Desiccant dehumidification and mechanical refrigeration compared
An exercise in determining which type of system is best suited for a given application.
By Guy A. Griesse and Kenneth E. Baker, Bry-Air
Both desiccant dehumidifiers and mechanical refrigeration systems can remove moisture from the air, so the question is: which type is best suited for a given application? There really are no simple answers to this question, but we can make several general conclusions based on our theoretical work.
To assist us in making such conclusions, let`s compare a desiccant-based dehumidification system at three different design conditions. For illustration purposes, the energy of the components in each system will be given in kilowatts.
For these comparisons a hypothetical manufacturing plant in Rio de Janeiro, Brazil, will be used. In all of the following examples, the physical facts and assumptions will remain unchanged.
Physical facts
a. Outside ambient condition: 94 degrees Fahrenheit db, 80 degrees F wb (1 percent summer design).
b. Area to be conditioned: 50 feet ¥ 50 feet ¥ 10 feet.
c. Construction: Frame construction, concrete floor.
d. Insulation: Walls and ceiling — “U” factor of .15.
e. Doors: 1 (3 feet ¥ 7 feet) opened 20/hr; 1 (8 feet ¥ 8 feet) opened 4/hr.
f. Number of people working in area — 10.
g. Lights in area — 5,000 watts.
h. Motors in area — 7.5 HP total.
i. Other equipment — 2 kW total.
j. Outside air required — 200 CFM.
Assumptions
a. The interior of the conditioned space is coated with two coats of vapor barrier paint.
b. All physical cracks are sealed.
c. All doors are adequately weather-stripped.
d. A vapor barrier is provided in or under the concrete floor.
e. All workers are engaged in light physical exertion.
f. There are no other openings in and out of the conditioned space.
g. All filters are 30 percent efficient.
h. Supply fan has an external static pressure capability of 1.0 inch water column (WC) to overcome supply and return air duct losses.
i. Reactivation fan on desiccant-based system has an external static pressure of 1.0 inch WC.
j. All fans assumed to have an 80-percent static efficiency.
k. All motors assumed to have an 80-percent motor efficiency.
l. DX cooling is by air-cooled condensing units.
 |
As the above comparison shows, at equal design conditions and equal latent and sensible loads, the cooling-reheat-based dehumidification system requires 57.2 percent less energy than the desiccant-based system. This clearly demonstrates that the cooling-reheat-based system is more economical in achieving and maintaining these desired design conditions.
 |
As the above comparison shows, at equal design conditions and equal latent and sensible loads, the desiccant-based dehumidification system requires 32.8 percent less energy than the cooling-reheat-based system. This indicates that the desiccant-based system is more economical in achieving and maintaining these desired design conditions. In addition, the cooling-reheat-based system requires a low saturated suction temperature that may be difficult to maintain to avoid icing of the evaporator coil.
 |
As the above comparison shows, at equal design conditions and equal latent and sensible loads, the desiccant-based dehumidification system requires 65.4 percent less energy than the cooling-reheat-based system. This clearly indicates that the desiccant-based system is more economical in achieving and maintaining these desired design conditions. In addition, the cooling-reheat-based system requires a saturated suction temperature that will be almost impossible to maintain to avoid icing of the evaporator coil.
Case Study 1
Desiccant-based dehumidification system (see Figure 1).
Design conditions of the space: 72 degrees F db @ 55 percent RH.
Internal sensible load: 60,943 BTUH
Internal latent load: 157,476 grains/hour
Energy required
Supply fan motor, 3.11″ S.P. 2.14 kW
Reactivation fan motor, 3.49″ S.P. 0.27 kW
Reactivation heater 24.00 kW
DX after cooling compressor, 45 degrees F SST 11.68 kW
Air-cooled condenser fan motor, 95 degrees F O.A. 2.82 kW
Total energy 40.91 kW
Cooling-reheat-based dehumidification system (see Figure 2).
Design conditions of the space: 72 degrees F db @ 55 percent RH.
Internal sensible load: 60,943 BTUH
Internal latent load: 157,476 grains/hour
Energy required
Supply fan motor, 2.75″ S.P. 1.90 kW
DX cooling compressor, 45 degrees F SST 10.41 kW
Air-cooled condenser fan motor, 95 degrees F O.A. 2.82 kW
Reheated heater 2.49 kW
Total energy 17.53 kW
Case Study 2
Desiccant-based dehumidification system (see Figure 3).
