Accurately calculating water vapor loads in industrial environments helps size and select systems with minimal operating costs.
The need to control the
amount of water vapor in
the air is felt in all industrial,
commercial, and institutional
facilities. Humidity control
is important to human health and
comfort. Humidity control also
improves the reliability of equipment,
production processes, and
materials by controlling static
electricity, corrosion, and other
factors. The purpose of this article
is to provide information on energy
usage, energy calculations,
and operating energy costs to help
engineers evaluate industrial dehumidification
systems. I use
mass-flow analysis, adiabatic
mixing, and thermodynamics to
evaluate the air- and water-vapor
mixture as it travels through each
component of the dehumidification
system. This procedure produces
an in-depth analysis of the
sensible and latent heat energy
flows. This analysis is also meant
to help the engineer with the concepts
of heat of condensation, reactivation heat, heat dump back,
and desiccant heat.
Some of the types of equipment
and their moisture-removal capabilities
are illustrated in Table 1.
Moisture can be removed from the
air by cooling it below the dewpoint
temperature so condensation
occurs by air-to-air heat exchangers,
which bring in dryer outside
air, or by chemical methods.
Chemical dehumidification is carried
out through the use of sorbent
materials, which are solids or liquids
that can extract moisture
from the air and hold it. There are
two classifications of sorbents:
• Adsorbents—which do not
experience a phase change. Moisture
is deposited on the surface of
the dry desiccant. Most adsorbents
are solids.
• Absorbents—which change
physically, chemically, or both
during the sorption process. Most
absorbents are liquids or solids
that become liquid as
they absorb
moisture.
Portable Dehumidifiers
When most people think of portable
dehumidifiers, they think of
equipment that is used in the
home, but this is no longer true.
These systems are being used in
many commercial and industrial
applications. A few examples of
their uses are: indoor pools, cleaning
and restoration, locker rooms,
pump stations, libraries, restaurants
and bars, film and tape storage,
bakeries, well houses, and
canning plants.
The types of portable equipment
available are direct
expansion (DX) reheat,
dry desiccant, and air-toair
heat exchangers—to
name a few. The water-removal
capacities of these
systems at ANSI B149.1
inlet air conditions (80 F
and 60 percent RH) are as
follows:
• cDX reheat: 0.65 to 4.35
lb per hr; 1.88 and 12.50
gal per day, respectively.
• Dry desiccant: 6.26 lb
per hr; 18 gal per day.
• Air changers: 15.30 lb
per hr; 44 gal per day.
(Building air is exhausted
through an air-to-air heat
exchanger that brings in
dry outside air. Removal
capacity is based on outside
air at 0 F and 60 percent
RH.)
Table 2 is a summary of
more than 30 portable dehumidifiers
reviewed for
this article. The table illustrates
both the waterremoval
capacity and the
operating cost of each system.
Notice that the energy-
inefficient systems
use up to four times more
energy than the efficient
systems. Table 3 presents
the variation of water-vapor
removal and the operating
cost at various inlet
air conditions. This table
was based on one of the energy-
efficient models.
Some DX-reheat units will
begin to form frost on the cooling
coils when the inlet air dry bulb
goes below 60 to 65 F. If the unit
has frost control, it will experience
significant capacity loss due
to the frost-control operation. Two
typical approaches to frost control
are listed below:
• A temperature-sensing thermostat
diverts the hot refrigerant
gas through the evaporator coil
until the ice is melted.
• An automatic de-ice sensor
shuts the compressor off when
evaporator-coil temperature approaches
freezing. The fan continues
moving warm air
across the
coil to defrost it.
Large Industrial Dehumidifiers
This section of the article is a
detailed review of three common
types of industrial dehumidifiers:
direct expansion, dry desiccant,
and liquid desiccant. Each system
has an air intake of 30,000 cu ft
per min at 70 F and 56 grains of
water vapor per lb of dry air. Also,
each system was required to dry
the air down to 35 or 36 grains of
water vapor per lb of dry air. I
worked with several dehumidification
companies to size the systems
and determine all energy
consumption. Using the brake
horsepower provided by the companies,
I sized the motors and calculated
the electrical consumption
(KW per hr) illustrated in this
article. Operating costs were then
developed using $0.06 per KWH for
electricity and $5.00 per 106 Btu
for natural gas. The operating
costs were then developed into operating
cost per 1000 lb of water
vapor removed. In some cases, I
left out proprietary information
on the systems at the request of
the manufacturers.
