Industrial Dehumidification

7/31/2019 Industrial Dehumidification

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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 also

improves 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, adiabatic

mixing, 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 1Types 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 39.4535 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) 12.89

25 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


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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 exchangersto

name 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 4.35

lb per hr; 1.88 and 12.50gal 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 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: Adsorbentswhich do not

experience a phase change. Mois-ture is deposited on the surface ofthe dry desiccant. Most adsor-bents are solids. Absorbentswhich change

physically, chemically, or bothduring the sorption process. Mostabsorbents are liquids or solidsthat become liquid as they absorbmoisture.


activation heat, heat dump back,and desiccant heat.

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 2Capacity 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 3Capacity 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


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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 due

to 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 sensor

shuts 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 adetailed 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 water

vapor removed. In some cases, Ileft out proprietary information

on 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 ft

per 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 brine

coil 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-


5

4

3 21

SP = 2.79 in. WC21.97 bhp

P = 0.76 in. WC Heat addition437,964 (Btu per hr)

Processair outlet

DXcooling

coil

Heat removal538,617

(Btu per hr)(44.88 ton)

Heatpipeexchanger

Heat transferefficiency = 81.31 percent

Filter Damper

Processair inlet(moist)

P = 0.72 in. WCP = 0.45 in. WC P = 0.08 in. WC

Waterremoval380.24(lb per hr)

Warm

freongas

35Fliquidfreon

Exhaust air

Air inlet Air inletAir-cooledcondenser

DX chiller withair-cooled condenser(1.5405 KW per ton) at 95 F

Highpressure

Refrigerationgas

Hotgas

Low pressure

Compressor

1 DX with heat pipe dehumidifier.

Process airanaylsis

Dry 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)Q sensible (Btu per hr): 437,075Q latent (Btu per hr): 416,688Q total (Btu per hr): 853,763 (71.147 ton)Water vapor removed (lb per hr): 380.24


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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 chilled

water 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 greaterdehumidification.

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 evaporator

coil, compressor, condenser, andthrottling valve (or expansion

valve). 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 expansion

valve. As the liquid passes throughthe valve, its pressure is suddenly

decreased 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


Reactivationblower (exhaust)SP =2.3 in.

(Dehumidification)+ 1.6 in. (Burner)

= 3.9 in. WC

Process air bypass blowerSP = 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 bhp

Reactivationair inlet

Process air

inlet (moist)

9.52 bhp

Desiccantwheel

Process airoutlet

Process air blowerSP = 2.0 in.

(Dehumdification)+ 0.5 in. (Filter) = 2.5 in. WC

Drive0.25 bhp

2.59 bhp

2 Rotating dry desiccantdehumidifier.

Process airanaylsis

Dry 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 Dry 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 4274.78

M (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


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$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 use

the 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 addedcost 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 vaporremovedat 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 the

desiccant 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. The

vapor-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 (sensible

heat) is energy that is passed to


TABLE 5Types of drysolid adsorbents.

1) Silica gel2) Titanium gel3) Dry lithium chloride4) Zeolites5) Synthetic zeolites (molecular sieves)6) Activated alumina

7) Synthetic polymers

TABLE 4DX 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 hr

2) 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 dischargeof 56.3 F and36 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 6Dry desiccant dehumidification operating costto 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 WV

7) 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 WV8) 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 WV

With a process air discharge of 93.6 Fand 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


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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. The

desiccant 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 the

vapor 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 cost

increases 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


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 watervaporper lb of dry air

Wpo = amount of moisture in the process airleaving desiccant wheel, grains of water

vapor per lb of dry airWri = amount of moisture in the reactivation air

(after the heater) entering the desiccant

wheel, grains of water vapor per lb of dry airWro = amount of moisture in the reactivation air

leaving the desiccant wheel, grains of watervapor per lb of dry air

tro = heated reactivation air temperatureentering the desiccant wheel (equalstemperature of reactivation air leaving theheater)

tri = reactivation air temperature entering theheater

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.

Qs = dscfmp 1.08 [0.625 (Wpi - Wpo]= dscfmp 0.675 (Wpi - Wpo)

Reactivation air heaterheat addition (approximate)QRA = dscfmr 1.08 (tro - tri)

Reactivation system heat leakage passedto process air (approximate)

This is desiccant heat from the reactivationsection passing heat to the process air.

