F (0°C) dew point line, an indication
that the circulation process through the storage cooler is essential one of sensible heating.
20
Upper IBAC RH
Storage
Cooler RH
Upper IBAC Temp, Tupper
Upper IBAC Dew Point Temp
Ambient Air Inlet Temp, Tin
Ambient Air Outlet Temp, Tout
Storage Cooler Temp
80
70
60
50
40
30
20
10
001000 20001500500 2500 3000
Minutes
Temperature, C
26.7
21.1
15.6
10.0
4.4
-1.1
-6.7
-12.2
-17.8
Temperature, F
Relative Humidity, %
Upper IBAC RH
Storage
Cooler RH
Upper IBAC Temp, Tupper
Upper IBAC Dew Point Temp
Ambient Air Inlet Temp, Tin
Ambient Air Outlet Temp, Tout
Storage Cooler Temp
80
70
60
50
40
30
20
10
0
80
70
60
50
40
30
20
10
001000 20001500500 2500 3000
Minutes
Temperature, C
26.7
21.1
15.6
10.0
4.4
-1.1
-6.7
-12.2
-17.8
Temperature, F
Relative Humidity, %
Figure 25. Temperature and humidity values during a test run with the storage cooler
connected to the IBAC.
Figure 26. The two read lines were obtained by connecting the data points given in Table 6.
Temperature values on this chart are in °F and the specific volume value of 13.0 is in ft3/lbm of
dry air.
21
Table 6. Example Test Data Taken from the Graph in Figure 25
Upper Air in IBAC Air in Storage Cooler
Point Dry Bulb
Temp.,
°C
Dry Bulb
Temp.,
°F
Dew
Point
Temp.,°C
Dew
Point
Temp.,°F RH, %
Dry Bulb
Temp.,
°C
Dry Bulb
Temp.,
°F RH, %
1 8.63 47.53 1.71 35.08 61.75 10.74 51.325 50.75
2 7.43 45.38 1.45 34.61 65.75 9.46 49.028 54.75
3 6.62 43.92 1.09 33.97 67.75 8.49 47.289 57.25
During a sensible heating process, a decrease in air relative humidity corresponds to an
increase in air temperature. It follows that if temperatures are kept closer to 0°C (by minimizing
heat loss through the walls of the storage unit), relative humidity levels would be closer to 100%.
In other words, saturated air would leave the IBAC and remain close to 100% RH if it is
maintained near 0°C.
Tests with the cooler box as part of the air conditioning loop showed the cooler box gained too
much heat as temperatures ran higher than desired in the system. The storage cooler was
subsequently removed from the system and a duct was installed that took air from the
conditioned air outlet to the conditioned air return on the IBAC (figure 27b). Using this new test
setup, we were still not able to achieve our desired level of cooling.
Fan
Model
Storage
Cooler
Water/Ice
Mix - 32°F
Conditioned Air
Return Air
Heat Gain
from Room
Fan
Model
Storage
Cooler
Water/Ice
Mix - 32°F
Conditioned Air
Return Air
Heat Gain
from Room
Fan
Water/Ice
Mix - 32°F
Conditioned Air
Return Air
Heat Gain
from Room
Fan
Water/Ice
Mix - 32°F
Conditioned Air
Return Air
Heat Gain
from Room
Fan
Water/Ice
Mix - 32°F
Conditioned Air
Return Air
Heat Gain
from Room
(a) (b)
Figure 27. Setups for conditioning air (a) with, and (b) without the model storage cooler.
To increase the heat transfer rate between water and the air conditioning stream, a set of fins
was placed in the pan (figure 28). These fins significantly improved heat transfer as shown in
figure 29. Future tests on the system will focus on increasing the heat exchange rate between
the ice/water and air.
22
35
37
39
41
43
45
47
49
0 100 200 300 400
Minutes
Conditioned Air Inlet Temp., °F
No fins
Fins
35
37
39
41
43
45
47
49
0 100 200 300 400
Minutes
Conditioned Air Inlet Temp., °F
No fins
Fins
Figure 28. Set of fins added to pan to
increase heat transfer between water and
conditioned air stream.
