INDUSTRIAL DEHUMIDIFACATION By WILLIAM ACKER, President, Acker & Associates, Green Bay, Wis. T he need to control the amount of water vapor in the air is felt in all indus- trial, commercial, and institu- tional facilities. Humidity control is important to human health and comfort. Humidity control also improves the reliability of equip- ment, production processes, and materials by controlling static electricity, corrosion, and other factors. The purpose of this article is to provide information on en- ergy usage, energy calculations, and operating energy costs to help engineers evaluate industrial de- humidification 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 dehumidifica- tion system. This procedure pro- duces an in-depth analysis of the sensible and latent heat energy flows. This analysis is also meant to help the engineer with the con- cepts of heat of condensation, re- March 1999 HPAC Heating/Piping/AirConditioning 49 Industrial Dehumidification: Water Vapor Load Calculations and System Descriptions Accurately calculating water vapor loads in industrial environments helps size and select systems with minimal operating costs TABLE 1—Types of dehumidification systems. Common inlet Min. discharge air Min. discharge air moisture* Type air flows dew point temp. @ 29.921 Hg atm. pressure 1) DX cooling 50 to 50,000 cfm 42 F normal 39.45 35 F special design minimum (or lower) 29.92 2) Portable DX 150 to 330 cfm 40 F 36.48 23 F with defrost control & reduced capacity 18.19 3) Chilled water 500 to 50,000 cfm 40 F 36.48 37 F with 5 F approach on coil temp. 32.41 4) Chilled brine/glycol 500 to 50,000 cfm 32 F (below 32 F with frosted coil or 26.51 defrost control) 5) Liquid desiccant 750 to 84,000 cfm -80 F with LiCl 0.063 Typical LiCl: 15 F (with 45 F CW) 12.89 25 F (with 55 F CW) 19.78 32 F (with 65 F CW) 26.51 40 F (with 75 F CW) 36.48 48 F (with 85 F CW) 49.68 6) Dry desiccant rotating bed 500 to 20,000 cfm 5 F 8.231 7) Dry dsiccant multiple vertical bed 500 to 25,000 cfm -30 F 1.429 8) Dry desiccant rotating wheel 10 to 40,000 cfm -80 F 0.063 9) Portable desiccant 200 to 350 cfm 5 F 8.231 Note: CW = Chilled water or cooling tower water *Grains of water vapor per lb of dry air
Transcript
INDUSTRIAL DEHUMIDIFACATION
By WILLIAM ACKER,President,Acker & Associates,Green Bay, Wis.
The need to control theamount of water vapor inthe air is felt in all indus-
trial, commercial, and institu-tional facilities. Humidity controlis important to human health and
comfort. Humidity control alsoimproves the reliability of equip-ment, production processes, andmaterials by controlling staticelectricity, corrosion, and otherfactors. The purpose of this articleis to provide information on en-ergy usage, energy calculations,and operating energy costs to helpengineers evaluate industrial de-humidification systems. I use
mass-flow analysis, adiabaticmixing, and thermodynamics toevaluate the air- and water-vapormixture as it travels through eachcomponent of the dehumidifica-tion system. This procedure pro-duces an in-depth analysis of thesensible and latent heat energyflows. This analysis is also meantto help the engineer with the con-cepts of heat of condensation, re-
March 1999 HPAC Heating/Piping/AirConditioning 49
Industrial Dehumidification: Water Vapor Load Calculations and
System DescriptionsAccurately calculating water vapor loads in industrial environments helps
size and select systems with minimal operating costs
TABLE 1—Types of dehumidification systems.Common inlet Min. discharge air Min. discharge air moisture*
Type air flows dew point temp. @ 29.921 Hg atm. pressure
1) DX cooling 50 to 50,000 cfm 42 F normal 39.45 35 F special design minimum (or lower) 29.92
2) Portable DX 150 to 330 cfm 40 F 36.4823 F with defrost control & reduced capacity 18.19
3) Chilled water 500 to 50,000 cfm 40 F 36.4837 F with 5 F approach on coil temp. 32.41
4) Chilled brine/glycol 500 to 50,000 cfm 32 F (below 32 F with frosted coil or 26.51defrost control)
5) Liquid desiccant 750 to 84,000 cfm -80 F with LiCl 0.063Typical LiCl:15 F (with 45 F CW) 12.8925 F (with 55 F CW) 19.7832 F (with 65 F CW) 26.5140 F (with 75 F CW) 36.4848 F (with 85 F CW) 49.68
6) Dry desiccant rotating bed 500 to 20,000 cfm 5 F 8.231
7) Dry dsiccant multiple vertical bed 500 to 25,000 cfm -30 F 1.429
8) Dry desiccant rotating wheel 10 to 40,000 cfm -80 F 0.063
9) Portable desiccant 200 to 350 cfm 5 F 8.231
Note: CW = Chilled water or cooling tower water*Grains of water vapor per lb of dry air
Portable dehumidifiersWhen most people think of por-
table dehumidifiers, they think ofequipment that is used in thehome, but this is no longer true.These systems are being used inmany commercial and industrialapplications. A few examples oftheir uses are: indoor pools, clean-ing and restoration, locker rooms,pump stations, libraries, restau-rants and bars, film and tape stor-age, bakeries, well houses, andcanning plants.
The types of portable equip-ment available are directexpansion (DX) reheat,dry desiccant, and air-to-air heat exchangers—toname a few. The water-re-moval capacities of thesesystems at ANSI B149.1inlet air conditions (80 Fand 60 percent RH) are asfollows:
●DX reheat: 0.65 to 4.35lb per hr; 1.88 and 12.50gal per day, respectively.
● Dry desiccant: 6.26 lbper hr; 18 gal per day.
● Air changers: 15.30 lbper hr; 44 gal per day.(Building air is exhaustedthrough an air-to-air heatexchanger that brings indry outside air. Removalcapacity is based on out-side air at 0 F and 60 per-cent RH.)
Table 2 is a summary ofmore than 30 portable de-humidifiers reviewed forthis article. The table il-lustrates both the water-removal capacity and theoperating cost of each sys-tem. Notice that the en-ergy-inefficient systemsuse up to four times moreenergy than the efficientsystems. Table 3 presentsthe variation of water-va-por removal and the oper-ating cost at various inletair conditions. This tablewas based on one of the en-ergy-efficient models.Some DX-reheat units will
materials, which are solids or liq-uids that can extract moisturefrom the air and hold it. There aretwo classifications of sorbents:
● Adsorbents—which do notexperience a phase change. Mois-ture is deposited on the surface ofthe dry desiccant. Most adsor-bents are solids.
● Absorbents—which changephysically, chemically, or bothduring the sorption process. Mostabsorbents are liquids or solidsthat become liquid as they absorbmoisture.
Types of dehumidification systemsSome of the types of equipment
and their moisture-removal capa-bilities are illustrated in Table 1.Moisture can be removed from theair by cooling it below the dew-point temperature so condensationoccurs by air-to-air heat exchang-ers, which bring in dryer outsideair, or by chemical methods.Chemical dehumidification is car-ried out through the use of sorbent
50 HPAC Heating/Piping/AirConditioning March 1999
TABLE 2—Capacity of portable dehumidifiers Equipment Type Water removal1 Operating cost2
Note: 1 Water removal is at ANSI B149.1. Inlet air conditions of 80F and 60 percent RH relative humidity.2 Operating cost is based on an electric rate of $0.06 per kwh
TABLE 3—Capacity of portable dehumidifier at variousinlet air conditions
Tdb Tdp RH W* Removal Operationg cost
Note: This particular unit is a high capacity energy efficient unit. Tdb is dry bulb temperature. Tdp is dew point temperature. RH is relative humidity. Operating cost is based on an electric rate of $0.06 per kwh*Grains of water vapor per lb of dry air
begin to form frost on the coolingcoils when the inlet air dry bulbgoes below 60 to 65 F. If the unithas frost control, it will experi-ence significant capacity loss dueto the frost-control operation. Twotypical approaches to frost controlare listed below:
● A temperature-sensing ther-mostat diverts the hot refrigerantgas through the evaporator coiluntil the ice is melted.