QRL = dscfmp 1.08 k (tro - tpi)

Process airdischarge temperature (approximate)tpo = tpi + 0.750 (Wpi - Wpo) + k (tro - tpi)

Reactivation air flow (approximate)dscfmr = dscfmp (tpo - tpi)/(tro - (tri)

Reactivation outlet moisture content (approximate)Wro = Wri + (dscfmp/dscfmr)(Wpi - Wpo)

Process air through the desiccant wheelenergy loses and gainsQ sensible heat gain = heat of sorption + reac-

tivation system leakageQ latent heat loss = heat of sorptionQ total = Q sensible + Q latentQ total = (heat of sorption + reactivation

system leakage) - heat of sorption = reactivationsystem leakage

Note: Heat of sorption is sometimes calledheat of condensation


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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 and

desiccant 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 to

remove 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.

Liquid desiccant dehumidifierLiquid desiccant dehumidifica-

tion operates on the principal ofchemical absorption of water va-por from the air. The absorbent ordesiccant solution will changephysically, chemically, or both

during the sorption process. Someof the liquid desiccant solutionsused for dehumidification are: Lithium chloride (LiCl) Lithium bromide (LiBr) Calcium chloride (CaCl2) Triethylene glycol (TEG) Propylene glycolLiquid absorption dehumidifi-

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 a

vapor 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 water

vapor (or add water vapor) is de-termined by the temperature and

concentration 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 very

close to the 80 percent relative hu-midity line. Therefore, it can besaid that higher solution concen-

trations give lower equilibriumrelative humidity and thus allow

the 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 the

desired 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 than

chilled 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 7Liquid desiccant dehumidificationoperating cost to remove 1000 lb of water vapor

Operating cost Electric usage

1) Hot water boilernatural 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 hr

7) 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



9/16


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 bulb

temperature, 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 heatload added by the regenerationprocess.

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 its

vapo r pr essure is still a littlehighuntil 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) as

well 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 operatingenergy 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 warm

desiccant 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 F

discharge air temperature).The outcome of this study will

change, 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 KW H. 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


TABLE 8Dehumidifier 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 wheeldirect-fired 93.6 F 35 $15.87 per 1000 lb WV removed4) Dry desiccant rotating wheeldirect-fired with added cooling 55.0 F 35 $34.24 per 1000 lb WV removed5) Dry desiccant vertical beddirect-fired 89.5 F 35 $14.55 per 1000 lb WV removed6) Liquid desiccantwith electric air-cooled chiller 55.0 F 35 $28.43 per 1000 lb WV removed7) Liquid desiccantwith 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


10/16

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, heat

wheels, 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 reactivation

air 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,Heat ing/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.



11/16

B y W IL L IA M A C K ER * ,P r e s i d e n t ,A c k e r & A ss o c i at e s ,G r e e n B ay , W i s.

In de s i gn i ng de hu mi d i f i c a t i on sys t e ms , on e o f t h emost importan t t a sks i s to quan t i fy the wa te r vaporl o a d s t h a t m u s t b e r e m o v e d b y t h e s y st e m . T w oqua l i fi e d i nd i v i dua l s ma y a r r i ve a t d i f fe re n t t o t a l

m o i s t u r e l o a d s fo r t h e s a m e b u i l d i n g . S o m e o f t h e s e

d i f fe r e n c e s o c c u r f r o m a b b r e v i a t e d o r a p p r o x i m a t eequa t ions tha t were developed to make th e ca lcula t ionse a s ie r . H o w e v e r , t h e s e a p p r o x i m a t e e q u a t i o n s lo s eaccuracy when air temperatures are higher or lower th anthe condi t ions assumed when the approximate equa t ionswere deve loped. T his a r t ic le wi l l c la r i fy some of thesee q u a t i o n s a n d p r e se n t a l t e r n a t i v e e q u a t i o n s t h a t a r e

more precise across a larger range ofcondi t ions .

T h is article will also explain th e airf lows use d in p roposa l s f rom de hu-midificat ion compan ies as well as thep ropose d f l ow d i a gra ms f rom de h u-m i d i f ic a t i o n e q u i p m e n t s u p p l ie r s.

I ha ve d i s c ove red a g re a t a moun t o f uncerta in ty over th e a ir flows in t hesediagrams.

DEHUMIDIFICATION INDUSTRYEQUATIONS AND TERMS

D e h u m i d i f ic a t i o n m a n u f a ct u r e r sl i ke t o d e ve l op a i r f l ow d i a gra ms o f

May 1999 HPAC Heating/Piping/AirConditioning 93

IndustrialDehumidification:A ir Flow D iagrams &

Water Vapor Load Equations

* W i ll ia m A c k e r i s a m e m b e r o f H P A C s

Editorial A dvisory 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

Systemfan

Coolingcoil

Makeup air

2400 cfm50 F/49 F50 GR

22,315 cfm55 F

20 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.


