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The Electronic Newsletter of The Industrial Refrigeration Consortium
Vol. 13 No. 1, 2013
PROSPECTS
FOR DESICCANTS IN
REFRIGERATED FACILITY APPLICATIONS
In the last edition of The Cold Front (Vol. 12 No. 4), we introduced the basic principles of operation
of desiccants and showed common configurations of the solid adsorbent media commonly used in air
systems applications. In this issue, we discuss the application of solid desiccants in a refrigerated warehouse
dock application.
INTRODUCTION
Refrigerated warehouses are designed to provide temperature-controlled spaces in order to store a wide
variety of food products. Coolers operate from temperatures that range from just above freezing to as high
as 50°F with some specialty spaces, such as fruit ripening rooms operating even warmer. Freezers are
designed to operate at temperatures well below 32°F and generally frozen products are held at temperatures
that can range from 0°F to -15°F for general frozen foods and -20°F or colder for ice cream. Chapter 21 of the
ASHRAE Refrigeration Handbook provides more details on the temperature and humidity requirements for a
IRC Staff
In This Issue
Director
Doug Reindl
• Prospects for Desiccants In
1-9
Refrigerated Facility Applications
608/265-3010
or 608/262-6381
dreindl@wisc.edu
Assistant Director
Todd Jekel
608/265-3008
tbjekel@wisc.edu
Research Staff
Dan Dettmers
• Upcoming Ammonia Classes
2
• Noteworthy
2
608/262-8221
djdettme@wisc.edu
IRC Contact Information
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FAX
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Web Address
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Vol. 13 No. 1, 2013
wide range of perishable and frozen products. Our focus
in this edition of the Cold Front is on space
conditioning for the dock areas associated with
refrigerated facilities storing general frozen goods similar
to the plan view diagram shown below in Figure 1.
The most common means for controlling the temperature
in a refrigerated freezer warehouse, regardless of its
operating temperature, is by using air cooling evaporators
in either a ceiling-hung or penthouse arrangement.
Because these evaporators typically operate at
temperatures well below 32°F, frost from the air will
accumulate on the evaporator surfaces in proportion to
the moisture (latent) load in the space (Reindl and Jekel
2009). As frost accumulates on the surfaces of an air unit,
its capacity decreases; thereby, necessitating periodic
defrosting. With higher moisture (latent) loads, more
frequent defrosts are required and increased defrost
frequency degrades the overall efficiency of the
refrigeration system. In addition to degrading
refrigeration efficiency and capacity, higher latent loads
can lead to the formation of condensation or ice on floors,
doors, walls, and ceilings with a corresponding increase in
worker hazard (Cleland 2005).
Where does the moisture within the freezer come from?
For tightly constructed refrigerated warehouse structures,
the primary source of moisture that makes its way into the
freezer is from warm moist outside air first infiltrating the
dock before finally moving into the freezer. Of course
dock evaporators will partially cool and dehumidify the
infiltrating outside but that infiltrating air will eventually
migrate into the freezer. Therefore, the dock of a
refrigerated warehouse serves not only as a staging area
for products being loaded/unloaded but also as an
“anteroom” or a vestibule that separates the outside
ambient environment from the freezer.
Appropriately conditioning a dock by maintaining its
temperature and humidity can mitigate the adverse effects
of moisture infiltrating to the freezer. Docks are normally
held in a temperature range between 35°F and 45°F. This
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Noteworthy
Noteworthy
•• Visit
Hansen
the IRC
Technologies
website tohas
access
issued
presentations
a recall for certain
made atmodels
the 2011
of theirIRC
safetyResearch
relief valves.
and
Technology
Forum.
Browse here for more information: Hansen Technologies Product Safety Notice.
•• Mark
now
fornewsletter
the 2012
IRC Research
and Technology
Send your
itemscalendars
of note for
next
to Todd
Jekel, tbjekel@wisc.edu.
Forum – May 2-3, 2012 at the Pyle Center in Madison, WI.
• Send items of note for next newsletter to Todd Jekel, tbjekel@wisc.edu.
2
Vol. 13 No. 1, 2013
intermediate operating temperature for the dock allows moisture from the infiltrating outside air to be partially
removed; thereby, reducing the latent load on the freezer. This operating strategy allows a larger fraction of
the ambient infiltrating moisture (latent load) to be removed using more efficient and less expensive highertemperature refrigeration. In general, it is desirable to remove as much moisture from the dock air as possible
to minimize the freezer latent load.
