Techno-economic analysis of waste heat recovery systems for wet-cooled combined cycle power plants

Accepted Manuscript Techno-economic analysis of waste heat recovery systems for Wet-Cooled Combined Cycle Power Plants Achyut Paudel, Todd Bandhauer PII: DOI: Reference:

S1359-4311(18)30111-X https://doi.org/10.1016/j.applthermaleng.2018.07.138 ATE 12496

To appear in:

Applied Thermal Engineering

Received Date: Revised Date: Accepted Date:

6 January 2018 5 July 2018 29 July 2018

Please cite this article as: A. Paudel, T. Bandhauer, Techno-economic analysis of waste heat recovery systems for Wet-Cooled Combined Cycle Power Plants, Applied Thermal Engineering (2018), doi: https://doi.org/10.1016/ j.applthermaleng.2018.07.138

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TECHNO-ECONOMIC ANALYSIS OF WASTE HEAT RECOVERY SYSTEMS FOR WET-COOLED COMBINED CYCLE POWER PLANTS

Achyut Paudel and Todd Bandhauer1

Interdisciplinary Thermal Science Laboratory, Department of Mechanical Engineering, Colorado State University, Fort Collins, CO, 80523

1

corresponding author email: [email protected] phone: 970-491-7357 fax: 970-491-3827 Address: Colorado State University Department of Mechanical Engineering 1374 Campus Delivery Fort Collins, CO 80523

ABSTRACT Increasing ambient temperature negatively impacts the performance of natural gas combined cycle (NGCC) power plants. There have been multiple approaches to mitigate this performance reduction by chilling the compressor inlet air of the gas turbine, as well as by recovering the waste heat emanating from the power plant to generate additional power. In the present study, a detailed techno-economic assessment on the application of different types of waste heat recovery systems used to chill the compressor inlet air has been assessed. A simplified thermodynamic and heat transfer model is developed to predict the performance of an evaporatively cooled NGCC at varying ambient conditions. By taking typical meteorological year (TMY3) hourly weather data for two different locations – Los Angeles, California and Houston, Texas – the yearly output for a 565 MW plant is predicted at a 100% capacity factor. The feasibilities of different waste heat recovery (WHR) systems including a flue gas driven absorption chiller, a steam driven absorption chiller, and an electrically driven vapor compression chiller are assessed by calculating the levelized cost of electricity (LCOE) for each scenario. The results of the analysis showed that, for a fixed WHR system costs (i.e., $ per kWth), the system powered by flue gas generated the smallest LCOE, followed by the mechanically-driven vapor compression, steam-heated chiller, and, finally, the gas turbine exhaust chiller for both the locations at all COP combinations. The analysis also investigated the impact of fixed investments cost, and the flue gas system again yielded the smallest LCOE, while still yielding a lower LCOE than at baseline case over a wide range of COP and tolerable cooling costs for both locations.

KEYWORDS:

Techno-economic analysis, waste heat recovery, combined cycle power

plants

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1

INTRODUCTION Natural gas combined cycle (NGCC) power plants contain both a gas turbine and

bottoming steam cycle that is powered by the high temperature exhaust emanating from gas turbine. At steady state and full load, the overall efficiency for conventional NGCC power plants is typically between 50-52% [1], while modern systems have been able to achieve thermal efficiencies higher than 60% [2]. At ISO operating conditions, the thermal efficiency of the gas turbine cycle for these power plants are nominally 30% [3]. Gas turbines are constant volume machines, and, hence, an increase in ambient temperature reduces the intake air density into the compressor. As a result, the air mass flow rate into the compressor reduces, which lowers both the efficiency and power output of the gas turbine, as well as the bottoming Rankine cycle [4]. For example, on a typical system, every 1°C temperature rise in ambient temperature above design condition can yield a loss of 0.1% in thermal efficiency and 0.5-0.9% reduction in gross output [5]. Unfortunately, these losses generally correspond with the time when there is peak power demand [6]. Compressor inlet air chilling has been employed to counteract these effects and to increase the power plant output during these peak demand periods [5]. Evaporative coolers or electrically driven vapor-compression chillers are used to provide compressor inlet air cooling. However, evaporative chillers suffer from performance degradation as the relative humidity increases, and electrically driven compression chillers consume a significant amount of electric power [7]. As an alternative, heat-driven absorption chillers are potentially a suitable choice, as they require a small fraction of the electrical energy consumed by vapor compression chillers, contain environmentally-friendly fluids (e.g., water and ammonia) and can be easily integrated within the existing facility [8]. Rahim compared evaporative cooling, fogging, absorption cooling and electrical chilling options for a 96 MW gas turbine plant and found that the most efficient option was to utilize an absorption chiller, which increased the 3

power output by 3.5 MW and the efficiency by 0.05% by cooling the inlet air to 10°C [9] . Similarly, Dawoud et. al [10] compared different gas turbine inlet air cooling techniques for a preinstalled gas turbine (39.62 MW) in Oman. The results of the analysis showed that fogging technique generated 11.4% more electricity compared to evaporative cooling, whereas LiBr-H2O absorption cooling generated 40% more electricity compared to fogging. Compared to Li-Br systems, a water-ammonia absorption system and a vapor compression system generated 39% and 46% more electricity for the same location, respectively. In a different experimental study performed on a preinstalled 336 MW combined cycle plant, Boonnasa et al. showed that installing a steam driven absorption chiller to cool compressor intake air can enhance the annual power production for both the gas turbine and the combined cycle by 10.6% and 6.24%, respectively. The cost of this absorption unit was $158 per kWth [11]. In addition, recovering the waste heat generated from a power plant and utilizing it to power these thermally activated cooling systems is an efficient option to boost the thermal efficiency. In a recent study, Popli et. al. [12] analyzed the effectiveness of utilizing a single effect LiBr-H2O absorption chiller powered by the gas turbine (8.96 MW) exhaust to cool the gas turbine compressor inlet air. The results showed that by utilizing 17 MW of the exhaust waste heat, the chiller was able to produce 12.3 MW of cooling to reduce the temperature of compressor inlet air to 10°C. This approach was able to generate additional 5263 MWh of electricity per year, and the payback period for this retrofitting option was estimated between 1.3 and 3.4 years. Similarly, Barigozzi et. al. [13] performed a techno-economic assessment of gas turbine inlet air cooling by utilizing an electrically driven chiller for a combined cycle unit at three different climatic conditions. The results showed that the cooling system was best suited for a location representative of a high temperature and low humidity weather conditions.

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The previous studies have primarily focused on comparing different gas turbine inlet air cooling techniques, with very few studies conducted on utilizing a waste heat powered cooling system to chill the compressor inlet air. Moreover, these studies were only focused on enhancing the performance and did not discuss the economic effectiveness of these techniques. Although some studies have performed a pay-back period analysis in comparing different inlet chilling systems, a detailed economic analysis that determines the impact of these techniques on the cost of electricity has not been done. In addition, there has not been any study that compares the effectiveness of different waste heat recovery scenarios in terms of the LCOE. In the present study, compressor inlet chilling by utilizing absorption cooling systems powered by different forms of waste heat in a wet-cooled NGCC power plant are compared with an electrically driven vapor compression cooling system. Three possible waste heat sources for the absorption chilling system are investigated: high grade heat from the gas turbine discharge (~600-650°C), low pressure steam (~5 bar) from the bottoming Rankine cycle, and low-grade waste heat from the power plant stack (~106°C). The impact of these different waste heat recovery (WHR) schemes and extracting mechanical power to drive a vapor compression system is determined for two locations (Los Angeles, CA, and Houston, TX) over the course of an entire year using variable weather conditions. Los Angeles represents a milder climate, with the ambient temperature ranging between 10°C and 20°C for ~60 % of the entire year and an average relative humidity of 73%. In contrast, Houston has a more extreme climate, with temperatures reaching as high as 40°C during the summer and as low as -7°C during the winter, with the average relative humidity of 73%. Therefore, it is anticipated that compressor inlet air chilling to increase the power output of the plant is more promising for locations similar to Houston during the summer months, than for milder climates like Los Angeles. 5

