l. S. Oindrvas
`Project Manager.
`
`D. A. Wilson
`Project Engineer.
`
`M. Kawamoto
`Engineer.
`
`Fluor Daniel Power,
`Irvine, CA 92730
`
`G. L. Haub
`Project Engineer,
`Kern River Cogeneration Company,
`Bakersfield, CA 93380
`
`Options in Gas Turbine Power
`Augmentation Using Inlet Air
`Chilling
`
`Gas turbine power augmentation in a cogeneration plant using inlet air chilling is
`investigated. Options include absorption chillers, mechanical (electric driven) chill-
`ers, thermal energy Storage. Motive energy for the chillers is steam from the gas
`turbine exhaust or electrical energyfor mechanical chillers. Chilled water distribution
`in the inlet air system is described. The overall economics of the power augmentation
`benefits is investigated.
`
`the chilled air temperature is below the ambient temperature,
`but especially during summer on—peak hours.
`0 GT exhaust heat should be used to generate cooling of
`the cooling media (chilled water), via absorption chilling; how-
`ever, other means (electrically driven chillers) may be used I‘or
`peaking.
`
`Application of Inlet Air Chilling
`The GT inlet air chilling, i.c., air cooling below the ambient
`wet bulb temperature, can increase the GT capacity needed
`during peak hours and improve the GT heat rate. Selection of
`the temperature of the chilled air is important. The temperature
`at the compressor inlet must be above 32°F to prevent ice
`buildup on the compressor blades, since the chilled inlet air
`shall be at 100 percent relative humidity clue to moisture con-
`densation during the air chilling process. Due to a major in-
`crease in air velocity in the compressor inlet
`the static
`
`DESlGN HEAT RATE
`
`DESIGN AIFIFLOW
`
`DESIGN OUTPUT 0.
`
`Introduction
`
`Recent rapid growth in summer electricity demand experi-
`enced by most of the US. utilities results in the need to build
`power plants that generate maximum output at summer am-
`bient temperature ratings. Due to their short installation time
`and low installation cost, gas turbines are often used to meet
`this peak demand. One disadvantage, however, penalizes the
`gas turbine peaking plant rating, namely the inversely pro-
`portional effect of the ambient temperature on the gas turbine
`output, which is depicted in Fig. l. The gas turbines typically
`produce 30 percent higher output at 20°F than at 95°F. Thus
`the cost of installing a gas turbine or combined cycle plant
`rated at 95°F is 20-30 percent higher than that rated at 20°F.
`This inherent disadvantage of reduced gas turbine output at
`high ambient temperatures can be mitigated by the reduction
`of the GT compressor inlet air temperature, which would result
`in increase of GT output at a given ambient temperature.
`The traditional way to decrease the compressor inlet air
`temperature is the incorporation of an inlet air evaporative
`cooler, which can reduce the inlet air dry bulb ambient tem-
`perature by up to 90 percent of the dry bulb—wet bulb tem-
`perature difference. To achieve additional inlet air temperature
`reduction, alternative methods of inlet air cooling must be
`used, referred here as inlet air chilling.
`
`Objectives
`0 Develop a GT power augmentation concept based on
`reducing the GT inlet air temperature below the ambient air
`dew point.
`0 The power augmentation may be available anytime when
`
`Contributed by the International Gas Turbine institute and presented at the
`35th International Gas Turbine and Aeroengine Congress and Exposition, Brus-
`sels, Belgium, June 11-14, 1990. Manuscript received by the International Gas
`Turbine Institute January 16, 1990. Paper No. 90-GT-250.
`
`120.
`
`110.
`
`‘l00.
`
`PERCENT 90.
`
`80.
`
`20.
`
`120‘
`100.
`8!).
`60.
`40.
`COMPRESSOR INLET TEMPERATURE DEG. F
`FRAME 7E GT OUTPUT VERSUS COMPRESSOR INLET TEMPERATURE
`
`140.
`
`Journal of Engineering for Gas Turbines and Power
`
`APRIL 1991, Vol. 1131203
`
`Dowraioacicd ‘$93 Que Edit to ?’:i.23?’.2»<t5.§-ti. git-C°sZ§§$‘K§‘i§Zti.¥§§=z§:§“50£§%§r€;€%tiE3
`
`tics so or copyright; see §§ttp:!.I‘www.asane.<n'g!tenusf‘E‘c:rrns____l.§sc.ctau
`1 bay ASME
`
`
`
`Fig. 1
`
`PAGE 1 of 9
`
`PETITIONER'S EXHIBIT 1023
`
`

`
`
`
`CxENEiU\TDR
`cnnamrv cur-WE lwnsnmua comm)
`(FIN F/«Ni
`
`
`
`or CAPABILITY
`wnnoui evnrlommvs.