● An automatic de-ice sensorshuts the compressor off whenevaporator-coil temperature ap-proaches freezing. The fan contin-ues moving warm air across thecoil to defrost it.
Large industrial dehumidifiersThis section of the article is a
detailed review of three commontypes of industrial dehumidifiers:direct expansion, dry desiccant,and liquid desiccant. Each systemhas an air intake of 30,000 cu ftper min at 70 F and 56 grains ofwater vapor per lb of dry air. Also,each system was required to drythe air down to 35 or 36 grains ofwater vapor per lb of dry air. Iworked with several dehumidifi-cation companies to size the sys-tems and determine all energyconsumption. Using the brakehorsepower provided by the com-panies, I sized the motors and cal-culated the electrical consump-tion (KW per hr) illustrated in thisarticle. Operating costs were thendeveloped using $0.06 per KWH forelectricity and $5.00 per 106 Btufor natural gas. The operatingcosts were then developed into op-erating cost per 1000 lb of watervapor removed. In some cases, Ileft out proprietary informationon the systems at the request ofthe manufacturers.
Each industrial dehumidifier in-cludes detailed air- and water-va-por mass-flow analysis, psychro-metrics, thermodynamics, andadiabatic mixing (Figs. 1 to 3) tohelp the readers understand theenergy flows. The psychrometricsand thermodynamics used in thediagrams follow the principals of
Zimmerman and Lavine. The air-and water-vapor flows are also il-lustrated in acfm (actual cu ft permin) and the dehumidification in-dustry dscfm (dry standard cu ftper min). I used dscfm to avoid con-fusion with the fan industry scfm.
Cooling-based dehumidificationMoisture can be removed from
the air by cooling the air below itsdew-point temperature. This can
be achieved through the followingsystems:
● Chilled water, glycol, or brinecoil system
● DX cooling coil system● Chilled water air-washer sys-
temThe first two systems accom-
plish dehumidification by passingthe air through a cooling coil witha coil-surface temperature belowthe dew point of the air. Water va-
March 1999 HPAC Heating/Piping/AirConditioning 51
54
3 21
DSP = 2.79 in. WC21.97 bhp DP = 0.76 in. WC Heat addition
437,964 (Btu per hr)
Processair outlet
DXcoolingcoil
Heat removal538,617(Btu per hr)(44.88 ton)
Heat
pip
e ex
chan
ger
Heat transferefficiency = 81.31 percent
Filter Damper
Processair inlet(moist)
DP = 0.72 in. WCDP = 0.45 in. WC DP = 0.08 in. WC
Waterremoval380.24(lb per hr)
War
m fr
eon
gas
35 F
liqu
id fr
eon
Exhaust air
Air inlet Air inletAir-cooledcondenser
DX chiller withair-cooled condenser(1.5405 KW per ton) at 95 F
High
pre
ssur
e
Refrigerationgas
Hot g
as
Low pressureCompressor
1 DX with heat pipe dehumidifier.
Process air anaylsis 1 2 3 4 5Dry bulb 70 F 53.40 F 39.73 F 42.73 F 56.3 FWet bulb 58.8 F 52.17 F 39.73 F 41.29 F 47.90 FDew point 51.26 F 51.26 F 39.73 F 39.73 F 39.73 FRelative humidity 51.56% 92.86% 100% 89.61% 53.95%W (grains WV 56 56 36 36 36
per lb dry air)acfm 30,000 29,059.79 28,157.15 28,326.30 29,091.51dscfm 29,574.27 29,574.27 29,574.27 29,574.27 29,574.27M (lb dry air per hr) 133,084.56 133,084.56 133,084.56 133,084.56 133,084.56M (lb WV per hr) 1064.68 1064.68 684.44 684.44 684.44Q sensible 1,215,006 684,051 246,976 342,866 776,787
(Btu per hr)Q latent 1,162,977 1,155,315 738,627 739,522 743,565
(Btu per hr)Q total 2,377,983 1,839,366 985,603 1,082,388 1,520,352
(Btu per hr)
Coil heat removal (information shows what occurs between Points 2 and 3)DQ sensible (Btu per hr): 437,075DQ latent (Btu per hr): 416,688DQ total (Btu per hr): 853,763 (71.147 ton)Water vapor removed (lb per hr): 380.24
INDUSTRIAL DEHUMIDIFACATION
por condenses on the coil surfaces.The amount of moisture removaldepends on how cold the air can bechilled. The lower the tempera-ture, the drier the air. The chilledwater air-washer system alsocools the air below its dew pointby using water that is colder thanthe dew-point temperature. Thewater vapor in the air condenseson the water spray or the nearestsurface. In this case, the use ofcolder water results in greater dehumidification.
The type of system chosen for il-lustration in this article is the DXcooling coil system. The basic com-ponents of this mechanical refrig-eration system are an evaporatorcoil, compressor, condenser, andthrottling valve (or expansionvalve). The system uses a refriger-ant that enters the evaporator coil(cooling coil) in a liquid state. Therefrigerant evaporates inside thecoil and, in doing so, absorbs heatfrom the process air movingthrough the coil. It then leaves thecoil in the form of a gas. The com-pressor takes the cold vapor fromthe evaporator and compresses itto a hot gas at high pressure. Whenthe refrigerant leaves the compres-sor, it is still a gas but at a muchhigher pressure (five to ten timesgreater) and a much higher tem-perature. The hot refrigerant gasis then pushed through a con-denser (in this case, an air-cooledcondenser) where the hot gas iscooled and condensed into a liquidby some substance, usually air orwater. The refrigerant then flowsfrom the condenser as a high-pres-sure liquid through the expansionvalve. As the liquid passes throughthe valve, its pressure is suddenlydecreased to the pressure in theevaporator coil. At the same time,the temperature of the liquid re-frigerant drops down from thewarm condenser temperature tothe cold evaporator temperature.This occurs because a smallamount of liquid suddenly flashes
52 HPAC Heating/Piping/AirConditioning March 1999
Reactivationblower (exhaust)
DSP =2.3 in. (Dehumidification)
+ 1.6 in. (Burner) = 3.9 in. WC
Process air bypass blowerDSP = 0.5 (Filter) + 0.25 in. (Duct)= 0.75 in. WC
C
1 2
4
B
3 5
A
Q sensible (Btu per hr) = 950,225.90Q total (Btu per hr) = 1,095,457.91
Natural gas
Reactivationsector
Natural gasdirect-firedheater Filter
Filter
Filter
4.3 bhpReactivationair inlet
Process airinlet (moist)
9.52 bhp
Desiccantwheel
Process airoutlet
Process air blowerDSP = 2.0 in.