To simulate the weather-dependent performance of the power plant, a representative NGGC cycle with performance characteristics given at a single operating condition is scaled using a method previously utilized to predict the flue gas exhaust temperature as the ambient temperature changes [14]. Modifications to this model are made to investigate the different heat recovery and compressor inlet air chilling scenarios, including the addition of an evaporative cooling tower model that accounts for the effect of relative humidity and ambient temperature on the performance. To assess the economic viability of these options, the Levelized Cost of Electricity (LCOE) is calculated using a range of possible additional capital costs associated with these additional systems. In the following sections, a detailed thermodynamic and heat transfer modeling approach to predict the performance of the wet cooled NGCC system at different ambient conditions is presented. Afterwards, a detailed economic approach to calculate the LCOE of the system is discussed. Finally, the results of the study are presented and the most feasible options among different WHR strategies are identified. 2

NGCC PERFORMANCE MODELING The method used to predict the hourly output of the wet-cooled NGCC power plant is

to scale the performance of a particular plant so that its output can be determined when the ambient temperature changes. The baseline performance for the NGCC power plant investigated in this study is taken from the U.S. Department of Energy (DOE) National Energy Technology Laboratory (NETL) Case 13, which is summarized in Table 1 [15]. This plant generates 565 MW of electrical power at an ambient temperature of 15°C, with 362 MW generated in an F-class gas turbine and 203 MW generated by the steam turbines. 10 MW is consumed by the auxiliary systems, resulting in the net power generation of 555 MW. Table 2 summarizes the lists of major assumptions used in this study to predict the performance as the ambient conditions change. In the following sections, the performance 6

prediction methods for the gas turbine cycle, Rankine cycle, as well as the performance prediction for the evaporative cooling tower, are discussed in detail. 2.1

Gas Turbine The baseline case in NETL Case 13 utilizes an F-class gas turbine. These gas turbines

have multi-stage axial flow compressors. In these compressors, the pressure ratio is a function of non-dimensional mass flow parameter and non-dimensional rotational speed [16]. At off design and part load conditions, the compressor is operated on a “working line” to achieve maximum possible efficiency. To maintain maximum efficiency, a decrease in the mass flow rate causes the operational speed and the pressure ratio to follow this working line, which can be predicted by equation (1) as follows [14]:

m CpT01 D 2 p01

1 (  1/2p )

p   C  0e   p01 

(1)

The Cp and D on the left side of the equation are held as constant at different operating conditions. The inlet pressure po1 and temperature To1 are varied according to the available TMY3 data at a given location. The Cp value for air is assumed to be constant within the operating temperature range. D represents the diameter at the exit of the compressor. The constant C on the right side of the equation is the product of the compressor non-dimensional mass flow at the exit during choked condition and the exit area ratio. For a small range of fluctuation of the incoming mass flow of air, a reasonable assumption is to assume a constant polytropic efficiency [16]. Hence, to simulate the effect of changing ambient temperature, the ratio of the design point (i.e., NETL Case 13) to the off-design point can now be simplified to yield a relation between the inlet temperature, air mass flow and pressure ratio given by equation (2) as follows:

m m

T01 T01

 

 p01 ODP  p01 DP ODP DP

7

1 (  1/2p )

 PRDP     PRODP 

(2)

In addition, by assuming a constant volumetric flow rate at off-design operating conditions, the mass flow rate, pressure ratio across the compressor and the corresponding isentropic efficiency of the compressor can now be determined. The combustion in gas turbines can be modeled by assuming a low NOx combustion chamber, which are nearly 100% efficient and can convert almost all fuel energy into heat energy with minimum losses [17]. Hence, an adiabatic combustion is assumed in the gas turbine combustor. The mass flow rate of natural gas is also adjusted such that the turbine inlet temperature always remains constant. This approach is practiced widely to maintain the turbine inlet temperature at a constant value because the turbine blades are exposed to a significant thermal stresses at higher inlet temperatures [18].

The turbine isentropic

efficiency as well as the turbine outlet pressure is assumed to be constant in the present study. 2.2

Steam Cycle This section outlines the major assumptions and processes used to model the off-

design Rankine cycle performance. The Rankine cycle consists of the heat recovery steam generator (HRSG), steam turbines, steam condenser, and condensate pumps. The HRSG is a complex set of heat exchangers that utilizes the waste heat from gas turbine exhaust to generate superheated steam for the steam turbines. The baseline NETL Case 13 used in this study utilizes a triple pressure HRSG, which consists of multiple sets of heat exchangers: economizers, evaporators, superheaters, reheaters, and preheaters, each divided into high pressure (HP), intermediate pressure (IP) and low pressure (LP) systems. Figure 1 shows the process flow diagram of the NGCC power plant with a triple pressure HRSG that has been used in the present study. The turbine exhaust gases passes from left to the right through each heat exchangers, starting from the Reheater and finally to the power plant stack. Superheated steam is produced in each superheater and reheater, which is then sent to the respective steam turbines where the steam is expanded and power is generated. The LP steam leaving the LP 8

steam turbine is condensed in the surface condenser. Nominally, the quality of the incoming steam to the condenser is limited to 88%, because deposition of wet steam in turbine blades can result in erosion of turbine blade material and result in significant damage [19]. In the present study, the quality of the condensing steam is above 90% at all operating conditions. The condensate formed in the condenser is pumped to a higher pressure and sent to the preheater where it absorbs the last remaining portion of heat energy from the flue gases. After getting preheated, the hot water is then sent to the feed water pumps where it gets pressurized into respective HP, IP and LP pressures. These streams are sent to the economizers where it gets heated nearly to the saturation temperature before being sent to the evaporators. The dry steam with quality of 1 exits the steam drum on the top of the evaporators, which is then sent to the superheaters to achieve the desired degree of superheat. The isentropic efficiency for feed water pumps and steam turbines are held constant in this study for all operating conditions. Figure 2 shows the temperature profile across the HRSG heat exchangers at the baseline operating case. For the heat exchangers, mass and energy balances are used to predict the temperatures across each of the heat exchanger modules. By varying the steam mass flow rates, the temperatures across each heat exchanger is adjusted that the effectiveness of all heat exchangers is less than 0.85 at the baseline case. By utilizing cross flow and counter-flow NTU-ɛ relationships for heat exchangers, the representative overall heat transfer conductance (UA) for all heat exchangers at the baseline operating case are determined. To determine the performance of the heat exchangers at ambient temperatures different than the baseline case, the UAs for the HRSG heat exchangers is scaled by using equation(3). (UA)DP  mg,DP    (UA)ODP  mg,ODP  9

0.63

(3)