12/16

th eir en t ire system s. Mass flow analysis is used in th esediagrams because mass flow does not change if there is atemperature or pressure change. Mass flows can also beadded an d subtracted. Th e actual a ir flows or acfm (cu ftp e r m i n ) w i l l c h a n g e i f t h e t e m p e r a t u r e o r p r e ss u r e

c h a n g e s. A l so , a c fm v a l u e s c a n n o t b e a d d e d o r s u b-t rac ted b ecause th ey a re a t d i f fe rent a i r den s i t i e s. Th ed e h u m i d i f i c a t i o n i n d u s t r y c h o s e a t y p e o f m a s s f lo wanalysis that is flow in dry standard cu ft per min (dscfm)at a com mon air densi ty of 0.075 lb dry air per dry stan-dard cu f t . Th e ac fm flows a re conver t ed to t hese f lowsfor i llustra t ion in t he a ir system diagrams (shown in Fig.1 and Fig. 2).

A s men tion ed earlier, dscfm air flow is a flow at a com-mon air density of 0.075 lb dry air per dry stan dard cu ft. I

prefer to use the term dscfm to avoid any possible confu-s ion wi th ac fm or t he fan in dus t ry te rm sc fm. In F ig. 1 ,

one m anufacturer uses cfm to represent th e dry stan dardair flow. Wh en you look at t his diagram, i t is easy to te l lth at th e flows are dry stan dard air flows and n ot acfm. Ifyou look a t the f low in an d out of th e dehumidi f ie r , theflows are both 7500 cfm. T h is is n ot po ssible with acfmbe c a use t h i s fl ow i s ma de o f bo t h a i r a nd wa t e r va po r ;th erefore, there is a loss (of water vapor or cfm) as it trav-els th rough th e deh umidifier. Listed below is an exampleof an acfm flow broken down into dry air flow and watervapor flow:

A ir pressure = 29.921 in. H gA ir tempera ture = 70 FRelat ive hum idity = 100 percentT otal a ir flow = 100,000 acfm

Dry air flow = 97,530 cfmW ater vapor flow = 2470 cfmA lso, note t hat t here is a chan ge in t emperature across

th e dehum idi fie r tha t would a l so cause a chan ge in t heacfm flow.

Fig. 2 i s a d iagram p repared by ano th e r deh umidi f ie rman ufacturer. T he air flows are in dscfm, but in t his case ,th ey use th e term scfm. Th is scfm should n ot be con fusedwith th e fan ind ustry scfm, which is a type of mass flowth at represent s the combin ed air and water vapor flow. Bylooking at Fig. 2, you can see that th e flow into t he deh u-

midifier (or con dit ion er) 20,833 scfm is th e same as theflow leavin g; therefore , th is is a dry a ir mass flow. If th een terin g flow is fan ind ustry 20,833 scfm, th e leaving flowwould have dropped to 20 ,754 scfm due t o th e removal ofwater vapor (79 scfm or 355.41 lb of water vapor per hr) .

T he n ext sect ion on fan in dustry scfm will help you to un-ders tand th e d i fference be tween fan and deh umidi f ica -t ion scfm terms.

Th e equa t ions used by the deh umidi fica t ion indus t ryto con vert t he act ual air flows to dry stand ard air flows arelisted b elow:

W he re :

T his is th e dry stan dard air flowuse d by t h e de hum i d i fi c a t i onindustry.

T his is th e actual air flow of dryair and water vapor.

T his is the a ir-specific volume.T he value can be found in psy-c h r o m e t r i c ch a r t s p r o v i d e dtha t th e ba romet r ic pressure of the ch ar t i s the same as the to-t a l p r e s su r e ( b a r o m e t r i c p l u ss t a t i c p r e s s u r e s ) o f t h e a c f mflow.

T his is th e standard air densi tyfor the deh umidificat ion indus-

try. Th e inverse of this value is13.3333.

Th is i s th e to ta l amount of dryair m ass flow.

FAN INDUSTRY EQUATIONS AND TERM SFans must be selected based on th e actual flow rate an d

th e actual density a t the inle t to t he fan. Some fan m anu-facturers prefer to specify th e flow rate based on stand ardin le t con di t ions . Fan pe rformance curves a re deve loped

94 HPAC Heating/Piping/AirConditioning May 1999

I N D U S T R I A L D E H U M I D I F I C A T I O N

4,200 acfmTo outside

19,833 scfm39 GR per lb.

75 F

95 F

1000 scfm118 GR per lb.

20,833 scfm42.8 GR per lb.

