| GAS TURBINE INLET-AIR CHILLING AT A COGENERATION FACILITY |
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By A.D. Hall; J.C. Stover; and Ralph Breisch, Member ASHRAE 
Abstract Introduction The plant operates all year at base load. The summer season is when power sales are most valuable. The plant makes almost 80% of its electrical revenues between May 1 and October 31, yet the plant power output is substantially reduced by the high ambient temperature. Inlet air chilling with TES was installed at the facility to increase output during critical "on-peak" hours in the summer when maximum unit performance is required. The chiller/TES system consists of a mechanical vapor-compression refrigeration cycle driving an ice "harvester" that is operated in the evening hours to stockpile ice in a thermal energy storage tank. Chilled water from the tank is circulated through cooling coils at the gas turbine air inlet during the heat of the day to increase the plant's electrical output. Plant Description In new and clean condition, the unit is rated for 36.24 MW gross output at the generator terminals with ambient air at ISO conditions (dry-bulb temperature, 59°F; relative humidity, 60%). Approximately 1 MW is used by the plant's station light and power requirements. At the rated output, 1,050,000 lb/h of air is consumed by the engine. Inlet Chilling Concept The power purchase agreement is structured so that the utility pays an energy payment for every kilowatt-hour (kWh) delivered, a capacity payment for delivering power at no less than 85% of a dedicated firm capacity level, and a bonus payment based on how well the plant meets the remaining 15% of the dedicated firm capacity during on-peak hours. No bonus is earned on kilowatts delivered above the dedicated firm capacity. The dedicated firm capacity level was set by a test of plant output at the time commercial operation began. That was, as it happened, in winter months when cool temperatures and the new and clean condition of the unit allowed the firm to set a 35.5 MW dedicated firm capacity level. The gas turbine air inlet system was originally equipped with an evaporative cooler to reduce inlet dry-bulb temperature. With the evaporative cooler operating at 85% effectiveness on a typical 95°F day with 20% relative humidity, the net output of the plant is at best 34 MW. Further reduction of output occurs due to unit degradation--the effect of fouling, erosion, corrosion, and foreign object damage that inevitably degrades performance by reducing compressor airflow. Typically, such degradation will advance very rapidly during the first two or three years of operation to as much as 6% of output capability. Hence, performance of the plant would not meet the 35.5 MW dedicated capacity level and was subject to the loss of a large share of potential bonus revenue. To compensate for temperature and degradation effects, inlet air chilling was installed. Figure 1 shows the effect of inlet temperature vs. gross generator output. Design temperature for the chilled inlet air is 42°F. In addition to increased output, chilled inlet air improves the gas turbine heat rate. The net plant heat rate is worse when additional station power is used to generate ice at night, but the effect is almost completely mitigated by the heat rate improvement when chilling. Figure 2 shows the effect of inlet temperature vs. heat rate. Overall, net plant output with inlet air chilled to 42°F during on-peak hours now satisfies the dedicated capacity level. Figure 3 compares inlet temperature vs. net plant output at both new and clean conditions and with a 6% degradation factor. Operating points are indicated for typical summer on-peak ambient conditions (95°F, 20% RH) with and without evaporative cooling and with chilling. 