Each industrial dehumidifier includes
detailed air- and water-vapor
mass-flow analysis, psychrometrics,
thermodynamics, and
adiabatic mixing (Figs. 1 to 3) to
help the readers understand the
energy flows. The psychrometrics
and thermodynamics used in the
diagrams follow the principals of
Zimmerman and Lavine. The airand
water-vapor flows are also illustrated
in acfm (actual cu ft per
min) and the dehumidification industry
dscfm (dry standard cu ft
per min). I used dscfm to avoid confusion
with the fan industry scfm.
Cooling-based Dehumidification
Moisture can be removed from
the air by cooling the air below its
dew-point temperature. This can
be achieved through the following
systems:
• Chilled water, glycol, or brine
coil system
• DX cooling coil system
• Chilled water air-washer system
The first two systems accomplish
dehumidification by passing
the air through a cooling coil with
a coil-surface temperature below
the dew point of the air. Water vapor condenses on the coil surfaces.
The amount of moisture removal
depends on how cold the air can be
chilled. The lower the temperature,
the drier the air. The chilled
water air-washer system also
cools the air below its dew point
by using water that is colder than
the dew-point temperature. The
water vapor in the air condenses
on the water spray or the nearest
surface. In this case, the use of
colder water results in greater
dehumidification.
The type of system chosen for illustration
in this article is the DX
cooling coil system. The basic components
of this mechanical refrigeration
system are an evaporator
coil, compressor, condenser, and
throttling valve (or expansion
valve). The system uses a refrigerant
that enters the evaporator coil
(cooling coil) in a liquid state. The
refrigerant evaporates inside the
coil and, in doing so, absorbs heat
from the process air moving
through the coil. It then leaves the
coil in the form of a gas. The compressor
takes the cold vapor from
the evaporator and compresses it
to a hot gas at high pressure. When
the refrigerant leaves the compressor,
it is still a gas but at a much
higher pressure (five to ten times
greater) and a much higher temperature.
The hot refrigerant gas
is then pushed through a condenser
(in this case, an air-cooled
condenser) where the hot gas is
cooled and condensed into a liquid
by some substance, usually air or
water. The refrigerant then flows
from the condenser as a high-pressure
liquid through the expansion
valve. As the liquid passes through
the valve, its pressure is suddenly
decreased to the pressure in the
evaporator coil. At the same time,
the temperature of the liquid refrigerant
drops down from the
warm condenser temperature to
the cold evaporator temperature.
This occurs because a small
amount of liquid suddenly flashes
to a vapor as it passes through the
restriction in the valve. Then the
liquid, with some bubbles of flash
vapor, enters the evaporator coil.
The liquid refrigerant in the coil
evaporates and, in doing so, absorbs
heat from the air passing
through a coil.
An example of a DX dehumidification
system is illustrated in Fig.
1. The system includes a heat-pipe
heat exchanger and an air-cooled
condenser system. The heat pipe
removes 538,616 Btu per hr (44.88
ton) of heat from the inlet process
air and passes it to the dehumidifier
process air leaving the system.
The advantage is reduced cooling
load and a leaving air condition
that is not at or near saturation.
The operating cost in Table 4 is
$20.03 per 1000 lb of water vapor
removed at a process-air-leaving
condition of 56.3 F and 36 grains
per lb (53.9 percent RH). Dehumidification
reheat systems will use
the high-pressure, high- temperature
gas leaving the compressor in
a coil system to reheat the process
air leaving the dehumidifier. Most
common residential dehumidifiers
use this configuration. It is important
for designers to consider
waste energy usage, heat wheels,
or heat pipes for reheat because reheat
can add a significant "added"
cost to the operating cost of a DX
system.
Rotating Dry Desiccant Wheel
Dry desiccants are adsorbent
materials that attract moisture because
of the electrical field at the
desiccant surface. The field attracts
water molecules that have a net opposite
charge. Some of the solid adsorbents
used in dry desiccant systems
are illustrated in Table 5.