(Dehumdification) + 0.5 in. (Filter) = 2.5 in. WC
Drive0.25 bhp
2.59 bhp
2 Rotating dry desiccant dehumidifier.
Process air anaylsis 1 2 3 4 5Dry bulb 70 F 70 F 115 F 70 F 93.64 FWet bulb 58.80 F 58.80 F 65.64 F 58.80 F 62.54 FDew point 51.26 F 51.26 F 19.63 F 51.26 F 38.96 FRelative humidity 51.56% 51.56% 3.62% 51.56% 14.98%W (grains WV 56 56 15.8 56 35
per lb dry air)acfm 30,000 15,728.59 16,909.30 4,271.41 31,189.04dscfm 29,574.27 15,505.38 15,505.38 14,068.89 29,574.27M (lb dry air per hr) 133,084.56 69,774.41 69,774.41 63,310,15 133,084.56M (lb WV per hr) 1064.68 558.20 157.49 506.48 663.97M (lb per total hr) 134,149.24 70,332.61 69,931.90 63,816.63 133,748.53Q sensible 1,215,006 637,011 1,393,584 577,995 1,971,579
(Btu per hr)Q latent 1,162,977 609,733 175,060 553,244 728,304
(Btu per hr)Q total 2,377.983 1,246,744 1,568,644 1,131,293 2,699,883
(Btu per hr)
(The following information shows what occurs between Points 2 and 3.)Heat of condensation (Btu per hr): 434,673Reactivation system sensible heat leakage (Btu per hr): 321,900Water vapor removed (lb per hr): 400.70
Reactivation air analysis U U UDry bulb 90 F 293 F 130 FWet bulb 78.22 F 113.75 F 104.96 FDew point 74.06 F 81.24 F 100.76 FRelative humidity 59.77% 0.88% 43.71%W (grains WV per lb dry air) 127.0 162.25 308.07acfm 4583.44 6310.45 5103.46dscfm 4285 4274.78 4274.78M (lb dry air per hr) 19,282.55 19,236.55 19,236.55M (lb WV per hr) 349.84 445.89 846.59M (lb per total hr) 19,632.39 19,682.44 20,083.14Q sensible (Btu per hr) 268,784 1,219,010 453,895Q latent (Btu per hr) 385,151 530,383 946,355Q total (Btu per hr) 653,935 1,749,393 1,400,250
(The following information shows what occurs between Points B and C.)Sensible heat loss total (Btu per hr): 765,115Heat of desorption (Btu per hr): 443,215Reactivation heat leakage to process air (Btu per hr): 321,900Water vapor pick-up (lb per hr): 400.70
A B C
to a vapor as it passes through therestriction in the valve. Then theliquid, with some bubbles of flashvapor, enters the evaporator coil.The liquid refrigerant in the coilevaporates and, in doing so, ab-sorbs heat from the air passing
through the coil.An example of a DX dehumidifi-
cation system is illustrated in Fig.1. The system includes a heat-pipeheat exchanger and an air-cooledcondenser system. The heat piperemoves 538,616 Btu per hr (44.88
ton) of heat from the inlet processair and passes it to the dehumidi-fier process air leaving the system.The advantage is reduced coolingload and a leaving air conditionthat is not at or near saturation.The operating cost in Table 4 is
March 1999 HPAC Heating/Piping/AirConditioning 53
0.7 bhpregen desiccantpump
Plate-and-frameheat exchanger
Cold drydesiccant
Cooler45 F
55 F
7.82 bhpchilledwaterpump
Condenser
Exhaust air
DX chiller withair-cooledcondenser
(1.10 KW per ton)Evaporator
Refrigeration gasCompressor
Air inlet
DQ = 1,006,00 (Btu per hr)
2
1
4
3
Driedprocess airoutlet Mist
eliminatorMisteliminator
Wet regenexhaust air Plate-and-frame
heat exchanger
Fluegas
845,500(Btuper hr)
Hotwaterboiler
Wetprocess air inlet
Air Filter
Filter
Reci
rcul
ated
des
icca
nt
Wetdesiccant
Heater180 F HWR
Natural gascombination air
Packingblocks
Cond. fan13.07 bhp
Packingspheres Regen
fan
0.648bhp
DQ = 676,400 (Btu per hr)220 F
1.48 bhpHot waterpump
7 bhpcond.desiccantpump
Weak wetdesiccantto regen
Regeneratorair inlet
Diffuser Wet desiccantto heater
Conditionerair DP = 1.3 in. WC
Hot dry desiccantto conditioner
Regeneratorair DP = 2.1 in. WC
Low pressure
3 Liquid desiccant dehumidifier.
Process air anaylsis 1 2 3 4Dry bulb 70 F 55 F 95 F 160.1 FWet bulb 58.80 F 47.00 F 78.40 F 132.12 FDew point 51.26 F 39.01 F 72.42 F 129.97 FRelative humidity 51.56% 55% 48.42% 46.74%W (grains WV per lb dry air) 56 35 120 775.14acfm 30,000 29,011 1021.65 1308.73dscfm 29,574.27 29,574.27 948 948M (lb dry air per hr) 133,084.56 133,084.56 4266.01 4266.01M (lb WV per hr) 1064.68 665.42 73.13 472.39DQ sensible (Btu per hr) 1,215,006 735,214 64,597 131,723DQ latent (Btu per hr) 1,162,977 722,535 80,669 553,896DQ total (Btu per hr) 2,377,983 1,457,749 145,266 665,619Water vapor removed (lb per hr): 399.26
These data sections represent what occursbetween Point 1 and Point 2 and whathappens between Point 3 and Point 4:
Process air heat removalDQ sensible = 479,792 (Btu per hr)DQ latent = 440,442 (Btu per hr)DQ total = 920,234 (Btu per hr)
Cooling coil loadProcess air = 920,234 (Btu per hr)Heat dump back = 85,766 (Btu per hr)Total = 1,006,000 (Btu per hr)
Regen air heat additionDQ sensible = 67,126 (Btu per hr)DQ latent = 453,227 (Btu per hr)DQ total = 520,353 (Btu per hr)
Heater loadRegen air = 520,353 (Btu per hr)Desiccant heat = 156,047 (Btu per hr)Total = 676,400 (Btu per hr)
INDUSTRIAL DEHUMIDIFACATION
$20.03 per 1000 lb of water vaporremoved at a process-air-leavingcondition of 56.3 F and 36 grainsper lb (53.9 percent RH). Dehumidi-fication reheat systems will usethe high-pressure, high- tempera-ture gas leaving the compressor ina coil system to reheat the processair leaving the dehumidifier. Mostcommon residential dehumidifiersuse this configuration. It is impor-tant for designers to considerwaste energy usage, heat wheels,or heat pipes for reheat because re-heat can add a significant “added”cost to the operating cost of a DXsystem.
A standard DX system with noheat pipe has an operating cost of$26.37 per 1000 lb of water vaporremoved—at a process-air-leavingcondition of 42.7 F and 36 grainsper lb (89.7 percent RH).
Rotating dry desiccant wheelDry desiccants are adsorbent
materials that attract moisture be-cause of the electrical field at thedesiccant surface. The field attractswater molecules that have a net op-posite charge. Some of the solid ad-sorbents used in dry desiccant sys-tems are illustrated in Table 5.
Sorption is the adsorption pro-cess by which a desiccant removeswater vapor directly from the air.The ability of an adsorbent to at-tract moisture depends on the dif-ference in vapor pressure betweenthe desiccant surface and air. Thevapor-pressure difference drivesmoisture from the high vapor-pressure area to the low vapor-pressure area. Dry desiccants typ-
ically have low vapor pressure attheir surface and, therefore, ad-sorb moisture from the air. Whenmoisture is removed from the pro-cess air stream, it produces heatof sorption (or heat of adsorption),which is composed of latent heatof condensation of the removedmoisture plus additional chemicalheat. The heat of sorption of themoisture removed from the air isconverted to sensible heat. Theamount of heat released is usuallyaround 1080 Btu per lb WV re-moved to 1312 Btu per lb WV re-moved. The actual amount de-pends on the type of desiccant.The heat of sorption (sensibleheat) is energy that is passed to
TABLE 4—DX with heat pipe dehumidificationoperating cost to remove 1000 lb of water vapor.