In equation 3, it has been assumed that the overall heat transfer coefficient for the steam, tube and gas side is nearly equal to the heat transfer coefficient on the gas side alone as it constitutes the significant majority of the heat transfer resistance. The heat transfer coefficient is primarily a function of Reynolds number and gas mass flow rate, and, therefore, the ratio of the UAs are modeled as a function of gas mass flow rate with an approximating exponent of 0.63, consistent with aligned tube banks in the HRSG heat exchangers [20]. Additional details of this scaling process can be found elsewhere [14]. This approach of scaling the UAs of the heat exchanger accounts for the change in the overall heat transfer caused due to the change in the gas turbine exhaust mass flow rate. Figure 3 shows the representative temperature diagram across the same heat exchangers at ambient temperature of 40°C. The temperatures across each heat exchanger increases slightly as the gas turbine exhaust temperature increases. However, a significant decrease in the gas flow rate results in the decrease in the heat exchanger UAs and a concomitant decrease in steam mass flow rates. The scaled UA values at different ambient temperature are then used to solve for the effectiveness of the heat exchangers which in-turn is used to solve for the HP, IP, and LP superheat temperatures and mass flow rates while keeping the inlet saturation pressures the same as in design case. Another crucial approach adopted in the modeling process is to vary the condenser saturation pressure at different ambient conditions. The condenser saturation pressure directly affects the LP turbine work at different operating conditions. At higher operating ambient temperatures, a fixed cooling tower size requires the condenser saturation pressure to increase, which results in an increase in the LP steam turbine back pressure. This increase in the turbine back pressure is accompanied by significant loss in power output from the LP steam turbine. To accurately assess the effects of ambient temperature and also to account for the effects of the changing relative humidity in cooling performance and overall output of the 10

power plant, an evaporative cooling tower model is utilized and coupled to the NGCC model, which is described in the following section. 2.3

Cooling Tower An evaporative cooling tower is used to remove heat from the condenser in the steam

cycle. These type of wet cooled condensers utilize circulating cooling water between the cooling tower, where water is partially evaporated and cooled down, and the condenser in the Rankine power cycle. The quantity of water that evaporates through the cooling tower is then resupplied in the form of make-up water, which is nominally 2% of the total water flow[19]. Air cooled condenser technology is quite new to power plant cooling and not yet widely adopted in the US: only on 1% of U.S. plants are dry cooled as of May 2015 [21]. As discussed above, the condenser saturation pressure varies to accommodate the change in the ambient air temperature. In the present study, this is done by maintaining a constant temperature difference between the condenser saturation temperature and cooling water temperature exiting the condenser. This temperature difference is also known as terminal temperature difference (TTD), and its value at baseline case is 11.7 °C. As a result, maintaining the same TTD will yield a higher condenser saturation pressure and correspondingly higher LP turbine back pressure as the ambient temperature increases. The warm cooling water that leaves the condenser is sent to the cooling tower, where it falls and cools down while the cooler air travels upwards. Figure 4 shows a typical wetcooled tower which can be broadly divided into a spray zone, a fill zone and a rain zone. The cooling water returning from the condenser is sprayed through the spray nozzles. In the spray zone, up to 15% of heat transfer occurs [22]. The sprayed water passes through the fill zone where it interacts with the cooler air coming from the bottom. Here, most of the heat and mass transfer occurs. The rain zone constitutes of the region where nearly 10-20% of heat and

11

mass transfer occurs [22]. It is this zone that first comes in contact with the cooler upwards moving air. To model the performance of the this evaporatively-cooled tower, the widely adopted theory from Merkel is used and the modeling approach can be found elsewhere [22]. Equation (4) gives the functional relationship between the dimensionless Merkel number (Me) and the temperature distribution across the different zones in the cooling tower:

Me 

Twi cpw dTw Cpwm (Twi  Two )  1 hd A 1 1 1         Two (i  i1 i 2 i3 i 4  mw 4 masw  ima )  

(4)

With this equation, the temperature of the returning cooled water as well as the temperature of the air-vapor mixture leaving the cooling tower could be determined. The spray zone, fill zone and rain zone each can be modeled by applying their respective Merkel’s correlation. Equation (5) gives the Merkel number specific to fill zone used in this study.

Mefi  0.605LfiGw0.35Ga0.35

(5)

The type of the fill selected for the present study is Ecodyne Shape 10. The heat and mass transfer characteristics and the pressure drop characteristics in the fill zone depends on the type and material of the fill installed. Similarly, the Merkel number specific to the rain zone used in this study is given by Equation(6) which is based on the data and study of Lowe and Christie [22].

G  Mesp  0.2 Lsp  a   Gw 

0.5

(6)

In this study, it is also assumed that the air leaving the wet cooled tower is saturated with water [22]. Therefore, the relative humidity at the outlet of the cooling tower is taken 1. By using the Merkel’s modeling approach, the temperature of the cooling water returning to the condenser as well as the temperature of the air-vapor mixture leaving the tower could be easily determined at varying ambient conditions. This result can then be used to calculate the effects of changing cooling water temperature on the performance of NGCC. For example, an 12

increase in the ambient temperature from 15°C to 40°C at a relative humidity of 0.6 creates a corresponding change in entering/leaving cooling water temperature from 16/27 °C to 33.2/43.7 °C and a concomitant change in the condenser saturation pressure from 6.89 kPa to 16.03 kPa. The corresponding decrease in the LP steam turbine output due to this change in turbine back pressure is nearly 20.7 MW. It can be concluded that this increase in ambient temperature has a huge impact (nearly 20% power reduction on the LP steam turbine output and 2% reduction in the Rankine cycle efficiency). In the following section, different waste heat recovery scenarios are discussed. These options could be utilized in the turbine inlet chilling which can potentially increase the power output as well as efficiency of both Rankine and Gas turbine cycles.

3

WASTE HEAT RECOVERY SCENARIOS Table 3 shows the list of all compressor inlet air chilling options that have been analyzed

in the present study, and Figure 5 shows the schematic of the for these options. The three WHR thermally activated cooling systems are driven by different sources: high temperature gas turbine exhaust (~628°C), low pressure steam (~5 bar, 333°C) and low temperature flue gas (~106°C). For the first option, the high temperature gases from the turbine exhaust are extracted directly and utilized in the desorber of the absorption cooling system. The outgoing flue gases are then sent to the HRSG in the Rankine cycle. For the second option, part of the LP steam entering the LP steam turbine is routed to run the cooling system, which, after being utilized, is then mixed with the major portion of steam in the LP condenser. For the third option, the low temperature flue gases exiting the HRSG is utilized to run the cooling system. For this third approach, the outlet flue gas temperature is limited to 80°C to avoid the potential deposition of acid condensates in the stack [23, 24], which limits the amount of waste heat that can be extracted from the power plant. The fourth option is to use turbine shaft power to drive a vapor compression chiller. The COPs of a typical electrically driven 13

vapor compression chiller with a cooling capacity of greater than 50% of the full load can be in the range of 3 to 6 [25], which is dependent on many factors, including the ambient temperature. However, in the present study, the COP of the vapor compression system is kept constant to 5 for simplicity. Future investigations will include a more detailed variation of system COP as a function of ambient temperature. A primary challenge for comparing these 4 options is the different cooling capacities that each system can deliver. Chilled water outlet temperatures that most commercially available absorption and vapor compression chillers deliver are nominally 3°C to 4°C. Therefore, the present study assumes that the incoming air to the compressor inlet can be cooled to a minimum temperature of 7°C. Furthermore, because it may not be effective to extract steam or use gas turbine exhaust at lower ambient temperatures, the incoming air is cooled only if the ambient temperature is above ISO operating conditions (i.e., 15°C). During the simulations, the hourly cooling load for a given location can be calculated as follows: Qcool  mda (ha1  ha2 )  1hg1  2 hg2  (2  1 )hw 

(7)