Figure 1: Refrigerated warehouse (freezer) with dock area.
MANAGING MOISTURE
The most effective strategy for decreasing the moisture load on a refrigerated facility is to reduce the
infiltration rate between spaces of different temperature (outdoors-dock and dock-freezer). Reducing the rate
of outdoor air infiltrating to the dock, can be accomplished by controlling (closing) truck load-out doors and by
ensuring that the seals between the load-out doors and trailers locked in their bay are installed, operational,
and well maintained.
Minimizing the infiltration of dock air to the freezer is accomplished by the use of protection devices installed
on the doors between the dock and freezer. Cleland, et al. (2004) field-measured infiltration rates for a range
of different interior door types and means of door air exchange protection devices (strip curtains, air curtains,
fast-acting doors, etc.). Cleland, et al. used this data to refine previously-published methods for predicting air
infiltration rates. Although these door protection devices can reduce air infiltration to the freezer, the
infiltration reduction measures mentioned above cannot be overemphasized because this is the ultimate
source of moisture in most refrigerated warehouses. High outdoor air infiltration rates to the dock coupled
with air exchange from the freezer to the dock (a cooling credit) makes the dock a difficult space to condition.
To achieve an acceptable level of humidity control on the dock, the use of reheat is often required.
Ultimately, two key factors dictate the moisture (latent) load on a freezer associated with infiltrating dock air
(1) the rate of air exchange between the dock and freezer and (2) the moisture content in the dock air. We
have already discussed the importance of reducing the rate of air infiltration from the dock to the freezer and
from outdoors to the dock. Now let’s look at ways we can manage the moisture content of dock air to mitigate
moisture effects on the freezer operation.
Although the most common approach for dehumidifying dock air is by the use of refrigerant-based air units, the
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Vol. 13 No. 1, 2013
use of solid desiccants as a means of moisture removal is a growing trend. From a design standpoint, the dock
temperature set point is bounded on the high end by the outdoor air temperature (unconditioned dock) and by
the freezer temperature (essentially, no dock). Frequently, we will see docks controlled to temperatures in the
40-45°F range but operating the dock at a lower set point temperature decreases the moisture content of dock
air and latent load on the freezer; however, a lower dock set point temperature will require higher refrigeration
costs for conditioning the dock.
Figure 2 shows how dramatic the moisture content of air decreases as the dock air temperature set point is
lowered. Assuming the freezer is maintained at -10°F, its moisture content would typically equilibrate at about
2.9 grains of moisture per lb of dry air (gr/lb). When the dock set point temperature is 50°F, the corresponding
moisture content would be 48 gr/lb. In this case, the net moisture load to the freezer would be 45.1 grains for
each 13.3 cubic feet of infiltrating air. If the dock were maintained just 10°F cooler with a 40°F set point, the
equilibrium moisture content for the dock air drops to 32.7 gr/lb and the moisture load on the freezer
decreases to 29.8 gr for each 13.3 cubic feet of infiltrating air resulting in a 34% decrease in freezer moisture
load.
Figure 2: Moisture content in dock air assuming 90% relative humidity.
As introduced in the last edition of The Cold Front, a solid desiccant can be viewed effectively as a
“heat activated water vapor pump.” Water vapor from air is adsorbed by the desiccant material as the air
passes through the desiccant matrix. When the desiccant adsorbs moisture on its surfaces, water gives up its
heat of sorption resulting in a discharge air condition of the desiccant being hotter and drier. As the desiccant
material itself becomes “saturated” with water vapor, it has to be regenerated (or reactivated) using a suitable
heat source to drive off adsorbed moisture. The regeneration air stream temperature must be relatively hot (in
the range of 200-275°F for this application) so a consequence of using a desiccant for dehumidifying dock air is
a net increase in sensible refrigeration load on the dock. The heat addition for regeneration can be electric,
steam, or gas-fired combustion.