`, cootcn IN OPERATION
`(N&W.CLEAN)
`
`
`
`
`soa
`
` ‘:73“i..ii
`
`
`ii
`
`E
`"s
`
`.ii
`
`iifl
`
`3‘iiiiIi¢";l
`
`55
`so
`
`\
`
`\
`00
`
`\
`
`“ ~. \ \
`
`‘ \.\
`
`
`
`“\
`
`Fig. 2 GT inlet air cooling process
`
`D
`1D
`55 lj.
`
`20
`
`30
`
`40
`
`70
`60
`50
`AMBIENT TEMPERATURE, °F
`FIGURE 2
`
`DD
`
`Fig. 3 Generator capability
`
`\\
`4-
`
`>
`
`temperature of air may drop as much as 10°F. To safeguard
`against this drop the chilled air temperature should be a min-
`imum of 45°F, which includes a 3°F margin.
`Figure 1 shows that the output can be increased by 0.36
`percent with each l°F inlet air temperature reduction. If the
`chilled air temperature is selected at 52°F, then the temperature
`drop from 95°F is 43°F or 15.5 percent power boost. If the
`GT unit is or could be equipped with an evaporative cooler,
`which could reduce the ambient (95°F, 20 percent R.H.) to
`72°F (85 percent R.H.) the benefit of air chilling is only (72 ~ 52)
`or 20°F or 7.2 percent power boost. The cost of the inlet air
`chilling equipment must be amortized by the economic benefits
`from additional power sales during peak or during operation
`when ambient temperature is above the chilled air temperature.
`
`inlet Air Heat Exchange
`The ambient conditions (temperature and relative humidity)
`are important factors. The typical summer average on-peak
`ambient temperature in the U.S. is 95°F dry bulb. In eastern
`and midwest locations the relative humidity (R.H.) may be at
`90 percent level (wet bulb temperature is ~90°F) and the air
`enthalpy is ~59 Btu/lb dry air. In the western United States
`(dry climate), the same 95°F dry bulb temperature shall be
`associated with 20 percent R.H. (66°F wet bulb) and enthalpy
`of 34.7 Btu/lb dry air. If the compressor inlet temperature is
`to be lowered to 52°F dry bulb, 95 percent R.H., the enthalpy
`is 20.9 Btu/lb dry air. The required heat removal at high
`humidity location shall be (59—20.9) = 38.1 Btu/lb and in
`the dry climate it will be (34.7—20.9) = 13.8 Btu/lb. Thus,
`the duty of the inlet cooling system in humid climates is much
`higher than in the dry climates. In either case, moisture sep-
`aration apparatus should prevent carryover. The cooling pro-
`cess is shown in Fig. 2.
`The temperature at the compressor inlet must be above 32°F
`to prevent ice buildup on the compressor blades, since the
`chilled inlet air shall be at 100 percent relative humidity due
`to moisture condensation during the air chilling process. Due
`to a major increase in air velocity in the compressor inlet, the
`static temperature of air may drop as much as 10°F, which
`includes temperature drop in the compressor bellmouth (At :-
`4°F) and in the (open) inlet guide vanes (variable or stationary)
`(AT = ~4°F). Closed variable inlet guide vanes would result
`in additional pressure drop (--5°F) but this would not apply
`while inlet air chilling is used. To safeguard against this drop
`the chilled air temperature should be a minimum of 45°F.
`Increased wetness in the compressor inlet may induce corro-
`sion, which may require more frequent inspections of com-
`pressor inlet.
`
`GT Inlet Air Cooling Concepts
`Several concepts may be offered. The GT operates either in
`the simple cycle or in the cogeneration or combined cycle.
`Simple cycle GT shall be able to use all exhaust heat for chilled
`water generation. For 21 GT in cogeneration or in combined
`cycle service, we may need to evaluate the tradeoff of using
`the exhaust heat for power augmentation in the GT inlet versus
`the economic benefits of exhaust heat sales.
`
`Generator Capability
`The generator rating has to be verified. If the generator is
`cooled by cooling water it can be provided from two sources:
`cooling tower/river, and dry coolers (fin fan type).