State 1 represents the incoming ambient air (>15°C) and state 2 represents the cooled air entering the compressor. For the two representative cities, the maximum cooling load are estimated as 53.80 MWth and 34.06 MWth for Houston and Los Angeles, respectively, and the total number of hours that cooling possible for these locations are calculated as 6467 and 5954 hours respectively. In the present study, two types of comparisons are made between the four systems, with major assumptions given in Table 4. The first comparison is for a fixed maximum cooling load. Because the flue gas outlet is limited to an 80°C outlet temperature, and there is a significant heat and/or electrical energy available in the other three options, the systems are first compared with the cooling load possible from the flue gas powered system, which is the product of the minimum heat extracted from the flue gas and its COP. The minimum flue gas 14

waste heat is used because this allows the system to operate consistently throughout the year, with larger amounts of available waste heat yielding a flue gas exhaust temperature greater than 80°C throughout the entire year. From the power plant model, the smallest amount of waste heat that can be extracted from the flue gas at a given hour for Los Angeles and Houston is 22.646 MWth and 22.763 MWth, respectively. This is multiplied by the COP of the cooling system to yield the compressor inlet air chiller load throughout the year (Table 5). Typical COPs for single effect LiBr-H2O absorption chillers range from 0.7 to 0.9, whereas double effect absorption chillers have COPs as high as 1.2 to 1.5 [26]. However, doubleeffect systems require a higher source temperature and cost more than single effect systems. Triple effect systems have also been commercialized recently and have the highest COP compared to single and double effect systems, but again at a higher cost. Therefore, in this study the COPs are varied between 0.2 and 2.0, while system cost is varied between $10 and $10000 per kWth of cooling. Moreover, at lower COPs it will not be possible to cool the inlet air to the target 7°C minimum, and, in those cases, the inlet temperature is simply reduced to its lowest value. Also, if the available cooling capacity is larger than required to cool the inlet air to 7°C, then the capacity is lowered in that hour so that the 7°C compressor inlet is maintained. In addition, for Los Angeles, a system with COP greater than 1.6 is not considered in this study because the cooling system size at higher COPs exceeds the maximum cooling load for this location. The second comparison between the four system options is for a fixed investment cost. A key outcome from this study is the determination of the range of possible COPs and capital costs for absorption systems that can result in a lower LCOE than the baseline case and the competitive mechanical vapor compression system. If a power plant has a fixed amount of capital they can invest, the system that provides the lowest LCOE will be the best option. In the following section, a detailed economic calculation approach is discussed. 15

4

ECONOMIC MODELING Most of the previous techno-economic analyses for compressor inlet chilling focus

merely on the calculation of pay-back period for different cooling options, without estimating the overall cost of electricity. In this study, the effect of installing the four options for compressor inlet chilling on LCOE has been determined over the entire life of the power plant. The NETL Cost Estimation Methodology for Power Plant Performance [27] has been used to estimate the global economic parameters and procedures to calculate the LCOE for the representative NGCC power-plant. A detailed description and method to calculate the LCOE is presented here. The Cost of Electricity (COE) is the revenue received by the generator per net megawatt-hour during the first year of operation for the power plant, which assumes that the COE escalates thereafter at a nominal annual rate equal to the general inflation rate [27]. Equation (8) gives the functional relationship of COE based on the first year operation costs (OC), total overnight capital (TOC), capital charge factor (CCF), capacity factor (CF) and the electricity generated (MWh):

COE 

CCF  TOC  OCFIX  CF  OCVAR CF  MWH

(8)

The CCF is the rate of return required on the invested capital over the entire lifetime of the project. It can be calculated based on the global economic parameters presented in Table 6, and the process can be found elsewhere [27]. The LCOE is the revenue received by the generator that assumes a nominal annual inflation rate of 0 percent, and it includes the inflation effects on the cost of electricity over the entire course of power plant life. Equation (9) can be used to calculate the LCOE as the product of the COE and the Levelization Factor (LF):

16

LCOE = COE  LF

(9)

The corresponding LF value is a function of operating life, internal rate of return (IRR) and the inflation rate as follows:

A(1- K LPn ) LF = (D - N)

(10)

In Equation (10) the variables D, N and LPn are the discount rate (i.e., the IRR), nominal escalation rate (i.e., inflation), and the levelization period (30 years), respectively. The values K and A are calculated based on equations (11) and (12), respectively: 1+ N 1+ D

(11)

D(1+ D)LPn (1+ D)LPn - 1

(12)

K=

A=

In this study, the COE value is initially assumed and varied until a fixed IRR of 12% is achieved, which is consistent with power plant financial analyses [27]. For different WHR schemes, at this fixed IRR value, the system that produces the lowest LCOE is assumed to be the most economical choice. The capital, operation, and maintenance costs associated with the wet-cooled NGCC power plant are also taken from NETL Case 13 [15]. This cost is then escalated by using the general inflation rate of 3% to the base year 2017. The natural gas costs and other variable costs are estimated on the basis of plant hourly data and the capacity factor, which is assumed to be 100%. For the costs of the additional equipment, including the process heat exchanger for compressor inlet air cooling, chilled water pumps, and piping, it is assumed that the cost is 25% of the capital cost for the cooling unit [13]. The annual operation and maintenance costs (OPEX) for the absorption chiller is sum of the annual costs of both the electricity and water required by the four compressor inlet chilling options. The parasitic electric load for the absorption chiller is nominal compared to power plants parasitic load. Based on one such 17

installation, the parasitic load for the absorption chiller can be predicted as 37 kWe for 2535 tons of cooling provided [28]. However, in this study the electricity and water related OPEX for the absorption system is assumed to be 10% of the installed cooling system cost [13]. Similarly, the annual OPEX for the vapor compression chiller other than electricity requirements is estimated to be 2.5% of the initial installation costs [13], and the electricity required to run the vapor compression chiller is simply subtracted from NGCC power plant output.

5 5.1

RESULTS AND DISCUSSION System Performance Figure 6 shows the baseline performance for the NGCC power plant for Los Angeles

and Houston, respectively. The average highest temperature for Los Angeles is 20°C during August, while the average highest temperature for Houston is 28°C in July. The averaged power output for these locations during these hottest months (i.e., July, August, and September) are 534.3 MWe and 509 MWe, respectively, at a 100% capacity factor. Because of the elevated average temperature during these months, this corresponds to 3.73% and 8.29% less power output than the rated baseline power output of 555 MWe at 15°C. However, in January, the average temperatures in Los Angeles and Houston for the month of January are 13.7°C and 10.5°C, respectively, and the average power output for these two locations for the same month increases to 559.3 MWe and 572.5 MWe, respectively. Over the entire course of the year, the average NGCC output for these two locations are 546.79 MWe and 536.18 MWe, respectively, which is a corresponding 1.48% and 3.39% less power output than the rated output of 555 MWe. This reductions in average power is due solely to the average temperatures for the two locations being above 15°C. To investigate the impact of compressor inlet air chilling, the COP of the absorption system is varied between 0.2 and 2, while the size of the chiller is given in Table 5 for these 18

two locations. As noted above, the size of the chiller is selected based on the COP of the system and the minimum waste heat that can be extracted from the flue gases. In all simulations, the minimum cooled inlet air temperature for the compressor is set to 7°C. The corresponding average yearly power output for both locations with the three different waste heat driven absorption systems are presented in Figure 7. For Houston, installing an absorption chiller with a higher COP has a significant impact. For example, increasing the COP from 0.2 to 2.0 increases the average yearly power output by 25.06, 24.98 and 26.02 MWe for gas turbine exhaust (GTX), steam (STM) and flue gas (FG), which results in an additional electricity generation of 219.5, 218.8 and 227.9 GWh per year, respectively. Furthermore, the WHR system powered by the FG generates the highest power output. This is because using GTX or STM at a given COP to run the cooling system negatively impacts on the power generated by the bottoming Rankine cycle. For example, the average yearly Rankine cycle output for a WHR cooling system with COP of 1.0 at location Houston are 192.4 MWe, 196.0 MWe and 200.1 MWe for GTX, STM and FG systems, respectively. However, it should be noted that the COP of all these WHR systems cannot be compared on a common basis. The gas turbine exhaust temperature can reach as high as nearly 637°C at 40°C ambient temperature (Figure 3), and the steam extraction pressure in the current study is taken as 5.1 bar. At these temperatures, the COP of the absorption systems can be much higher than at low temperature of the flue gases (106°C), because the COP of absorption chillers is highly dependent on the desorber temperature [26]. However, it is possible that these higher temperatures could yield additional cost penalties, and, hence, additional efforts should explore on the dependence of the WHR system COP (and cost) on the waste heat temperature. Compared to Houston, the power output for Los Angeles continues to increase until the COP reaches 1.2, after which the output increases minimally. Beyond a COP of 1.6 for 19