Design conditions of the space: 72 degrees F db @ 40 percent RH.
Internal sensible load: 60,943 BTUH
Internal latent load: 218,718 grains/hour
Energy required
Supply fan motor, 3.22″ S.P. 2.22 kW
Reactivation fan motor, 3.64″ S.P. 0.43 kW
Reactivation heater 36.50 kW
DX after cooling compressor, 45 degrees F SST 15.51 kW
Air-cooled condenser fan motor, 95 degrees F O.A. 2.82 kW
Total energy 57.48 kW
Cooling-reheat-based dehumidification system (see Figure 4).
Design conditions of the space: 72 degrees F db @ 40 percent RH.
Internal sensible load: 60,943 BTUH
Internal latent load: 218,728 grains/hour
Energy required
Supply fan motor, 2.94″ S.P. 4.63 kW
DX cooling compressor, 38 degrees F SST 29.75 kW
Air-cooled condenser fan motor, 95 degrees F O.A. 4.94 kW
Re-heat heater 46.20 kW
Total energy 85.52 kW
Case Study 3
Desiccant-based dehumidification system (see Figure 5).
Design conditions of the space: 72 degrees F db @ 30 percent RH.
Internal sensible load: 60,943 BTUH
Internal latent load: 268,834 grains/hour
Energy required
Supply fan motor, 3.10″ S.P. 2.14 kW
Reactivation fan motor, 3.62″ S.P. 0.62 kW
Reactivation heater 53.50 kW
DX after cooling compressor, 45 degrees F SST 17.50 kW
Air-cooled condenser fan motor, 95 degrees F O.A. 2.82 kW
Total energy 76.58 kW
Cooling-reheat-based dehumidification system (see Figure 6).
Design conditions of the space: 72 degrees F db @ 30 percent RH.
Internal sensible load: 60,943 BTUH
Internal latent load: 268,834 grains/hour
Energy required
Supply fan motor, 3.21″ S.P. 9.47 kW
DX cooling compressor, 33 degrees F SST 66.23 kW
Air-cooled condenser fan motor, 95 degrees F O.A. 7.98 kW
Re-heat heater 137.50 kW
Total energy 221.18 kW
All of the comparisons made here are for illustration purposes only, and in no way are they meant to be definitive or misleading. As part of any analysis of the best approach to use in a particular case, a cost of operation evaluation is desirable. Since the specific operation costs of any of the systems is a function of the availability of energy sources, the relative energy costs, the ambient conditions, the seasonal weather variations, the operating range in the controlled areas, and the load producing elements of this analysis must be specific by application, and should be done by someone who is able and experienced in the dehumidification disciplines.
Conclusion
We can make several general conclusions based on our theoretical work. If refrigeration will do the job and provide adequate control without the need for reheat or defrost, it will generally be the most economical approach. This condition will be generally true for high-moisture loads in space conditions that do not have to be below 75 degrees F, 45 percent RH.
Desiccant-based systems are more economical than refrigeration systems at lower temperatures and lower moisture levels. Typically, a desiccant dehumidification system is utilized for applications below 45 percent RH down to less than 1 percent RH. Thus, in many applications, a DX or chilled water pre-cooling coil is mounted directly at the air inlet of the desiccant dehumidifier. This design allows for removal of much of the initial heat and moisture prior to entering the desiccant system where the moisture level is reduced even further. If cool, dry air is required in the conditioned space, an after-cooling coil may be required at the air outlet of the desiccant system.
Both desiccant-based and cooling-based dehumidification systems work most efficiently when used together. The advantages of each compensate for their individual limitations. If outside air is to be introduced, or if the internal moisture load is high, a hybrid combination of these first two systems will allow the most economical elements of each system to be used in the ranges of application where they are best suited. A supplier specializing in the “complete system approach” may build the combination into one “packaged” system.
The difference in the costs of electrical power and thermal energy (i.e., gas or steam) will determine the ideal mix of desiccant-based to cooling-based dehumidification in a given application. If thermal energy is inexpensive and electric power costs are high, a desiccant-based system may be the most economical to remove the bulk of the moisture from the air. If electrical power is inexpensive and thermal energy is costly or not available at all, a cooling-based system is probably the most efficient choice.