Operating cost Electric usage
1) Process air fan $2.74 per 1000 17.36 KW per hr2) Compressors $16.17 per 1000 102.49 KW per hr3) Condenser fans $1.12 per 1000 7.11 KW per hr
4) Total $20.03 per l000 lb WV 126.96 KW/hr@ process air discharge of 56.3 F and 36 grains WV per lb dry air
a) Electricity price: $0.06 per KWHb) DX cooling with no energy recovery has a cost of $26.37 per 1000 lb WV with a process air discharge
of 42.7 F and 36 grains WV per lb dry air.
TABLE 6—Dry desiccant dehumidification operating cost to remove 1000 lb of water vapor.
Direct fired Indirect fired Electricnat. gas heat Electric usage nat. gas heat reactivation heat
1) Reactivation heat $13.67 per 1000 lb — $16.33 per 1000 lb $43.19 per 1000 lb2) Reactivation blower $0.54 per 1000 lb 3.62 KW per hr $0.54 per 1000 lb $0.54 per 1000 lb3) Process air blower $1.25 per 1000 lb 8.37 KW per hr $1.25 per 1000 lb $1.25 per 1000 lb4) Process air bypass blower $0.35 per 1000 lb 2.34 KW per hr $0.35 per 1000 lb $0.35 per 1000 lb5) Desiccant wheel-drive motor $0.06 per 1000 lb 0.41 KW per hr $0.06 per 1000 lb $0.06 per 1000 lb
6) Sub-total operating cost $15.87 per 1000 lb WV 14.74 KW per hr $18.53 per 1000 lb WV $45.39 per 1000 lb WV7) Cost to cool air to 55 F $18.37 per 1000 lb WV 122.66 KW per hr $18.37 per 1000 lb WV $18.37 per 1000 lb WV
8) Total operating cost $34.24 per 1000 lb WV 137.40 KW per hr $36.90 per 1000 lb WV $63.76 per 1000 lb WVWith a process air discharge of 93.6 F and 35 grains per lb
Notes:a) Electricity price: $0.06 per KWH (or $17.584 per 106 Btu)b) Natural gas price: $5.00 per 106 Btuc) Direct fired natural gas usage: 1,095,458 Btu per hrd) Indirect fired natural gas usage: 1,308,550 Btu per hre) Electric coil heat provided: 984,128 Btu per hrf) Heat removal to cool air to 55 F: 1,247,418 Btu per hr—(103.95 ton)g) Cooling energy requirement: 1.18 KW per ton (air cooled-rotary liquid chiller and pump)h) Modular vertical bed system operating cost: $14.55 per 1000 lb WV @ 89.5 F and 35 grains WV per lb dry air
continued on page 56
INDUSTRIAL DEHUMIDIFACATION
the process air stream, whichraises the discharge air tempera-ture of the process air stream.
As the moisture content of thedesiccant rises, so does the water-vapor pressure at the desiccantsurface. At some point, the vaporpressure at the desiccant surfacewill be the same as the air, andmoisture adsorption will end. Thedesiccant is then taken out of theprocess air stream and is placedinto the reactivation air stream (ascavenger air stream consisting ofoutside air or building air). Thereactivation air stream is typi-cally heated to a temperature of190 to 375 F. The combination ofthe heat and moisture raises thevapor pressure at the desiccantsurface. When the surface vapor
pressure exceeds the vapor pres-sure of the reactivation air, mois-ture leaves the desiccant. Thisprocess is called reactivation.
The reactivation section, whichconstitutes less than half of thedesiccant wheel, uses flexibleseals to seal it from the adsorptionor process side to minimize crosscontamination. The typical leak-age rate is 1 to 2 percent.
Following reactivation, the hotdesiccant rotates back into theprocess air where the process aircools the desiccant which lowersthe desiccant vapor pressure, so itcan collect more moisture fromthe balance of the process airstream. Some of the equationsused by the rotating dry desiccantmanufacturers are shown in theaccompanying sidebar.
A typical layout of a rotary drydesiccant system is illustrated inFig. 2. The operating costs of thesystem are listed in Table 6. Theoperating cost for the example inFig. 1 is $15.87 per 1000 lb of wa-ter vapor removed with a leavingair condition of 93.6 F and 35grains per lb. If the air is too hotand has to be cooled to 55 F and 35grains per lb, the operating costincreases to $34.24 per 1000 lbWV. Reactivation heat represents86 percent of the operating costfor the direct-fired unit and 88percent for the indirect-fired unit.This represents a great opportu-nity for the designer to use low-cost hot water from cogeneration;the use of low-cost steam; or con-densate, refrigeration reject heat,or waste exhaust to preheat the
56 HPAC Heating/Piping/AirConditioning March 1999
continued from page 54
Nomenclaturedscfmp = process air flow through the desiccant
wheel in dry standard ft3 per mindscfmr = reactivation air flow through the desiccant
wheel in dry standard ft3 per minQS = heat of sorption, Btu per hr
QRA = reactivation heat addition, Btu per hrQRL = reactivation system heat leakage passed to
the process air Btu per hrtpo = process air temperature leaving desiccant
wheel, Ftpi = process air temperature entering desiccant
wheel, FWpi = amount of moisture in the process air
entering desiccant wheel, grains of water vaporper lb of dry air
Wpo = amount of moisture in the process air leaving desiccant wheel, grains of water vapor per lb of dry air
Wri = amount of moisture in the reactivation air (after the heater) entering the desiccant wheel, grains of water vapor per lb of dry air
Wro = amount of moisture in the reactivation air leaving the desiccant wheel, grains of water vapor per lb of dry air
tro = heated reactivation air temperature entering the desiccant wheel (equals temperature of reactivation air leaving the heater)
tri = reactivation air temperature entering the heater
k = a factor that varies from 0.038 to 0.11. (Some books recommend a value of 0.07.)
Heat of sorption equation (approximate)The water vapor removed from the process air
(latent heat) was adsorbed by the desiccant,which released that heat as sensible heat to theprocess air.
Process air through the desiccant wheel—energy loses and gains
DQ sensible heat gain = heat of sorption + reac-tivation system leakage
DQ latent heat loss = heat of sorptionDQ total = DQ sensible + DQ latentDQ total = (heat of sorption + reactivation
system leakage) - heat of sorption = reactivationsystem leakage
Note: Heat of sorption is sometimes called heat of condensation
reactivation air. These optionscould significantly reduce the op-erating cost. Keep in mind that insome cases it may be more eco-nomical to combine cooling anddesiccant dehumidification. Thetechnologies do complement eachother since the refrigeration con-denser reject heat from the cool-ing process can be used to preheatthe reactivation air.