Los Angeles, the power output does not increase at all. The primary reason is that the maximum cooling load requirement is met for the moderate climate conditions in Los Angeles. When limiting the compressor inlet temperature to a minimum of 7°C, the yearly maximum cooling load at any particular hour in Los Angeles is 34.06 MWth, and a cooling system with a COP of 1.6 operating on flue gas generates 36.23 MWth of cooling. Also, by installing a flue gas driven absorption chiller with a COP of 1.2, a cooling load of 27.18 MWth is generated, which meets the cooling requirements for 96.7% of the total cooling hours (i.e., 5760 of 5954 hours above 15°C ambient). Therefore, by installing a system with COP greater than 1.2 generates minimal additional compressor inlet chilling, and has a minimal impact on average yearly power output. 5.2

Techno-economic Analysis The LCOE converts the system performance into economic value. Although the cost

of a thermally activated chilling system with the higher COP is most likely higher, it is able to enhance more power output, which can potentially reduce the LCOE. Figure 8 shows the relationship between the cooling system cost, COP, and LCOE for the three waste heat recovery options and the electrically driven chiller (VC) used to chill the inlet air of the compressor. In these figures, the installed costs for each system is expressed in terms of $ per kWth of cooling to estimate the LCOE. For comparison, the LCOE of the power plants in Los Angeles and Houston without compressor inlet chilling are $62.499 and $63.049 per MWh, respectively. At cooling system sizes lower than 9.10 MWth (COP = 0.4) for Houston (Figure 8a), the vapor compression chiller is cheaper than powering an absorption chiller with exhaust gases or steam if the cost of all these systems is nominally the same per kWth of cooling provided. Los Angeles has a similar trend (Figure 8b), except that the LCOE for each of these systems is slightly lower than at Houston because the average yearly output is higher. At 20

large cooling system sizes (Figures 8c and 8d), similar patterns emerge for the LCOE, with the GTX system yielding the highest LCOE at a given cooling system cost. In addition, at both cooling loads, an absorption system powered by flue gas yields a lower LCOE at a fixed cooling system cost, with a larger difference with the VC at the higher cooling load. For example, in Houston, at a system cooling cost of $200 per kW th, the FG and VC systems yield LCOEs of $62.55 and $62.62 per MWh, respectively at a maximum cooling load of 9.1 MWth, while these are $62.20 and $62.47 per MWh, respectively, when the maximum cooling load increases to 36.4 MWth. Moreover, the difference between the steam-power chiller and the electrically driven vapor compression chiller also decreases as the cooling load increases. For example, in Los Angeles, at a system cooling cost of $200 per kWth, the STM and VC systems yield LCOEs of $62.46 and $62.11 per MWh, respectively, at a maximum cooling load of 9.04 MWth, while these are $62.44 and $62.30 per MWh, respectively, when the maximum cooling load increases to 36.23 MWth. Clearly, the cost effectiveness of each of these systems are controlled by the size of the chiller and the cost of the system. Furthermore, the decision to select a particular compressor inlet air chilling system is dependent on the actual cost of the system. For example, in Houston at a cooling load of 9.10 MWth, the LCOE of a STM system at a cooling cost of $1000 per kWth ($63.42 per MWh) is lower than the LCOE of a FG system with a cooling cost of $2000 per kWth ($63.49 per MWh). A similar result can be seen for STM and VC systems at these costs as well. Therefore, to understand which system would yield a lower LCOE, a better understanding on the costs of each of these different systems is required. One possible method for understanding the impact of system cost is to determine the “tolerable cost” of implementing a particular system, which is the cooling system cost that yields an LCOE that is less than or equal to at the baseline case. Cooling system costs less than the tolerable cost yield a lower LCOE than the baseline, and, as a result, are more 21

likely to be implemented. Figure 9 shows the tolerable cost for the different cooling systems as a function of cooling capacity for Houston and Los Angeles, respectively. These figures show that the tolerable costs for the FG system is highest at all chiller sizes, followed by VC, STM and GTX systems respectively. At lower chiller sizes, this limiting cost for the GTX system is negative until the chiller size is nearly 18 MWth for Houston and 16.3 MWth for Los Angeles. Hence, installing a GTX system with a lower COP and cooling size would have negative impacts on the LCOE. However, as the chiller size increases, the tolerable costs for the GTX and the STM system increases to a maximum, after which the tolerable costs decrease. In all cases, the amount of heat extracted from the exhaust is the same, and, therefore, lower chiller sizes correspond to a lower COP. Because the GTX and STM systems reduce power output from the Rankine cycle, the corresponding low enhancement of the gas turbine at low chiller outputs is not enough to accommodate the reduction in power by diverting heat available for the bottoming steam cycle. In contrast, a different trend is observed for the FG and VC systems, and the tolerable cost is much higher for these two options at low chiller loads. The VC system has a tolerable cost less than the FG system because the VC system is extracting electrical power from the power plant. In both cases, and for higher chiller sizes, the total capital cost increases with a diminishing increase in performance, which explains the reduction in tolerable cost. With these simulations, a range of possible COPs and chiller system cost that yield a lower LCOE than the baseline can be determined. Furthermore, if the power plant plans to make an investment but only has a particular amount of capital available, it is possible to use these results to determine the best choice based on the available equipment COP and cost per kWth. This will allow different systems with various COP vs. cost tradeoffs to be evaluated on a common basis. For example, Figure 10 shows the LCOE for the waste heat and mechanically driven chiller systems for a fixed investment of $10M in Houston, and a fixed 22

amount of heat removed from the system equal to the maximum flue gas heat rejection (limited by an outlet temperature of 80°C). Clearly, it can be seen that with an investment of $10M, all the four systems yield at point where they would generate a LCOE less than the baseline case. For example, a FG system would be able to generate the LCOE less than at the baseline case if the COP of the system were higher than 0.36 and cost less than $1217.3 per kWth of cooling. The breakeven costs for the VC system is $939.9 per kWth of cooling for a fixed COP of 5. With the STM system, the breakeven LCOE is only possible at higher value of COP at 1.04, and the systems can tolerate less costs targets ($423.6 per kWth) compared to the VC and the FG system. Failure to achieve any of these targets would result in an LCOE that is higher than at the baseline case. The corresponding value for the GTX system is a COP of 2.06 and a cost of lower than $213.6 per kWth. This is a much more difficult target than the STM system, which is likely due to the significant reduction in power plant performance by diverting the high quality heat emanating from the gas turbine. Figure 10b shows the results for Los Angeles. These corresponding COP and $ per kWth values for location of Los Angeles are more stringent compared to the location of Houston for the same investment because the baseline LCOE is smaller for Los Angeles. None of the GTX or the STM systems would be feasible within the specified range of COP and cost targets. Also, for all the absorption systems, the LCOE ceases to decrease further with the increase in the cooling system COP/size beyond COP 1.2. One of the main reasons behind this trend is that the minimum temperature the inlet air can be cooled has been set to 7°C. As mentioned earlier, for a milder climate condition like that of Los Angeles, a cooling system with a COP of 1.2 would generate enough cooling to meet almost 97% of the total cooling hours load. The FG system needs to have a minimum COP of 0.4 and a maximum costs of $1115.7 per kWth to have the same LCOE as the baseline, while the maximum cost for a VC system is $906.5 per kWth. 23