Figure 29. Placing fins in pan water resulted
in better mixing. System operated 2 °F cooler
with top fins in place.
Discussion
Model Performance
The model IBAC performed up to expectations with respect to heat exchange between ambient
air and the pan, but underperformed with respect to heat exchange between the conditioned air
stream and the ice/water in the pan. Tests with fins extending out of the ice showed that heat
exchange between the ice/water and conditioned air stream could be improved by designing the
topside of the pan more like the underside of the pan.
0
-5
-10
-15
-20
-25
30 60 90 120 150
0Hours to Freeze Pan of Water at 0C (32 F)
32
23
14
5
-4
-13
Ambient Air Temperature, C
Ambient Air Temperature, F
50 gallons
40 gallons
30 gallons
0
-5
-10
-15
-20
-25
30 60 90 120 150
0Hours to Freeze Pan of Water at 0C (32 F)
32
23
14
5
-4
-13
Ambient Air Temperature, C
Ambient Air Temperature, F
50 gallons
40 gallons
30 gallons
0
-5
-10
-15
-20
-25
30 60 90 120 150
0Hours to Freeze Pan of Water at 0C (32 F)
32
23
14
5
-4
-13
Ambient Air Temperature, C
Ambient Air Temperature, F
50 gallons
40 gallons
30 gallons
Figure 30. Time to form ice in the model IBAC for different quantities of liquid water at 0C.
Figure 30 contains a plot of the time required for the model IBAC to turn different quantities of
liquid water at 0C to solid ice at 0C as a function of ambient outdoor air temperatures. These
23
rates were based on a hpan x Apan, value of 31.4 W/C (59.4 Btu/(hF)). They assume no heat
transfer between the ice/water and air above the pan.
The IBAC will only function when ambient conditions are below 0°C for an extended period of
time. To run the system during periods above freezing or when temperatures fluctuate above
and below freezing, cooling coils could be placed directly in the water to form ice. In this case
money is saved by making ice only during off-peak electrical usage periods. Table 7 lists
operational modes for the IBAC as a function of ambient weather conditions.
Table 7. IBAC Mode of Operation
Ambient Weather Conditions System Operation
Above Freezing Cooling coils run during night
Below Freezing Ambient air used continuously
Day > 0°C
Night <0°C Ambient air used at night only, cooling coils ran only if needed to
complete freezing.
In many respects, the IBAC acts as a dehumidifier as it will help maintain a dew point
temperature for the conditioned air at a level near 0C anytime there is some ice in the pan.
This means that any humidity added to storage room air by something other than the IBAC will
end up as condensation in the IBAC pan. This also means that the only way to achieve a
relative humidity near 100% in the storage facility without active humidification is to maintain a
temperature near freezing in the facility. This, as previously noted, will require an IBAC design
with more effective heat transfer between the conditioned air and ice/water in the pan.
It is important to note that using conventional cooling coils or outdoor air to directly cool storage
room air will also dehumidify storage room air if they are below the dew point temperature of the
storage room air. The problem with using conventional cooling coils or outdoor air to directly
cool storage room air is that you run the risk of the freezing produce when they (i.e., the coils or
outdoor air) are at dry bulb temperatures below freezing. By using ambient air or cooling coils
to make ice, the IBAC eliminates this concern.
Commercial Use
In practice, the IBAC would not have to be a stand-alone unit but could be incorporated into the
floor of a cold storage facility as shown in Figures 31 and 32. In figure 31, a steel pan is
supported on fins/baffles and ambient air is brought in under the pan. In figure 32, the ice/water
is contained in an insulated concrete channel and outside air is circulated through piping laid in
the ice/water. In both cases, heavy duty steel grating is placed over the ice/water. This grating
is "thermally connected" to the ice/water with steel supports (figure 31) or steel fins (figure 32).
