The temperature of the air entering a gas turbine prime mover has a dramatic effect on its performance, including output, heat rate, and exhaust gas temperature (EGT). These variations are easily observed in actual operation and by reference to generic gas turbine (GT) performance curves. The gross capacity increase of a GT operating at 40F(8C) inlet compared to operation at 102F(70C) is 28%. The gross reduction in heat rate for this 62F(16.7C) differential is 6%, and the exhaust gas temperature is reduced 5%. Since the overall mass flow through the GT is increased through the cooling process, the added energy available in the heat recovery steam generator (HRSG), is increased 8% The significant improve- ments in GT output and efficiency which can be achieved by maintaining lower inlet air temperatures encourage the manufacturer, systems engineer, owner, and operator of GT facilities to consider seriously the implementation of a gas turbine inlet air cooling (GTIAC) system.
GTIAC systems have proven to produce some very excellent economic paybacks due to increased power output, EG mass flow, and reduced heat rates. Generic gross performance factors are plotted (See Figure 1) against inlet air temperature compared to International Standards Organization (ISO) conditions.
This new power enhancing technique is investigated and modeled on a Dallas, Texas site and based on the American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. (ASHRAE) and U. S. International Station Meteoro-logical Climate Summary (US ISMCS) ambient temperature data. This gives a very broad range of temperatures and humidities for analysis and comparison. A gas turbine generator (GTG) system using a Westinghouse 501D5 open cycle combustion turbine, is the exemplar GTG package producing a nominal ISO base rating of 118.5 mW at a heat rate of 10,023 btu/kWh(10,568KJ). The com-parison will be based on an uncooled turbine, an evaporative cooled GT and a refrigerated GTIAC system with the inlet air cooled to 40F(4.4C).
Comparison will be made between a traditional single temperature refrig-erated coil system with the new higher efficiency, 3-stage/temperature, cascade, (patented) system. Primary refrigerant systems (evaporating refrigerant within air cooling coils) and secondary refrigerant systems (recirculated chilled water/brine within cooling coils) will be reviewed.
For primary refrigerant systems, anhy-drous ammonia is the recommended refrigerant since it is the single most efficient industrial refrigerant available and is environmentally benign. Both its Global Warming Potential (GWP) and Ozone Depletion Potential (ODP) are zero. Other usable refrigerants are surveyed, including Propane (R290), HCFC 22, and R134a. In addition to this, various other cooling techniques, including air washers, absorption (steam, gas fired) systems, off-peak thermal energy storage (TES) options, and a variety of drivers (steam/gas turbine, IC engine) were reviewed.
A simple cycle, base-load model was chosen not only to simplify the analysis, but to focus on the advantages of GTIAC systems and their transparency to any other system configuration. Cooling the inlet air to 40F(4.4C), or lower, provides the net advantages described for the entire range of cycles and configurations including combined cycles and water/steam injection systems. The absolute increases in the system will be constant and independent of down-stream system configurations. The rela-tive percentage increases compared in each configuration will, of course, vary.
The observations of wide variations in power output due to the site ambient air temperatures are commonly realized in any gas turbine facility. Traditionally, the inlet air stream is cooled by an open evaporative cooler consisting of recirculated water spray over a saturating extended surface media which is mounted downstream from the air filter system, (See Figure 2). The heat transfer efficiency of this commonly available media provides a dry bulb "depression" or cooling "approach" to the ambient wet bulb temperature in the range of 80-90% of the difference between the dry bulb air temperatures. The level and efficiency of the evaporative cooling stream is limited to this approach to the wet bulb temperature of the site area.
Sensible cooling takes place as the process proceeds along the constant wet bulb temperature line and at constant enthalpy, progressing towards the saturation line. This adds evaporated moisture and reduces the dry bulb temperature. These two processes together increase the air density, which increases the total mass flow through the GT, enhancing the turbine output.
