VA-89-21-4
Two-Stage Desiccant Dehumidification
in Commercial Building HVAC Systems
G. Meckler, P.E. Member ASHRAE
ABSTRACT
A two-stage desiccant dehumidifier has been developed that significantly increases the thermal coefficient of performance (COP) of desiccant regeneration in commercial applications. The first stage is a desiccantimpregnated enthalpy exchange wheel that accomplishes 30% to 50% of the dehumidification task without requiring external, utility-generated heat to reconstitute the desiccant.
Described are the advantages of desiccant dehumidification for commercial facilities, the heat requirement that has made it difficult to develop costeffective commercial applications, and the two-stage desiccant system's contribution to thermal efficiency. Also described are alternative HVAC systems incorporating two-stage desiccant dehumidification for both new and existing commercial facilities such as office buildings, supermarkets, and restaurants; and the system arrangement, dehumidification process, and regeneration heat requirement and sources in an actual application in a twostory, 90,000 ft2 building in Washington, DC.
INTRODUCTION
Air-conditioning includes two basic functionsmoisture control and temperature control (dehumidification and sensible cooling) . Dehumidification typically represents 20% to 40% of the total air-conditioning load on a building's refrigeration system.
In conventional HVAC systems, the two functions are combined: vapor compression machines powered by onpeak utility electricity chill the air deeply to condense out adequate moisture. In terms of the sensible cooling requirement only, this process overcools in many systems, thereby consuming more utility energy than would be necessary for temperature control alone In these conventional systems. the "overcooled" air must be reheated in some way prior to distribution.
The use of a moisture-absorbing desiccant to dehu· midify the air, in place of deep chilling for condensation, creates the following advantages and opportunities
1. Reduces the electric utility demand for refrigeration by 36% to 52% based on:
(a) 20% to 40% reduction due to elim1nat1ng tile con-
densation task and shifting dehumidification to a desiccant system: and
(b) 20% reduction (in many applications) in the energy required for the remaining task-sensible coolingbecause it now can be done at an appropriate higher temperature, 58°F instead of 42°F
2. Desiccant dehumidification dries the air more deeply than is practical with conventional refrigeration systems. Therefore, a smaller quantity of superdry primary air can provide 100% of a facility 's dehumidification requirement. This makes it possible to do the following:
(a) distribute 0.1 to 0.3 cfm/ft2 of superdry primary/ ventilation air from the central plant (providing sensible cooling separately);
(b) reduce the size of the primary air distribution system by 70% to 80% and the cost by 40% to 50%:
(c) apply this significant fitst-cost saving to desiccant equipment; and
(d) also reduce substantially the fan energy required for primary air distribution.
3. Increases HVAC design options and flexibility. Separating the two functions of dehumidification and sensible cooling makes it possible to handle each task optimally, in terms of energy sources. distribution, etc .. for the specific facility and location. (Design options and integrated system examples are presented in subsequent sections.)
HEAT REQUIREMENT OF DESICCANT DEHUMIDIFICATION
Desiccant systems reduce a facility"s electric utility demand for refrigeration, but they also require heat to regenerate the desiccant. In past years desiccant regeneration was a thermally inefficient process. with a thermal COP in the range of 0.4 to 0.7. In addition. systems were configured to use high-temperature heat. usually steam, and were installed primarily in industrial and special· purpose applications where low humidity and stringent humidity control were essential. Because of the low thermal COP and high-temperature regeneration requirement in previous desiccant systems. desiccant dehumidification was not cost·compet1t1ve with conventional vapor com· oression HVAC systems in commercial facilities.