Sorption is the adsorption process
by which a desiccant removes
water vapor directly from the air. The ability of an adsorbent to attract
moisture depends on the difference
in vapor pressure between
the desiccant surface and air. The
vapor-pressure difference drives
moisture from the high vaporpressure
area to the low vaporpressure
area. Dry desiccants typically
have low vapor pressure at
their surface and, therefore, adsorb
moisture from the air. When
moisture is removed from the process
air stream, it produces heat
of sorption (or heat of adsorption),
which is composed of latent heat
of condensation of the removed
moisture plus additional chemical
heat. The heat of sorption of the
moisture removed from the air is
converted to sensible heat. The
amount of heat released is usually
around 1080 Btu per lb WV removed
to 1312 Btu per lb WV removed.
The actual amount depends
on the type of desiccant.
The heat of sorption (sensible
heat) is energy that is passed to
the process air stream, which
raises the discharge air temperature
of the process air stream.
As the moisture content of the
desiccant rises, so does the watervapor
pressure at the desiccant
surface. At some point, the vapor
pressure at the desiccant surface
will be the same as the air, and
moisture adsorption will end. The
desiccant is then taken out of the
process air stream and is placed
into the reactivation air stream (a
scavenger air stream consisting of
outside air or building air). The
reactivation air stream is typically
heated to a temperature of
190 to 375 F. The combination of
the heat and moisture raises the
vapor pressure at the desiccant
surface. When the surface vapor
pressure exceeds the vapor pressure
of the reactivation air, moisture
leaves the desiccant. This
process is called reactivation.
The reactivation section, which
constitutes less than half of the
desiccant wheel, uses flexible
seals to seal it from the adsorption
or process side to minimize cross
contamination. The typical leakage
rate is 1 to 2 percent.
Following reactivation, the hot
desiccant rotates back into the
process air where the process air
cools the desiccant which lowers
the desiccant vapor pressure, so it
can collect more moisture from
the balance of the process air
stream. Some of the equations
used by the rotating dry desiccant
manufacturers are shown in the
accompanying sidebar.
A typical layout of a rotary dry
desiccant system is illustrated in
Fig. 2. The operating costs of the
system are listed in Table 6. The
operating cost for the example in
Fig. 1 is $15.87 per 1000 lb of water
vapor removed with a leaving
air condition of 93.6 F and 35
grains per lb. If the air is too hot
and has to be cooled to 55 F and 35
grains per lb, the operating cost
increases to $34.24 per 1000 lb
WV. Reactivation heat represents
86 percent of the operating cost
for the direct-fired unit and 88
percent for the indirect-fired unit.
This represents a great opportunity
for the designer to use lowcost
hot water from cogeneration;
the use of low-cost steam; or condensate,
refrigeration reject heat,
or waste exhaust to preheat the
reactivation air. These options
could significantly reduce the operating
cost. Keep in mind that in
some cases it may be more economical
to combine cooling and
desiccant dehumidification. The
technologies do complement each
other since the refrigeration condenser
reject heat from the cooling
process can be used to preheat
the reactivation air.
In process-drying applications,
dry desiccant dehumidifiers are
sometimes used without added
cooling because the increase in
temperature caused by the heat of
adsorption is helpful in the drying
process. However, in some applications,
a provision must be made to
remove the excess sensible heat
from the process air after dehumidification.
For this reason,
Table 6 provides an added cost section
for cooling the air to 55 F as an
example of the possible added cost.
Liquid Desiccant Dehumidifier
Liquid desiccant dehumidification
operates on the principal of
chemical absorption of water vapor
from the air. The absorbent or
desiccant solution will change
physically, chemically, or both
during the sorption process. Some
of the liquid desiccant solutions
used for dehumidification are:
• Lithium chloride (LiCl)
• Lithium bromide (LiBr)
• Calcium chloride (CaCl2)
• Triethylene glycol (TEG)
• Propylene glycol
Liquid absorption dehumidification
is very similar to a chilled
water air-washer system. When
the air passes through the
washer, its dew point approaches
the temperature of the water supplied.