This cooling technique, while the lowest in first cost and easily accomplished compared to refrigerated systems, has a downside. It is only effective in the less humid areas of the world and at ambient temperatures greater than 50-55F (10-12.8C). The added inlet air pressure drop due to the resistance of the media and water spray (loading) further reduces turbine output. This is a constant output penalty imposed throughout all of the GT unit operating hours whether or not cooling is taking place. Nominally speaking, for each inch water gauge of added air inlet pressure drop, GTG capacity is reduced approximately 1/2%.
In addition, the evaporative system is a direct, open, wet system which acts as an air scrubber and also adds moisture to the air. Scaling caused by the evaporation process requires additional maintenance and water treatment.
Conventional water chilling techniques (compression or absorption/chillers) can also utilize an alternate direct form of air cooling which does not require conventional indirect finned cooling coils. Finned coils are replaced by direct spray air/coil air washers which use chilled water sprayed onto saturating media through which the air is passed, much like the evaporative units. In a similar fashion, this media increases the heat transfer efficiency and reduces the size of the air washers. (See Figure 2) This air washer technique is a mature one, taken from the industrial refrigeration industry for fruit and citrus precooling, and other food processing cooling systems. One system uses chilled hygroscopic brine to reduce the air temperature below the freezing point of water to take advantage of increased air density. A brine re-concentration system is needed to remove the condensed moisture and regenerate the brine to operating strength. The increased maintenance of a direct system and the increased size (volume) of the inlet air plenum as compared to a finned coil system tend to make the cost somewhat greater by comparison.
One major method of inlet air cooling is through the use of primary refrigerants circulated through extended surface (finned) coils mounted in the air stream of the GTG inlet. This conventional technique is similar to the coils and compression refrigeration systems used in the entire range of traditional air conditioning systems.
This inlet air cooling system, (See Figure 3), consists of a free standing central refrigeration system to cool the inlet air using direct evaporation of refrigerant within the finned air cooling coils located in the inlet air stream. These coils are usually mounted downstream of the inlet air prefilter/final filter section. The cooling coils are served by adjacent temperature dedicated refriger-ant recirculation systems including vessels and pumps. One of the advantages of a primary refrigerant system, is that the refrigerant is fed directly to the coils instead of a secondary refrigerant (chilled water, brines), thus utilizing only a single heat exchange, and reducing pump recircu-lation horsepower.
Refrigerant condensing in large indus-trial refrigeration systems is traditionally handled with evaporative condensers which condense the refrigerant inside the tube bundle located in the water/air stream. This arrangement reduces para-sitic pump power and saves one heat transfer process as compared to traditional recirculated cooling tower water with S&T condensers. Air cooled condensing is available when water supplies are unavailable or zero discharge systems are mandated
The consideration and selection of the material for the tubes and fins for the inlet air cooling coils is of extreme importance. The GTIAC system cools a "process" air stream for the GT on a continuous basis. This constant ambient air is naturally corrosive to some fin/tube materials.
Historically in the industrial refrig-eration industry, and particularly with the ammonia refrigeration systems, the materials of choice for cooling coils have been steel tube/steel fin, hot dip galvanized after fabrication (HDGAF). The heavy coating of zinc and the compatibility of mild steel with ammonia has been the most reliable combination over the years. Aluminum coils have also been effectively used in ammonia systems. With proper choice of thicknesses for the fin material (minimum of 0.014"(0.356mm) thick-ness), and heavy wall tubes and return bends, these are believed to be satisfactory. In addition, it is important that the aluminum fin, tube, and return bends be of identical alloys to prevent dielectric corrosion caused by the dissimilarity.
Cooling through the use of secondary chilled water/brine inside finned air cooling coils is a commonly utilized and mature air cooling technique, (used in all large air conditioning systems.) The most prevalent larger systems currently use a secondary refrigerant (chilled water or brine) produced by conventional HVAC water chillers. In a GTIAC system using chilled water or brine, multi-temper-ature coils would be used and connected to dedicated chillers operating at several temperatures similar to the coil arrangements shown in Figure 3. These units in the larger sizes are usually centrifugal compressor/chillers, absor-ption chillers utilizing HRSG steam, or are direct gas fired. Cooling tower water and shell and tube condensers are util-ized for the refrigerant condensing.