G. Meckler is President. Gershon Meckler Associates. P.C. , Herndon, VA.
THIS PREPRINT IS FOR DISCUSSION PURPOSES ONLY. FOR INCLUSION IN ASH RAE TRANSACTIONS 1989: V 95. Pl 2 No110 oe reprinled in whole or in parl wilhoul wri11en permission of lhe American Soc1e1y ol He31_, ng. Relrigerating and Air·Cond111 on•ng Engineers. Inc 1791 Tull1e Circle. NE Ailanla GA 30329 Opinions. find ing s. conclus ions or recommenda11ons e•or~ ssed 1n 1n1s paper are lhose of lhc au1norrs1 and do nol necessa,,ly rel/eel lhe views ol ASH RAE •
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GENERATED ELECTRICITY POWERS MOST OF HVAC
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RECOVERED EXHAUST HEAT (2 .5 BTUH/SF)
f::. RECOVERED JACKET WATER HEAT
COOLING TOWER
PUMP
GAS HX
COGENERATION ENGINE/ GENERATOR
HX OUTSIDE AIR
(0 .1 CFM / SF)
RELIEF AIR (0. 1 CFM / SF)
(5 .5 BTUH / SF)
DUMP COIL
SUPPLY
DESICCANT DEHUMIDIFIER (4 BTUH/SF)
I REGENERATION COIL (8 BTUH/SF)
REGENERATION AIR INTAKE
IND. £YAP. COOLER (4 BTUH/SF)
HX ENTHALPY /
EXCHANGER RETURN
PRIMARY AIR (0.3 CFM/SF')
(CLG. CAP. = 3 BTUH/SF' LAT. ) PUMP
ROOF LOAD TO PLEN .= 5 UNITARY HEAT PUMPS (CLC. CAP.= 28 BTUH/SF)
POWER = 2.4 W/SF) r--""""i'i'"""":;:;..._ __ LTG. LOAD TO PLEN.= 7
FIRE ANNUNCIATOR
PLEN. LOAD TO SPACE= (3) PLEN . LOAD, BTUH / SF 9
FIRE HOSE STATION
(TYP.)
FIRE WATER SPRINKLER HEADS
INTEGRATED
t
( 0. 8 CF M /S F") ri:i::==4J.I-~~~ SUPPLY AIR
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PLENUM LOAD TO SPACE = 3 SPACE AT 75 FDB/42?. RH LIGHTING LOAD TO SPACE = 10
FIRE ALARM VALVE OCCUPANT LOAD = .3SENS.+.3LAT.= __§_ U.S. PATENT NO. 4.7'2J.417 & PATENTS PENDING TOTAL SPACE LOAD, BTUH/SF' 19
Key:
1. Outside air entering enthalpy wheel, 8B"FDB, 7 4"FWB. 2. Outside air leaving enthalpy wheel. 86"FOB, 67"FWB 3. Mixture of outside and return air, 85'F'D8, 65"FW8. 4. Primary air leaving supply fan, 87"FDB. 65"FWB. 5. Primary air leaving desiccant wheel. 99"FDB. 65"FWB. 6. Primary air leaving indirect evop . cooler,
85"FD8. 61"FW8. 7. Plenum air, 85"FD8, 64"FWB. B. Mixture of primary air and plenum air in
terminal unit, 85"1"08, 63"FWB. 9. Supply air leaving teminal unit.
61"FDB. 54"FWB. I 0 . Space condition, 75"FDB, 61 'FWB. 427.RH.
ao.,__--------.