In this manner, the steel grating becomes an effective heat transfer element. In figure 32,
perforated pipes located just below the grating are used to blow cooler room air down onto the
surface of the ice/water. This or a similar air moving system is needed to enhance heat and
mass transfer between the ice/water and storage room air.
A complete system would need to incorporate temperature sensors, fans, dampers and a
programmable logic controller. Ideally the system would be programmed to take into account
time of day, weather trends with respect to time-of-day, and even weather forecasts. Such an
intelligent system would delay ice making to forecasted colder time periods when ice making is
more efficient (or even likely) or to times when electrical power usage may be less costly.
24
Air Intake Air Outlet
Heavy Duty
Steel Floor
Grate
Cooler Insulation
Steel Grate
Supports
Steel Pan Supports
Pan
Air Intake Air Outlet
Heavy Duty
Steel Floor
Grate
Cooler Insulation
Steel Grate
Supports
Steel Pan Supports
Pan
Figure 31. Lengthwise view of IBAC incorporated into the floor of a cold storage facility.
Cooler Insulation
Heavy Duty
Steel Floor
Grate Steel Fins
on grate
project into
water
Perforated Pipe
used to blow cooler
room air down onto
surface of ice/water
Ambient outdoor air
blown through pipes
submerged in water/ice
Concrete
Floor
Cooler Insulation
Heavy Duty
Steel Floor
Grate Steel Fins
on grate
project into
water
Perforated Pipe
used to blow cooler
room air down onto
surface of ice/water
Ambient outdoor air
blown through pipes
submerged in water/ice
Concrete
Floor
Figure 32. Cross-sectional view of IBAC incorporated into the floor of a cold storage facility.
Instead of circulating ambient air under the ice/water, ambient outdoor air is circulated through
tubes submerged in the ice/water.
25
A commercial system would also need to incorporate a system for maintaining a proper water
level. Anything that would increase the dew point temperature of the cooler room air above 0°C
is likely to increase the amount of water in the IBAC. Conversely, moisture would be lost
anytime air with a dew point temperature less than 0°C is brought into the storage facility.
Contamination of the water in the IBAC should probably be avoided as any non-volatile solute
(e.g. salts, sugars, etc) will lower the freezing point of the water. This in turn would likely mess
up a control system set to make ice at a temperature neat 0°C and/or could cause damage to
storage room contents.
Durability is a key to sustainability, and any IBAC design should minimize potential corrosion to
steel, decay of wood, and damage to concrete. One of the concerns with systems featuring
water in direct contact with concrete is how well the concrete can handle (both internally and
externally) the forces associated with expansion of water upon freezing. This may require that
the concrete channel shown in figure 32 be well reinforced and lined with a membrane that does
not allow large quantities of liquid water to diffuse into the concrete.
In lieu of manufacturing a special steel pan and then galvanizing it for containment of water,
consideration should be given to the use of widely-available, inexpensive, galvanized livestock
watering tanks. For an enclosed system similar to the model IBAC, these galvanized tanks
could be placed on a steel plate that features multiple steel supports for enhancement of heat
exchange between the tank(s) and ambient air. Alternately, livestock watering tanks could
simply be placed on the floor inside the storage cooler, and cool ambient air circulated through
pipes submerged in the tank water (this would be a system similar to that in figure 32). Grating
placed over the tanks could support crates of produce. Note that if left completely exposed,
condensation would form on the sides of such tanks.
As previously noted, cooling coils could also be placed in the water to make ice when ambient
conditions are above freezing, or replace the need to draw in ambient air in totality by using
them in a manner similar to Zall et al. (1981).
Conclusion
On site refrigerated storage is a very important asset for small scale vegetable farms. Much of
this storage takes place when ambient conditions can be used to save energy. A model Ice
Bank Air Conditioner (IBAC) was developed to take advantage of ambient air for winter storage
of vegetables. This model showed that:
1. A system that effectively forms ice using ambient air can be inexpensively fabricated.
2. Ice needs to be formed around steel fins or other similar devices that are capable of
enhancing the transfer of heat between the storage room air and the ice.