The use of secondary refrigerants (chilled water or brines) requires a considerable amount of recirculation pumping horse-power as compared to primary refri-gerants. This total requirement is approx-imately 6 gpm/ton(0.0018ltr.per min/KJ) of refrigeration. For example, the 501D5 GT requires a refrigeration capacity of approximately 6,000 tons(75.9mmKJ) of refrigeration. This then requires 36,000 gpm(136.3cu.m/min) of recirculated chilled and cooling tower condensing water. This volume calls for approxi-mately 300kW of recirculation pumping parasitic power. The same pumping requirement is present for any secondary refrigerant used regardless of type of refrigeration system utilized, i.e. centri-fugal chiller, reciprocating chiller, absor-ption chiller, TES ice maker/storage, or TES chilled water storage.
Secondary chilled water/brine cooling coils offer a different challenge. Traditionally, these types of chilled water and brine coils used in air conditioning systems utilize copper tubes and aluminum fins. Since the thicker aluminum fins described earlier do not allow very close fin spacing, lighter gauge aluminum fins are usually used. These lighter thicknesses do not stand up well in the 100% outside air environ-ments found in GTIAC systems. It is therefore recommended that copper tube, copper fin coils be utilized, which would have very high compatibility and long life when operating with chilled water/brine and compatible primary refrigerants, i.e., Propane, HCFC 22.
The utilization of thermal energy storage (TES) techniques as a way of reducing the parasitic refrigeration load during the on-peak period operation on the GTIAC system can be viable. Consideration of all the economic factors of the power sales, dispatch, and firm capacity guarantees need rigorous review, analysis, and modeling for each site situation.
If the plant has available HRSG steam or low grade hot water/steam/condensate from process, these can be utilized in steam turbine drives, steam absorption water chilling, and ice making systems. Depending on the time/availability and grade of the HRSG energy, this can be used in TES systems for creating and storing chilled water or ice off-peak, for usage on-peak. These TES techniques are mature in the HVAC and food industry. Traditionally, the TES systems are smaller and use more total energy than instantaneous systems, but since the storage is created off-peak, the operating cost is usually less, due to lower off-peak energy rates.
Any TES system has the disadvantage of needing an off-peak "build" time to create the "storage battery" of cooling energy. This makes it unavailable, then, to provide GTIAC during this "build" period without diminishing the peaking capacity of the system at a later time. In addition, the cooling coils in the airstream continuously add to the air pressure drop on the inlet side which reduces the GT capacity during off-peak periods when no air cooling takes place.
The added costs for the TES system over the conventional full time cooling system vary greatly based on system type. Studies in this area have indicated a cost of approximately $100/ton-hr($0.008/KJ-hr) of TES system capacity.
The four major industrial refrigerants which are in use in the industry today are: Anhydrous Ammonia (R717), Propane (R290), HCFC 22 and Refrigerant 134a (replacement for R12).
In summary, our studies of large GTIAC systems have showed that the most efficient refrigerant for this (as well as most other industrial refrigeration systems) is anhydrous ammonia. It was also shown that ammonia was the least in capital cost and lowest in overall operating cost. Relative system capital costs were estimated using the ammonia system as the base system. The propane system was 35% higher and consumed 20% more energy; the R22 system cost 30% more using 10% more energy; the R134a system cost 50% more than the ammonia system and consumed 40% more energy.
Using the Dallas 501D5 example for a 6,000 ton(75.9mmKJ) GTIAC system, the following refrigerant charge quantities and relative costs are estimated. Ammonia: 80,000 lbs(36,364Kg)/$25,000; Propane: 150,000 lbs(68,182Kg)/$75,000; R22: 200,000lbs(90,909Kg)/$500,000; R134a: 500,000 lbs(227,273Kg)/$3,000,000.