JO 50 60 70 80 0
90 100 TEMP DEC. F
Figure 1 Two-stage desiccant dehumidification HVAC system for office buildings and shopping centers
In actual commercial and institutional applications designed by the author, the energy efficiency of desiccant regeneration has been improved significantly in two ways. resulting in cost-effective desiccant systems for commercial buildings
1. The systems have been configured to use relativeiy low-temperature heat for regeneration. Usable energy sources in these systems include recovered waste heat. cogenerated heat . and low-temoe1 atu rc inermal storage (130r: - 140°F) heated by solar e:1ergy or off-;)eak electricity
2 1':... two-stage desicca nt de!lu:rnd1f1cat1on!re~:wnera-11on process has been deve:oped an(i ?..l)Oi1ecJ wil1ch
significantly reduces the quantity of external heat required
TWO-STAGE DESICCANT DEHUMIDIFICATION/REGENERATION
The two-stage system. shown schematically in Figure 1. includes two rotating desiccant-impregnated wt1eels The first-stage wheel is an enthalpy exchanger that 11andies 30% to 50°10 of the building's dehumidification :ask without the need ior external heat to regenerate the rJes1ccant Tl1is wheel absorbs both heat and moisture I rom :11e 111corning outside airstream and transfers tl:ern to ttie clrier exhaust a 1rstroam
As a result. the second wheel. which completes the dehumidification process. has a lighter task of moisture removal and requires 30% to 50% less external heat for desiccant regeneration than a one-stage system. Based on the following configuration. the thermal COP of the twowheel regeneration process is 1.5 to 2.0 at design conditions_ In the specific application described subsequently, the thermal COP at design conditions is 1.75.
Dehumidification and regeneration occur as follows. Each desiccant-impregnated wheel rotates through two separate airstreams: the moist. incoming outside airstream and the regeneration airstream that is exhausted to the outside. As the wheels rotate, the desiccant absorbs moisture from the outside airstream and then gives it up to the regeneration airstream.
In first-stage regeneration, the outgoing (regeneration) air is the relatively dry building relief air. The second-stage regeneration airstream may be either relief air or moist outside air; the air is heated prior to flowing through the regeneration chamber. The desiccant in the second-stage dehumidifier is more concentrated than in the first-stage wheel .
The desiccant used in salt form is lithium chloride, which is a nontoxic, bactericidal, inorganic material often used in hospitals.
An additional benefit of the two-wheel desiccant system is that rain or a very humid outside condition does not cause supersaturation in the second wheel; in a onewheel system, such saturation can significantly impair performance and require a much higher temperature regeneration heat.
ALTERNATIVE TWO-STAGE DESICCANT HVAC SYSTEMS
Fully integrated, cost-effective HVAC systems incorporating two-stage desiccant dehumidification have been developed by the author for two commercial building categories:
1. Office buildings, shopping centers, institutional facilities
- New construction - Retrofit of existing building system
2. Supermarkets - New construction - Retrofit of existing store system
Office Building/Shopping Center Applications
In office building and shopping center applications. only outside air is dehumidified in the desiccant conditioner, as shown in Figure 1. In these medium- and largesize buildings a principal goal is to minimize primary air distribution to reduce the size and cost of the primary air handler and distribution ductwork.
The incoming outside air is dried deeply in the twostage system to 33 to 40 gr/lb (grains of moisture per pound of dry air) . The resulting superdry condition of the air permits a reduction 1n the quantity of primary air that must be distribuled throughout the building to 0.1 to 0.4 cfm!W or the m1r11mum established by codes. This quantity handles 100% of the building's ventilation and dehum1dif1cat1on requirements.
3
The desiccant-dried primary air 1s distributed in variable volume (determined by space conditions) to faninduct1on terminal units that incorporate either unitary heat pumps or chilled water coils to provide sensible cooling.
Medium-size Facilities. In medium-size commercial buildings the superdry primary air is distributed to faninduction terminals that incorporate high-efficiency unitary heat pumps. These UHPs have been specially developed to provide sensible cooling only, not dehumidification. As a result, they cool at a higher temperature level and require on the order of 35% less utility energy than conventional UHPs_
Superdry primary air is mixed in the terminals with faninduced room air, then cooled as required and supplied at a constant volume to occupied areas.
UHPs are small air conditioners that provide a separate, self-contained cooling and heating capability at each terminal . Some UHPs can be in the cooling mode while others are heating or off . Where there are different tenants, diverse activities. or differing hours in a building. UH Ps permit separate cooling without activating a central chiller. UHPs are also desirable where equipment room space is extremely limited _
Terminal UHPs are joined by a closed water loop that is maintained at 70°to 90°F UHPsdraw heat from or reject heat to the water loop. When the temperature of the water loop exceeds 80° to 90°F. heat is re1ected via the cooling tower. When the temperature drops below 70°F, the water loop draws heat from the hot thermal storage tank.