`Since the generator capability is indirectly proportional to
`the cooling water temperature, it is possible in the case of dry
`cooling that the GT output may be increased via inlet air
`chilling while the generator capability would not increase. This
`could result in GT exceeding the generator capability with
`potential overtemperature on the generator winding. The gen-
`erator rating must be carefully checked. The same applies for
`air-cooled generators. Examples of the GT and generator ca-
`pability are shown in Fig. 3.
`
`Systems Options
`The complete inlet air cooling system consists of several
`systems, such as:
`
`0 Water chilling system
`0 Absorption chiller thermal source
`0 Chilled water system
`- Cooling system
`The following options exist for each system;
`Chilling System
`0 Absorption chiller utilizing:
`(a) Lithium bromide absorbent.
`(b) Ammonia absorbent.
`
`0 Mechanical chilling utilizing compression with freon re-
`frigerant.
`Steam jet refrigeration.
`- Hybrid chilling system utilizing both absorption and me-
`chanical chilling systems.
`0 TES (Thermal Energy Storage)
`
`204 I Vol. 113, APRIL 1991
`
`Transactions of the ASME
`
`Downioaded ‘tit fies EM? to ?’:i.23?’.2»<ir5.§-ti. i-Eedéstrihzition subject to .=‘-‘tfiiiifi Eiet-Erase or copyright; see §tttp:!.I‘www.asine.orgitermef‘E‘errr;s____l.§se.ctm
`
`
`
`
`
`3:UI
`
`
`
`uon
`
`~ro
`
`
`
`GT,GENERATORCABIUTY,NW ane
`
`\osu:nA1on
`i
`WATER mo.....,
` or can IVY.
`cootmo rowan)
`nzomne
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`mm: or
`/\\ e
`CDNCERN
`\
`
`v$¢y<:NILLlNG
`"'41/4,9
`\
`\
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`or cmnanurv
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`WITH !;VAPLJllAI|Vh
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`COOLER [N Of’ENA‘flGN
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`
`PAGE 2 of 9
`
`PETITIONER'S EXHIBIT 1023
`
`

`
`
`
`'
`
`Indirect evaporative cooling for distribution of chilled
`water within the inlet air filter.
`
`Cooling System. Study of the following cooling water sys-
`tems:
`' Mechanical draft cooling tower
`.
`Closed—loop water to air cooled system (fin fan).
`
`Absorption Chiller Thermal Source. A thermal source to
`provide operating force to absorption chillers was studied based
`
`on usmg:
`° Low—pressure steam generated in a new HRSG installed
`in a Simple cycle GT exhaust‘
`I Tie-in to existing steam export line in cogeneration ap-
`plications to generate steam through a low—pressure heat
`exchanger (i.e., reboiler).
`0 Generation of hot water from the above methods for the
`ammonia absorption chiller.
`0 Direct integration of I-IRSG preheat coil with ammonia
`absorption chiller to provide desorption heat in simple
`cycle GTs.
`
`Chilled Water Distribution System. Review of the follow—
`ing schemes to provide chilled water to/ from the chilled gas
`turbine inlet air:
`
`0
`
`0
`
`0
`
`Installation of PVC media similar to the evaporative
`cooler media downstream of the inlet air filter, using
`both the evaporative and direct Contact principle.
`Installation of chilled water coils using the convection
`principle (indirect cooling).
`Installation of freon—cooled coils using the convection
`principle.
`
`REF\FI{AC;Eo|;ANT
`
`
`
`
`
`CONCENTRATED
`SOLUTION
`vs vuo-vv
`CONCENTRATOR
`
`40F
`AIR CONDITIONING
`LOAD
`55F
`
`CONDENSING
`
`REFRIGERANT
`
`WATER ABSORBINGSOLUTION
`
`
`SOLUTHBN PUMPS
`
`REFRIGERANT PUMF
`
`
`
`
`
`DILUTESOLUTION
`
`Fig. 4 Absorption chiller schematics (courtesy of Trane Corp.)
`
`‘
`‘
`‘
`
`:
`.
`.
`.
`_
`_
`_
`
`.
`
`65,000 kW
`12,200 Btu/kWh—LHV
`70 GPM
`Natural gas
`14.3] psia
`95°F
`70.8°F
`30 percent
`34.7 Btu/lb d.a.
`700 ft.
`2,074,982 lb/h each
`
`5 in. WC
`
`Description of Considered System
`The system evaluation is based on retrofitting four General
`Electric Frame 7E GTGS with inlet air chilling system.