For reasons of lowest first cost, highest operating efficiency, the fact that ammonia is environmentally benign, the refrigerant of choice around the world for larger industrial refrigeration systems, coupled with the safety and uncertainty issues with the other candidates, make ammonia the surviv-ing ideal candidate.
To provide the necessary refrigerated cooling capacity downstream to balance the air cooling coils at the GT inlet, the systems, techniques, and component variety available to the systems engineer are broad when coupled to the various prime mover drivers available. The industrial refrigeration and HVAC industry has standardized equipment packages available using all of these varieties and options.
Selection of refrigeration compressors, similar to the application of any gas compression system, covers a range of compressors including screw, centrifugal, and reciprocating compressors. Each has its advantages depending on size, system location, and application. These comp-ressors are used in primary refrigerant systems as well as secondary water/brine chillers.
The prime mover drives utilized for the refrigeration systems, again, are many. The simplest installation and lowest first cost driver, is the electric motor. If HRSG steam is available and economies permit, then consideration should be given to steam turbine drive. Another option for the steam turbine drive is its use in conjunction with steam/hot water absorption chilling systems by letting down through the steam turbine and condensing into the absorption units as a hybrid direct compression/steam absorption system.
Natural gas fueled IC engines or small gas turbines may also be used effectively. Consideration could be given to combining the exhaust from the IC engine/small gas turbine into the main turbine exhaust for combined heat recuperation and emissions treatment.
If the GTG has available low grade HRSG heat, absorption refrigeration water chillers may be utilized along with the finned cooling coils. These can utilize either lithium bromide/water or ammonia/water absorption cycles. The absorption systems have a Coefficient of Performance (COP) ranging around 1.0 and commonly require approximately 10-18 lbs/hr/ton(0.33-0.65g/hr/KJ) of refrigeration of 25-100psig(172-690KPa) steam . Similarly these options are avail-able in direct fired natural gas absorption chillers. Typically these absorption chillers require approximately 7 gpm/ton (0.002 ltr. per min/KJ) of recirculated chilled and C.T. water, and a parasitic load of 0.175 kW/ton of chilling capacity. The advantage of the absorption option is the reduction of parasitic load, which is somewhat less than electric driven compressor systems.
The refrigeration system load (process air load) is the difference between the combined sensible and latent cooling of the ambient inlet air down to the lowest temperature recommended either by the GT manufacturer or the mindset of the system operator. Generally speaking, this minimum temperature design point is 40F(4.4C). And since the air leaving the cooling coils is saturated (100%RH), any pressure drop in the air duct into the bell mouth will cause further cooling and condensation of the remaining moisture in the air stream. This phenomenon is similar to high altitude contrails or the vapor trails one would see off the wing flaps during an airplane landing on a cool moist day. One term for this temperature reduction due to the air duct/bellmouth pressure drop is the "dynamic temperature reduction" and is in the range of 5-10F(2.8-5.6C). Therefore in order to prevent the condensed moisture from freezing at 32F(0C), the so called "ice point", the cooled air temper-ature is limited to approximately 40F(4.4C), or 8F(4.4C) above freezing. Conversely, when the outside ambient conditions begin to cause this "icing" which is usually in the fall or springtime, considerations for inlet air heating become necessary. Some aeroderivative turbines are temperature limited above this ice point for other reasons and require inlet heating to maintain proper output levels.
Let's look at the air cooling process path on the psychrometric chart and follow through several air cooling and dehumidifying/humidifying paths to the ultimate condition. (See Figure 4)
In Case A, evaporative cooling begins at inlet condition, point 1, of 102F(38.9C) DBT, 75F(23.9C) WBT, 29% RH, enthalpy 38.6 btu/lb(89.5KJ/Kg). The evaporative cooling adds moisture to the air as it travels up along the wet bulb/constant enthalpy line. With an 85% effective evaporative cooling media, the approach to the wet bulb temperature, point 2, will be to approximately 102-[(102-75) x 0.05)] = 79F(21.6C) DBT, 75F(23.9C) WBT leaving air temperature, 83% RH.