With a terminal unitary heat pump system, there is no need for both hot and chilled water piping. To save piping costs. these systems use the integrated sprinkler piping to circulate water between the UHP water loop and the central plant.
Large Facilities. When the building is very largesay, more than 500,000 ft2 -terminals incorporate chilled water coils instead of UHPs. Either a gas absorption chiller or vapor compression chiller may be used to provide sensible cooling. The chiller operates at an energy-efficient temperature level producing 55°F chilled water. Chilled water is distributed via integrated sprinkler piping to the fan-induction coil terminals.
Retrofit of Office Building Unitary Heat Pump System. UHP systems as currently configured are highly energy intensive. The addition of a two-stage desiccant dehumidification system can reduce the utility electricity required for air-conditioning by up to 50%. The percentage varies by region, based on the humidity level. In the retrofitted system. there will no longer be condensation in the UHP terminals.
Source of Regeneration Heat. To provide heat for second-stage desiccant regeneration. these systems incorporate either a gas-energized heat source (cogenerator or direct-fired heater) or an electric heat source (off-peak resistance heater or electric heat pump). whichever is most cost-effective 1n the specific locale. The electric demand charge is a key factor.
Supermarket Applications. Supermarkets are highly energy-intensive operations. About 500-\:i of the electr1c1ty used annually goes to maintain food cases: anotl1er 15% goes to air condition the store.
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REUEF AIR
COOLING TOWER
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COIL
HEATING COIL
BUILDING RETURN AIR
SUPPLY l>JR TO SP4CE
U.S PATlMT NO 4 72Jool 1 7 6: PAT[M'f'S P(NOIHC
Figure 2 Two-stage desiccant dehumidification HVAC system for new supermarket construction
A major portion of the energy feeding these supermarket compressors is consumed unnecessarily. The root cause of the inefficiency is the high humidity level (50% RH) maintained in the stores by conventional supermarket HVAC systems. In the special environment of supermarkets, with their numerous cold food cases. 50% RH is highly inappropriate. At this humidity level, frozen food cases condense moisture out of the surrounding air, a process that wastes utility energy as heat is absorbed during condensation. The process also causes frost buildup in frozen food cases.
More energy is wasted as frost buildup on food-case compressor coils reduces compressor efficiency and as more frequent defrost cycles are required.
Desiccant dehumidification maintains store humidity at 35% RH instead of 50% RH, which is not practical with conventional vapor-compression HVAC systems. Compared with a conventional supermarket system. a twostage desiccant HVAC system-in a new store or added to an existing store system-ach;eves the following:
(a) eliminates frost buildup in food cases and on compressor coils
(b) reduces defrost cycles. · (c) cuts store energy costs for case refrigeration and
air-conditioning on the order of 30%, and (d) eliminates the chilly, uncomfortable condition often
experienced by shoppers in aisles near cold food cases. which result when food-case compressors "overcool" as they work hard to combat excess humidity and frost buildup on coils.
Supermarket System Configuration. As shown in Figure 2. first-stage dehumidification is similar for supermarket ;rnd office building systems a desiccantimpregnated enthalpy exchange wheel transfers both heat and moisture from the incoming outside air to the build ing exhaust air. The supermarket system differs thereafter, however. 1n that buildina return air is added :o the partiall y der1umidi! ied outside a7r prio' to second-s~age dehum;dilicat1on This configuration eff 1c1entiy ma1 :·:a1ns the lowe: l1um;.j 1: /desired 1n the supe'~a rket (35°: RHi.