Air that is more humid is
dehumidified, and air that is less
humid is humidified. In a similar
manner, the liquid absorption dehumidifier
sprays the air with a
desiccant solution that has a
lower vapor pressure than the vapor
pressure of the entering process
air stream. The liquid has a
vapor pressure lower than water
at the same temperature, and the
air passing over the solution approaches
this reduced vapor pressure.
The ability to remove water
vapor (or add water vapor) is determined
by the temperature and
concentration of the solution. The
conditioner can be adjusted so
that the conditioner delivers air at
the desired relative humidity.
The vapor pressure of a given
concentration of absorbent solution
approximates the vapor-pressure
values of a fixed relative humidity
line on a psychrometric
chart. For instance, a 40 percent
concentration of lithium chloride
closely approximates the 20 percent
relative humidity line. Also,
a 15 percent concentration is very
close to the 80 percent relative humidity
line. Therefore, it can be
said that higher solution concentrations
give lower equilibrium
relative humidity and thus allow
the absorbent to dry air to lower
levels. Temperature also affects
the absorbents' ability to remove
moisture. For instance, a 25 percent
solution lithium chloride has
a vapor pressure of 0.37 in Hg at
70 F (same as air at 70 F and 50
percent RH). When the solution is
heated to 100 F, the vapor pressure
climbs to 0.99 in Hg. Therefore,
the warmer the desiccant,
the less moisture it can absorb.
Also, if the solution vapor pressure
is higher than the surrounding
air, the water vapor will
transfer to the air and dry the desiccant
solution.
A typical system diagram is illustrated
in Fig. 3. In the operation,
warm, moist air is sprayed
with a solution of chilled lithium
chloride, which was cooled with
chilled water in a plate-and-frame
heat exchanger. The air is cooled
and dehumidified by heat and
mass transfer to the lithium chloride
solution. A chiller with an
air-cooled condenser section provides
the chilled water to cool the
lithium chloride solution. If the
desired dehumidified air-moisture
content is 50 grains per lb (or
48 F dew point), the water used to
cool the desiccant can be 85 F cooling-tower water rather than
chilled water. Consult Table 1 for
additional information.
When moisture is removed from
the air, the reaction liberates
heat. This is the reverse of evaporation,
when heat is consumed by
the reaction. The heat that is generated
is the latent heat of condensation
of the water vapor plus
the heat of solution (or the heat of
mixing of the water and desiccant). In desiccant dehumidification,
this heat (approximately
1080 to 1320 Btu per lb water vapor
removed) is transformed to
the air, raising the air dry bulb
temperature, and therefore, increasing
the load on the chilled
water system. The chilled water
system must be sized to remove
the latent heat of condensation
(heat of sorption), the air sensible
heat, and the residual heat
load added by the regeneration
process.
To remove the water extracted
from the air and keep the liquid
desiccant at a fixed concentration,
a small percentage of the conditioner-
desiccant pump flow (typically
around 15 percent) is transferred
to the regeneration system.
The weak desiccant solution is
pumped up to a heating system
(plate-and-frame heat exchanger),
which raises the temperature
and vapor pressure of
the liquid desiccant. The hot desiccant
is then sprayed at a scavenger
air stream (outside air or
building air) with a lower vapor
pressure that forces the water vapor
out of the desiccant and into
this air, which is exhausted outside.
The dry desiccant returns to
the regenerator sump. The desiccant
is still a little warm, and its
vapor pressure is still a little
high—until it flows back to the
conditioner and is cooled by the
chilled water heat exchanger.
Therefore, the cooling system
must be sized to include this
residual heat load added by the
regeneration process (sometimes
called heat dump back). The
amount of heat and dump back is
typically in the range of 50 to 350
Btu per lb water vapor removed.
Table 7 is a review of the operating
costs associated with the
system in Fig. 3. The total operating
cost is $28.43 per 1000 lb of
water vapor removed from the
process air. You will note that the
natural gas cost is 37 percent of
the total cost. Therefore, it pays to
find a source of waste heat to reduce
these costs. Some possible
sources are condenser-rejected
heat, solar heat, or a natural gas
engine-driven chiller, which produces
hot water (engine heat) as
well as chilled water at a reduced
cost. The operating cost of a liquid
desiccant dehumidification system
combined with a natural gas
engine-driven chiller is $19.85 per
1000 lb of water vapor removed.
This combination represents a 30
percent reduction in operating
energy costs.