3. Using ice to condition storage room air will maintain the dew point temperature of the
storage room air near 0C regardless of the temperature of the storage room air.
Efficiency of an IBAC could be enhanced by using a programmable logic controller that switches
fans on and off and controls dampers based on the time of day, weather forecasts, and weather
trends with respect to time-of-day.
An IBAC could be augmented with cooling coils to freeze water at off-peak energy prices. This
could extend the operating months for the IBAC into earlier fall and later spring.
26
Acknowledgements
The authors wish to acknowledge graduate research assistant Andrew Holstein for his
assistance in the set-up of the summer tests.
References
Arora, C. 2000. In Refrigeration and Air Conditioning. Tata McGraw-Hill Education. p. 678.
Biermacher, J., S. Upson, D. Miller, & D. Pitmman. 2007. Economic Challenges of Small-Scale
Vegetable Production and Retailing in Rural Communities: An Example from Rural
Oklahoma. Journal of Food Distribution Research. pp. 1-13.
EERE. 2011. Earth Cooling Tubes. DOE Office of Energy Efficiency and Renewable Energy
(EERE) at
http://www.energysavers.gov/your_home/space_heating_cooling/index.cfm/mytopic=124
60. Accessed April 29, 2011.
James, S., & C. James. 2010. The food cold-chain and climate change. Food Research
International, 1944-1956.
Lung, R. B. 2006. The Role of Emerging Technologies in Improving Energy Efficiency.
Proceedings of the Twenty-Eighth Industrial Energy Technology Conference. New
Orleans.
Maynard, D. N. & G. L. Hochmuth . 2007. Knott's Handbook for Vegetable Growers 5th ed. New
York: John Wiley & Sons, Inc.
Pincus, S. 2010. Personal interview on September 18, 2011 of Tipi Produce Farm owner Steve
Pincus. Interviewed by L. Syse, A. Gardebrecht, M. Muehlbauer, A. Lofy, & J. Banach.
Silvetti, B. 2007. Thermal Energy Storage. In B. L. Capehart, Encyclopedia of Energy
Engineering and Technology. CRC Press. pp. 1412-1422.
Zall, R. R., Jordan, W. K., Ludington, D. C., & Chen, J. H. 1981. Using Winter Coldness to
Provide Refrigeration. Transactions of the ASAE. pp. 1073-1076.
27
Appendix A - Vegetable Storage Requirements
Table A. Storage Temperature, Relative Humidity and Storage Life of Vegetables
Vegetable Temp.,
°F RH,
% Storage Life Vegetable Temp.,
°F
RH,
% Storage Life
Artichoke 32 95-100 2-3 weeks Kohlrabi 32 98-100 2-3 months
Asparagus 32-35 95-100 2-3 weeks Leek 32 95-100 2-3 months
Bean, lima 37-41 95 5-7 days Lettuce 32 98-100 2-3 weeks
Bean, snap 40-45 95 10-14 days Melon, honeydew 45 90-95 3 weeks
Bean sprouts 32 95-100 7-9 days Melon, muskmelon 32-36 95 5-14 days
Beet, bunched 32 98-100 10-14 days Melon, watermelon 50-60 90 2-3 weeks
Beet, topped 32 98-100 4-6 months Mushroom 32 95 3-4 days
Bitter Melon 53-55 85-90 2-3 weeks Onion, dry 32 65-70 1-8 months