Another effective way to compare the WHR systems with mechanically driven vapor compression system is to determine the minimum COP and maximum cost per kWth that yields an equivalent LCOE for a VC system for a fixed investment. VC systems are typically commercially available, and this method allows the different technologies to be compared on a common basis to determine what technoeconomic targets the WHR systems need to meet to be competitive on the market. In this example, the cost of the VC system is assumed to be $500 per kWth, which is consistent with the cost of a similar system ($ 426.4 for an 1.27 MW system) escalated to 2017 dollars [13]. At the assumed VC COP of 5 and a chiller size of 20 MWth, this cost yields a lower LCOE than the baseline case for both Houston and Los Angeles (i.e., $62.84 and $62.36 per MWh, respectively). The minimum COP and maximum cost requirements for the WHR systems to generate the same LCOE for the VC system are shown in Table 7 for the Houston and Los Angeles with different investment levels. For example, with the investment of $2M at a cost of $500 per kWth, the VC chiller size is a 4 MW, which is the maximum cooling available for this system. However, the capacity of the WHR system is allowed to change depending on the cost per kWth and the fixed investment cost. For example, if the maximum cost for a WHR system is $200 per kWth and the investment is $2M, then the maximum cooling load is 10 MW. (However, as with all the modeling in this paper, the maximum cooling load for the WHR system is still limited by the flue gas energy content when cooled to 80°C, and the cooling load is still limited by the required energy to cool the ambient air to 7°C.) As shown in this table, the minimum COP requirement for the FG system is lower than the STM, which is in turn lower than the GTX system, to generate the same LCOE. In addition, the FG can tolerate a significantly higher cost than the other two systems. For example, in Houston at a fixed investment cost of $5M, the minimum COPs required for the FG, STM, and GTX systems are 0.34, 1.0, and 1.97, respectively, while their maximum allowable costs are 24

$630.5, $218.4, and $111.3 per kWth, respectively. If any of these systems can meet these targets at the same investment, then it would be more economical to install them than the VC system. However, it is anticipated that the STM and GTX are unlikely to be less than the VC system cost of $500 per kWth, which operates at lower temperatures than either the STM or GTX systems, likely reducing the material costs for the VC system. In contrast, the FG system likely has the best chance to compete with the VC system, with the relatively low COP and high cost targets for the FG system. The difficulty for the STM and GTX systems to be competitive is likely because the energy normally utilized by the medium pressure steam and gas turbine exhaust is used to produce power, thus lowering the output of the NGCC power plant unless the compressor inlet chilling can overcome this effect. This is not the case for the FG system, which utilizes heat that is otherwise wasted. Similar results can be seen for Los Angeles, with even lower COPs required and higher cost thresholds for the WHR cooling systems. For example, at a fixed investment of $2M, the minimum COPs for the FG system are 0.19 and 0.16, while the maximum cost is $455.5 and $546.7 per kWth for Houston and Los Angeles, respectively. The reason that these targets are easier for Los Angeles is because the number of cooling hours are less and the ambient temperatures are less extreme than for Houston. As a result, a comparatively smaller chiller and a lower COP can be used to generate similar compressor inlet temperatures for most of the year. However, for the STM and GTX systems in Los Angeles, the cost and COP targets cannot be achieved, except for the STM system at a low investment of $2M. This is because a GTX or a STM system with even a higher COP cannot generate an LCOE value that is less than that for a VC system at $500 kWth. For example, if the investment is $5M in Los Angeles, then the LCOE for the GTX and STM system even with a high COP of 1.6 are $62.64 and $62.31 per MWh, respectively, which are both higher than the VC LCOE of $62.26 per MWh. At higher investments, the relative LCOE different is even higher for these 25

two WHR options compared to the VC system. Clearly, because of the low cooling requirements in Los Angeles, the relative power increase by using turbine exhaust or medium pressure steam to drive an absorption system that chills this compressor inlet air is not enough to overcome the additional capital cost and power reduction associated with diverting this heat from the power plant, and the VC system would be a much better option. However, if an absorption system can be installed with the COP and cost target while utilizing the normally wasted flue gas heat, then it can compete with the VC system. 6

CONCLUSIONS In this study, a detailed techno-economic assessment was conducted on a 555 MW

NGCC power plant subjected to different compressor inlet chilling scenarios. In contrast to prior investigations, these systems were compared on the basis of LCOE, which allows all of the costs – capital and operating – to be considered alongside the performance of various system combinations. The results show that compressor inlet air chilling is more ideally suited for locations where the temperature is high during the summer months (i.e., Houston vs. Los Angeles). Amongst the different WHR options, for a given cooling system cost, the absorption cooling system running on the low temperature flue gases produced the lowest LCOE compared to the systems running on the high temperature gas turbine exhaust or moderate pressure steam from the bottoming Rankine cycle. This is mostly likely because utilizing the wasted flue gas heat achieves a higher performance boost than diverting some of the higher quality heat in the NGCC cycle. However, at higher cooling system costs, the differences in the LCOE of different WHR systems tend to decrease for both locations, with the LCOE increase rapidly for all systems. These systems were compared with the baseline NGCC, and, at a fixed maximum compressor inlet cooling load, a flue gas driven system can tolerate the highest system cost to yield the same LCOE as the baseline, followed by the mechanical vapor compression, steam 26

driven, and gas turbine exhaust driven cooling systems. For the gas turbine driven system, the system always yields a lower LCOE than the baseline at low maximum cooling capacities. As the chiller size is increased, this system becomes more economical because the power reduction from diverting the gas turbine exhaust is offset by the power boost from as compressor inlet chilling. However, as the capacity increases further, the tolerable cost begins to reduce again because the capital cost of the system increases its significance. Similar trends are observed for the steam-driven system. If the owner of the power plant has a fixed amount available for investment, it is likely that mechanical vapor compression system used to chill the inlet air will be more economical than the two other heat diversion options of steam and gas turbine exhaust. The flue gas driven system has the potential to be more economical than the vapor compression system because it does not require a high COP, and it can tolerate a higher cost per kW th. These results appear to be more favorable for locations with lower cooling loads (i.e., Houston and Los Angeles), as the enhancement from the vapor compression system is reduced. Because the results presented here were for only two locations a single NGCC configuration. Further studies that investigate a wider range of weather patterns and different power plant system sizes and configurations are warranted.

ACKNOWLEDGEMENTS The authors acknowledge the Department of Energy Advanced Research Project Agency-Energy (ARPA-e) for their support under contract DE-AR0000574.