Another option to save energy on
liquid desiccant systems is to install
a liquid-to-liquid-type heat
exchanger (some call this an interchanger)
placed between the warm
desiccant leaving the regenerator
and the cool desiccant entering the
regenerator. By doing so, less energy
is needed to regenerate the
desiccant because it is warmer
than when it left the regenerator.
The heat exchanger will typically
reduce the heat dump back to the
conditioner-cooling consumption
by about 65 percent and reduces
the regenerator heat consumption
by about 15 percent.
Conclusion
Table 8 is a summary of the operating
costs for each of the large,
industrial dehumidifiers using
the electricity cost of $0.06 per
KWH and the natural gas cost of
$5.00 per 106 Btu. One can see
that the dry desiccant systems
have the lowest energy operating
cost of all the systems, but they
also have the highest discharge
air temperatures.
If elevated process air temperature
is not acceptable, the liquid
desiccant or DX with heat pipe
would be the choice (at 55 to 56 F
discharge air temperature).
The outcome of this study will
change, depending upon the actual
energy costs for that area. In
the United States, natural gas
costs vary from $2.40 to $7.90 per
106 Btu and electricity goes from
$0.018 to $0.15 per KWH. What
this says is that every system
must be evaluated based on the
energy costs for that region. Also,
the evaluation of operating costs
should include maintenance and
capital amortization.
Each system is unique, and
they all offer ways to reduce energy
operating costs. For example,
the liquid desiccant system operating
cost dropped 30 percent
(from $28.43 to $19.85 per1000 lb
WV removed) by incorporating a
natural gas engine-driven chiller
and using the hot water jacket
heat to preheat the regenerator
system. Also if the required dew
point is 48 F (49.68 grains per lb)
or above, cooling-tower water can
be used in the summer. The DX
systems can use heat pipes, heat
wheels, and/or heat reject from
the process cooling for reheat. Using
purchased energy for DX reheat
can raise the operating costs
significantly. The major operating
cost for dry desiccant systems is
the reactivation heat that represents
86 percent of the total operating
cost for the system in Table
6. Using alternative low-cost energy
sources can reduce the operating
costs significantly. DX system
condenser reject heat could be
used to preheat the reactivation
air as well as desuperheater coil
heat. For this reason, the most
economical system may be a combination
DX system and desiccant
(solid or liquid) system. The technologies
do complement each
other since the refrigeration condenser
reject heat from the air
cooling process can be used to prevent
the air entering the reactivation
or regeneration section of the
desiccant dehumidifier. In industrial
plants, there are many
sources of cheap, low-grade heat
(low temperature) that can be
used in the reactivation or regeneration
system. It is up to the engineer
involved in the equipment
selection to consider these sources
before a decision can be made on
the type of system.
I would like to thank the following companies
for assistance in the preparation
of this article. Without their assistance,
it would not have been possible:
Kathabar, Des Champs Laboratories,
Munters Cargocaire, Bry-Air, Inc.,
Desert Aire, Therma-Stor Products,
Drieaz, Dectron, Inc. I would also like
to thank Craig Pekarek for the CAD
work and Jodi Cavil for the typesetting
of this article.
Bibliography (Part 1)
1) "Water Vapor Migration and
Condensation Control in Buildings,"
by W. Acker, Heating/Piping/Air
Conditioning Magazine, June 1998.
2) Zimmerman, O.T., and I. Lavine,
Industrial Research Services: Psychrometric
Tables and Charts, 2nd
Ed., Industrial Research Services,
Inc., Dover, New Hampshire, 1964.
Industrial Dehumidification Part 2:
Air Flow Diagrams & Water Vapor Load Equations
Air Flow Diagrams &
Water Vapor Load Equations. Accurate water vapor load
equations for industrial
dehumidification systems
design are difficult to find,
and terminology is not
standard. This article
provides a thorough
review of both.
In designing dehumidification systems, one of the
most important tasks is to quantify the water vapor
loads that must be removed by the system. Two
qualified individuals may arrive at different total
moisture loads for the same building. Some of these
differences occur from abbreviated or approximate
equations that were developed to make the calculations
easier. However, these approximate equations lose
accuracy when air temperatures are higher or lower than
the conditions assumed when the approximate equations
were developed.