''' r.etrof1t svstems (Figure 3). th e dN1•J'T11dii1ed air is supo:1ed to the pre -exist111g cer:tra: a;r ha1'ch~ ' for se11s11J!c cooling 111 new consi ruc!10:1 r:=.0urc 2) tiv" conf iguri:1t• o ~'
F'RESH AIR
RE Liff AJR
ENTHALPY EXCHANCER
L nEs1cCANT DEHUUIOIFIER
us P ... run ,..0 4,71) 4 17 • PA.TENTS PCNOINC
~ H(.t.Jli'G COIL
~ AnE.RCOOUrlG COIL (CHILLED WATER OR DX)
BUILDING SUPPLY AIR RETURN AIR TO SPACE
Figure 3 Two-stage desiccant dehumidification system for retrofit of existing supermarket system
is changed by incorporating the vapor compression chiller or direct expansion refrigeration system into the dehumidification unit and by providing precooling (prior to secondstage dehumidification) as well as final cooling.
In both retrofit and new construction, the dehumidified . sensibly cooled air is distributed at a constant volume through low-pressure ductwork to air outlets.
Source of Supermarket Regeneration Heat. As in office building systems, energy source options for desiccant regeneration include:
(a) gas-energized cogeneration, which would simultaneously provide regeneration heat and electricity for store use:
(b) a direct-fired gas heater: (c) an off-peak electric resistance heater; or (d) an electric heat pump.
Restaurant Applications
Restaurants are another highly appropriate, costeffective applicat ion for the two-stage desiccani dehumidification HVAC system. due to the large amount of moisture generated from cooking. As in the supermarket system . recirculated air from the space is added to the partially dehumidified outside air prior to second-stage dehumidification.
NURSING HOME/CLINIC APPLICATION, WASHINGTON, DC*
Introduction
A two-stage desiccant dehumidification system was incorporated into the HVAC system designed by the aut~1o r for a two-story add1t1on to the Veteran's Administrat1or, Hospital 1n Washington. DC Desiccant dehumidific3t1on and a solar collection system were included because of their potential to significantly reduce the facility's electric utility demand and costs for cooling . Based on design data. desiccant dehumidif1catio:-i v11ii reduce the mecr1 -anical refrigeration load by 2oq/o. and solar energy will p1 o·J1 d e 70% to 80°·o of the he3t recu 1red !or des1cc;:n: regenerat1011 011 an annual basis /... 1 equ1red emerge 01c, generator is conl 19urcd as a c.'.) :~c: 1era 1 o r io 1; rO\"d' .. recovered heo.t \\' ! 1 1~ 1 ' 11c·2occ! as a :;acku1) s011rc;e
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Figure 4 Clinic/nursing home application: Two-stage desiccant dehumidificationlsolar/cogeneration HVAC system
In this facility an all-air system was a requirement; therefore, energy and first-cost savings are less than would be the case in a well-integrated air-water system. In a desic· cant dehumidification air-water HVAC system. in which the quantity of superdry primary air approximately equals the ventilation requirement. the size and cost of the primary air handler and ductwork are significantly less. as is the fan energy required for air distribution.
This is the first large-building application. to our knowledge. of a solar desiccant dehumidification system that uses low-temperature heat, 130° to 140°F, to regenerate the desiccant. The two-story addition has a separate HVAC system. The addition includes a 120-bed nursing home care unit, which requires 24-hour air-condrtioning. and clinical services on the first floor that operate during the daytime only. Occupied floor areas are as follows:
First Floor (clinical facility) - 50.394 ft2
Second Floor (nursing home) - 38,662 ft 2
Total Air-Conditioned Area - 89.056 ft2
The facility's inside design conditions. outside design parameters. and daytime building cooling load at design conditions in summer are as follows:
Outside conditions Summer- 89°F DB. 74°F WB Winter - + 17°F DB
Inside conditions Summer- 78°F DB. 50% RH Winier - 72°F FOB. 30% RH
Cooling loacJ (day) . Mecl1anic::i! relr1gera11011 d1splaceci b/ desiccant deilu: 1Hd1f1cat1on sys!e1n (1s: Fl 19 tons 2nd Fl 35 :ons) 54 tons
First-floor mechanical refrigeration cooling load 114 tons
Second-floor mechanical refrigeration cooling load 101 tons
Total Building Cooling Load (Day) 269 tons
Desiccant/Solar/Cogeneration HVAC System Overview
Figure 4 gives a schematic overview of the desiccant. solar, and cogeneration subsystems In general. their lune· tion and relationships are as follows.