`Existing System: 4 X Frame 7E GTGS
`I
`Power output
`Heat rate
`NOX control——water injection
`Fuel
`Barometric air pressure
`Ambient dry bulb temperature
`Ambient wet bulb temperature
`Relative humidity
`Inlet air enthalpy
`Site elevation (ASL)
`Inlet air flow rate
`Inlet system pressure drop (filter,
`evaporativc cooler, silencer)
`Openings available:
`Inlet duct, 1 opening (W X H) 25’ X25 ’
`Filter house, 6 openings 25 ’ X 5 ’
`80 percent
`Evaporative cooler efficiency
`72°F
`Temperature after evaporative cooler
`80 percent
`Humidity
`Heat Recovery: 80 percent steam quality HRSG for EOR
`Inlet temperature
`970°F
`Stack temperature
`260 ° F
`Feedwater temperature
`150 ° F
`Steam output @ 800 psig, 80 percent quality
`418,000 lb/h each
`10 in. WC
`
`Gas side pressure drop
`Retrofit With Inlet Air Chilling:
`Retrofitted inlet air temperature (dry bulb)
`Retrofitted inlet air humidity
`Inlet air enthalpy
`Air chilling duty
`or
`
`OOII
`
`52°F
`95 percent
`21 Btu/lb da.
`28 mm Btu/H each
`2333 tons each
`
`CMILLED \-IMER CDJLS
`
`rim : nu curse PUMP
`
`Iu wmr
`\/Mm
`
`3____________
`
`
`
`l3n'/ aumn I1———————————————
`
`
`AsSl.IRPr|ON CNILLER rm um. «ms
`
`run HAG:
`L llNlU)»I~BF<|lKlDF
`
`rmmmm: arm-ma
`Tnux
`
`l
`g
`:I
`
`«mam:
`-»[»4A><:—uP mum
`
`.4 mm...
`
`rm st: ucmemnrn:
`s
`1
`cnunzusnln: n:ruLzM1..,._,T@i _,?______M. .
`cuunrus/us Pum-
`£2
`
`~
`
`J
`|lL|.lVlrllslN
`
`Fig. 5 Absorption chilllng diagram
`
`Journal of Engineering for Gas Turbines and Power
`
`APRIL 1991, Vol. 113 I 205
`
`Dowriieaded ‘$93 Qee 20% to ?’:i.23?’.2»<t5.§-ti. Fieriéstribziiiori subject to .=‘-‘ifiiiifi §§€:t-‘:3’iS={3 or copyright; see §ittp:!.I‘www.asane.orgltermsf‘E‘errns____i.§se.etm
`
`
`
`PAGE 3 of 9
`
`PETITIONER'S EXHIBIT 1023
`
`

`
`
`
`um nnaunos BAFFLE
`
`coursruszn wsvzn our
`
`
`
`Water Chilling via Absorption Chillers (Fig. 5)
`Absorption Chilling System.
`‘Absorption chilling systems
`operate with a low-grade thermal source to provide chilled
`water at 44°F. The system contains refrigerant and absorbent
`that cycles at low internal pressures. This pressure in the evap-
`orator section will allow evaporation of the refrigerant liquid
`at a low temperature providing the chilling effect for the in-
`coming chilled water. At slightly higher pressures, the refrig-
`erant Vapor and absorbent liquid are combined in the absorber
`section by their strong affinity for combining. As the vapor
`is condensed, heat is released into the condenser water coils
`(cooling water). The resulting solution will be a “strong” ab-
`sorbing solution of refrigerant and absorbent. This solution
`is pumped into the absorbing section where the refrigerant is
`evaporated from the absorbent by a thermal source, either
`low-pressure steam or hot exhaust gases. in the condenser, the
`refrigerant vapor is condensed to liquid by cooling water, and
`the refrigerant liquid is pumped back into the evaporator. The
`“weak” absorbing solution is recirculated into the absorbing
`. section. An example is shown in Fig. 4.
`Ammonia and lithium bromide (LiBr) are two types of ab-
`sorption chiller that are available for this chilling service.
`Lithium Bromide Absorption Chilling Systems. The lith-
`ium bromide absorptiort chilling system can be designed as a
`single or two-stage chiller. These chillers utilize a design with
`a lithium bromide water absorption refrigeration cycle. The
`chilling systems typically provide chilled water at 44°F with
`return water at 54°F. Consideration was made to lower the
`chilled water temperature to 42°F; however, this would have
`reduced the chiller’s cooling capacity and efficiency. Chilled
`water temperatures less than 42°F are not considered practical
`for absorption units. This is due to the water-based refrigerant
`used in the chiller, which, when operating between 34°F and
`36°F, provides 42°F chilled water and may potentally freeze
`the chilling unit.