27

NOMENCLATURE A Cp D G h i, I L LF m Me p Pr PR R Re T UA

Heat transfer surface area (m2) Specific heat, constant pressure (kJ kg-1 K-1) Diameter (m),Discount rate Mass flux (kg/s-m2) Heat transfer coefficient (kW m-2 K-1) enthalpy (kj/kg) Inflation Rate Length (m) Levelization Factor Mass flow rate (kg/s) Merkel Number Pressure (kPa) Prandtl Number Pressure ratio Real Escalation Rate Reynolds Number Temperature (K) Overall Conductance (kW K-1)

Symbols γ Ratio of specific heat capacities p Polytropic efficiency Subscripts 01 stagnation property at inlet 0e stagnation property at exit a air DP Design point fi fill zone g Gas masw saturated air-water ma air sp spray zone ODP Off-design point o outside w water

28

REFERENCES [1] Mohan, G., S. Dahal, U. Kumar, A. Martin, and H. Kayal, Development of Natural Gas Fired Combined Cycle Plant for Tri-Generation of Power, Cooling and Clean Water Using Waste Heat Recovery: Techno-Economic Analysis. Energies, 2014. 7(10): p. 6358-6381. [2] Robb, D. CCGT: Breaking the 60 percent efficiency barrier. Power Engineering International, 2010. 18(3). [3] Cohen, H., H.I. Saravanamuttoo, and G. F. Rogers, Gas Turbine Theory. 4th ed. 1996, Harlow Essex CM20 2JE, England: Longman Group Limited. 442. [4] Elhadik, A.A., The Impact of Atmospheric Conditions on Gas-Turbine Performance. Journal of Engineering for Gas Turbines and Power-Transactions of the Asme, 1990. 112(4): p. 590-596. [5] Eveloy, V., P. Rodgers, and S. Popli, Power generation and cooling capacity enhancement of natural gas processing facilities in harsh environmental conditions through waste heat utilization. International Journal of Energy Research, 2014. 38(15): p. 1921-1936. [6] Gobran, M.H., Off-design preformance of solar Centaur-40 gas turbine engine using Simulink. Ain Shams Engineering Journal, 2013. 4(2): p. 285-298. [7] Al-Ibrahim, A.M., and A. Varnham, A review of inlet air-cooling technologies for enhancing the performance of combustion turbines in Saudi Arabia. Applied Thermal Engineering, 2010. 30(14-15): p. 1879-1888. [8] Kuehn, T. H., J.W. Ramsey, and J. L. Threlkeld, Thermal Environmental Engineering. 3rd ed. 1998, NJ, USA: Prentice-Hall. [9] Rahim, M.A., Performance and sensitivity analysis of a combined cycle gas turbine power plant by various inlet air-cooling systems. Proceedings of the Institution of Mechanical Engineers Part a-Journal of Power and Energy, 2012. 226(A7): p. 922-931. [10] Dawoud, B., Y.H. Zurigat, and J. Bortmany, Thermodynamic assessment of power requirements and impact of different gas-turbine inlet air cooling techniques at two different locations in Oman. Applied Thermal Engineering, 2005. 25(11-12): p. 1579-1598. [11] Boonnasa, S., P. Namprakai, and T. Muangnapoh, Performance improvement of the combined cycle power plant by intake air cooling using an absorption chiller. Energy, 2006. 31(12): p. 2036-2046. [12] Popli, S., P. Rodgers, and V. Eveloy, Gas turbine efficiency enhancement using waste heat powered absorption chillers in the oil and gas industry. Applied Thermal Engineering, 2013. 50(1): p. 918-931. [13] Barigozzi, G., A. Perdichizzi, C. Gritti, and I. Guaiatelli, Techno-economic analysis of gas turbine inlet air cooling for combined cycle power plant for different climatic conditions. Applied Thermal Engineering, 2015. 82: p. 57-67. [14] Paudel, A. and T.M. Bandhauer, Simulation of Natural Gas Combined Cycle Power Plants at Elevated Temperatures, in American Society of Thermal and Fluids Engineering 2016: Las Vegas. [15] U.S. Department of Energy, National Energy Technology Laboratory.Cost and Performance Baseline for Fossil Energy Plants. 2013. [16] Dixon, S.L., and C.A. Hall, Fluid mechanics and thermodynamics of turbomachinery. Seventh edition. ed. 2014, YBP Print DDA. xviii, 537 pages. [17] Kiameh, P., Power generation handbook : selection, applications, operation, and maintenance. McGraw-Hill handbooks. 2003, New York: McGraw-Hill. [18] Weber, P.T., Modeling Gas Turbine Engine Performance at Part-Load. 2011, Electric Power Research Institute. [19] El-Wakil, M.M., Powerplant technology. 1984, New York: McGraw-Hill. xv, 861 p. [20] Bergman, T. L., A. S. Lavine, F. P. Incropera, and D. P. DeWitt, Fundamentals of heat and mass transfer. 7th ed. 2011, New Delhi: Wiley. xxiii, 840 pages. [21] Bustamante, J.G., A.S. Rattner, and S. Garimella, Achieving near-water-cooled power plant performance with air-cooled condensers. Applied Thermal Engineering, 2016. 105: p. 362-371. [22] Kroger, D.G., Air-Cooled Heat Exchangers and Cooling Towers. I ed. Vol. II. 2004, Oklahoma: PennWell Corporation. 29

[23] Bahadori, A., Estimation of combustion flue gas acid dew point during heat recovery and efficiency gain. Applied Thermal Engineering, 2011. 31(8-9): p. 1457-1462. [24] Kren, C., Flue gas fired absorption chillers. 2008. p. XXII, 253 S. [25] Pereira, G. d. S., A. R. Primo, and A. A. O. Villa, Comparative study of air conditioning systems with vapor compression chillers using the concept of green buildings. Acta Scientarium Technology 2015. 37(4): p. 339-346. [26] Al-Tahaineh, H., M. Frihat, M. Al-Rashdan, Exergy Analysis of a Single-Effect Water-Lithium Bromide Absorption Chiller Powered by Waste Energy Source for Different Cooling Capacities. Energy and Power, 2013. 3(6): p. 106-118. [27] U.S. Department of Energy, National Energy Technology Laboratory., Cost Estimation Methodology for NETL Assessments of Power Plant Performance. 2011, The Energy Lab. [28] Maridat Ed, Braddock Cliff, and L. Chris, Performance Results and Lessons Learned from Austin Energy's Packaged COoling-Heating-Power System. 2005.

30

LIST OF FIGURES Figure 1. Detailed process flow diagram of the modeled NGCC including the steam cycle heat exchangers Figure 2. Temperature profile across the HRSG Heat Exchangers at 15°C ambient Figure 3. Temperature profile across the HRSG Heat Exchangers at 40°C ambient Figure 4. Wet Cooled Mechanical Tower Figure 5. Options for waste heat recovery and mechanical vapor compression for gas turbine inlet chilling systems in the present study Figure 6. Monthly average dry bulb temperature and power output a) Los Angeles-CA b) Houston-TX Figure 7. Output Vs. COP of the absorption system: Houston and Los Angeles Figure 8. LCOE comparison for different cooling systems at variable costs Figure 9. Tolerable costs for the different chilling systems a) Houston b) Los Angeles Figure 10. Investment based results for different cooling systems a) Houston b) Los Angeles

LIST OF TABLES Table 1 .

Baseline performance characteristics of the NGCC power plant at 15 ºC

Table 2.

List of major assumptions to predict elevated temperature performance

Table 3.

Compressor inlet chilling scenarios investigated in the present study

Table 4.

Summary of the cooling system assumptions and specifications

Table 5.

Chiller size (MWth) based on location and COP

Table 6.

Economic modeling parameters in the present study

Table 7.