This article will clarify some of these
equations and present alternative equations that are
more precise across a larger range of
conditions.
This article will also explain the air
flows used in proposals from dehumidification
companies as well as the
proposed flow diagrams from dehumidification
equipment suppliers.
I have discovered a great amount of
uncertainty over the air flows in these
diagrams.
Dehumidification Industry Equations and Terms
Dehumidification manufacturers
like to develop air flow diagrams of their entire systems. Mass flow analysis is used in these
diagrams because mass flow does not change if there is a
temperature or pressure change. Mass flows can also be
added and subtracted. The actual air flows or acfm (cu ft
per min) will change if the temperature or pressure
changes. Also, acfm values cannot be added or subtracted
because they are at different air densities. The
dehumidification industry chose a type of mass flow
analysis that is flow in dry standard cu ft per min (dscfm)
at a common air density of 0.075 lb dry air per dry standard
cu ft. The acfm flows are converted to these flows
for illustration in the air system diagrams (shown in Fig.
1 and Fig. 2).
As mentioned earlier, dscfm air flow is a flow at a common
air density of 0.075 lb dry air per dry standard cu ft. I prefer to use the term dscfm to avoid any possible confusion
with acfm or the fan industry term scfm. In Fig. 1,
one manufacturer uses cfm to represent the dry standard
air flow. When you look at this diagram, it is easy to tell
that the flows are dry standard air flows and not acfm. If
you look at the flow in and out of the dehumidifier, the
flows are both 7500 cfm. This is not possible with acfm
because this flow is made of both air and water vapor;
therefore, there is a loss (of water vapor or cfm) as it travels
through the dehumidifier. Listed below is an example
of an acfm flow broken down into dry air flow and water
vapor flow:
Air pressure = 29.921 in.
Hg
Air temperature = 70 F
Relative humidity = 100 percent
Total air flow = 100,000 acfm
Dry air flow = 97,530 cfm
Water vapor flow = 2470 cfm
Also, note that there is a change in temperature across
the dehumidifier that would also cause a change in the
acfm flow.
Fig. 2 is a diagram prepared by another dehumidifier
manufacturer. The air flows are in dscfm, but in this case,
they use the term scfm. This scfm should not be confused
with the fan industry scfm, which is a type of mass flow
that represents the combined air and water vapor flow. By
looking at Fig. 2, you can see that the flow into the dehumidifier
(or conditioner) 20,833 scfm is the same as the
flow leaving; therefore, this is a dry air mass flow. If the
entering flow is fan industry 20,833 scfm, the leaving flow
would have dropped to 20,754 scfm due to the removal of
water vapor (79 scfm or 355.41 lb of water vapor per hr).
The next section on fan industry scfm will help you to understand
the difference between fan and dehumidification
scfm terms.
The equations used by the dehumidification industry
to convert the actual air flows to dry standard air flows are
listed below:
Fan Industry Equations and Terms
Fans must be selected based on the actual flow rate and
the actual density at the inlet to the fan. Some fan manufacturers
prefer to specify the flow rate based on standard
inlet conditions. Fan performance curves are developed (from a series of laboratory tests) at these standard conditions.
Listed below are the equations that allow designers
to convert from the actual conditions to the standard
conditions:
acfm (cu ft per min) =
Actual cubic feet per minute. It represents the volume
of dry gas and water vapor flowing at a specified point in a
system. In fan sizing, this would be the flow entering the
fan.
Density of moist gas (lb total per cu ft) =
The ratio of the mass of a substance to its volume. The
fan gas density or fan air density is the total density at the
fan inlet.
Standard density (lb total per std cu ft) =
Some fan manufacturers like to develop fan performance
curves based on a standard gas or air density of
0.075 lb total per std cu ft. The Air Movement and Control
Association (AMCA) indicates that this density is
substantially equivalent to air at a temperature of 68 F,
50 percent RH, and a pressure of 29.92 in. of mercury.
scfm (std cu ft per min) =
The standard gas or air flow rate entering the fan at an
inlet density of 0.075 lb total per std cu ft.
MW wet mix (lbm per mole) =
The molecular weight of the dry gas (or dry air) and
water vapor mixture.