`Typical flow calculations for each of the absorption chiller
`units are based on the following:
`
`0 Leaving chilled water temperature = entering temper-
`ature — temperature drop
`ll
`0 Leaving condenser water temperature
`perature + temperature rise
`
`entering tem-
`
`=:
`
`Cooling load (tons) X 24
`t ‘fl
`Ch'll d
`temperature drop
`OW (gpm)
`I e wa 61
`.
`ll
`' Steam Flow (lb/hr)
`Cooling load (tons)><9.9 (two
`stage)
`Cooling load (tons) X 19 (single
`stage)
`
`ll
`
`' Condenser water flow (gpm) :
`
`Cooling load (tons) X 44
`temperature drop
`
`Single Stage (LiBr). The chiller design utilizes a single shell
`hermetic construction that will enhance the integrity of the
`unit. These chillers operate on low—pressure saturated steam
`at 15 psig, consuming approximately 19 lb/hr per ton of chill-
`ing. Hot water at 270°F can also be used as the thermal energy
`source. The single-shell design helps prevent air leakage, which
`can cut the capacity and promote corrosion. A pre-packaged
`system can be delivered completely assembled with available
`size range from 101 to 1,660 tons. The single-stage system will
`operate at a vacuum and may be susceptible to air leakage,
`which could cause some damage. However, if the proper op-
`erating and maintenance procedures are followed, i.e. , periodic
`purging of the system and operation above 42°F, there should
`be no problems during normal operation. A total loss of power
`would cause the solution of the absorption chiller to solidify
`into a solid mass within one hour. Manufacturers that currently
`produce the single-stage absorption chillers are Trane, Carrier,
`
`cmmzussa watifi m
`
`.
`
`r-c-‘LEVEL ccmrnm.
`
`,mr rusrn mus
`coursrssow
`
`worm
`
`,,\ .
`
`nsrma now VALVE
`
`-—+ WAVEV1 FLDV4
`M. urrnuesamr Ftow
`
`uwmu ummamicn BAFFLE
`2 LJDUDD HLPKIGERANY
`It v/won nsrmszxuur
`
`|:| WAYER
`
`can Lin wnrra our
`LLE
`AYIR nc
`P05
`.SID£l
`
`Fig. 6 Mechanical chiller schematics (courtesy Carrier Air Conditioning
`Co.)
`
`and York. Cost may be estimated at $250 per ton of refrig-
`eration.
`
`Two Stage (LiBr). The two-stage lithium bromide absorp-
`tion chiller system will operate with saturated steam pressures
`of approximately M4 to 128 psig allowing the unit to consume
`less steam than a single—stage chiller at 15 psig. The two-stage
`steam consumption is estimated at 9.9 lb per ton versus 19 lb
`per ton for the single stage; thus, less steam will be consumed
`during chiller operation, resulting in a net savings of thermal
`energy over the single-stage chillers.
`The cost estimate for the two-stage chillers is $330 per ton,
`which is approximately 30 percent more than the single stage
`and twice the cost of a centrifugal chiller. The chiller size is
`limited to 1500 tons for the larger chillers. Manufacturers of
`the two-stage absorption chillers are Hitachi Paraflow, Sanyo-
`Bohn, and The Tranc Company.
`
`Ammonia Absorption Chilling System. The ammonia ab-
`sorption system is an engineered refrigeration system specifi-
`cally designed for the tons of cooling required. The ammonia
`system can produce lower chilled water temperature than lith-
`ium bromide systcms, from +50°F through —50°F chilled
`medium temperatures. Low-pressure steam (175 to 265 psig)
`or turbine exhaust gases can be used as the thermal source for
`the ammonia absorption cycle. However, the ammonia system
`requires higher capital costs, approximately $650/ton of cool-
`ing. Also, larger plot space is required, approximately 30’ X 30 ’
`for a 1,000 ton chiller installation. The ammonia system field
`installation requires significant structural steel to support the
`various ammonia storage tanks, heat exchangers, and other
`vessels. Manufacturers of the Ammonia Absorption Chillers
`are: Lewis Refrigeration Co., Houston, Texas, and Borsig
`GmbH.
`
`Water Chilling Via Mechanical Chillers (Fig. 7)
`The Mechanical Chiller is a vapor compression cycle with a
`compressor, liquid cooler (evaporator), condenser, and com-
`pressor drive. Water at 54°F returning from the inlet cooling
`coils enters the chiller (evaporator) where it is cooled to 44°F
`by the refrigerant liquid evaporating at a lower temperature.