COP and costs targets for WHR systems compared to Vapor Compression system costing 500 $ per kWth: Houston and Los Angeles

31

Figure 1. Detailed process flow diagram of the modeled NGCC including the steam cycle heat exchangers

32

Figure 2. Temperature profile across the HRSG Heat Exchangers at 15°C ambient temperature

33

Figure 3. Temperature profile across the HRSG Heat Exchangers at 40°C ambient temperature

34

Figure 4. Wet Cooled Mechanical Tower

35

Figure 5. Options for waste heat recovery and mechanical vapor compression for gas turbine inlet chilling systems in the present study

36

560

15

550

10

540

5

530

30

520 Month

Average Temperature Average Output

605 590

25

575

20

560

15

545

10

530

5

515

0

500

Output [MWe]

20

0

35

570

Temperature [ C]

Average Temperature Average Output

Output [MWe]

Temperature [ C]

25

Month

(a)

(b)

Figure 6. Monthly average dry bulb temperature and power output a) Los Angeles-CA b) Houston-TX

37

580

0

0.2

0.4

Power Output [MWe]

STM

0.6

FG

0.8

1

1.2

1.4

1.6

1.8

2

2.2

580

GTX

570

570

560

560

550

550 HOUSTON

540

LOS ANGELES

530

540

530

0.0 0.2 0.4 0.6 0.8 1.0 1.2 1.4 1.6 1.8 2.0 2.2 Absorption System COP

Figure 7. Output Vs. COP of the absorption system: Houston and Los Angeles

38

69.00

LCOE [$ MWh-1]

68.00 67.00

66.00

GTX (0.4)

STM(0.4)

FG(0.4)

VC(5.0)

LCOE [$ MWh-1]

69.00

Location: Houston Chiller Size: 9.10 MWth

65.00

64.00 63.00

68.00

GTX (0.4) STM(0.4)

67.00

FG(0.4)

66.00

Location: Los Angeles Chiller Size: 9.04 MWth

65.00 64.00

63.00 62.00

62.00

61.00

61.00 1

10 100 1,000 10,000 Cooling System Cost [$.kWth-1]

1

100,000

10

100

1,000

10,000

100,000

Cooling System Cost [$.kWth-1]

(a)

(b)

69.00

69.00

68.00

GTX (1.6)

STM (1.6)

FG (1.6)

VC(5.0)

68.00

67.00 66.00

LCOE [$ MWh-1]

LCOE [$ MWh-1]

VC(5.0)

Location: Houston Chiller Size: 36.42 MWth

65.00 64.00

63.00 62.00

GTX (1.6)

STM (1.6)

FG (1.6)

VC(5.0)

67.00 66.00

Location: Los Angeles Chiller Size: 36.23 MWth

65.00

64.00 63.00

62.00

61.00

61.00 1

10 100 1,000 10,000 Cooling System Cost [$.kWth-1]

100,000

1

(c)

10 100 1,000 10,000 Cooling System Cost [$.kWth-1]

(d)

Figure 8. LCOE comparison for different cooling systems at variable costs

39

100,000

Houston, TX

1500

FG

VC

GTX

Tolerable Cost [$. kWth-1]

Tolerable Cost [$ kWth -1]

STM

1000

500

0

-500

-1000

Los Angeles, CA

1500

GTX

9.11

18.21 27.32 36.42 Chiller Size [MW]

STM

FG

VC

1000

500

0

-500

-1000

45.53

9.05

(a)

18.12 27.18 Chiller Size [MW]

36.23

(b)

Figure 9. Tolerable costs for the different chilling systems a) Houston b) Los Angeles

40

(a)

(b)

Figure 10. Investment based results for different cooling systems a) Houston b) Los Angeles

41

Table 1. Baseline performance characteristics of the NGCC power plant at 15 ºC Parameter (Unit) Ambient Temperature (ºC ) Atmospheric Pressure(kPa) Net Total Electric Power (MW) Fuel Natural Gas, LHV/HHV(kJ/kg) Supply Condition (MPa/ºC) Gas Turbine Parameters Compressor Isentropic Efficiency (%) Turbine Isentropic Efficiency (%) Turbine Pressure Ratio Turbine Inlet/Outlet Temperature (ºC) Turbine Outlet Pressure (kPa) Steam Turbine Cycle High-Pressure Steam(MPa/ºC) Intermediate Pressure Steam (MPa/ºC) Reheat Pressure Steam (MPa/ºC) Low Pressure Steam (MPa/ºC) Approach Temperature (HP,IP,LP) ºC Steam Turbines and Pumps HP, IP, LP Turbine Isentropic Efficiency (%) Condenser Inlet Steam Quality Pump Isentropic Efficiency (%) Condenser Condenser Temperature/Pressure(C/kPa) Condenser Heat Duty (MWth)

Value 15 101.32 555 47454/52581 3.1 / 37.8 80.2 91.5 18.5 1371, 628.6 104.8 16.5/525 2.5/510 2.5/570 0.51/330 10,10,5 85, 93, 93.1 0.93 71 38.71/6.89 324.97

42

Table 2. List of major assumptions to predict off-design performance Component Compressor

Combustor

Turbine

HRSG

Steam Turbines Feed Water Pumps Condenser Cooling Tower

Major Assumptions Constant air volumetric flow rate Pressure ratio dependence on speed and intake mass flow Mechanical efficiency: 95 % Constant polytropic efficiency Combustor pressure loss: 5 % Lean and complete combustion Adiabatic combustion: Negligible heat loss Constant isentropic efficiency Constant firing temperature: 1371 ºC Constant discharge pressure: 104.8 kPa Mechanical efficiency: 95 % Generator electrical efficiency: 97.7% Negligible pressure loss on gas stream Negligible heat loss in heat exchangers Maximum heat exchangers effectiveness ≈ 85% Constant isentropic efficiency (85%,93%,93%: HP,IP,LP) Outlet steam quality LP turbine > 90% Negligible pressure loss in piping- feed lines Constant pump isentropic efficiency: 71% Constant cooling water flow rate: 7062.2 kg/s Constant degree of sub-cooling : 4.4 ºC Constant air volumetric flow rate

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Table 3. Compressor inlet chilling scenarios investigated in the present study Energy Source

System

Gas Turbine Discharge (High Grade) Low Pressure Steam

Absorption

Flue Exhaust Gases (Low Grade) Electricity

Absorption

Absorption

Vapor Compression

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Table 4. Summary of the cooling system assumptions and specifications Description Cooling System COP Cooling System Costs Minimum Cooled Air Temperature Cooling System Operating Temperature Minimum Flue Gas Temperature Maximum Cooling Load Minimum Available Waste Heat Absorption system parasitic load

Values 0.2 to 2.0 $10 to $1000 per kWth of cooling 7°C >15°C ambient 80°C Houston: 53.8 MWth, Los Angeles: 34.06 MWth Houston: 20.11 MWth, Los Angeles: 21.67 MWth 37 kWe per 2535 Tons of cooling

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Table 5. Chiller size (MWth) based on location and COP Location Los Angeles Houston

0.2 4.53 4.55

0.4 9.05 9.11

0.6 13.59 13.66

0.8 18.12 18.21

COP 1.0 1.2 22.65 27.18 22.76 27.32

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1.4 1.6 31.70 36.23 31.87 36.42

1.8

2.0

40.97

45.53

Table 6. Economic modeling parameters in the present study Parameter Income Tax Rate Capital Depreciation Repayment term of Debt Capital Expenditure Period Plant Operational Period Economic Analysis Period Capital Costs Escalation During Capital Expenditure Distribution of Capital over construction years Interest/Discount Rate Annual Inflation Rate Escalation of COE, O&M costs and Fuel costs Desired Internal Rate of Return (IRR) Finance Structure

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Value 38 % 20 years, 150% DB 15 years 3 years 30 years 33 years 3.6% 10%,60%,30% 5.5% 3% 3% 12% 45% Debt,55% Equity

Table 7. COP and costs targets for WHR systems compared to Vapor Compression system costing 500 $ per kWth: Houston and Los Angeles Houston STM

GTX Investment ($M) 2 5 10

Min. COP

Max. Cost ($/kWth)

1.38 1.97 3.46

63.7 111.3 126.9 GTX

2 5 10

-

-

Min. COP

FG

Max. Cost ($/kWth)

0.73 120.2 1.0 218.4 1.4 313.1 Los Angeles STM 0.55 161.5 -

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Min. COP

Max. Cost ($/kWth)

0.19 0.34 0.65

455.5 630.5 668.7

FG 0.16 546.7 0.32 684.8 0.56 747.4