CASE STUDY
Referring again to the refrigerated freezer as shown in Figure 1, assume it is storing general frozen goods
at -10°F with a connected dock being maintained at 35°F. The dock has 8,000 square feet of floor area per
freezer door and 1,000 square feet of dock area per truck bay door. The freezer doors in this case are 10 feet
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Vol. 13 No. 1, 2013
wide and 14 feet tall and assumed to be open 6 minutes per hour on average while the truck bay doors are 9
feet wide by 9 feet tall and assumed to be open 1.5 minutes per hour on average.
Sensible Load [tons per door]
The corresponding sensible load to the freezer resulting from the infiltration of dock air when the freezer door
protection effectiveness values range from 85% to 95% is shown in Figure 3. using the methodology used
in the 2010 ASHRAE Refrigeration Handbook Chapter 24, which is the method developed by Downing and
Meffert (1993). A fast acting door that is open 9 minutes per hour corresponds to a door effectiveness of 85%,
if it were only open 3 minutes per hour the effectiveness would be 95%. A horizontal recirculating air curtain
protecting the door would have an effectiveness of 85-90%.
12
10
Warehouse: -10 oF, 90% RH
Dock: 80% RH
h = 85%
8
h = 90%
6
4
h = 95%
2
0
20
25
35
30
40
45
50
Dock Temperature [o F]
Figure 3: Freezer sensible load due to effects of infiltrating air from the dock.
As expected, higher dock dry bulb temperatures result in higher freezer sensible loads. The latent load on the
freezer attributable to infiltration will depend on the traffic through the door, the effectiveness of the freezer
door protection and the moisture content of dock air infiltrating. Figure 4 shows the latent load on the
freezer for varying dock temperatures and freezer door effectiveness.
Two alternatives are analyzed to maintain identical dock space conditions. The base case involves using airunits for conditioning the dock to maintain a 35°F set point along with hot-gas reheat to maintain the 80%
relative humidity. The alternate case meets the same dock space conditions, but uses a solid desiccant
dehumidifier that takes dock air into the unit for dehumidification and recirculates the dry air back to the dock
as illustrated in Figure 5. For this example, the outdoor air is at a design condition of 95°F dry bulb and
80°F wet bulb (52.5% RH and 131 gr/lb humidity ratio). The dock envelope insulation values are: R-30 on the
roof, R-10 on the floor, and R-25 on both the exterior and interior walls. The roof and exterior walls are light
colored to reduce solar heat gain. The dock lighting level is assumed to be 0.5 W/ft2 and there are four (4)
persons in the dock operating two (2) fork trucks and two (2) pallet jacks. The load associated with infiltration
from the ten (10) truck load-out doors assumed to be fully open 1.5 minutes per hour corresponding is
estimated at 2.1 tons per door. The infiltration between the freezer and the dock assumes a 6 minute per hour
equivalent open time with the freezer conditions of -10°F and 90% RH. For the design dock load, the freezer
infiltration, a credit, is assumed to be zero. The electric and gas utility rates are fixed at $0.10/kWh for
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Vol. 13 No. 1, 2013
60
6
Warehouse: -10 oF, 90% RH
Dock: 80% RH
h = 85%
40
4
h = 90%
20
2
h = 95%
0
20
0
25
30
35
40
45
50
Dock Temperature [o F]
Figure 4: Latent load on freezer due to infiltrating dock air.
Intake
from
ambient
Conditioned
Supply Air
Desiccant
wheel
Intake air
from dock
Ambient
35 F
Freezer
-10 F
Exhaust to
ambient
Figure 5: Illustration of a desiccant unit providing dehumidification of dock air.
6
mw [lb/hr]
Latent Load [tons per door]
electricity and $0.50/therm for natural gas (to regenerate the desiccant).
Vol. 13 No. 1, 2013
The capacity of the dock evaporators is based on manufacturers’ ratings and assumes that the rated capacity
includes both the sensible and latent heat removal capability. The latent capacity of the dock evaporator is
estimated using the straight-line principle. That is, the outlet condition from the evaporator is assumed to lie
on a straight “line” connecting the dock conditions and the apparatus dew point (i.e. saturated evaporating
temperature) on the psychrometric chart. Reheat is added to ensure the latent capacity of the evaporators will
maintain the dock at the specified 80% RH. The selected dock evaporators have 12.5 tons of total capacity with
a 15°F TD and require 3-hp of fan per unit. The evaporator fans are assumed to be always on in order to ensure
that the dock air is well mixed and the risk of localized condensation is minimized.