Bok Choy 32 95-100 3 weeks Onion, green 32 95-100 3-4 weeks
Broccoli 32 95-100 10-14 days Parsley 32 95-100 8-10 weeks
Brussels Sprouts 32 95-100 3-5 weeks Parsnip 32 98-100 4-6 months
Cabbage 32 98-100 3-6 weeks Pea 32 95-98 1-2 weeks
Cabbage, Chinese 32 95-100 2-3 months Pea, snow 32-34 90-95 1-2 weeks
Carrot, bunched 32 95-100 2 weeks Pepper, sweet 45-55 90-95 2-3 weeks
Carrot, mature 32 98-100 7-9 months Potato, early 40 90-95 4-5 months
Carrot, immature 32 98-100 4-6 weeks Potato, late 50-60 90-95 5-10 months
Cauliflower 32 95-98 3-4 weeks Radicchio 31-32 90 2-3 months
Celeriac 32 97-99 6-8 months Radish, spring 32 95-100 3-4 weeks
Celery 32 98-100 2-3 months Radish, winter 32 95-100 2-4 months
Chard 32 95-100 10-14 days Rhubarb 32 95-100 2-4 weeks
Chinese Long Bean 40-45 90-95 7-10 days Rutabaga 32 98-100 4-6 months
Collards 32 95-100 10-14 days Salsify 32 95-98 2-4 months
Cucumber 50-55 95 10-14 days Scorzonera 32-34 95-98 6 months
Eggplant 46-54 90-95 1 week Spinach 32 95-100 10-14 days
Eggplant, Japanese 46-54 90-95 1 week Squash, summer 41-50 95 1-2 weeks
Endive/Escarole 32 95-100 2-3 weeks Squash, winter 50 50-70 Variable
Garlic 32 65-70 6-7 months Sweet corn 32 90-98 5-14 days
Haricot Verts 40-45 95 7-10 days Tomatillo 55-60 85-90 3 weeks
Horseradish 30-32 98-100 10-12 months Tomato, ripe 46-50 90-95 4-7 days
Kale 32 9 5-100 2-3 weeks Turnip 32 95 4-5 months
From: Maynard, D. N., & Hochmuth, G. L. (2007). Knott's Handbook for Vegetable Growers 5th ed. New York: John Wiley & Sons,
Inc.
28
Appendix B - Preliminary Calculations
Heat Load Calculation
Carrot Mass = 3630 kg = 8000 lbm
Respiration Rate = 2.19 kJ/(kg·day) = 0.94 Btu/(lbm·day)
Wall Area, A = 55.7 m2 = 600 ft2
Wall R-Value, R = 4.40 °K·m2/W = 25 h·°F·ft2/Btu
T Across Walls = 11.1°K = 20°F
qrespiration = (Respiration Rate)(Carrot Mass) = 7950 kJ/day = 92 W = 7500 Btu/day
qconduction = (A T)/ R = 12,100 kJ/day = 140 W = 11,500 Btu/day
qtotal = qrespiration + qconduction = 20,000 kJ/day = 232 W = 19,000 Btu/day
Heat Sink Calculation
Size pan (heat sink) for triple the daily heat loss = 60,000 kJ = 57,000 Btu
Energy Required to Melt Ice, Qmelt = (Hfusion)(Mwater ) = 60,000 kJ = 57,000 Btu
Latent Heat of Fusion, Hfusion = 335 kJ/kg =144 Btu/lbm
Mass of Water, Mwater = Qmelt /Hfusion = 179 kg = 395 lbm
Density of Water,

water = 1000 kg/m3 = 62.4 lbm/ft3
Volume of Water, Vwater = Mwater /
water = 0.179 m3 = 6.33 ft3 = 53 gallons
29
Appendix C - Pan Part Drawings
30
Appendix D - IBAC Material and Labor Costs
Table D. Expenses for the Ice Box Air Conditioner Prototype
Description Quantity Unit Price Total
Extruded polystyrene insulation - 2 in. x 4 ft x 8 ft, R-10 3 $21.50 $64.50
Oriented Strand Board (OSB) - 5/8 in. x 4 ft x 8 ft 3 $10.46 $31.38
Lumber - nominal 2- by 4-inch x 10 ft 15 $1.89 $28.35
