FLEXIBLE COMBINED HEAT AND POWER SYSTEMS FOR OFFSHORE OIL AND GAS FACILITIES WITH CO 2 BOTTOMING CYCLES

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1 Proceedings of the ASME 2014 Power Conference POWER2014 July 28-31, 2014, Baltimore, Maryland, USA POWER FLEXIBLE COMBINED HEAT AND POWER SYSTEMS FOR OFFSHORE OIL AND GAS FACILITIES WITH CO 2 BOTTOMING CYCLES Marit J. Mazzetti Yves Ladam Harald T. Walnum Brede L. Hagen Geir Skaugen Petter Nekså ABSTRACT In this work different concepts are investigated for combined heat and power production (CHP) from offshore gas turbines. Implementation of such technology could improve energy efficiency of offshore oil and gas production and lead to reduced fuel consumption and resulting CO 2 emissions. Offshore electric power is in most cases generated by gas turbines operating in a simple cycle. However it would be desireable to increase energy efficiency by adding steam or CO 2 bottoming cycles to produce power from the exhaust heat. However part of the heat from the gas turbine exhaust is normally used for onboard process heat for the oil/water separation process among others, this must be taken into consideration when estimating capacity for additional power production. Different CHP concepts will be evaluated at different operating conditions while running the turbines in both design and off-design mode The results show that it is possible to produce an additional 6-8 MW of electrical power from a 32 MW turbine(depending on the conditions) while using 15 MW of heat from the exhaust for on-board processing. NOMENCLATURE P Pressure (Pa) R Volume flow ratio (-) T Temperature (K) h Specific enthalpy (J/ kg) s Specific entrophy (J/kg K) v Tip speed velocity ratio (-) η Efficiency (-) Subscripts DP in out INTRODUCTION Design point Inlet Outlet Improved energy efficiency is the only "fuel" that simultaneously meets economic, energy security and environmental objectives according to the 2013 IEA report "World Energy Outlook" [1]. This is also the case for oil and gas production where it is gaining importance. Particularly as offshore fields are ageing, the energy needed to produce a barrel of oil and gas increases significantly. Offshore facilities are normally designed for maximum production (or plateau ) rates [2]. In many cases, declining production results in increased power demand. One example is water injection in order to maintain reservoir pressure. This is a common energy intensive process which is often necessary as the platform goes into tail production. Improved energy efficiency will lead to reduced fuel consumption and resulting CO 2 emissions and help meet the world's climate goals, as well as improve offshore process economics by reducing fuel cost and CO 2 taxes where applicable. This is the case for Norway where the government has introduced an offshore CO 2 tax in order to accelerate implementation of CO 2 reduction measures. Offshore oil and gas platforms are in most cases generated by gas turbines operating in a simple cycle. However, on three offshore installations on the Norwegian continental shelf(ncs) a steam bottoming cycle has been installed that recovers the 1 Copyright 2014 by ASME

2 heat from the hot exhaust of the gas turbine increasing the efficiency of the electricity production on the platform [3]. Alternative cycles for more compact bottoming cycles saving weight and footprint on the platform have been discussed by Walnum et al. [4]. Those cycles are based on the use of CO 2 as working fluid. Lately, supercritical CO 2 cycles have received much attention, especially for nuclear applications [2, 3, 5]. The main reason is the potential for weight, size and cost reduction. These characteristics are also transferable to bottoming cycles for gas turbines [6]. For most oil and gas production platforms, part of the heat from the gas turbines is recovered for use in the energy intensive onboard oil/water separation process, among others. Depending on the plant layout, compressor driver concepts and power production strategy, a waste heat recovery unit (WHRU) may be needed on multiple gas turbines. The ratio of heat to power demand will typically vary and heat recovery systems are needed that can combine heat integration and power production (CHP) over a wide range of operating conditions. For steam systems this is typically performed with steam extraction or backpressure turbines where steam is extracted at an intermediate pressure and utilized for process heat. For CO 2 systems this is not an option as the CO 2 is typically expanded in a single stage turbine and far away from the two phase boundary. In this work several concepts for CHP will be evaluated at different operating conditions. The concepts are modelled using in-house tools, enabling the use of detailed component models taking real off-design effects into account. This is necessary due to the large load changes, enabling operation of the cycles with power production only and all the way to pure heat production. The work will extend the investigation presented previously [4], and present alternative system layouts to combine power production from a CO 2 bottoming cycle with heat generation. Southern climate: the cooling water temperature is higher (25C) and the CO 2 remains in the gas phase all along the cycle. The CO 2 is compressed (at point "g" in) before the waste heat recovery unit. This is called the Brayton cycle. ALTERNATIVE LAYOUT AND BOUNDARY CONDITIONS Two options are investigated for heat production integrated with a CO 2 bottoming cycle, Fig. 1. The first layout simply uses a secondary WHRU after the bottoming cycle WHRU. The mass flow of CO 2 is controlled to make sure that the necessary heat is available for the secondary WHRU. It will be referred to as the dual waste heat recovery unit (DWHRU). The second layout exploits the large amount of super-heat available at the CO 2 turbine outlet to produce process heat. It will be referred to as the internal heat recovery unit (IHRU). Two locations are considered: -Northern climate: for this location cold cooling water (10C is available) and the CO 2 bottoming cycle is able to condense at sub-critical temperature (31 C). This is the standard Rankine cycle layout. A pump (at point "g" in) is used to compress the liquid CO 2 before waste heat recovery. FIGURE 1 LAYOUT OF THE TWO CONCEPTS FOR COMBINED HEAT AND POWER (CHP) PRODUCTION. UPPER: DUAL WASTE HEAT RECOVERY UNIT (DWHRU). LOWER: THE INTERNAL HEAT RECOVERY UNIT (IHRU) 2 Copyright 2014 by ASME