In process-drying applications,dry desiccant dehumidifiers aresometimes used without addedcooling because the increase intemperature caused by the heat ofadsorption is helpful in the dryingprocess. However, in some applica-tions, a provision must be made toremove the excess sensible heatfrom the process air after dehu-midification. For this reason,Table 6 provides an added cost sec-tion for cooling the air to 55 F as anexample of the possible added cost.
tion operates on the principal ofchemical absorption of water va-por from the air. The absorbent ordesiccant solution will changephysically, chemically, or bothduring the sorption process. Someof the liquid desiccant solutionsused for dehumidification are:
cation is very similar to a chilledwater air-washer system. Whenthe air passes through thewasher, its dew point approachesthe temperature of the water sup-plied. Air that is more humid isdehumidified, and air that is lesshumid is humidified. In a similarmanner, the liquid absorption de-humidifier sprays the air with adesiccant solution that has alower vapor pressure than the va-por pressure of the entering pro-cess air stream. The liquid has avapor pressure lower than waterat the same temperature, and the
air passing over the solution ap-proaches this reduced vapor pres-sure. The ability to remove watervapor (or add water vapor) is de-termined by the temperature andconcentration of the solution. Theconditioner can be adjusted sothat the conditioner delivers air atthe desired relative humidity.
The vapor pressure of a givenconcentration of absorbent solu-tion approximates the vapor-pres-sure values of a fixed relative hu-midity line on a psychrometricchart. For instance, a 40 percentconcentration of lithium chlorideclosely approximates the 20 per-cent relative humidity line. Also,a 15 percent concentration is veryclose to the 80 percent relative hu-midity line. Therefore, it can besaid that higher solution concen-
trations give lower equilibriumrelative humidity and thus allowthe absorbent to dry air to lowerlevels. Temperature also affectsthe absorbents’ ability to removemoisture. For instance, a 25 per-cent solution lithium chloride hasa vapor pressure of 0.37 in Hg at70 F (same as air at 70 F and 50percent RH). When the solution isheated to 100 F, the vapor pres-sure climbs to 0.99 in Hg. There-fore, the warmer the desiccant,
the less moisture it can absorb.Also, if the solution vapor pres-sure is higher than the surround-ing air, the water vapor willtransfer to the air and dry the des-iccant solution.
A typical system diagram is il-lustrated in Fig. 3. In the opera-tion, warm, moist air is sprayedwith a solution of chilled lithiumchloride, which was cooled withchilled water in a plate-and-frameheat exchanger. The air is cooledand dehumidified by heat andmass transfer to the lithium chlo-ride solution. A chiller with anair-cooled condenser section pro-vides the chilled water to cool thelithium chloride solution. If thedesired dehumidified air-mois-ture content is 50 grains per lb (or48 F dew point), the water used to
cool the desiccant can be 85 F cool-ing-tower water rather thanchilled water. Consult Table 1 foradditional information.
When moisture is removed fromthe air, the reaction liberatesheat. This is the reverse of evapo-ration, when heat is consumed bythe reaction. The heat that is gen-erated is the latent heat of con-densation of the water vapor plusthe heat of solution (or the heat ofmixing of the water and desic-
Table 7—Liquid desiccant dehumidification operating cost to remove 1000 lb of water vapor
Operating cost Electric usage
1) Hot water boiler—natural gas $10.59 per 1000 lb —2) Regen air fan 0.09 per 1000 lb 0.63 KW per hr3) Process air fan 1.67 per 1000 lb 11.14 KW per hr4) Regen desiccant pump 0.09 per 1000 lb 0.64 KW per hr5) Cond. desiccant pump 0.85 per 1000 lb 5.69 KW per hr6) Hot water pump 0.27 per 1000 lb 1.83 KW per hr7) Chilled water pump 1.02 per 1000 lb 6.82 KW per hr8) Chiller compressors 12.33 per 1000 lb 82.07 KW per hr9) Chiller condenser fans 1.52 per 1000 lb 10.14 KW per hr
10) Total $28.43 per 1000 lb WV 118.96 KW per hrwith a process air dischargeof 55 F and 35 grains per lb
a) Electric price: $0.06 per KWHb) Natural gas price: $5.00 per 106 Btuc) Hot water boiler efficiency: 80 percentd) Hot water boiler natural gas usage: 845,500 Btu per hre) Operating cost of liquid desiccant system with a natural gas engine driven chiller using jacket heat to preheat
regenerator air: $19.85 per 1000 lb WV
March 1999 HPAC Heating/Piping/AirConditioning 57
INDUSTRIAL DEHUMIDIFACATION
cant). In desiccant dehumidifica-tion, this heat (approximately1080 to 1320 Btu per lb water va-por removed) is transformed tothe air, raising the air dry bulbtemperature, and therefore, in-creasing the load on the chilledwater system. The chilled watersystem must be sized to removethe latent heat of condensation(heat of sorption), the air sensibleheat, and the residual heat load added by the regeneration process.
To remove the water extractedfrom the air and keep the liquiddesiccant at a fixed concentration,a small percentage of the condi-tioner-desiccant pump flow (typi-cally around 15 percent) is trans-ferred to the regeneration system.The weak desiccant solution ispumped up to a heating system(plate-and-frame heat ex-changer), which raises the tem-perature and vapor pressure ofthe liquid desiccant. The hot des-iccant is then sprayed at a scav-enger air stream (outside air orbuilding air) with a lower vaporpressure that forces the water va-por out of the desiccant and intothis air, which is exhausted out-side. The dry desiccant returns tothe regenerator sump. The desic-cant is still a little warm, and itsvapor pressure is still a littlehigh—until it flows back to theconditioner and is cooled by thechilled water heat exchanger.Therefore, the cooling systemmust be sized to include thisresidual heat load added by the
regeneration process (sometimescalled heat dump back). Theamount of heat and dump back istypically in the range of 50 to 350Btu per lb water vapor removed.
Table 7 is a review of the oper-ating costs associated with thesystem in Fig. 3. The total operat-ing cost is $28.43 per 1000 lb ofwater vapor removed from theprocess air. You will note that thenatural gas cost is 37 percent ofthe total cost. Therefore, it pays tofind a source of waste heat to re-duce these costs. Some possiblesources are condenser-rejectedheat, solar heat, or a natural gasengine-driven chiller, which pro-duces hot water (engine heat) aswell as chilled water at a reducedcost. The operating cost of a liquiddesiccant dehumidification sys-tem combined with a natural gasengine-driven chiller is $19.85 per1000 lb of water vapor removed.This combination represents a 30percent reduction in operating energy costs.
Another option to save energy onliquid desiccant systems is to in-stall a liquid-to-liquid-type heatexchanger (some call this an inter-changer) placed between the warmdesiccant leaving the regeneratorand the cool desiccant entering theregenerator. By doing so, less en-ergy is needed to regenerate thedesiccant because it is warmerthan when it left the regenerator.The heat exchanger will typicallyreduce the heat dump back to theconditioner-cooling consumptionby about 65 percent and reduces
the regenerator heat consumptionby about 15 percent.
ConclusionTable 8 is a summary of the op-
erating costs for each of the large,industrial dehumidifiers usingthe electricity cost of $0.06 perKWH and the natural gas cost of$5.00 per 106 Btu. One can seethat the dry desiccant systemshave the lowest energy operatingcost of all the systems, but theyalso have the highest dischargeair temperatures.
If elevated process air tempera-ture is not acceptable, the liquiddesiccant or DX with heat pipewould be the choice (at 55 to 56 Fdischarge air temperature).
The outcome of this study willchange, depending upon the ac-tual energy costs for that area. Inthe United States, natural gascosts vary from $2.40 to $7.90 per106 Btu and electricity goes from$0.018 to $0.15 per KWH. Whatthis says is that every systemmust be evaluated based on theenergy costs for that region. Also,the evaluation of operating costsshould include maintenance andcapital amortization.
