PERFORMANCE TESTING OF HYBRID WET-DRY COOLING SYSTEMS: INSIGHTS FROM NUMERICAL MODELING

Transcription

1 Proceedings of the ASME 2014 Power Conference POWER2014 July 28-31, 2014, Baltimore, Maryland, USA POWER PERFORMANCE TESTING OF HYBRID WET-DRY COOLING SYSTEMS: INSIGHTS FROM NUMERICAL MODELING R. W. Card CB&I Stone & Webster Inc. Boston, Massachusetts, USA ABSTRACT A hybrid wet-dry cooling system can reduce water consumption at a power plant while minimizing the performance penalty of an air-cooled condenser (ACC). Automatic allocation of turbine exhaust steam among the wet and dry sections provides robust performance. However, the performance test for the unit must be carefully designed to prove the guarantees for water conservation and thermal performance with minimal uncertainty. A hybrid wet-dry cooling system of the parallel type is modeled based on recently-constructed power plants. Effects of typical off-design test conditions are demonstrated. Techniques are recommended for designing an effective performance test for a hybrid wet-dry cooling system based on the use of existing Performance Test Codes (PTC). INTRODUCTION All thermal power plants release waste heat to the environment. The waste heat ranges from about 70% of the energy in the fuel for a traditional Rankine-cycle nuclear power plant to about 40% for a modern combined-cycle gas turbine plant. Most recent power plants release the waste heat by the surface condenser and cooling tower, as shown in Figure 1 [8]. The exhaust steam condenses on the outside surface of a bundle of tubes which are cooled by a continual flow of circulating water inside the tubes. The circulating water (CW) is cooled by evaporation and sensible heat transfer in the cooling tower, known as a wet tower. The lowest water temperature achievable is limited by the wet-bulb temperature of the air entering the tower, T WB. This system provides a reasonable exhaust pressure for the turbine, but consumes water by evaporation whenever the plant is operating. (Evaporation provides 80 to 90% of the heat transfer duty at design conditions.) Figure 1: Conventional Wet System Many recent power plants release the waste heat by the aircooled condenser (ACC). The exhaust steam condenses inside an array of finned tubes which are cooled by a continual flow of air outside the tubes. The heat transfer occurs by convection, with no evaporation, so there is no consumption of water at any time; it is dry. However, the lowest steam temperature achievable is limited by the dry-bulb temperature of the air entering the tower, T DB. The T DB always equals or exceeds the T WB, so the ACC system generally provides a somewhat higher exhaust pressure for the turbine. This penalizes the cycle efficiency, particularly on hot days with low humidity (high T DB, low T WB ). Also, the equipment for dry cooling is much more costly than the equipment for wet cooling. A hybrid cooling system combines wet & dry cooling. The typical goal is to reduce water consumption by using dry cooling for as great a portion of the duty as possible, while 1 Copyright 2014 by ASME

2 minimizing the penalties to cycle efficiency. The optimum dry duty fraction depends on the Owner s strategy and the relative costs of water consumed and power generated [1]. If water consumption is limited by permit, as is usually the case, a hybrid system allows the available water to be consumed in the most strategic manner. HYBRID COOLING SYSTEM This paper considers a hybrid cooling system of the direct or parallel design, as shown in Figure 2 [8]. The exhaust steam is divided between a steam surface condenser (SSC) and an air-cooled condenser (ACC). The SSC is usually located in the turbine building beside or below the turbine. The ACC is usually in the yard, so the condenser neck includes a branch to the ACC duct, which may be long. The SSC and the ACC are each sized for only a fraction of the exhaust steam flow, but they are otherwise conventional. The SSC is cooled by a wet cooling tower and circulating water pumps (CW), and the hotwell provides suction for conventional condensate pumps. The ACC should be located as close to the turbine building as possible, to minimize pressure drop in the ACC duct. The ACC may also include its own condensate collection tank, which may drain either by gravity or by forwarding pump to the hotwell in the SSC. Other arrangements are possible, and may be advantageous for the steam turbine drains, steam line drains and steam turbine bypass, but they do not change the basic performance of the hybrid system. In this example, the ACC is sized for 70% and the SSC for 30% of the cooling duty at the design condenser pressure, T DB and T WB. It is convenient to describe the allocation of duty as the dry duty fraction, DDF, which is defined in equation (1): DDF Q ACC (1) QSSC QACC The DDF of 70% is nominal; there is no steam valve or control system which assures it. The steam flow is balanced between the SSC and the ACC in proportion to their condensing capacity, as modified by the pressure drop in the condenser neck and ACC duct. Under some off-design conditions, the duty will reallocate without operator actions. Some operator actions may also influence the duty fraction. This allows the operators to trade-off water consumption and power generation when it is strategic [1]. At least nine direct hybrid systems have been installed (new or converted), mostly for combined-cycle plants with cooling duties 30 to 300 MW [13]. Afton combined-cycle, in Las Cruces NM, was originally conceived with a wet cooling tower but was reconfigured to include a direct hybrid system [10]. Figure 2: Hybrid Parallel Cooling System PERFORMANCE TESTING Hybrid cooling for power plants is relatively new. As for other new technologies, it is not yet clear what is required for a robust system that satisfies both the Owner and the contractors. There is no ASME performance test code for hybrid cooling systems. However, there are test codes for each of the components: the ACC [2], SSC [3], CW pump and wet cooling tower [4] & [5]. The test engineer for a hybrid cooling system may use some features from each of these codes to design a practical test with minimal uncertainty. The challenge for testing the hybrid cooling system is determining the overall duty, Q. The ST exhaust steam is wet, with a typical quality x = 85 to 90%. The ST exhaust enthalpy h exh cannot be determined from temperature and pressure which can be measured. However, testing an ACC presents the same challenge, and PTC 30.1 gives two reasonable techniques for computing the h exh from data recorded for the ST: the heat balance method, and the expansion line method [2]. The heat balance method is more certain, if all the required data can be collected. ST exhaust steam flow rate ṁ exh can be determined from the condensate flow rate (total of SSC and ACC), with the typical data for flowing or isolated drains and hotwell makeup. The condensate temperature must be measured [2] and its enthalpy h cond can be determined from the steam tables. The overall cooling duty Q can be estimated by equation (2): exh exh cond Q m h h (2) The duty in the wet section, Q SSC, is relatively easy to compute from the CW flow rate and temperatures T HOT and T COLD, perhaps as described in PTC 12.2 [3]. The duty of the dry section is the difference, as shown in equation (3): Q Q (3) ACC Q SSC 2 Copyright 2014 by ASME