P (lbf per sq ft) =
The total pressure at the inlet to the fan. The pressure
represents the barometric pressure plus the static pressure
(or gauge pressure).
1545.43 ft - lbf per mole - °R =Universal gas constant.
T(°R) =
Absolute gas or air temperature at the inlet to the fan
in degrees Rankine (° R = ° F + 459.67).
DSCFM Flow Versus SCFM Flow
The following examples show how the dehumidification
industry dscfm flow compares to the fan industry
scfm flow.
Example 1:
Air pressure = 29.921 in. Hg
Air temperature = 70 F
Relative humidity = 100 percent
Humidity ratio = 109.93 grains WV per lb dry air
Moist air density = 0.074190 lb wet air per cu ft wet air
Specific volume = 13.6906 cu ft wet air per lb dry air
acfm flow = 40,000 cu ft wet air per min
Dehumidification industry dscfm = 38,956 dry std cu ft
per min
Fan industry scfm = 39,568 std cu ft per min
Example 2:
Air pressure = 29.921 in. Hg
Air temperature = 200 F
Relative humidity = 50 percent
Humidity ratio = 2809 grains WV per lb dry air
Moist air density = 0.051216 lb wet air per cu ft wet air
Specific volume = 27.3599 cu ft wet air per lb dry air
acfm flow = 40,000 cu ft wet air per min
Dehumidification industry dscfm = 19,493 dry std cu ft per min
Fan industry scfm = 27,315 std cu ft per min
Example 1 shows that at 70 F, the dscfm flow is very
close to the fan industry scfm flow (only 1.6 percent variation).
However, at more elevated temperatures, such as
in Example 2, the fan industry scfm flow is 40 percent
higher than the dscfm flow.
Water Vapor Loads Due to Air Flow
Water vapor loads on industrial buildings come from
many sources. Listed below are a few of these sources of
water vapor:
People
Permeation through walls, roofs, and floors
Moisture from products and packaging materials
Evaporation from open tanks or wet surfaces
Product dryer leakage
Open combustion
From air flow
• Air leakage through cracks and holes
• Air leakage through conveyor openings
• Intermittent door openings
• Building-to-building air infiltration
• Makeup air
In many cases, water vapor loads by air flow are a major
contributor to the total building vapor load. In the
research work for this article, I came across a number of
approximate equations that are used to calculate the water
vapor load from air flow. Approximate equations
can be fairly accurate as long as the air conditions are
close to 70 F air temperature.
Accurate water vapor load equations for industrial dehumidification
systems design are difficult to find, and terminology is not standard.
This article provides a thorough review of both.
The first article in this
series, which covered air flow, can be read in
HPAC Engineering's May 1999 issue.
WATER VAPOR LOADS DUE TO AIR FLOW
Water vapor loads on industrial buildings come from
many sources. Listed below are a few of these sources of
water vapor:
❒ People
❒ Permeation through walls, roofs, and floors
❒ Moisture from products and packaging materials
❒ Evaporation from open tanks or wet surfaces
❒ Product dryer leakage
❒ Open combustion
❒ Air flow
• Air leakage through cracks and holes
• Air leakage through conveyor openings
• Intermittent door openings
• Building-to-building air infiltration
• Makeup air
In many cases, water vapor loads by air flow are a major
contributor to the total building vapor load. In the research
work for this article, I came across a number of approximate
equations that are used to calculate the water vapor load
from air flow. Approximate equations can be fairly accurate
as long as the air conditions are close to 70 F air temperature.
For more information on water vapor permeation loads,
consult the June 1998 issue of HPAC Engineering. The next
few sections will compare approximate and exact equations
for selected water vapor sources.
MOISTURE FROM AIR LEAKAGE
Equations (8) to (12) can be used to calculate moisture for
air leakage through cracks, holes, and conveyor openings.
Equations (8) to (11) were taken from engineering books or
from manuals prepared by dehumidification companies. Notice
that the engineering units do not properly cancel out,
which is why they are considered approximate equations.
Equations (11) and (12) are applied and compared in the
example below to show how results from approximate and
exact equations can vary under different conditions. Equation
(11), which is approximate, calculates a water vapor
load 20.18 percent over the exact equation (12). Equation
(11) is more accurate if the entering air flow is close to 70 F.