When dehumidification is required in summer and intermediate seasons. the desiccant conditioner dehumid· 1fies all incoming outside air (19,300 cfm)-the minimum quantity required for ventilation. The desiccantdehumidified ventilation air provides 100% of the latent cooling required by the facility.
The dry air is distributed to two primary air handlers (one serving the first floor and one serving the second floor). where it is mixed with a variable volume of return air from the occupied spaces and sensibly cooled . The cold mixed air is distributed from the primary air handlers to faninduction terminals in each separate zone on each floor. This primary air handles all of the space latent and sensible cooling. The terminals mix the cold air with locally induced air and supply a constant volume to the rooms
Based on design data. the facility's total mechanical refrigeration load 1s reduced by 20% (54 tons) as a result of shifting the latent cooling task from the chillers to the desiccant system .
The solar collection subsystem provides hea; tor aes1ccant regenera tion (summer/intermediate seasons) and space heating (winter) as well as flir domestic 110t
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Figure 5 Two.stage desiccant dehumidification subsystem clinic/nursing home application
water. On ::;unny days during the dehumidificauon season solar energy satisfies the design regenefation requirement for daytime operation ; when less heat is needed tor regeneration. the balance of collected solar energy goes to the domestic hot water loop. On an annual basis, solar energy can provide 70% to 80% of the heat required for desiccant regeneration. In the winter. all solar energy contributes to space heating and/or domestic ho· water.
The diesel engine/cogeneration subsys erri meets the desiccant regeneration requirement plus the domestic hot water requirement . when there 1s no sun. When solar heat is insufficient , cogenerat1on provides the balance. The co generator operates only when thermal energy rs required. The simultaneously generated electricity 1s sufficient to power the desiccant system's fans and pumps. and to help power the facility 's chillers. In the winter. the cogenerator provides all space heating plus domestic hoi water heating. while displacing utility electrtc1ty with cogenerated electricity
Solar and cogenerated heat provide winter space heating via heating coils in fan·induction HVAC terminals (Fl Us). Heating coils are included in all second·floor Fl Us and in perimeter Fl Us on the first floor. Heating coils are served by the low-temperature (sola r) storage tank and then by the h1gh·temperature (diesel engine/steam) storage tank.
Desiccant Dehumidification Subsystem
Introduction. The desiccant dehumid:fication sub· system is located in the basement mechanical equipment room The system includes the following as shown ;n F1g~ire 5
2 rota ry w~1eei condrt:oner s (des;cca rY- •:-'' :xegnateo nonevcor.1b wheel s rnour:ted 10 cyl1:1cJricz, ; cas,ngs)
·- enthalpy exchanger - d~siccant cond1l1onerlfegcnera10'
2 filters - outside air filter (entering desiccant subsystem) - exhaust air filter (entering desiccant subsystem)
2fans - OA fan 6·F5-20,000 elm (two-speed, day/night) - EXH fan 6-F6-19.300 cfm
(two-speed, day/night) Recovery pump 6-P3-160 gpm Coils
- aftercool coil (cooling tower water) - recovery coils (two) - solar hot water coil - engine hot water coil
5 dampers-0·9, 0·10, D-11. D-12. 0·13 3 valves-V-19, V-28, V·29
Basic Concept. Figure 5 shows the dehumidification/regeneration process for outside air (OA) entering at the subsystem's design conditions, 93°F DB, 105 gr/lb. The basic concept is as follows:
1. Moisture is removed from the outside air in two stages: first in the enthalpy-exchange wheel. where heat and moisture are transferred from the incoming air to the building exhaust airstream; Hien 1n the second-stage desic· cant wheel, where moisture is absorbed by the desiccant for transfer to the hot exhaust airstream. At design conditions. approximately half of the moisture-removal job is done in each wheel .