76 F

Conditioner 20,833 scfm16.1 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)

mi n

dry air mass flow(lb dry air)

mi n

.lb dry air

dry std cu ft

(1)

acfm(cu ft wet air)

mi n

. lb dry airdry std cu ft

(cu ft wet air)lb dry air

(2 )

acfm(cu ft wet air)

mi n.

dry std cu ft

lb dry air

(cu ft wet air)

lb dry air

(3)

=

=

=

0 0 75

0 0 75

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 0 75.lb dry air

dry std cu ft=

Dry air mass flow(lb dry air)

min=


13/16

(from a series of laboratory tests) at th ese stan dard con di-t ions. Listed below are th e equations th at a l low designerst o c o n v e r t fr o m t h e a c t u a l c o n d i t i on s t o t h e s t a n d a r dcondit ions:

W he re :

acfm ( cu ft per min) =A ctual cubic feet per minut e . It represent s the volume

of dry gas and water vapo r flowing at a specified point in asystem. In fan sizing, this would be the flow entering thefan.

Den sity of m oist gas ( lb t otal per cu ft ) =Th e rat io of the mass of a substance to i ts volume. Th e

fan gas densi ty or fan air densi ty is the t ota l densi ty a t th efan in le t .

Stan dard density ( lb total per std cu ft) =S ome fa n m a nufa c t u re r s l ike t o de ve l op fa n p e r fo r -

man ce curves based on a s tand ard gas or a i r dens i ty of 0 .075 lb to ta l pe r std cu f t . The A ir Movement an d C on-t ro l Assoc ia t i on (A MC A ) i n d i ca t e s t ha t t h i s de ns it y i ssubstan t ia l ly equiva lent to a i r a t a t em pera ture of 68 F ,50 pe rcent R H , and a pressure of 29.92 in . of mercury.

scfm ( std cu ft per min) =Th e standard gas or a ir flow rate entering th e fan at an

inlet den sity of 0.075 lb tota l per std cu ft .M W wet m ix ( lbm per mole) =T he m olecula r we ight of the dry gas (or dry a i r ) and

water vapor mixture .P ( lbf per sq ft) =Th e total pressure a t the in le t to th e fan. T he pressure

represent s th e baromet ric pressure plus th e stat ic pressure(or gauge pressure).

154 5. 43 ft - lbf per mole - R =U niversal gas constant .T ( R ) =A bsolute gas or a ir temp erature a t th e inle t to th e fan

in degrees Rankin e ( R = F + 459.67).

DSCFM FLOW VERSUS SCFM FLOWTh e fo llowing examples show how th e dehu midi fica -

t i on i ndus t ry d sc fm fl ow c ompa re s t o t h e f a n i ndus t ryscfm flow.

Example 1:A ir pressure = 29.921 in. HgA ir tempera ture = 70 FRelat ive hum idity = 100 percent

H umidity ra t io = 109.93 grains WV per lb dry airMoist a ir density = 0.074190 lb wet a ir per cu ft wet a irSpecific volum e = 13.6906 cu ft wet air per lb dry airacfm flow = 40,000 cu ft wet air per minDeh umidificat ion indu stry dscfm = 38,956 dry std cu ft

per minFan industry scfm = 39,568 std cu ft per min

Example 2:A ir pressure = 29.921 in. HgA ir tempera ture = 200 FRelat ive hum idity = 50 percentH umidity rat io = 2809 grains WV per lb dry a irMoist a ir density = 0.051216 lb wet a ir per cu ft wet a irSpecific volume = 27.3599 cu ft wet air per lb dry airacfm flow = 40,000 cu ft wet air per minDeh umidificat ion industry dscfm = 19,493 dry std cu ftper minFan industry scfm = 27,315 std cu ft per min

Example 1 shows th a t a t 70 F , th e dsc fm f low i s ve ryclose to th e fan in dustry scfm flow (on ly 1.6 percent vari-a t ion) . However, a t more e levated temperatures, such asin Exam ple 2 , th e fan in dus t ry sc fm flow i s 40 pe rcen thigher th an t he dscfm flow.

WATER VAPOR LOADS DUE TO AIR FLOWW ater vapor loads on indus t r ia l bui ld ings come from

man y sources. Listed below are a few of these sources ofwater vapor: People Permeation th rough walls, roofs, an d floors Moisture from products and packaging mat eria ls

Evaporat ion from open tan ks or wet surfaces Product dryer leakage O pen combust ion From air flow

A ir leakage through cracks and h oles A ir leakage through conv eyor openin gs In te rmi t tent door openings Building-to-building air infiltration Makeup air

In m any cases, water vapo r loads by air flow are a majorc o n t r i b u t o r t o t h e t o t a l b u i ld i n g v a p o r lo a d . I n t h eresearch work for th is art ic le , I came across a numb er ofapproximate equations th at are used to calculate th e wa-t e r v a p o r l o a d f r o m a i r fl o w . A p p r o x i m a t e e q u a t i o n s

can b e fa ir ly accura te a s lon g as th e a i r condi t ion s a reclose to 70 F air temperature .