For the alternative case with a solid desiccant dehumidification system, the desiccant unit’s energy use is based
on manufacturers’ performance (www.munters.com). At the specified dock conditions, for each ton of latent
load reduction to the dock attributable to the desiccant (~11 lb water/hr removal), there is an additional 1.9
tons of sensible load added to the dock, and approximately 3 kW of process and regeneration fan energy use.
The desiccant unit also requires 10 kW (or 0.34 therm/hr) of energy use associated with regeneration for each
ton of dehumidification. The use of a desiccant changes the dock load’s sensible heat ratio by removing latent
load and adding sensible load. The desiccant unit was sized to reduce or eliminate the need for reheat on the
dock while being able to maintain the target 80% RH on the dock.
Loads
The first thing to consider is the design loads for the dock space. These loads will dictate the sizing of
equipment for both cases. Although the dock design loads assume no cooling credit from air that infiltrates
from the freezer to the dock, the dock conditioning equipment will still need to provide good space
temperature and humidity control when there is freezer air infiltration.
For both systems, the space loads are assumed to be identical. The space loads consist of infiltration, heat gain
through the structure, lights, people and equipment. The equipment related load like fan power, desiccant unit
power, and reheat will be added separately to each alternative.
The aggregate space load for the design conditions consists of 22 tons sensible and 11.9 tons latent for a total
of 33.9 tons at a sensible heat ratio of 65%. The breakdown of the load in tons is shown in Figure 6.
Design Loads
5.4
1.4
0.3
11.9 tons
Envelope
Lights
People
5.5
Equipment
Infiltration (S)
Infiltration (L)
9.4
Figure 6 Breakdown of space loads in dock.
7
Vol. 13 No. 1, 2013
Mechanical Refrigeration-Only Design
For the mechanical refrigeration-only system, five (5) evaporators are installed and provide a total of 62.5 tons
of cooling capacity. To achieve the 80% space relative humidity target on the dock, a total of 140 MBH (11.75
tons) of reheat is required. The reheat is achieved using coils fed with hot-gas from the refrigeration system.
The number of evaporators specified will allow this design load to be met even if one of the evaporators is in
defrost.
The total load on the refrigeration system for this case is 43.6 tons at the design condition. This load translates
to 230 ft2/ton or 0.175 tons per 1,000 cubic feet load and that figure of merit agrees with well with the typical
range for a dock space given by Stoecker (1998).
Assuming the portion of the refrigeration system serving the dock operates with a 20°F saturated suction (34
psig), a reasonable approximation for the refrigeration system efficiency is about 1.15 hp/ton. Combining the
refrigeration system power with the fan power from the evaporators results in a total electric demand of 51.8
kW and assuming $0.10/kWh, the hourly cost of this option at the design condition is $5.18.
Desiccant-assisted Mechanical Refrigeration Design
As noted above, the basis for sizing the desiccant system is to eliminate the need for reheat while maintaining
the dock at 80% relative humidity. With this assumption, the desiccant system must be capable of removing 30
lb water/hr. A desiccant system with this moisture removal capacity will yield a 2.7 tons of latent load
reduction to the dock but result in 5.2 tons of added sensible load. The elimination of reheat allows the
designer to meet the dock space conditioning requirements with only four (4) evaporators.
The total load on the dock portion of the refrigeration system in this case is 33.6 tons at the design condition.
This represents a 300 ft2/ton or 0.135 tons per 1,000 cubic feet load and that still fits well within the typical
range for a dock space given by Stoecker.
Combining the total refrigeration system for the dock, fan, and desiccant unit power (48.5 kW) and the natural
gas use of the regeneration of the desiccant at $0.50/therm, the hourly cost of this option at the design
condition is $5.30. This is a modest $0.12/hr (2.3%) increase in system hourly cost. If electric resistance is used
to regenerate the desiccant rather than the less costly natural gas, the hourly cost increases to $5.85 (+13%).