2. Each wheel rotates continuously through two cham· bers. Incoming outside air flows through one side or chamber of eact1 wheel . for moisture removal.
3. Exhaust air flows through the opposite chamber o: each wheel. ac;ing as a sink to collect and carry oi: moisture dllci 1:1us regenerate ti·,e desiccant Only the second stage requires external heat for desiccant regeneration
TABLE 1 Outside Air Dehumidification Process
Outside-Air Moisture Air Quantity Stage Condition Grain Drop CFM
Entering first-stage wheel 93°FDB, 105 gr/lb 20,000
Leaving first stage/ Entering second ·stage 84°FDB, 78 gr/lb (27 gr/lb) 19,300
Leaving second stage 104°FDB, 56 gr/lb (22 gr/lb) 19,300
Leaving aftercool coil (cooling-tower water) 91°FDB, 56 gr/lb'~- 19,300
Total Moisture Drop (49 gr/lb)
wcondition of OA supplied to primary air handlers.
4. External energy sources for cooling/heating are : (a) Cooling tower water, to remove heat released by the
desiccant moisture-absorption process in the secondstage wheel.
(b) Solar energy and, when needed as backup, heat recovered from the diesel engine, to heat the exhaust air entering the second-stage desiccant regenerator chamber. (Heat is also recovered from air leaving the regenerator and transferred back to the entering air upstream of the solar coil and engine coil.)
Dehumidification Process. The desiccant subsystem operates both night and day when the outside-air moisture content exceeds 56 gr/lb. At night it serves only the second-floor primary air handler (6-AHU2).
As shown in Figure 5, when valve D-9 is open and OA fan 6-F5 operates at full speed (day). 20,000 elm of outside air enters the desiccant subsystem (19,300 cfm leaves the first-stage wheel) . The OA is filtered and then passes consecutively through the first-stage enthalpy exchanger and the second-stage dehumidifier. Table 1 shows the dehumidification process at subsystem design conditions (93°F DB, 105 gr/lb).
A small portion of the outside air (16%) bypasses the second-stage wheel and then rejoins the main airstream The air then passes over the aftercool coil , which circulates 75 gpm of 81°F cooling tower water.
Dehumidified OA leaving the aftercool coil at 91°F DB, 56 gr/lb is supplied to primary air handlers 6-AHU1 (12,400 cfm) and 6AHU2 (6900 cfm) .
First-Stage Regeneration (Enthalpy Exchange). Exhaust-air fan 6-F6 circulates 19,300 cfm of building exhaust air through a filter and then through the enthalpyexchange wheel for firststage regeneration (enthalpy exchange). The resulting change in air condition is shown in Table 2. (When the desiccant subsystem is not operating, exhaust air bypasses to the outside.)
Leaving the enthalpy wheel, the airstream divides: one-fourth of the air (5000 cfm) passes through an orifice plate to be exhausted to the outside; 15,000 cfm flows to the second-stage regenerator.
Second-Stage Regeneration (Desiccant). Prior to entering the second-stage wheel, as shown in Figure 5. the exhaust air is heated-in series-by the recovery coil. the solar hot-water coil (on sunny days), and the engine hotwater coil (when solar is insufficient) . Table 3 gives the entering and leaving air temperatures.
The recovery coil transfers heat to the entering air from the wheel's hot leaving air via a closed hot-water recovery loop. Recovery pump 6-P3 pumps 160 gpm of 99°F hot water through the entering recovery coil.