`The refrigerant gas produced is compressed to a higher pressure
`and temperature so that it may be condensed by the cooling
`water in the condenser. The condensed refrigerant is returned
`to the evaporator through an expansion/metering valve.
`Mechanical chillers may be provided with different types of
`compressors: reciprocating, screw, or centrifugal. Recipro-
`cating and screw type compressors are generally used for sizes
`less than I000 tons. Centrifugal liquid chillers are used for
`sizes greater than 1000 tons, with factory—assembled units up
`
`206 I Vol. 113, APRIL 1991
`
`Transactions of the ASME
`
`Downioacled ‘iii fies wit to ?’:i.23?’.2ai.5.§-ti. §-Eeriéstribution subject to .=‘-‘tfiiiifi Eicense or copyright; see §ittp:!.I‘www.asine.orglternisf‘E‘ern*;s____i.§se.cim
`
`
`
`PAGE 4 of 9
`
`PETITIONER'S EXHIBIT 1023
`
`

`
`
`
`CHILLED VVWER CDILS
`
`LHILLLD VAHZR mass
`
`——————————
`
`‘
`\
`
`\
`
`TU V/\SlL 3,,
`s/Arte
`
`lwiiiiéwl
`i
`INLET ———————»;
`EFWLTEE’
`’
`,
`.
`IT ;
`//
`-\
`j
`\\
`.
`r//
`TURBINE carmussas CUMFPESSDR
`,/’
`7 "7 \\
`./
`/,/
`\\\
`
`\
`
`_,,
`
`_
`<"""'"" *
`
`'
`
`\
`/‘ \
`
`cv
`
`ELECIRIC DRIVEN
`cr.mRxruc.Ai_ (ZIIILLER
`
`
`
`cnnm: vnunz
`luzvnnvn
`
`Fig. 7 Mechanical chiliing diagram
`
`to 1600 tons, and field-assembled machines to about 10,000
`tons.
`Chillers from 100 tons to l500 tons are packaged off—the—
`shelf units that are hermetically sealed. Chillers from 1500 tons
`to 3000 tons are generally the open-drive centrifugal water
`chillers. These units are not hermetically sealed because the
`compressor and motor drive are mounted outside the chiller
`unit. Three thousand tons and larger are customized units.
`Manufacturers such as Carrier, Trane, and York are capable
`of providing chillers in the 1000 ton to 3000 ton size. Each
`chiller will operate in parallel, to increase availability of the
`system.
`Typical flow calculations for each centrifugal chiller unit
`are based on the following:
`
`0 Leaving chilled water temperature = entering temper-
`ature — temperature drop
`0 Leaving condenser water temperature
`perature + temperature rise
`
`entering tem-
`
`0 Chilled water flow (gpm)=
`
`Cooling load (tons) X 24
`temperature drop
`
`Cooling load (tons) X 22
`' Condenser water flow (gpm) =“"""
`
`The temperature rise and drop is generally limited to 10°F.
`This value can be increased at the expense of derating of the
`chiller unit. For example, the increase temperature of the cool-
`ing water from 10°F to 15°F will derate the chilling capacity
`of the chiller by approximately 5 percent.
`The selection of the 1500 ton centrifugal chiller was based
`on the cost per ton, delivery, and electrical load requirements.
`Based on manufacturer’s information, the costs per ton for
`chillers were:
`
`10004500 ton chillers
`1500-5000 ton chillers
`Over 5000 ton chillers
`
`$l33/ton
`$320/ton
`$l45/ton
`
`The smaller centrifugals would be the most economical pur-
`chase based on capital cost; however, the cost of the installation
`and required area would be much greater than for the large
`chillers. The larger tonnage chillers were either too costly or
`required a large motor driver for their operation. A 5000 ton
`chiller would require a motor size of approximately 4300 hp.
`
`This type of motor would be custom built and associated with
`long delivery. In addition,
`these units would require high—
`voltage electrical devices.
`The mechanical chillers need only electricity and condenser
`water to provide all chilled water requirements; there is no
`need to provide steam or another thermal energy source. Also,
`the mechanical chillers provide more tons of cooling in one
`machine than in other types of chillers. This will result in lower
`installation costs and less plot space required.