Design considerations
Comparison summary of the two (2) design alternatives is shown below:
Mechanical only
Mechanical + desiccant
Number of
Evaporators
5
Refrigeration
Load
43.6 tons
4
33.6 tons
Reheat or Desiccant
Dehumidification
140 MBH reheat
30 lb water per hour
dehumidification
Hourly Cost at Design
conditions
$5.18
$5.30 (gas regen)
$5.85 (elect regen)
The economics will depend on the capital cost difference between the two alternatives. The desiccant system
option requires the installation of the desiccant unit and ducting to the space but requires one fewer
evaporator than the air unit-only option.
Operational considerations
The hourly cost at design conditions in only one consideration, what about off design conditions? How will the
performance and operation costs for these two system options compare throughout the year? The first
consideration that must be addressed is the sensible and latent credit from the freezer infiltration. These
credits reduce the hourly cost of conditioning the dock, but effect how the systems are run to maintain the
dock conditions. Analyzing an off-design condition at the same outdoor ambient state but accounting for
8
Vol. 13 No. 1, 2013
freezer infiltration will be the first case considered.
For the mechanical refrigeration only alternative assuming that the freezer doors are open 6 minutes per hour,
that provides a 15.44 ton sensible credit and a 5.29 ton latent credit to the dock. The resulting dock space
relative humidity balances out at 82%. The total resulting refrigeration load is 22.8 tons with all the reheat coils
operating and only two (2) evaporators in refrigerating mode. The other three (3) evaporators have their fans
running, but are not actively refrigerating (i.e. cooling). The resultant hourly cost at this off-design condition is
$3.30.
For the desiccant alternative assuming the same freezer door operation, the freezer infiltration credit to the
dock is 15.45 ton sensible and 5.35 ton latent. The resulting space relative humidity balances out at 83% with a
total refrigeration load of 12.7 tons and a desiccant unit running along with only one (1) evaporator in
refrigerating mode. Just like with the mechanical only alternative, the other three (3) evaporators have their
fans running, but are not feeding refrigerant (i.e. cooling). The resultant hourly cost with natural gas
regeneration is $3.42, and $3.96 with electric regeneration. Comparatively, the natural gas-regenerated
desiccant option increases the hourly operating cost by $0.12 (+4%).
Comparison summary of the two (2) design alternatives with freezer credits is shown below:
Mechanical only
Mechanical + desiccant
Active
Evaporators
2
Refrigeration
Load
22.8 tons
Dock Relative
Humidity
82%
1
12.7 tons
83%
Hourly Cost with
freezer credit
$3.30
$3.42 (gas regen)
$3.96 (elect regen)
Notice in the above summary that the system with the desiccant balances out at a slightly higher relative
humidity in the dock when the freezer credits are considered. That is due to the fact that only one dock
evaporator is running to maintain the dock temperature, thus only one evaporator is providing
dehumidification and the desiccant is doing most of water removal. The relative humidity could be reduced if
one of the evaporators were operated in a modified defrost mode (i.e. defrosting with the fans on) thus
requiring another evaporator to be in refrigeration mode.
CONCLUSIONS
At both design and the single off-design condition analyzed, a desiccant unit used for space condition a dock
area led to increased operating costs compared to a system using air-cooling units with hot-gas reheat for
humidity control. Whether or not there is a first cost savings associated with the desiccant option will depend
on the cost impact of the refrigeration system design requiring the additional evaporator and reheat coils
versus the cost of the desiccant unit and installation of natural gas service for regeneration.
Regardless of whether or not a desiccant is applied, both systems can successfully maintain the dock at the
specified temperature and humidity conditions in the case considered.
REFERENCES
Cleland, D.J., Chen, P., Lovatt, S.J., and Bassett, M.R., “A modified model to predict air infiltration into refrigerated facilities
through airways”, ASHRAE Transactions, Vol. 110, No. 1, pp. 58-66, (2004).
Cleland, D.J., “Implications of coil frosting on system designs for low-temperature applications”, ASHRAE Transactions, Vol.
111 No. 1, pp. 336-345, (2005).
Downing, C.C. and Meffert, W.A., “Effectiveness of cold-storage door infiltration protective devices,” ASHRAE Transactions
99(2): 356-366, (1993).
Reindl, D.T. and Jekel, T.B., “Frost on air-cooling evaporators”, ASHRAE Journal, Vol. 51, February, pp. 27-33, (2009).
Stoecker, W.F., Industrial Refrigeration Handbook, McGraw Hill, (1998).
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