Often the question is raised as to the efficacy of evaporative cooling when used in series and in conjunction with refrigeration. Observe the process of evaporative cooling following the path up the wet bulb/constant enthalpy line. The dry bulb temperature is reduced and the moisture content is increased. The enthalpy (heat content) and wet bulb temperature remain constant. Since the refrigeration load is a direct function of the enthalpy difference between inlet and outlet, evaporative cooling offers no advantage. In fact, since moisture is added to the air, the refrigeration coil has an increased cooling load due to the added condensation of water at the dew point.
In Case B, refrigerated cooling, the inlet condition point 1 is the same, and the process path followed is horizontal at constant absolute humidity until point 5 at the saturation curve (dew point temperature) is reached, at which time condensation begins. Further cooling follows down the saturation curve, point 6, to the lowest refrigerant coil temper-ature (30F(-1.1C), point 7), approaching it to within approximately 10F(5.6C). The air condition at point 6 is 40F(4.4C) DBT, WBT, 100% RH, enthalpy 15.2 btu/lb (35.2KJ/Kg).
The current trend for equipping gas turbines with dry low NOx combustors is driven by the need to reduce water/steam injection, increase SCR catalyst life, and reduce heat rate. The adverse effect of this change causes a reduction in turbine output of 4-5% which can markedly affect the existing site firm capacity and reduce annual energy production. Refrigerating the inlet air with a GTIAC system provides the means for not only recovering this loss of output, but can provide additional capacity as well.
The control of power output of a GT is currently maintained in several ways. These include the controlling of the fuel input rate, the adjustment of the inlet guide vanes, and the rate of steam injection for power enhancement, or a combination of all three.
A GTIAC system offers an additional method of controlling the output of a GT by the expedient of linking the turbine control to the GT inlet air temperature, which is controlled by the refrigeration system. With a given setting of turbine output, the inlet air cooling system can be modulated so that the temperature of the air delivered to the turbine is adjusted upward to reduce the turbine capacity and downward to increase the turbine output. This modulation technique has the potential of allowing the inlet guide vanes to be positioned at their highest efficiency point while optimizing the fuel and steam injection rates. (Additional patents are pending on these features)
Carnot established that the highest efficiency is achieved when any heat transfer process is carried out at the highest temperature consistent with achieving the desired result. Hence, the approach to cooling the inlet air stream with multiples of decreasing refrigerant temperatures in multiple, separate coil sections aims directly at this goal. As a practical matter, multi-stage, multi-temperature, coil/compressor configur-ations have proven to be most efficient and economically viable for GTIAC systems.
With any refrigeration system using primary or secondary refrigerants, there are several feasible coil configurations. The traditional method is to use a single cooling coil operating at a single temperature sufficiently low enough to create the final air temperature (delivered to the mouth of the turbine). This system uses the most refrigeration energy which encourages the investi-gation, development, and use of more efficient systems.
The multi-stage, cascaded, multiple temperature/coil arrangement is a GTIAC system which carries out the air cooling through several in-line cooling coils operating as separate coil sections at dedicated refrigerant temperatures. In the Dallas example, see Figure 5, the 102F(38.9C) incoming air passes through the first coil and is initially cooled to 71F(21.7C) with the refrigerant at 61F(16.1C) The air then continues through the second coil section and is cooled to approximately 55F(12.8C) with 45F(7.2C) refrigerant. The third and final coil section cools the air down to its lowest temperature of 40F(4.4C) with 30F(-1.1C) refrigerant. Each section requires separate compressors and associated piping.