`However, the consumption of electricity will reduce the po-
`tential output of the plant, reducing the electrical energy gen-
`erated. Although chlorinated fluorocarbons (CFC) have been
`in use for years as part of refrigerants used in direct expansion
`refrigeration such as centrifugal chillers, these CFCS may po-
`tentially be banned from use because of their damaging effect
`on the ozone layer in the atmosphere. The mechanical chillers
`will require high—voltage electrical devices to support the chiller
`motors, MCCS, relays, motor starters, etc. This additional cost
`’will increase the overall price of the chillers. Manufacturers
`of the Mechanical Chillers are: Carrier Corp., The Trane Com-
`pany, and York, Division of Borg-Warner.
`
`Thermal Energy Storage Systems
`The GT can take advantage of on-peak and midpeak energy
`costs by using mechanical chillers and the Thermal Energy
`Storage Systems or TES. The TES system is based on a storage
`medium with high specific or latent heat, ice, water, or eutectic
`salts. This storage will contain the cooling produced during
`the low-cost off—peak hours for utilization during the high-
`cost on—peak hours.
`
`Chilled Water Storage. The design basis requires chilled
`water to be provided during the peak months of the year, from
`March to November, during the peak hours, averaging 10 hours
`per day. During the off-hours, the operator can take advantage
`of the lower electrical rates by operating the mechanical chill-
`ers, producing 44°F chilled water for the TES system. The size
`of the storage capacity will depend on an economic evaluation
`of the chilling profile. A full—size TES tank would provide all
`of the cooling load during the rnidpeak and peak hours. A
`partial, load-leveling TES tank will reduce the overall size of
`the peak cooling load profile and levelize the production of
`chilled water over the 10 hour period. Estimated size of the
`TES tank will be:
`
`Journal of Engineering for Gas Turbines and Power
`
`APRIL 1991, Vol. 118 I 207
`
`Dowaioaried ‘$93 Qec Eat? to ?’:i.23?’.2aii5.§-ti. Fietiéstriiiziiioti subject to .=‘-‘ifiidfi Eieerise or copyright; see §rttp:!.I‘www.aatria.orgltarmsf‘E‘errr;s____i.§se.ctm
`
`
`
`PAGE 5 of 9
`
`PETITIONER'S EXHIBIT 1023
`
`

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`Fig. 8 Hybrid chilling diagram
`
`Full
`size (1)
`15,120,000
`360,000
`7,500
`105,000
`
`Partial
`size (2)
`8,820,000
`210,000
`4,375
`61,250
`
`Tank Capacity (gal.)
`(bbl)
`Chiller size (tons)
`Stored cooling (ton/hr)
`Notes:
`1 These values assume that the mechanical chillers will not
`operate during midpeak and on—peal<.
`2 These values assume mechanical chiller operation 24 hr/
`day at an average load.
`The size of the tank that will be required for the TES system
`may result in space constraints and large costs.
`
`Ice Storage. The ice storage system can also be used in a
`scenario similar to the chilled water storage system. The volume
`of ice required will be from 7 to 14 times less than the water
`system (based on 144 Btu/lb ice latent heat of fusion versus
`10 to 20 Btu/lb water heat capacity). However, the equipment
`to provide the 85,000 ft3 to 168,000 ft3 of ice will require large
`real estate.
`
`Hybrid System (Mechanical/Absorption) (Fig. 8)
`The hybrid system is a combination of the lithium—bromide
`(LiBr) and the mechanical centrifugal chiller systems. The LiBr
`system will chill the return water from 54°F to 44°F and the
`mechanical chiller will further reduce the temperature to 38°F.
`This will be a series type system: two 50 percent capacity trains
`of absorption and mechanical chillers (reference drawing Fig.
`8).
`This system will initially take advantage of the steam avail-
`able at the facility and will enable the mechanical chillers to
`consume less power during the production of colder chilled
`water. The system is sized for one—half of the cooling by ab-
`sorption and the other by centrifugal chilling, which will con-
`sume approximately one-half of the total required power than
`the mechanical chiller system. The 38°F chilled water supply
`will increase the temperature differential of the cooling coil
`inlet and outlet (38°F to 54°F). The larger differential tem-
`perature will reduce the overall chilled water flow rate, reducing
`pumping costs as well as piping requirements.
`The lower chilled water temperature can be used in the GTG
`
`
`
`Fig. 9 Chilled water coils (courtesy of Coil Co. Inc.)
`
`inlet air coils to cool the air from 95°F to 52°F or to 46°F.
`Application of the hybrid system results in lower chilled water
`temperatures in the chilled water system associated with less
`piping and pumping costs. Also, the cooling coils will be more
`effective, decreasing the overall size requirements.