3 MODELS AND METHODOLOGY The main purpose of this study is investigate how the two proposed layouts for a combined heat and power bottoming cycle manage varying load ratio between process heat production and power production. Realistic off-design evaluation is needed which imposes advanced geometry based component models (as opposed to performance based model which cannot provide off design information) Gas turbine model The gas turbine performance was calculated separately with GT MASTER from Thermoflow Inc. [7]. The chosen model is a GE LM2500+G4 with the dry low emission (DLE) setup. Compressor and turbine maps relating corrected inlet air mass flow to compressor pressure ratio and efficiency were utilized. The gas turbine is supposed to run at 100% load for all power/heat ratios for the bottoming cycles. TABLE 1: GAS TURBINE MODEL ASSUMPTIONS Ambient Northern Southern Temperature [ C] 9 30 Pressure [bar] Relative humidity [%] Sea water temperature [ C] Gas Turbine: GE LM2500+G4DLE Fuel Methane Methane P inlet [bar] GT gross power output [MW] Net efficiency [%] Exhaust mass flow [kg/s] T exhaust [ C] Heat exchanger models An in-house framework is used to model the heat exchangers. The models use geometrical input data to calculate parameters such as hydraulic diameters, perimeters and cross sectional areas for each fluid pass. Based on the geometry specification and the fluid inlet conditions, the outlet conditions are found through integration of the fluid passes (with a 4 th order Runge Kutta routine) and iteration on the wall temperature profile (with DNSQE from SLATEC [8]). Relevant heat transfer and pressure drop correlations are obtained from the literature, see TABLE 2. More details on the heat exchanger framework can be found in Skaugen et al. [9]. TABLE 2: HEAT TRANSFER AND PRESSURE DROP CORRELATIONS Correlation WHRU Fin side heat transfer [10] Fin side pressure drop [10] Tube side heat transfer [11] Tube side pressure drop [12] Condenser Single phase heat transfer [13] Condensing heat transfer [14] Single phase pressure drop [13] Condensing pressure drop [14] Recuperator and IHRU Single phase heat transfer [11] Single phase pressure drop [12] Waste heat recovery unit (WHRU) The WHRU is modelled as a cross-flow finned tube heat exchanger with serrated fins. The tube passes are arranged in horizontal serpentines inside a rectangular vertical gas duct, approaching a counter flow configuration. The WHRU is designed for a 3 kpa pressure drop on the exhaust side. Condenser The condenser is modelled as a plate heat exchanger. The high condensing pressure of CO 2 (55 60 bar) makes it unsuitable for standard plate-and-frame configurations. However plate-and-shell configurations could be an option. In this work, the performance characteristics of the heat exchanger at off-design conditions is of most interest, and this will be relatively independent of the type of heat exchanger. The condenser is assumed to be cooled with sea water. At the design point, the cooling water flow rate is set to give a 10 C increase in cooling water temperature, and this flow rate is kept constant also at off-design conditions. Recuperators and process heat generator (d in) for the IHRU The recuperator model is based on stacked layers of multiport tubes with counter-current flow, and is meant to represent a generic compact heat exchanger. Due to the high operating pressure (200 bar), diffusion bonded printed circuit heat exchanger might be the most relevant solution currently available. Pump and turbine for the bottoming cycle The pump is modelled with constant isentropic efficiencies defined as follows: η Pump = h(p out, s in ) h in h out h in (1) 3 Copyright 2014 by ASME