Each system is unique, andthey all offer ways to reduce en-ergy operating costs. For example,the liquid desiccant system oper-ating cost dropped 30 percent(from $28.43 to $19.85 per1000 lbWV removed) by incorporating anatural gas engine-driven chillerand using the hot water jacketheat to preheat the regenerator
58 HPAC Heating/Piping/AirConditioning March 1999
TABLE 8—Dehumidifier operating cost comparison.Dry bulb Water vapor*
System type temp in the air Operating cost (energy cost only)
1) DX with heat pipe 56.3 F 36 $20.03 per 1000 lb WV removed2) DX without heat pipe 42.7 F 36 $26.37 per 1000 lb WV removed3) Dry desiccant rotating wheel—direct-fired 93.6 F 35 $15.87 per 1000 lb WV removed4) Dry desiccant rotating wheel—direct-fired with added cooling 55.0 F 35 $34.24 per 1000 lb WV removed5) Dry desiccant vertical bed—direct-fired 89.5 F 35 $14.55 per 1000 lb WV removed6) Liquid desiccant—with electric air-cooled chiller 55.0 F 35 $28.43 per 1000 lb WV removed7) Liquid desiccant—with natural gas engine-driven chiller 55.0 F 35 $19.85 per 1000 lb WV removed
Notes: 1) Electricity cost: $0.06 per KWH 2) Natural gas cost: $5.00 per 106 Btu*Grains of water vapor per lb of dry air
system. Also if the required dewpoint is 48 F (49.68 grains per lb)or above, cooling-tower water canbe used in the summer. The DXsystems can use heat pipes, heatwheels, and/or heat reject fromthe process cooling for reheat. Us-ing purchased energy for DX re-heat can raise the operating costssignificantly. The major operatingcost for dry desiccant systems isthe reactivation heat that repre-sents 86 percent of the total oper-ating cost for the system in Table6. Using alternative low-cost en-ergy sources can reduce the oper-ating costs significantly. DX sys-tem condenser reject heat could beused to preheat the reactivationair as well as desuperheater coilheat. For this reason, the mosteconomical system may be a com-bination DX system and desiccant(solid or liquid) system. The tech-nologies do complement eachother since the refrigeration con-denser reject heat from the aircooling process can be used to pre-vent the air entering the reactiva-tion or regeneration section of thedesiccant dehumidifier. In indus-trial plants, there are manysources of cheap, low-grade heat(low temperature) that can beused in the reactivation or regen-eration system. It is up to the en-gineer involved in the equipmentselection to consider these sourcesbefore a decision can be made onthe type of system. HPAC
Bibliography1) “Water Vapor Migration and
Condensation Control in Buildings,”by W. Acker, Heating/Piping/AirConditioning Magazine, June 1998.
2) Zimmerman, O.T., and I. Lavine,Industrial Research Services: Psy-chrometric Tables and Charts, 2ndEd., Industrial Research Services,Inc., Dover, New Hampshire, 1964.
I would like to thank the following com-panies for assistance in the preparationof 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 liketo thank Craig Pekarek for the CADwork and Jodi Cavil for the typesettingof this article.
Mr. Acker has 25 years of experience inindustrial HVAC, hygieneology, co-generation, and environmental engi-neering. He is considered an expert in
the analysis of air and water vapor atatmospheric pressure or compressed.Questions or comments on this articlemay be directed to the author at 902-465-3548.
March 1999 HPAC Heating/Piping/AirConditioning 59
By WILLIAM ACKER*,President,Acker & Associates,Green Bay, Wis.
In designing dehumidification systems, one of themost important tasks is to quantify the water vaporloads 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 calculationseasier. However, these approximate equations lose accuracy when air temperatures are higher or lower thanthe conditions assumed when the approximate equationswere developed. This article will clarify some of theseequations and present alternative equations that are
more precise across a larger range ofconditions.
This article will also explain the airflows used in proposals from dehu-midification companies as well as theproposed flow diagrams from dehu-midification equipment suppliers. I have discovered a great amount ofuncertainty over the air flows in thesediagrams.
DEHUMIDIFICATION INDUSTRYEQUATIONS AND TERMS
Dehumidification manufacturerslike to develop air flow diagrams of
May 1999 • HPAC Heating/Piping/AirConditioning 93
IndustrialDehumidification:
Air Flow Diagrams & Water Vapor Load Equations
*William Acker is a member of HPAC’s Editorial Advisory Board.
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.
P A R T 0 N E
17,215 cfm55 F20 GR
Conditioned space55 F DB, 30 percent RH,
20 GR per lb
5100 cfm55 F20 GR
System fan
Cooling coil
Makeup air 2400 cfm50 F/49 F50 GR
22,315 cfm 55 F20 GR
24,71542 F15 GR
Dehumidifier
7500 cfm53 F30 GR
24,715 cfm68 F15 GR
7500 cfm96 F4 GR
FIGURE 1 Dry desiccant systems flow diagram.
their entire systems. Mass flow analysis is used in thesediagrams because mass flow does not change if there is atemperature or pressure change. Mass flows can also beadded and subtracted. The actual air flows or acfm (cu ftper min) will change if the temperature or pressurechanges. Also, acfm values cannot be added or sub-tracted because they are at different air densities. Thedehumidification industry chose a type of mass flowanalysis that is flow in dry standard cu ft per min (dscfm)at a common air density of 0.075 lb dry air per dry stan-dard cu ft. The acfm flows are converted to these flowsfor illustration in the air system diagrams (shown in Fig.1 and Fig. 2).
As mentioned earlier, dscfm air flow is a flow at a com-mon air density of 0.075 lb dry air per dry standard cu ft. I
prefer to use the term dscfm to avoid any possible confu-sion with acfm or the fan industry term scfm. In Fig. 1,one manufacturer uses cfm to represent the dry standardair flow. When you look at this diagram, it is easy to tellthat the flows are dry standard air flows and not acfm. Ifyou look at the flow in and out of the dehumidifier, theflows are both 7500 cfm. This is not possible with acfmbecause this flow is made of both air and water vapor;therefore, there is a loss (of water vapor or cfm) as it trav-els through the dehumidifier. Listed below is an exampleof an acfm flow broken down into dry air flow and watervapor flow:
Air pressure = 29.921 in. HgAir temperature = 70 FRelative humidity = 100 percentTotal air flow = 100,000 acfmDry air flow = 97,530 cfmWater vapor flow = 2470 cfmAlso, note that there is a change in temperature across
the dehumidifier that would also cause a change in theacfm flow.
Fig. 2 is a diagram prepared by another dehumidifiermanufacturer. The air flows are in dscfm, but in this case,they use the term scfm. This scfm should not be confusedwith the fan industry scfm, which is a type of mass flowthat represents the combined air and water vapor flow. Bylooking at Fig. 2, you can see that the flow into the dehu-
midifier (or conditioner) 20,833 scfm is the same as theflow leaving; therefore, this is a dry air mass flow. If theentering flow is fan industry 20,833 scfm, the leaving flowwould have dropped to 20,754 scfm due to the removal ofwater vapor (79 scfm or 355.41 lb of water vapor per hr).The next section on fan industry scfm will help you to un-derstand the difference between fan and dehumidifica-tion scfm terms.
The equations used by the dehumidification industryto convert the actual air flows to dry standard air flows arelisted below:
Where:
This is the dry standard air flowused by the dehumidificationindustry.
This is the actual air flow of dryair and water vapor.
This is the air-specific volume.The value can be found in psy-chrometric charts providedthat the barometric pressure ofthe chart is the same as the to-tal pressure (barometric plusstatic pressures) of the acfmflow.
This is the standard air densityfor the dehumidification indus-try. The inverse of this value is13.3333.
This is the total amount of dryair mass flow.
