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  • 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 devel- oped that significantly increases the thermal coefficient of performance (COP) of desiccant regeneration in commercial applications. The first stage is a desiccant- impregnated enthalpy exchange wheel that accom- plishes 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 cost- effective 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 arrange- ment, dehumidification process, and regeneration heat requirement and sources in an actual application in a two- story, 90,000 ft2 building in Washington, DC.

    INTRODUCTION

    Air-conditioning includes two basic functions- moisture control and temperature control (dehumidifi- cation 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 on- peak 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 conven- tional 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 cooling- because it now can be done at an appropriate higher tem- perature, 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 re- quirement. 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 inte- grated 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 regen- eration 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 •

    J

  • GENERATED ELECTRICITY POWERS MOST OF HVAC

    -------

    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

    ( 1 . 1 CFM/SF)

    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 commer- cial buildings

    1. The systems have been configured to use relative- iy 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 two- wheel regeneration process is 1.5 to 2.0 at design condi- tions_ 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 out- side. 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 out- side 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 sup

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