For systems that are essentially low temperature refrigeration systems for the maintenance of spaces at temperatures below freezing and requiring continual operation, (and therefore requiring continual defrosting of coils), a desiccant dehumidifier can be added to the system in place of a spray brine system and make the refrigeration system more efficient by allowing a higher suction temperature.
In some cases, the use of a desiccant-based system can reduce the operating costs of the existing cooling-based system. For example: When treating ventilation air in building HVAC systems, the dehumidification of the fresh air with the desiccant-based system decreases the installed cost of the cooling system and eliminates deep coils with high air and liquid-side pressure drops. This saves considerable fan and pump energy as well.
Guy A. Griesse is vice president of marketing at Bry-Air Inc. (Sunbury, OH), a manufacturer of industrial drying equipment. He is a member of the Society of Plastics Industry (SPI), the Society of Plastics Engineers (SPE) and the American Society of Heating, Refrigeration and Air Conditioning Engineers (ASHRAE).
Kenneth E. Baker is director of engineering at Bry-Air, Inc. He has over 25 years` experience in design, calculation and selling of desiccant dehumidifiers for industry.
References:
1. ASHRAE: 1989 ASHRAE Handbook of Fundamentals (American Society of Heating, Refrigeration and Air Conditioning Engineers Inc., Atlanta, GA, 1989).
2. American Air Filter: American Air Filter extended surface pleated panel filters. Bulletin No. RFP-1-165-G-April-91 (Snyder General Corp., Louisville, KY, 1991).
3. Bry-Air Inc.: Applications Engineering Manual (Bry-Air Inc., Sunbury, OH, 1989).
4. Chicago Blower Corp.: Chicago Blower Engineering Guide (Chicago Blower Corp., Glendale Heights, IL).
5. Coils Plus Inc.: Coils Plus Coil Sizing Program Performance, Ver. 4.0 (Coils Plus Inc., Longview, TX ,1990).
6. Carrier Corp.: Carrier 1994/1995 Commercial Products and Systems catalog (Carrier Corp., Syracuse, NY, 1994).
7. Guy A. Griesse, ed.: Bry-Air Application Update 94-3 (Bry-Air Inc., Sunbury, OH, 1994).
8. L. Harridan, ed.: The Dehumidification Handbook, second edition (Munters Cargoaire, Amesbury, MA, 1990).
Mechanical refrigeration
The principle of condensation is a very simple concept. If air is chilled below its dewpoint temperature, moisture will condense on the cold surface. In other words, the air becomes dehumidified through the process of cooling and condensation. The amount of moisture removed depends upon how cold the air can be cooled. A rule of thumb is the lower the air temperature, the drier the air.
A typical mechanical refrigeration system cools the air, removes some of the moisture as condensate, and sends the cooler, drier air to the conditioned space.
However, cooling the air just to dry it may not be practical for many applications. An exception might be when cool air is required along with some limited dehumidification, which satisfies the desired conditions. Normally, this method is reserved for applications in which outdoor air is being dried to levels only slightly lower than the incoming ambient or system air, and separate control of temperature and humidity is not a strict requirement.
For a large reduction of the moisture content of the air by cooling, over-cooling and subsequent reheating may be required. But such procedures typically have problems with operation and maintenance, as well as cycle and control; this method can be costly for producing low dewpoint air. Another limitation to this technique is the freezing point of water. When air is dried via mechanical refrigeration, the cooling surfaces of the coils may reach sub-freezing temperatures. This causes ice to form, which, in turn, reduces the efficiency of the cooling system, so anti-icing devices or tandem coils and defrost cycles may be required.
Abbreviations:
BTUH=British thermal unit/hour
BHP= Brake horse power
CFM=cubic feet/minute
db = dry bulb
DX= direct expansion
° Fahrenheit = degrees Fahrenheit
gr=grains/pound of dry air
kW=kilowatts
O.A.= outside air
RH = relative humidity
S.P. = Static pressure
SST= Saturated suction temperature
“U” = coefficient of transmission
wb = Wet bulb
WC = Water column