This technique called "sectionalization" or "temperature staging", optimizes the thermodynamics by dedicating the refrigerant temperatures to separate compressor "stages". In addition, this temperature staging is coupled with pre- cooling/cascading of the refrigerant for each stage by the preceding stage for maximum thermodynamic efficiency. This allows each stage to operate at a higher, more efficient temperature; each with less energy per unit of refrigeration produced than a single lower temperature coil system.
Sectionalization then provides a higher thermodynamic cooling efficiency. This technique is common to both primary and secondary refrigerants. The comparison of a primary single stage GTIAC cooling system to a 3-stage system for the Dallas example indicates a reduc-tion of 24% in the parasitic energy con-sumption when compared to the traditional single temperature system.
Capital cost for a 3-stage coil system is slightly higher in the smaller systems than its single temperature counterpart by approximately 10% for a typical installation. However, the 18-25% net increase in system efficiency offsets this higher first cost and continues to pay over the life of the project, in terms of reduced parasitic electrical power consumption, providing additional power to sell.
As the GTIAC systems get larger, as in the Dallas 501D5 system, the multiple temperature/stage system capital cost is less than the single temperature system. This combination of temperature staging and cascading along with computer controlled load modulation by ambient temperature demand is the subject of the author's system patents now approved by the U.S. Patent Office.
The control system for a 3-stage GTIAC system is designed to maintain the temperature of the air leaving the third stage cooling coil at 40F(4.4C). The air temperature leaving the third stage coil is the process variable by which the refrigeration system is controlled. This temperature will be used to control the slide valves of the compressors to modulate their capacity and maintain the unit inlet air temperature at the design point of 40F(4.4C).
In order to prevent icing conditions at the turbine inlet, "freezestats" (safety shutdown sensors) are mounted at the turbine inlet flange to measure and limit the minimum air temperature. These will back up the operating thermostat to prevent freezing of the moisture in the airstream and inlet bell/struts to prevent the ingestion of ice particles or "rim ice" calving off into the turbine. The entire equipment control function and sequencing logic will reside within a PC based control system which will communicate with the current turbine control system using standard protocols.
HRSG production is affected by the total mass throughput and temperature of the turbine exhaust gas. The exhaust gas temperature (EGT) decrease resulting from GTIAC is approximately 5%. However, the EG mass flow is increased approximately 12%. This combination results in an approximate 8% increase in HRSG steam production. This added energy is available for use in the combined cycle or other process use, or reduced HRSG size if the increased steam produced is not usable.
The increase in GTG output resulting from the GTIAC system (particularly in retrofit) requires a new look at the capacities of the existing generator, transformers and electrical wiring capacities of these existing components. This to make sure that there is sufficient capacity in each. In come cases the generator can be upgraded with additional generator cooling which can be integrated within the GTIAC system.
In order to summarize the benefits of GTIAC, it is necessary to provide a rigorous analysis of the site specific factors including all of the owning and operating costs of the various system configurations coupled to the complete economics analysis considering all of the time related power sales agreements and dispatch commitments. These need to be completely modeled for the full 8,760 annual hours in order to provide the developed options sufficient for a final decision. Today we have the technology and the computer tools to provide these answers.
The computer model for the 501D5 CT is based on Dallas, Texas ASHRAE 1% maximum design ambient air temperature conditions with annual weather data taken from the US ISMCS data for that location. The 8,760 hour model interactively involves the critical GT variables with the inlet air cooling system refrigeration variables and predicts all of the necessary output of performance increase, parasitic power, and full economics. This master model has then been truncated to develop a usable preliminary screening "snapshot" model enabling a more rapid analysis of each GT configuration. This tool is being used routinely for these purposes.
Figures 6 and 7 show the Dallas, Texas average monthly temperatures and the resultant 501D5 GTG output and heat rate response.
The plot plan (See Figure 6) for the Dallas system shows a generic layout of one Westinghouse 501D5 Combustion Turbine with a 3-stage/temperature inlet air cooling system nestled alongside.