FAN INDUSTRY EQUATIONS AND TERMSFans must be selected based on the actual flow rate and
the actual density at the inlet to the fan. Some fan manu-facturers prefer to specify the flow rate based on standardinlet conditions. Fan performance curves are developed
94 HPAC Heating/Piping/AirConditioning • May 1999
INDUSTRIAL DEHUMIDIF ICAT ION
4,200 acfmTo outside
19,833 scfm39 GR per lb.
75 F
95 F1000 scfm118 GR per lb.
20,833 scfm42.8 GR per lb.
76 FConditioner 20,833 scfm
16.1 GR per lb.
78.8 tons45 F chilled water
Cooler
Heater743,000 Btuh200 F hot water
3,150 scfmOutside air
Regen
55 F
FIGURE 2 Liquid desiccant system flow diagram.
dscfm(dry std cu ft)
min
dry air mass flow(lb dry air)
min
.lb dry air
dry std cu ft
(1)
acfm(cu ft wet air)
min
.lb dry air
dry std cu ft(cu ft wet air)
lb dry air
(2)
acfm(cu ft wet air)
min.
dry std cu ftlb dry air
(cu ft wet air)lb dry air
(3)
=
=×
=×
0 075
0 075
13 3333
v
v
dscfm(dry std cu ft)
min=
acfm(cu ft wet air)
min=
v(cu ft wet air)
lb dry air=
0 075.lb dry air
dry std cu ft=
Dry air mass flow(lb dry air)
min=
(from a series of laboratory tests) at these standard condi-tions. Listed below are the equations that allow designersto convert from the actual conditions to the standardconditions:
Where:
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 asystem. In fan sizing, this would be the flow entering thefan.
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 thefan inlet.
Standard density (lb total per std cu ft) = Some fan manufacturers like to develop fan perfor-
mance curves based on a standard gas or air density of0.075 lb total per std cu ft. The Air Movement and Con-trol Association (AMCA) indicates that this density issubstantially 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 FLOWThe following examples show how the dehumidifica-
tion industry dscfm flow compares to the fan industryscfm flow.
Example 1:Air pressure = 29.921 in. HgAir temperature = 70 FRelative humidity = 100 percent
Humidity ratio = 109.93 grains WV per lb dry airMoist air density = 0.074190 lb wet air per cu ft wet airSpecific volume = 13.6906 cu ft wet air per lb dry airacfm flow = 40,000 cu ft wet air per minDehumidification industry dscfm = 38,956 dry std cu ftper minFan industry scfm = 39,568 std cu ft per min
Example 2:Air pressure = 29.921 in. HgAir temperature = 200 FRelative humidity = 50 percentHumidity ratio = 2809 grains WV per lb dry airMoist air density = 0.051216 lb wet air per cu ft wet airSpecific volume = 27.3599 cu ft wet air per lb dry airacfm flow = 40,000 cu ft wet air per minDehumidification industry dscfm = 19,493 dry std cu ftper minFan industry scfm = 27,315 std cu ft per min
Example 1 shows that at 70 F, the dscfm flow is veryclose to the fan industry scfm flow (only 1.6 percent vari-ation). However, at more elevated temperatures, such asin Example 2, the fan industry scfm flow is 40 percenthigher than the dscfm flow.
WATER VAPOR LOADS DUE TO AIR FLOWWater 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 majorcontributor to the total building vapor load. In the research work for this article, I came across a number ofapproximate equations that are used to calculate the wa-ter vapor load from air flow. Approximate equations can be fairly accurate as long as the air conditions areclose to 70 F air temperature.
For more information on water vapor permeationloads, consult the June 1998 issue of HPAC magazine.The next few sections will compare approximate and exact equations for selected water vapor sources. HPAC
This article will continue in HPAC’s July issue. Part II willcover moisture from air leakage, water vapor load frommakeup air, and water vapor removed by dehumidifiers.
May 1999 • HPAC Heating/Piping/AirConditioning 95
acfm (cu ft per min) =total mass flow (lb per min)
density of moist gas (lb total per cu ft) (4)
scfm (std cu ft per min) =total mass flow (lb per min)
std density 0.075 (lb total per std cu ft) (5)
acfm (cu ft per min) = scfm (std cu ft per min)
0.075 (lb total per std cu ft)
density of moist gas (lb total per cu ft) (6)
Density of moist gas =MW wet mix (lb per mole) P (lbf per sq ft)
1545.43 lb per mole T( R) (7)
(lb total per cu ft)
m
f
×
×− × 7
By WILLIAM ACKER*, President,Acker & Associates, Green Bay, Wis.
This is the second article of a two-part series on in-dustrial dehumidification. The first article in theseries, which covered air flow, can be read inHPAC Engineering’s May 1999 issue.
WATER VAPOR LOADS DUE TO AIR FLOWWater 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 majorcontributor to the total building vapor load. In the researchwork for this article, I came across a number of approximateequations that are used to calculate the water vapor load
from air flow. Approximate equations can be fairly accurateas 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 nextfew sections will compare approximate and exact equationsfor selected water vapor sources.
MOISTURE FROM AIR LEAKAGEEquations (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 orfrom manuals prepared by dehumidification companies. No-tice that the engineering units do not properly cancel out,which is why they are considered approximate equations.
Where:Wo = The higher moisture level of the air outside the
IndustrialDehumidification:
Air Flow Diagrams & Water Vapor Load Equations
*William Acker is a member of HPAC Engineering’s Editorial AdvisoryBoard.
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.
PART TWO
Heating/Piping/AirConditioning • July 1999 59HPACENGINEERING
M(grains WV)
hrVelocity
ftmin hr
Area (sq ft)
d(lb wet air)
(cu ft wet air)grains WVlb dry air
(8)
= acfm(cu ft wet air)
minminhr
d(lb wet air)
(cu ft wet air)
(grains WVlb dry air
(9)
= acfm(cu ft wet air)
minminhr
(lb wet air)
(cu ft wet air)( )
grains WVlb dry air
(10)
M(grains WV)
hracfm
(cu ft wet air)min
(
= × ×
× × −
× ×
× −
×
× × −
= × ×
60
60
60
0 075
4 5
min
( )
)
.
.
W W
W W
W W
W
o i
o i
o i
o −− Wi)grains WVlb dry air
(11)
room that is entering with the air flow. It is an absolute hu-midity term. (ASHRAE calls this the Humidity Ratio.)
Wi = Moisture level of the air inside the room which inthis case is at a lower absolute humidity. (ASHRAE callsthis the Humidity Ratio.)
Is the density of moist air entering the room
Equation (12) can be used for the same calculations. It isan exact equation because the engineering units convert tograins of water vapor per hr.
Specific volume of the air flowing into the room. This value can be found in psychro-
metric charts or calculated. When using the charts, thecombined air barometric and static pressure should equal thetotal air pressure of the chart.
Equations (11) and (12) are applied and compared in theexample below to show how results from approximate andexact equations can vary under different conditions. Equa-tion (11), which is approximate, calculates a water vaporload 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 psychro-metric chart to obtain the entering air specific volume, or apsychrometric computer program that can calculate the airmixture properties.Example conditions:Room conditionsAir pressure: 29.921 in. HgDry bulb temp: 70 FMoisture level: 35 grains WV/lb dry airRelative humidity: 32.38 percent
Entering air flow conditionsAir pressure: 29.921 in. HgDry bulb temp: 120 FMoisture level: 420 grains WV/lb dry airRelative humidity: 76.44 percentSpecific volume: 16.02379 cu ft wet air/lb dry airAir flow acfm: 200 cu ft per min
Example calculations:
WATER VAPOR LOAD FROM MAKEUP AIREquation (12) can also be used to calculate the water va-
por load from makeup air. Let’s compare it to approximate
Equations (13) and (14). In this example, equations (13)and (14) produced water vapor loads that were 7.27 and12.63 percent above the exact equation (12) water vaporload. The error is a direct result of the assumed entering airspecific volume. Note that the makeup air specific volumewill vary with the entering air psychrometric properties.Therefore, you cannot select a standard value for specificvolume and expect the equation to be exact. For this reason,equations (13) and (14) are approximate equations. Equa-tions (13) and (14), which can be found in many engineer-ing books and dehumidification manuals, will be fairly accu-rate as long as the entering makeup air is close to theselected specific air volume.