Design Considerations
where V = 350 fps, g = 32.2 ft . lbm/lbf . s2, and 778 is the conversion factor (ft . lbf/Btu). From Equation 1, the resulting temperature drop is then:
where cp = .24 Btu/lbm . R for air at low pressures. With this limitation in mind, three types of inlet air chilling were considered:
The overall effect of net power produced for each scheme is shown in Figure 4. Direct mechanical refrigeration consists of a vapor-compression refrigeration system to chill inlet air to 42°F during the on-peak hours without benefit of thermal storage. This system has the added benefit of chilling capability during all hours of the day. However, the incremental revenue to be gained from this capability is not significant. Furthermore, there is a 1,300-kW penalty associated with running the refrigeration compressor while chilling. The compressor load penalty would lower the net plant power output during on-peak hours. Referring back to Figure 3, it is obvious that, using the degraded engine curve, the plant would not meet 35.5 MW with 42°F inlet air if an additional 1,300-kW station load were subtracted from the net output. Furthermore, compared with thermal energy storage, the installed refrigeration capacity required for direct refrigeration would be three times larger. Direct refrigeration would have to be sized to deliver the full instantaneous chilling duty for a turbine generator. With TES, the chilling duty for 6 hours is accumulated and stored over 18 hours, thus reducing the refrigeration required by a factor of 6/18, or one-third the size required for direct refrigeration. Absorption refrigeration, such as a lithium bromide system, was considered. It would have required modifying the existing HRSG to pick up low-pressure steam to drive the absorption system, thereby avoiding the power penalty associated with a direct mechanical refrigeration system. A drawback, however, is that there is no cooling water available at the site. A closed-loop cooling system would have been required, with the heat rejected to the air with an aerial cooler. As a result, during the on-peak hours when the ambient temperature is above 90°F, it would be difficult to chill the inlet air below 51°F. This system would not maximize power output during on-peak hours. Thermal energy storage was selected because it allows power output to be maximized during the on-peak hours. The TES configuration allows operation of the refrigeration system when the value for power is lowest. In turn, the refrigeration system is turned off during on-peak hours to minimize station load, and inlet chilling is accomplished with the stored energy. The size of the refrigeration equipment is optimized with TES since it is allowed 18 hours of operation to store only 6 hours of chilling capacity. The method of energy storage with ice was evaluated against cold water and cold brine storage. The latent heat of fusion for ice, 143.5 Btu/lbm, substantially reduces the mass required to store energy, providing a more compact and economical system. After ice storage design considerations, such as air space and circulating water space, are accounted for, the volume required to store a given amount of energy with ice is about one-fifth the volume required to store an equivalent amount of energy with liquid. With ice storage, chilled water can be consistently supplied to the inlet air coil at 32°F to 34°F, whereas if storing chilled liquid, the temperature of the liquid gradually increases during the six-hours-per-day use cycle, creating a transient heat transfer problem at the air inlet. System Design Basis The system was not designed to fully accommodate extreme weather conditions of more than 100°F with greater than 15% relative humidity, which only periodically occurs at the site. Under these conditions, the system may be operated to chill inlet air to a point above 42°F in order to maintain enough ice storage to last the week. More ice storage capacity, or refrigeration tonnage, necessary to accommodate extreme conditions was deemed too costly for optimum economic return on investment. Another option, reducing the size of the refrigeration equipment to a capacity of 300 to 350 tons, was considered. With this scenario, an ice storage tank 1.7 times larger than the one selected would be required. The savings in the size of the refrigeration equipment did not offset the cost of the increased storage. For this reason, as well as fear of ice/water distribution problems in a tank this size, this option was not selected. 
Inlet Air Chiller Coil Design The resultant design in three sections of horizontal tube bundles, with each bundle having a tube length of 26.6 feet. The bundles are aluminum finned tubes with a diameter of 1.25 inches; the fin diameter is 2.5 inches with 12 fins per inch. Tubes are arranged four deep in a triangular pitch with 176 tubes per bundle. Each bundle is set in a galvanized steel framework that mounts on a concrete ring wall built around the outside perimeter of the inlet housing. The coils are supported by the ring wall on the front and sides of the filter house. Aluminum sheet metal was installed along the back of the filter house and between the coil housings and the filter house to prevent airflow from bypassing the coils. Chilled water flow to the coils is controlled by a single flow control valve that regulates on a desired inlet air temperature setpoint. The plant's distributed control system manages valve position with feedback from air temperature sensors positioned at the compressor inlet. Water flows in parallel to the three tube bundle sections. The coil arrangement allowed for ease of installation and a minimum reconfiguration of the existing plant. Pressure loss across the coils is only 0.5 inch H2O due to the low air velocity. There was concern about placing the coils upstream of the inlet air filters when chilling below the dew point since this would expose the filters to possible carry-over of condensed water droplets from the coils. Carryover has not proved to be a problem; condensate forms on the coils and falls down into the basin formed inside the ring wall and is drained away.