The flow diagram, (Refer back to Figure 3) for the refrigeration system shows three cooling coils installed in the inlet air plenum behind the air filter section and weather louvers. Three ammonia refrigerant pump/recirculation vessels are interconnected to the coils along with three screw compressors and three evaporative condensers, completing the system. By comparison, the single stage system would require approximately 25% more capacity, (4) screw compressors and (4) evaporative condensers and a single temperature, inlet air cooling coil with somewhat fewer rows deep to the airstream.
Generally speaking, the normal approach to the author's snapshot computer screening analysis is to look at two design points. The first is a maximum design point which is established by the ASHRAE 1% weather data for the site. These are the maximum temperature conditions which the GT would experience at the site for all but 87 hours, or 1% of the total annual operating hours. These conditions establish the maximum refrigeration load conditions, which were calculated to be 6,000 tons (75.9mmKJ). This cooling provides constant 40F(4.4C) air to the turbine inlet during the hottest summer days, providing maximum GTG output.
The minimum system design point considered is the lower or minimum design point that is established by the US ISMCS average summer day temperatures for the site. This minimum capacity GTIAC system would provide 40F air to the turbine at all times for temperatures at or below this point. When the ambient temperature increases above this average, the air temperature to the GT will increase proportionately. For the Dallas site, this refrigeration design point is 5,200 tons(65.8mmKJ). This lower capacity would produce 44F(6.7C) air to the GT on the maximum (ASHRAE 1%) day.
This minimum/maximum design/cost "window" becomes the "performance envelope" made possible by the GTIAC system. The final economic performance design point and decision for the entire system would fall within this envelope and would be ultimately based on a site specific design and full 8,760 hour model with complete economic analyses. The Dallas 501D5 case study is based on the larger refrigeration system which provides maximum system output and continuous 40F(4.4C) air.
The comparative analysis of GTIAC systems for a 501D5 Westinghouse turbine sited in Dallas, Texas in the various configurations of "Non-Cooled", "Evaporative Cooled", "Single Stage", and "3-Stage" Refrigerated Cooling, provide some definitive answers for the relative advantages of each system.
Table 1, "Conclusions/Final Recap Com-parison" provides the summary of the advantages of GTIAC compared to a non-cooled base example. The net power output is increased by 26 mW (24%). Gross heat rate is reduced 666 BTU/kW(702 KJ/kW) (6%). Revenue from additional annual power sales is increased by $ 3 mm, providing an actual payback of the $6 mm of added capital cost of 2 years.
The Net Present Value (NPV) of the savings @ 12%/10yrs, is $ 14.4 mm. Finally, this results in an adjusted capital cost per added kW of capacity of $89/kW (Adjusted to ISO and by crediting the capital cost by the first year savings.)
The advantages of GTIAC on the individual GT is independent of other system features of combined cycle, duct firing, etc. This makes it transparent to any of those other features. The overall system thermal efficiency increase derived by GTIAC is approximately 1%.
The gains in output and fuel efficiency derived from GTIAC make the use of some type of system almost mandatory. Certainly evaporative cooling, considering its low first cost, is almost always a consideration in all but the most humid areas. Prudent engineering design would dictate that, as a minimum, sufficient structure and enclosure size should be included for the ultimate addition of refrigerated GTIAC coils. When cooling coils are installed, the evaporative system media would be removed to reduce the media added inlet air pressure drop.
Given the increased performance of the 3-Stage refrigerated system over the single stage system, and since both system costs are equal, decision for the 3-Stage system is the obvious choice.
This final iteration within any engineering study which includes engineering economics down to the so called "irreducibles", then forces the final decision to an emotional one and is termed "the luxury of management". It has been said that "Our job as engineers, ultimately, is to so sharpen the available options for management that their making the decision based on all of the economic rules and parameters is easy".
This dilemma can be reconciled by the appropriate application of the economic variables which the computer program models. Once the economic rules are agreed to and established, thanks to the computer simulator and modeling tools available to us, an absolute answer within that framework can be achieved.