F o r m o r e i n fo r m a t i o n o n w a t e r v a p o r p e r m e a t i o nloads , consul t th e June 1998 i ssue of H PA C magaz ine .T h e n e x t fe w se c t i o n s w i ll c o m p a r e a p p r o x i m a t e a n dexact equ ation s for selected water vap or sources. HPAC

T his article w ill continu e in H P A C s July issue. Part II w illcover 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) =tot al mass flow (lb per min)

density of moist gas (lb total per cu ft)(4)

scfm (std cu ft per min) =tot al mass flow (lb per min)

std density 0.075 (lb to tal 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 c u ft)(6)

Den sity of moist gas =M W w et m ix ( lb p er m ol e) P ( lb f p er sq ft )

1545.43 lb per mole T ( R )(7)

(lb total per cu ft)

m

f


14/16

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 Engineerings 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 ofwater vapor: People Permeation through walls, roofs, and floorsMoisture from products and packaging materials Evaporation from open tanks or wet surfaces

Product dryer leakageOpen combustionAir 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 Engineerings 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.

PART TWO

Heating/Piping/AirConditioning July 1999 59HPACENGINEERING

M(grains WV)

hr

Velocityft

min hr

Area (sq ft)

d(lb wet air)

(cu ft wet air)

grains WV

lb dry air(8)

=acfm(cu ft wet air)

min

min

hrd

(lb wet air)

(cu ft wet air)

(grains WV

lb dry air(9)

=acfm(cu ft wet air)

min

min

hr

(lb wet air)

(cu ft wet air)( )

grains WV

lb 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 WV

lb dry air(11)


15/16

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 calls

this 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 theroom. 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 a

psychrometric 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 percent

Specific 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. Lets 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 air

specific 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.

Example conditions:Inside room conditions

Air pressure: 29.921 in. HgDry bulb temperature: 70 FMoisture level: 35 grains WV/lb dry airRelative humidity: 32.38 percent

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

I N DU S T R I AL DEHU MI DI F I C AT I ON

M(grains WV)

hracfm

(cu ft wet air)

min hr (cu ft wet air)

(lb dry air)

( )grains WV

lb 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 air

min

grains WV

lb dry air

grains WV

hr(11)

cu ft wet air

min

min

hr

lb dry air

16.02379 cu ft wet air

grains WV

lb dry air

grains WV

hr(12)

=

=

. ( )

,

( ) ,

M(grains WV)

hracfm

(cu ft wet air)

min

min

hr

lb dry air

14 cu ft wet air( )

grains WV

lb dry air(13)

M(grains WV)

hracfm

(cu ft wet air)

min hr

lb

cu ft(

grains WV

lb dry air(14)

M(grains WV)

hracfm

(cu ft wet air)

min

min

hr

(cu ft wet air)

(lb dry air)

( )grains WV

lb dry air(12)

=

=

=

60

60

0 075

60

1

W W

W W

v

W 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 60

1

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)=


16/16

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 equationaccuracy 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.

Not e: Th is equ at ion can be fou nd in ex ample s 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 sure

that 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 authors 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 flow

and flue gas flow, and boiler efficiency and boiler emissions to mention afew. Questions or comments about this article may be directed to theauthor 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 airper cu ft wet air

Specific volume 27.3599 cu ft wet airper lb dry air

acfm flow 40,000 cu ft wet airper 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 watervapor 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

min

hr

lb wet air

std cu ft air( )

grains WV

lb dry air(15)

= 27, 314.98 60 0.075 (2809

=

=

60

0 075

35

340 972 895

.

)

, ,

W Wo i

M(grains WV)

hr

acfm(cu ft wet air)

min

min

hr

lb wet air

cu ft std air( )

grains WV

lb dry air(10)

M(grains WV)

hracfm

(cu ft wet air)

min

( )grains WV

lb 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 WV

lb 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)

min60

min

hr=

M(grains WV)

hrM

(lb dry air)

hr( )

grains WV

lb dry air(17)

= 87,719.78lb dry air

hr

=

=

W Wo i

( )

, ,

2809 35

243 334 670

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