TES Design As ice falls into the tank, it forms a mound similar to sand in an hourglass. The shape of the angle of the mound is called the angle of repose. The ice built on the ice harvester develops a 15° angle of repose. The location and size of the ice opening relative to the tank-top area is critical. The evaporator plates for this application were positioned to minimize tank void volumes and optimize tank utilization. The size of the ice storage tank was determined based on 57.2 lbm of ice per ft3 and the 1,235,520 lbm of maximum ice storage required for the design condition shown in Figure 5. Assuming a 50% water/ice ration below the water level and a 50% air/ice ratio above, the tank minimum volume requirement was as follows: 2 . 1,235,520 57.2 = 43,200 ft3 Additional space allowance is added for the ice mounding at the top of the tank. The final internal dimensions of the tank are 45 feet long by 38 feet wide by 26 feet high. The ice storage tank is cast-in-place concrete. The tank is installed partially in ground on a hillside. One end of the tank is exposed from the hillside; this is where circulating pumps are installed, avoiding the need for a pump vault below grade. The tank is initially filled with water to the 60% level. As ice is dropped into the tank it floats, with 91.7% of the ice below the water level and the remainder above. The water level remains constant during the initial stages of the charge cycle. When the ice level meets the bottom of the tank, the water is displaced and the water level starts to drop. When the high ice level has been reached, the water is at the 20% level. During the charging (ice-making) cycle, 2,000 gpm is pumped over the evaporator. During the discharge (melting) cycle, 2,000 to 4,000 gpm will be pumped through the system as needed to meet the chilling demand. During the discharge cycle, when the chilled water feeds the turbine inlet air coil, the water will bypass the evaporator and flow directly into the tank. The "warm" return water is distributed in the tank via a spray header. The spray header is mounted at the top of the tank and evenly distributes the water over the ice pile. This even distribution of return water over the ice is necessary to maintain a constant supply water temperature to the coil through the entire discharge cycle. Refrigeration Design Due to the large size of the system, a single screw compressor was chosen over multiple reciprocating compressors. While in the ice-making mode, the compressor operates with a coefficient of performance (COP) of 2.8. Ammonia was chosen as the refrigerant due to the operators' familiarity with it (ammonia is stored on-site for a selective catalytic reduction system) and because it has environmentally safe qualities: zero ozone depletion and greenhouse effect factors. System Operation A sister plant, a cogeneration plant situated two miles away, is identical to this facility in every respect except that it has evaporative cooling instead of an inlet chilling/TES system. The sister plant experiences the same ambient air conditions and thus provides an ideal yardstick by which to measure the benefit of inlet chilling at the original facility. During June, July, and August, the sister plant generated an average net output of 34.26 MW compared to the facility's average net output of 36.57 MW during on-peak hours. The inlet air temperature averaged 67°F at the sister plant with the evaporative cooler. The inlet air temperature averaged 51°F at this facility with the chiller/TES system. In future summer periods, the output of the sister plant will continue to degrade without the ability to make up the loss with chilled inlet air. The inlet temperature at this facility will be depressed further with the chiller/TES system, thus overcoming degradation in order to keep output above 35.5 MW. Conclusion 
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ombustion gas turbines are constant-volume engines for which shaft horsepower is proportional to the combustion air mass flow. Engine output improves if the air temperature is depressed at the compressor inlet to increase the air density. When a combustion turbine generator is used in a power plant, increased engine output increases the electrical generating capacity. That is the concept presented in this case study of an inlet air-chilling system installed in a cogeneration plant in California. The plant also uses a thermal energy storage (TES) system with the inlet air chiller to optimize the plantŐs economic performance. This application has improved the hot weather generating capacity by 10%.
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