4 The efficiencies are set to 80 % through-out the load range studied. It is assumed that the pump is equipped with variable frequency drive (VFD). The turbines are calculated using the inlet guide vanes (VIGV) model [15]. An efficiency η DP of 85% was assumed at design. Off design performance was evaluated using manufacturer efficiency charts. The efficiency factor as function of volume flow ratio R (to design volume flow) is obtained from Atlas Copco [16]: η R = 0.34R R (2) The Efficiency factor as function of tip velocity ratio (to design tip velocity) is obtained from [17]: η v = 3. 11v v (3) The turbine efficiency was then calculated as product of the design point efficiency and the two efficiency coefficients and is used in the iterative cycle calculations: η = η DP η R η v (4) Calculation procedure The design of the bottoming cycle model was performed in Aspen HYSYS [18] to enable simple modifications and tuning of the processes. The design case is based on 10 MW heat production. The design point parameters shown in. TABLE 3 were used to define the HYSYS model. Then advanced geometry-based models were designed for individual components using the in house code to match the HYSYS model. The advanced model will allow for realistic off-design calculations. TABLE 3: BOTTOMING CYCLE PROCESS AND COMPONENT DESIGN POINT PARAMETERS Component Position Value WHRU (exhaust/co 2 ) UA [kw/k] b 400 WHRU (exhaust/ hot fluid) UA g 500 [kw/k] Recuperator UA [kw/k] a 700 Process heat generator (CO 2 /hot h 850 fluid) UA [kw/k] condenser d 3400 Max pump/compressor outlet f 200 pressure [bar] Pump/compressor efficiency [%] f 80 Expander efficiency [%] c 85 Motor/generator efficiency [%] 95 Stationary solutions for the whole bottoming cycle is solved using a sequential quadratic programming (SQP) method (NLPQL [19]). More details about the calculation procedure can be found in [4]. For off-design simulation, a control strategy must be chosen for the bottoming cycle. The condensation pressure could to some degree be controlled by the flow of cooling water, but in these simulations it was decided to keep the cooling water flow constant for the DWHRU system. For the northern conditions, where the system operates in trans-critical mode, the condensation pressure will be controlled by the heat rejected in the condenser. For the IHRU system the cooling water flow rate is variable in order to obtain necessary control flexibility. For the southern conditions, when the system is operating fully in supercritical mode, pressures are controlled by - mass repartition between the low- and high pressure side. The mass flow and pump outlet pressure is controlled by the turbine and pump operation. The VFD of the CO 2 pump/compressor enables a high efficiency in a wide range of flow rates and pressure ratios. The VIGV allows the turbine to operate with constant pressure ratios over a broad flow range [20]. The mass flow rate and high pressure are therefore considered free variables and are optimized during simulation. RESULTS AND DISCUSSION The produced power is shown as a function of produced heat for the different cases in Fig. 2. As expected, the power output is in general higher for the Northern case compared to the Southern. This is due to the lower cooling water temperature and the higher exhaust mass flow. For the dual WHRU system, the power output is not effected by the heat production up to 5 MW, since this heat is available anyway. If the heat demand is increased further, the power output drops relatively steadily. For the IHRU system, there is a significant power drop from 0 to 5 MW heat production. This is mainly due to the control strategy applied. The mass flow and pressure levels are controlled to heat the hot fluid to 170 C. For moderate process heat production, the mass flow of CO 2 is not optimized for power production. The mass flow is increased to lower the CO 2 temperature at the turbine inlet such that hot fluid is produced at 170 C. It would be beneficial to produce the hot fluid at a higher temperature so that the CO 2 mass flow is optimized for power production. The hot fluid could then be reduced to the desired temperature by mixing. For heat production between 15 and MW the IHRU system performs equally or better than the dual WHRU system. For the Northern case, the power from the IHRU system drops quite drastically for heat production above 20 MW. To make more heat available for the hot water, the work extracted from the expander must be reduced. This is done by reducing the pressure ratio. The pressure ratio was decreased by reducing the heat uptake pressure (as condensing pressure is controlled by heat exchange with the cooling water). The high pressure is an important parameter for cycle efficiency. For the Southern case, operating as a Brayton cycle, the power production reduction is less drastic. Here the low pressure is free to 4 Copyright 2014 by ASME