Insulated flexile duct - 8-inch diameter, 25 ft, R-6 1 $29.97 $29.97
Polyurethane adhesive (PL Premium) - quart tube 1 $6.87 $6.87
Deck screws #8 x 1-3/4 in. - 5 lbm 0.25 $18.98 $4.75
Galvanized elbow - 8 inch diameter, 90° 2 $4.10 $8.20
Flex duct connector - 8 inch diameter 2 $2.49 $4.98
Spray foam insulation -12 oz bottle 2 $2.50 $5.00
Self-adhesive foam tape - 1/8 in. x 1/4 in. x 10 ft 3 $2.97 $8.91
Polystyrene foam adhesive -qt tube 1 $5.99 $5.99
Latches, 4-1/2" 2 $3.39 $6.78
Steel butt hinges - 3 in. x 3 in. 3 $1.29 $3.87
Galvanized duct caps - 8 inch diameter 4 $4.92 $19.68
Permanent split capacitor (PSC) blower -1500 cfm 2 $148.05 $296.10
Sheet steel, 16 gauge, 4 ft x 8 ft 1.5 $57.20 $85.07
Misc supplies (weld wire, drain line, stopper) $24.60
Material Total $635.00
Pan fabrication labor (1-hr unit) 6 $30.00 $180.00
HX box fabrication labor (1-hr unit) 10 $15.00 $150.00
Labor Total $330.00
Total $965.00
31
Appendix E – Summer Test 3 Data and Calculations
Variable Value
Ice added to pan, kg (lbm) 28.37 (62.56)
Water added to pan, kg (lbm) 62.06 (136.85)
Ice and water added to pan, kg (lbm) 90.43 (199.41)
Airflow in exhaust duct based on vane anemometer measurements, m3/min (ft3/min) 10.35 (365.7)
Airflow in exhaust duct based on pitot-static tube measurements m3/min (ft3/min) 11.06 (390.8)
Airflow at input duct based on static pressure drop across blower, m3/min (ft3/min) 12.54 (443)
Wet bulb temperature of room air at beginning of test, C (F) 23.6 (74.5)
Dry bulb temperature of room air at beginning of test, C (F) 29.4 (85.0)
Wet bulb temperature of room air at end of test, C (F) 24.4 (76.0)
Dry bulb temperature of room air at end of test, C (F) 31.7 (89.0)
Elapsed
Test Time Tupper Tout Tin and
Troom Tfloor Tw&p Tlower Wout Win
h C C C C C C kg H2O/kg da kg H2O/kg da
0.0 16.7 24.6 29.6 28.4 0.5 27.1 0.0121 0.0136
0.2 14.1 24.1 29.8 28.5 1.1 27.0 0.0122 0.0136
0.4 13.5 24.1 29.8 28.5 3.2 27.0 0.0123 0.0136
0.6 14.3 24.6 29.9 28.5 6.9 27.2 0.0128 0.0138
0.8 15.8 25.3 30.0 28.4 10.6 27.6 0.0131 0.0139
1.0 17.3 26.0 30.1 28.4 13.6 28.1 0.0136 0.0139
1.2 18.6 26.6 30.2 28.4 16.1 28.4 0.0140 0.0138
1.4 19.7 27.1 30.3 28.4 17.9 28.7 0.0142 0.0140
1.6 20.6 27.6 30.5 28.4 19.4 29.1 0.0144 0.0140
1.8 21.5 28.1 30.6 28.4 20.6 29.4 0.0141 0.0136
2.0 22.2 28.5 30.7 28.5 21.6 29.6 0.0140 0.0136
2.2 22.9 28.8 30.7 28.5 22.6 29.8 0.0138 0.0136
2.4 23.6 29.0 30.8 28.6 23.6 29.9 0.0135 0.0132
2.6 24.2 29.3 30.8 28.6 24.4 30.0 0.0132 0.0128
2.8 24.8 29.6 30.9 28.7 25.2 30.2 0.0123 0.0122
3.0 25.3 29.9 31.1 28.7 25.9 30.5 0.0117 0.0116
3.2 25.9 30.0 31.1 28.7 26.5 30.6 0.0110 0.0111
3.4 26.5 30.2 31.2 28.7 27.1 30.7 0.0107 0.0108
3.6 27.1 30.4 31.3 28.8 27.7 30.9 0.0109 0.0109
3.8 27.9 30.6 31.3 28.9 28.2 31.0 0.0106 0.0107
4.0 28.3 30.7 31.4 29.0 28.7 31.1 0.0106 0.0106
32
Appendix E – Summer Test 3 Data and Calculations, cont.