`
`Chilled Water Distribution System
`The chilled water distribution system will circulate 44°F
`chilled water from the chilled water plant and into each GTG
`inlet filter house and return with 54°F water. The inlet cooling
`system will distribute the chilled water through an indirect or
`a direct contact heat exchanger that should be installed across
`the inlet air flow cross section.
`The installation of the inlet cooling system will result in
`permanent increased pressure loss to the inlet of the GTG.
`General Electric’s performance curves have indicated that a
`4" water column pressure drop will result in a 1.45 percent
`decrease in power output, 0.45 percent increase in heat rate,
`and an increase in exhaust gas temperature of 2°F. We antic-
`ipate that the inlet pressure loss should not exceed 2" water
`column.
`
`208 I Vol. 113, APRIL 1991
`
`Transactions of the ASME
`
`Dewnieaded ‘$93 Qec mil ta ?’:i.23?’.2ail5.§-ti. Fiedéstrihatiori subject to .=‘-‘ifiiiéfi Eiet-Erase or copyright; see §§ttp:!.I‘www.asane.ergltermsf‘E‘errr;s____l.§se.c‘im
`
`
`
`PAGE 6 of 9
`
`PETITIONER'S EXHIBIT 1023
`
`

`
`
`
`ysis utilizing PVC cooling tower media in direct contact with
`44°F chilled water for gas turbine inlet cooling was performed.
`The medium is cooling tower packing placed in the straight
`crossflow position and sprayed with chilled water to create a
`direct contact chilled water evaporative cooler with an ap-
`proximately 90 percent relative humidity saturation capacity,
`shown in Fig. 10.
`A computer multiple alternative analysis examined three
`flow rate conditions consisting of 8, 12, and 16 gallons per
`square foot of face area and comparing two face velocities of
`500 and 750 feet per minute, respectively.
`To obtain the 52°F inlet temperature requirement necessi-
`tates a face velocity of 500 feet per minute and 12 gallons per
`square foot, i.e., 12,000 gallons per minute and 1000 square
`feet of surface area or 750 feet per minute and 16 gallons per
`square foot, i.e., 650 square feet of surface area and 10,000
`gallons per minute, respectively.
`The advantage of direct contact cooling is lower first cost
`and relatively low pressure drop. The disadvantage is the need
`for extensive water treatment associated with continuous blow-
`down. Potential for biological buildup exists, resulting in higher
`maintenance than a closed system. Also, large circulation water
`quantities are required, depending on performance, from 8 to
`16 gallons per square foot.
`
`Indirect Evaporative Cooling. The indirect evaporative
`cooling method lowers the dry bulb temperature without the
`increased moisture content of the GTG inlet supply air, ef-
`fectively rcducing the wet bulb temperature of the air and the
`overall energy content.
`The indirect method is accomplished by using a plate-type
`air-to—air heat exchanger, shown schematically in Fig. ll:
`- On one side, outside air is directed to one end of the
`exchanger as scavenger air. Water is sprayed onto the plate
`surface in contact with this air stream where evaporation takes
`place, cooling the plate approaching the wet bulb temperature.
`The water is collected in a sump and is recirculated back into
`the exchanger. The air is then exhausted above the unit into
`the atmosphere by an axial fan unit.
`- On the other side, the GTG inlet air supply passed across
`the plate and is cooled by convection to near the wet bulb
`temperature and passes into the gas turbine.
`This type of cooling is effective when there is adequate water
`available. The dry bulb air temperature can be reduced without
`increasing the latent heat at minimal cost of water and elec-
`tricity.
`For example, if the outside ambient air temperature is 95°F
`dry bulb and 71°F wet bulb, the indirect evaporative method
`of cooling will reduce the GTG inlet dry bulb air temperature
`to approximately 74”F, effectively reducing the total cooling
`load by 40 percent. This would translate into approximately
`4000 tons of cooling.
`Although the indirect cooling system will reduce the inlet
`air temperature to 75 "F, cooling coils will be required to reduce
`the air temperature further to 52°F. The pressure drop through
`the indirect cooler is estimated to be from 3 ” to 4” of water,
`which would reduce the gas turbine power output by approx-
`imately l.45 percent. In addition, the cooling coils will add
`another I" to l.5 ” of water pressure drop.
`~ Due to the overall size of the cooling unit, there may be
`some problems of system installation. There is limited space
`available within the filter house for an installation of this type
`and significant GTG downtime.
`The saturated air exhaust from the indirect cooler is located
`at the filter house, which would defeat the purpose of this
`system; the air will potentially recirculate back into the inlet,
`add