5 increase resulting in a reduced pressure ratio without decreasing the high pressure as much. FIGURE 5: EXPANDER EFFICIENCY VS. PRODUCED HEAT FIGURE 2: PRODUCED POWER VS. PRODUCED HEAT The temperature-enthalpy diagrams for the dual WHRU system with no heat production and 20 MW heat production (Northern) are shown in Fig. 6. The pinch point (minimum temperature difference) moves from the low temperature side to the high temperature side when the mass flow of CO 2 is reduced. FIGURE 3: CO2 MASS FLOW RATE VS. PRODUCED HEAT FIGURE 4: PRESSURE RATIO VS. PRODUCED HEAT The 10 MW heat production case is used as the design case. The resulting expander efficiency is shown in Fig. 5. The expander efficiency is relatively constant for all cases, except for the IHRU system at the highest heat production case. The low efficiencies experienced here indicate that the operating conditions are outside the range of the expander. FIGURE 6: TEMPERATURE-ENTHALPY DIAGRAM FOR DUAL WHRU SYSTEMS (NORTHERN). NO HEAT PRODUCTION (UPPER), AND 20 MW HEAT PRODUCTION (LOWER). PURPLE LINES SHOW THE CYCLE, RED THE GT EXHAUST (DOTTED 5 Copyright 2014 by ASME

6 REPRESENTS THE SECONDARY WRHU) AND BLUE THE COOLING WATER. The temperature-enthalpy diagrams for the IHRU system with 5 MW heat production and 20 MW heat production (Northern) are shown in Fig. 7. The corresponding diagram for no heat production is approximately equal to the one for the dual WHRU system shown in Fig 6. For the 5 MW case (left) the increase in CO 2 mass flow rate creates a very uneven temperature difference in the WHRU, with a very small temperature difference in the low temperature part and high temperature difference in the high temperature part. This indicates a non-ideal utilization of the heat transfer and large exergy losses. The pinch point (minimum temperature difference) moves from the low temperature side to the high temperature side in the WHRU when the mass flow of CO 2 is reduced. The opposite happens in the IHRU, as the water flow rate is increased. Heat and power production with variable gas turbine load So far the gas turbine load has been held constant at 100 % load. This is not a realistic operating condition for real oil platforms, where the turbine load varies according to the heat and power demand at the platform. The combined heat and power cycle was therefore analyzed at different gas turbine loads. To perform this study, off-design data for the gas turbine is required. The compressor of the LM2500+G4 GT has variable guide vanes in order to increase part load efficiency. This again affects the combined heat and cycle performance at part load conditions. In addition, the DLE setup has fuel staging, where for each stage the flame temperature will range between its maximum value for allowable NO x emissions and its minimum for allowable CO 2 emissions or flame blowout. Because of this, the turbine exhaust temperature will vary up or down with decreasing load, as shown in FIGURE 8 FIGURE 8 RELATIVE EXHAUST MASS FLOW AND TEMPERATURE VS GAS TURBINE LOAD FIGURE 7 TEMPERATURE-ENTHALPY DIAGRAM FOR IHRU SYSTEMS (NORTHERN). 5 MW PRODUCTION (UPPER), AND 20 MW HEAT PRODUCTION (LOWER). PURPLE LINES SHOW THE CYCLE, RED THE GT EXHAUST, GREEN THE HOT WATER AND BLUE THE COOLING WATER. In this study the heat demand was held constant at 15 MW while the gas turbine load was varied between 50 and 100% with 10% intervals. The resulting net power output and the CO 2 mass flow are shown in FIGURE 9. A large drop in power production is observed when the turbine load is reduced from 100 to 90%. For lower loads down to 60% the power production is almost constant around 5.5 MW. In this load region the higher exhaust temperature almost outweighs the reduced exhaust gas mass flow. The variation in the mass flow is quite similar to the variations in the net power output. This illustrates that the CO 2 mass flow is the main control parameter to ensure that enough heat is left in the exhaust gas to deliver the desired heat in the secondary WHRU. 6 Copyright 2014 by ASME