Elapsed
Test Time qt quw qlw qf qws hin hout Specific Volume
of input air
h W W W W W kJ/kg da kJ/kg da m3/kg da
0.0 6 10 2 1 16 64.56 59.37 0.8768
0.2 8 12 2 1 20 64.78 58.83 0.8775
0.4 8 12 2 1 20 64.70 58.79 0.8774
0.6 8 12 2 1 19 65.37 59.92 0.8778
0.8 7 11 2 0 18 65.66 60.88 0.8782
1.0 6 10 1 0 16 65.84 61.61 0.8787
1.2 6 9 1 0 14 65.56 61.85 0.8787
1.4 5 8 1 0 13 66.22 62.90 0.8794
1.6 5 7 1 0 12 66.35 63.41 0.8798
1.8 5 7 1 -1 11 65.64 63.10 0.8797
2.0 4 6 1 -1 11 65.74 63.52 0.8799
2.2 4 6 1 -1 10 65.76 63.76 0.8801
2.4 4 5 1 -1 9 64.78 63.02 0.8796
2.6 3 5 1 -1 8 63.78 62.20 0.8791
2.8 3 5 0 -1 8 62.35 60.93 0.8787
3.0 3 4 0 -1 7 61.06 59.79 0.8783
3.2 3 4 0 -1 6 59.59 58.49 0.8776
3.4 2 3 0 -1 6 59.10 58.09 0.8775
3.6 2 3 0 -1 5 59.36 58.48 0.8779
3.8 2 3 0 -1 4 58.91 58.13 0.8777
4.0 2 2 0 -1 4 58.69 58.02 0.8778
33
Appendix E – Summer Test 3 Data and Calculations, cont.
Elapsed
Test Time
Volume
of Input
Air qin qout qin - qout
Tw&p/
t
Qp&w qp
(eq. 4) qp
(eq. 3)*
h m3/min kW kW W C/min W W W
0.0 12.5 15.34 14.11 1234 0.0071 46 30 1236
0.2 12.5 15.38 13.97 1412 0.0960 624 604 1415
0.4 12.5 15.36 13.96 1404 0.2432 1581 1561 1407
0.6 12.5 15.51 14.22 1292 0.3319 2158 2138 1294
0.8 12.5 15.58 14.44 1134 0.2817 1832 1814 1136
1.0 12.5 15.61 14.61 1002 0.2272 1477 1461 1003
1.2 12.5 15.54 14.67 879 0.1781 1158 1143 880
1.4 12.5 15.69 14.90 787 0.1386 901 888 788
1.6 12.5 15.71 15.02 697 0.1093 711 698 698
1.8 12.5 15.54 14.94 600 0.0896 582 571 600
2.0 12.5 15.56 15.04 526 0.0863 561 551 526
2.2 12.5 15.57 15.09 472 0.0814 529 519 472
2.4 12.5 15.34 14.93 419 0.0740 481 472 419
2.6 12.5 15.11 14.74 374 0.0671 436 428 373
2.8 12.5 14.78 14.45 336 0.0615 400 392 336
3.0 12.5 14.48 14.18 303 0.0580 377 370 302
3.2 12.5 14.15 13.89 261 0.0519 337 331 260
3.4 12.5 14.03 13.79 238 0.0474 308 303 237
3.6 12.5 14.09 13.88 208 0.0428 279 273 207
3.8 12.5 13.98 13.80 184 0.0391 254 250 184
4.0 12.5 13.93 13.77 159 0.0378 246 242 158
* The rate calculated using equation 3 is incorrect because it does not adjust for
condensation and subsequent evaporation of water from beneath the pan.
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