Entering makeup airAir pressure: 29.921 in. HgDry bulb temperature: 100 FMoisture level: 280 grains WV/lb dry airRelative humidity: 93.67 percentSpecific volume: 15.01722 cu ft wet air/lb dry airAir flow acfm: 2000 cu ft per min
Example calculations:
WATER VAPOR REMOVED BY DEHUMIDIFIERSDehumidification systems remove water vapor from the
process air that travels through the unit. This section looksat the equations used to determine the humidity ratio of theprocess air entering and leaving the unit as well as equationsused to estimate the amount of water vapor removed whenthe inlet and discharge humidity ratios and air flow areknown. Note that some of the equations are the same as theequations used in preceding sections. The equations listed as
INDUSTRIAL DEHUMIDIF ICAT ION
M(grains WV)
hracfm
(cu ft wet air)min hr (cu ft wet air)
(lb dry air)
( )grains WVlb dry air
(12)
(cu ft wet air)(lb dry air)
= × ×
× −
=
601min
v
W W
v
o i
200 4 5 420 35
346 500
200 60
420 35 288 321
cu ft wet airmin
grains WVlb dry air
grains WV
hr (11)
cu ft wet airmin
minhr
lb dry air16.02379 cu ft wet air
grains WVlb dry air
grains WVhr
(12)
× × −
=
× ×
× − =
. ( )
,
( ) ,
M(grains WV)
hracfm
(cu ft wet air)min
minhr
lb dry air
14 cu ft wet air ( )
grains WVlb dry air
(13)
M(grains WV)
hracfm
(cu ft wet air)min hr
lb
cu ft(
grains WVlb dry air
(14)
M(grains WV)
hracfm
(cu ft wet air)min
minhr
(cu ft wet air)
(lb dry air)
( )grains WVlb dry air
(12)
= ×
× × −
= ×
× × −
= ×
× × −
60
60
0 075
60
1
W W
W W
vW W
o i
o i
o i
min
. )
2000 280 35 2 100 000
2000 35 2 205 000
2000 601
15 01722280 35
1 957 753
cfm 601
14 grains WV per hr (13)
cfm 60 0.075 (280 grains WV per hr (14)
cfm
grains WV per hr (12)
× × × − =
× × × − =
× × × −
=
( ) , ,
) , ,
.( )
, ,
HPACENGINEERING
60 July 1999 • Heating/Piping/AirConditioning
d(lb wet air)
(cu ft wet air)=
approximate can be very accurate if the process air flow tem-perature is close to 70 F (Table 1).
In this series of equations and calculations, approximateequations (15) and (11) produced water vapor removals thatwere 40 and 105 percent above the exact equations (16) and(17). The approximate equations can be fairly accurate ifthe air entering the dehumidifier is close to 70 F. Engineersthat desire greater accuracy can use the psychrometric chartto get the specific volume needed to make the conversionfrom acfm to dscfm or purchase psychrometric programs thatcan calculate the value for them.
Note: This equation uses the fan industry scfm flow enteringthe dehumidifier. The engineering units do not properly cancelout on the right side of the equation. This is why the equation accuracy is listed as approximate.
Note: This equation appears in some dehumidification publica-tions. Because the engineering units do not properly cancel out on theright side of the equation, it is considered an approximate equation.
Note: Uses the dehumidification industry dscfm flow enteringthe dehumidifier.
Note: This equation can be found in examples of theASHRAE Handbook of Fundamentals; in Modern Heatingand Ventilating Systems Design, by George E. Clifford; andin Fan Engineering, by Buffalo Forge.
CONCLUSIONMany of the flow diagrams presented in articles, books,
engineering manuals, and proposals from dehumidificationcompanies do not indicate the engineering units for theflows in the diagrams. As indicated in this article, the flowsare illustrated in cfm or scfm with no explanation. In mostcases, the flows are in dry standard cubic feet per minute.
Over the years, I have been contacted by many engineersover the issue of calculated water vapor load variances fromdifferent equations. In most cases, the approximate equa-tions are equations that have been shortened to make thecalculations easier for engineers. If you have any questionson your equations, check the engineering units to make surethat they properly cancel out to grains of water vapor per hr,or lb of water vapor per hr.
BIBLIOGRAPHY1) 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 Engineer-ing, March 1999.
3) Clifford, G. E., Modern Heating and Ventilating SystemsDesign, 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, AmericanSociety of Heating, Refrigerating, and Air-ConditioningEngineers, Inc., Atlanta, Ga.
The author would like to thank his wife, Sandra, for her patience and as-sistance during the preparation of these articles. He would also like tothank Nels Strand, the author’s mentor and close friend for over 20years.
Mr. Acker has 25 years of experience in industrial HVAC. He has de-veloped many computer programs, which include psychrometrics andthermodynamics of air flows, water vapor permeation, and condensationanalysis of walls and roofs, ductwork heat loss and heat gain for air flowand flue gas flow, and boiler efficiency and boiler emissions to mention afew. Questions or comments about this article may be directed to the author at 920-465-3548.
HPAC
TABLE 1 Into dehumidifier Dehumidifier discharge
Air pressure 29.921 in. Hg 29.921 in. Hg
Air temperature 200 F
Relative humidity 50 percent
Humidity ratio 280 grains WV 35 grains WV per lb dry airper 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
Dry air mass flow 87,719.78 lb dry air per hr 87,719.78 lb dry air per hr
Water vapor flow 35,197.64 lb water vapor per hr
Fan industry scfm 27,315 std cu ft per min
Dehumidification 19,493 dry std cu ft industry dscfm per min 19,493 dry std cu ft per min
M(grains WV)
hrscfm
(std cu ft air)min
minhr
lb wet air
std cu ft air( )
grains WVlb dry air
(15)
= 27,314.98 60 0.075 (2809
= ×
× × −
× × × −=
60
0 075
35
340 972 895
.
)
, ,
W Wo i
M(grains WV)
hracfm
(cu ft wet air)min
minhr
lb wet air
cu ft std air( )
grains WVlb dry air
(10)
M(grains WV)
hracfm
(cu ft wet air)min
( )grains WVlb dry air
(11)
= ×
× × −
= ×
× −
= × × −=
60
0 075
4 5
40 000 4 5 2809 35
499 320 000
.
.
, . ( )
, ,
W W
W W
o i
o i
lb dry air
dry std cu ft( )
grains WVlb dry air
(16)
= , . (
× × −
× × × −=
0 075
19 493 60 0 075 2809 35
243 334 670
.
)
, ,
W Wo i
M(grains WV)
hrdscfm
(dry std cu ft)min
60minhr
= ×
M(grains WV)
hrM
(lb dry air)hr
( )grains WVlb dry air
(17)
= 87,719.78lb dry air
hr
= × −
× −
=
W Wo i
( )
, ,
2809 35
243 334 670
Heating/Piping/AirConditioning • July 1999 61HPACENGINEERING