7 PETROMAKS. The authors acknowledge the partners Statoil, TOTAL E&P Norway, Shell Technology Norway, PETROBRAS, and the Research Council of Norway (203310/S60) for their support. REFERENCES FIGURE 9. POWER PRODUCTION AND CO 2 MASS FLOW FOR THE DWHRU SYSTEM (NORTHERN) VS GAS TURBINE LOAD. HEAT DEMAND IS SET CONSTANT AT 15 MW. CONCLUSIONS Compact CO 2 cycles could be an interesting alternative for additional power generation on platforms equipped with gas turbines [4]. On many installations, both power and process heat has to be provided from the fuel burned in the gas turbines. A bottoming cycle added to increase power production would have to be controlled such that the heat demand is also satisfied. The ratio of power to heat demand is expected to vary during operation which adds complexity to the operation of the bottoming cycle. Advanced models able to provide realistic off design calculations for two bottoming cycles have been implemented. The calculations have shown that both proposed CO 2 processes are able to produce both heat and power, both in Northern and Southern climates in a wide range of power to heat demand ratios. The dual waste heat recovery system has also been demonstrated to deliver heat and power at lower loads of the gas turbine, at least for Northern climates. Initial evaluations indicate that the expanders are able to operate in a large range of conditions and are able to handle large variations in the ratio of power to process heat demand. ACKNOWLEDGMENTS This publication forms a part of the EFFORT project, performed under the strategic Norwegian research program [1] International Energy Agency, 2013, "World Energy Outlook 2013." [2] Vanner, R., 2005, "Energy Use in Offshore Oil and Gas Production: Trends and Drivers from 1975 to 2025," Policy Studies Institute (PSI), London, United Kingdom. [3] Kloster, P., "Energy Optimization on Offshore Installations with Emphasis on Offshore Combined Cycle Plants," Proc. Offshore Europe Oil and Gas Exhibition and Conference. [4] Walnum, H. T., Nekså, P., Nord, L. O., and Andresen, T., 2013, "Modelling and simulation of CO2 (carbon dioxide) bottoming cycles for offshore oil and gas installations at design and off design conditions.," Energy 59, pp [5] Johnson, G. A., McDowell, M. W., O Connor, G. M., Sonwane, C. G., and Subbaraman, G., 2012, "Supercritical CO2 cycle development at Pratt & Whitney Rocketdyne," ASME Turbo Expo, ASME, Copenhagen, Denmark. [6] Kimball, K. J., and Clementoni, E. M., 2012, "Supercritical carbon dioxide brayton power cycle development overview," ASME Turbo Expo, ASME, Copenhagen, Denmark. [7] Thermoflow, 2011, "GT MASTER 21.0." [8] 1993, "SLATEC - Common Mathematical Library," Netlib Repository, U. T. Computer Science Dept, and O. R. N. Laboratory, eds. [9] Skaugen, G., Kolsaker, K., Walnum, H. T., and Wilhelmsen, Ø., 2013, "A Flexible and Robust Modelling Framework for Multi-Stream Heat Exchangers," Computers & Chemical Engineering, 49, pp [10] Næss, E., 2007, "An Experimental Study of Heat Transfer and Pressure Drop in Serrated-Fin Tube Bundles and Investigation of Particulate Fouling in Waste Heat Recovery Heat Exchangers," Dr.Ing, NTNU,. [11] Gnielinski, V., 1976, "New Equations for Heat and Mass Transfer in Turbulent Pipe and Channel Flow," 16(April), pp [12] Selander, W. N., 1978, "Explicit Formulas for the Computation of Friction Factors in Turbulent Pipe Flow," No. AECL-6354, Chalk River Nuclear Laboratories, Chalk River, Ontario CANADA. [13] Martin, H., 1996, "A theoretical approach to predict the performance of chevron-type plate heat exchangers," Chemical Engineering and Processing, 35(4), pp [14] Han, D.-H., Lee, K.-J., and Kim, Y.-H., 2003, "The Characteristics of Condensation in Brazed Plate Heat Exchangers with Different Chevron Angles," J. Korean Phys.Soc, 43(1), pp Copyright 2014 by ASME

8 [15] Marcuccilli, F., 2006, "Kalina & Organic Rankine Cycles: How to Choose the Best Expansion Turbine?," ENGINE Workshop 5: Electricity generation from Enhanced Geothermal Systems. Strasbourg (France). [16] Atlas Copco, 2012, "Driving Expander Technology," Atlas Copco Gas and Process Solutions(B03/004/24/0512). [17] GE, 2012, "Turboexpander-Generators. Company catalogue. energy.com/businesses/ge_oilandgas/en /literature /en/downloads/turbo_generators.pdf." [18] Aspen Technology Inc, 2010, "Aspen HYSYS v7.2." [19] Schittkowski, K., 1986, "NLPQL: A fortran subroutine solving constrained nonlinear programming problems," Annals of Operations Research, 5(1), pp [20] Meitner, P. L., and Glassman, A. J., 1980, "Off-Design Performance Loss Model for Radial Turbines With Pivoting, Variable-Area Stators," Cleveland, Ohio. 8 Copyright 2014 by ASME

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