Engineers preferring the exact equation will need a psychrometric
chart to obtain the entering air specific volume, or a
psychrometric computer program that can calculate the air
mixture properties.
Example conditions:
Room conditions
Air pressure: 29.921 in. Hg
Dry bulb temp: 70 F
Moisture level: 35 grains WV/lb dry air
Relative humidity: 32.38 percent Entering air flow conditions
Air pressure: 29.921 in. Hg
Dry bulb temp: 120 F
Moisture level: 420 grains WV/lb dry air
Relative humidity: 76.44 percent
Specific volume: 16.02379 cu ft wet air/lb dry air
Air flow acfm: 200 cu ft per min
Equations (13) and (14). In this example, equations (13)
and (14) produced water vapor loads that were 7.27 and
12.63 percent above the exact equation (12) water vapor
load. The error is a direct result of the assumed entering air
specific volume. Note that the makeup air specific volume
will vary with the entering air psychrometric properties.
Therefore, you cannot select a standard value for specific
volume and expect the equation to be exact. For this reason,
equations (13) and (14) are approximate equations. Equations
(13) and (14), which can be found in many engineering
books and dehumidification manuals, will be fairly accurate
as long as the entering makeup air is close to the
selected specific air volume.
Example Conditions
Inside room conditions
Air pressure: 29.921 in. Hg
Dry bulb temperature: 70 F
Moisture level: 35 grains WV/lb dry air
Relative humidity: 32.38 percent Entering makeup air
Air pressure: 29.921 in. Hg
Dry bulb temperature: 100 F
Moisture level: 280 grains WV/lb dry air
Relative humidity: 93.67 percent
Specific volume: 15.01722 cu ft wet air/lb dry air
Air flow acfm: 2000 cu ft per min
Water Vapor Removed by Dehumidifiers
Dehumidification systems remove water vapor from the
process air that travels through the unit. This section looks
at the equations used to determine the humidity ratio of the
process air entering and leaving the unit as well as equations
used to estimate the amount of water vapor removed when
the inlet and discharge humidity ratios and air flow are
known. Note that some of the equations are the same as the
equations used in preceding sections. The equations listed as
approximate can be very accurate if the process air flow temperature
is close to 70 F (Table 1).
In this series of equations and calculations, approximate
equations (15) and (11) produced water vapor removals that
were 40 and 105 percent above the exact equations (16) and
(17). The approximate equations can be fairly accurate if
the air entering the dehumidifier is close to 70 F. Engineers
that desire greater accuracy can use the psychrometric chart
to get the specific volume needed to make the conversion
from acfm to dscfm or purchase psychrometric programs that
can calculate the value for them.
Conclusion
Many of the flow diagrams presented in articles, books,
engineering manuals, and proposals from dehumidification
companies do not indicate the engineering units for the
flows in the diagrams. As indicated in this article, the flows
are illustrated in cfm or scfm with no explanation. In most
cases, the flows are in dry standard cubic feet per minute.
Over the years, I have been contacted by many engineers
over the issue of calculated water vapor load variances from
different equations. In most cases, the approximate equations
are equations that have been shortened to make the
calculations easier for engineers. If you have any questions
on your equations, check the engineering units to make sure
that they properly cancel out to grains of water vapor per hr,
or lb of water vapor per hr.
The author would like to thank his wife, Sandra, for her patience and assistance
during the preparation of these articles. He would also like to
thank Nels Strand, the author's mentor and close friend for over 20
years.
Bibliography (Part 3)
1) Acker, W., Water Vapor Migration and Condensation
Control in Buildings, HPAC Engineering, June 1998.
2) Acker, W., Industrial Dehumidification Water Vapor
Load Calculations and System Descriptions, HPAC Engineering,
March 1999.
3) Clifford, G. E., Modern Heating and Ventilating Systems
Design, Prentice Hall, Englewood Cliffs, N.J., 1992.
4) Fan Engineering, 8th Ed., Edited by Robert Jorgensen,
Buffalo Forge, Buffalo, N.Y., 1983.
5) 1997 ASHRAE Handbook: Fundamentals, American
Society of Heating, Refrigerating, and Air-Conditioning
Engineers, Inc., Atlanta, Ga.