With valve V-28 open, the solar hot water coil circulates 210 gpm of 135°F water from storage tank 6-ST1 . On sunny days. tank 6-ST1 provides sufficient heat to raise the
TABLE 2 First-Stage Regeneration (Enthalpy Exchange)
Entering/Leaving Air Quantit:t Exhaust-Air Condition
Entering first-stage enthalpy wheel 19, 300 cfm 8l°FDB, 70 gr/lb
Leaving first-stage enthalpy wheel 20,000 cfm 90°FDB, 97 gr/lb
.. . ., ..
TABLE 3 Exhaust-Air Heating for Desiccant Regeneration·
Heat Source Exhaust-Air Temperature Entering Leaving
90°FDB 97 ° FOB
97 °FDB 130°FDB
Recovery coi 1-:H<Solar HW coil (sunny days) Engine HW coil (only as solar backup)
*Prior to entering second-stage wheel. **Recovered from air leaving second-stage wheel.
TABLE 4 Second-Stage Moisture Absorption for Desiccant Regeneration
Exhaust-Air Condition Item Entering Leaving
Desiccant wheel regenerator
Heat recovery coil*
Exhausted to outside
130°FDB, 97 gr/lb
107°FDB, 126 gr/lb
107°FOB, 126 gr/lb
99°FOB, 126 gr/lb
99°FDB, 126 gr/lb
*Recovers heat from air leaving the wheel, transfers it back to air entering the wheel.
exhaust-air temperature to the design regeneration requirement (130°F DB). On these days, the engine hot water coil is not needed and valve V-29 remains closed until night, when the collected solar energy is depleted.
When solar heat is insufficient and with valve V-'2.':j open, the engine hot-water coil circulates 210 gpm (at design conditions) of 175°F hot water from the hightemperature storage tanks (6ST2 . 6-ST3).
Hot exhaust air flows through the regerlerator chamber of the desiccant wheel and collects moisture, to be exhausted to the outside. The process is shown in Table 4.
Solar Collection for Desiccant Regeneration. The solar collection subsystem includes 4563 ft2 of flat plate collector area consisting of 234 collector panels (3 ft x 7 ft) with a tilt angle of 40° facing south Other principal components include a 10,000-gal solar (low-temperature) hot water storage tank, a plate/frame heat exchanger to transfer solar energy to the tank. a solar hot water coil that heats the regeneration airstream prior to its entry into the second-stage desiccant regenerator. pumps, and two shell-and-tube heat exchangers to transfer solar energy for hot water space heating and domestic hot water. The circulating solar energy collection fluid is a 350/o glycol/65% water solution
f11e solar system is designed to provide all daytime regeneration hea! required by ~he des1cca;it subsystem
operating at design capacity. On sunny days during the dehumidification cycle (summer/intermediate seasons). collected solar heat is on the order of 6 x 106 Btu per day. The maximum heat requirement of the desiccant equipment 1s at or below 600,UOO Btu per hour.
CONCLUSIONS A two-stage desiccant dehumidification system has
been developed in which the first-stage enthalpy exchange wheel handles 300/o to 500/o of a building's dehumidification load at design conditions without requiring external, utility-generated heat to regenerate the desiccant.
As a result . the second-stage desiccant wheel, which completes the dehumidification process, requires significantly less regeneration heat than a single-stage desiccant dehumidification system. Thus the two-stage process increases the thermal efficiency of desiccant regeneration .
When a two-stage desiccant system also is designed to use relatively low-temperature heat for desiccant regeneration (130° to 140°F) and 1s efficiently integrated into a building 's HVAC system, the desiccant HVAC system can be a cost-ef~ect1ve alternative to conventional HVAC systems for commercial buildings such as office buildings. supermarket<. .1nd restaurants as ·..veil as for inst1!ut ion3: IY.Jild1ngs suc:1 as hosp11;:i ls