3 The performance test should include the measurements and computations required to determine the Q, Q ACC and Q SSC at the test conditions with the usual rigor of the PTC. The dry duty fraction DDF may then be computed from equation (1). As is typical for test parameters, the test DDF may not be the same as the design DDF. The test condenser pressure of the SSC is adjusted to reference conditions by a formula [3]. For the ACC, the test pressure is adjusted by the manufacturer s curve [2]. Both adjustments require the actual steam flow to the device. But in the hybrid system, if the actual performance of the SSC or ACC falls short of the expected, some thermal duty and some steam flow will be diverted to the other device. The test DDF should be used to determine the correct steam flows for adjustment of the test condenser pressures. COMBINED-CYCLE POWER PLANT The behavior of the hybrid cooling system was examined by considering a typical combined-cycle power plant, recently designed and constructed. The original design included a conventional steam surface condenser and wet cooling tower. The heat balance models were prepared with the Thermoflex software [8], and included manufacturer s performance data for the gas turbines (GT), steam turbine (ST), condenser, cooling tower, and other major equipment. The key features are shown in Table 1. For this study, the maximum output case was used, with the duct firing assumed in operation in order to maximize ST generation and duty in the ST cooling system. This would be typical of a corrected net power test as described in PTC 46 [6]. Table 1: Example Power Plant Gas turbine combined cycle 2 x 1 F-class GT (GTCC) single reheat ST Mode Maximum output Net generation, nominal 645 MW Net efficiency, nominal 51.3% Net heat rate, nominal (LHV) 7013 kj/kwh (6647 Btu/kwh) GT gross generation, total MW Duct firing (LHV), total 1173 x 10 3 kj/hr (1112 x 10 3 Btu/hr) Heat-recovery steam generator Triple pressure (HRSG) Throttle pressure, nominal 132 bar abs (1900 psig) Throttle temperature, nominal C (1025 F) Reheat temperature, nominal C (1030 F) generation MW Elevation meters (670 feet) MODELING THE HYBRID COOLING SYSTEM This example of a hybrid cooling system was designed with the same condenser pressure, steam flow rate and ambient conditions as the original condenser and cooling tower. The original steam turbine model was tuned to match the vendor heat balance, so it represents a reasonable ST for these conditions, and it can be used in the example model of the plant with no change. The flowsheet for the heat balance model was modified to include the SSC, cooling tower, ACC, ACC duct and condensate forwarding pump [8]. The entering air for the cooling tower and the entering air for the ACC were assumed to have the same properties as the ambient air entering the gas turbine (T DB, T WB and pressure). The requirements of the cooling system are shown in Table 2. Table 2: Hybrid Cooling System ST exhaust steam flow t/hr (ṁ exh ) ( x 10 3 lb/hr) ST exhaust steam enthalpy kj/kg (h exh ) ( Btu/lb) ST exhaust steam quality 93.0% Dry Step Efficiency of 91.96% Last Blade Group Condenser pressure bar abs (P COND ) (1.052 psia, 2.14 inch Hg A) Condensate enthalpy kj/kg (h cond ) (71.19 Btu/lb) Overall cooling duty (Q) MW Air temperature, entering 14.4 C dry-bulb (58 F) Air temperature, entering 9.0 C wet-bulb (46.4 F, 40% RH) Air pressure 0.99 bar abs (14.35 psia) The exhaust steam flow was divided, 30% to the SSC and 70% to the ACC. The Thermodynamic Design features of Thermoflex were used to model the performance of the ACC, SSC, cooling tower and CW pumps for these duties [8]. The Engineering Design features were used to select hardware details of the various components. Then, the components were switched to Off-Design which fixes their configuration and allows the model to respond in a realistic manner. The resulting heat balance model does not represent any particular plant, actual or proposed, but will demonstrate the behavior of the hybrid cooling system for a large combined-cycle plant. The wet section includes the steam surface condenser (SSC), wet cooling tower, and CW pumps. The condenser pressure is the same as the ST exhaust pressure, because the typical test boundary is the ST exhaust flange. According to PTC 12.2, the pressure should be measured at a location from 0.3 to 0.91 m above the tube bundle (1 to 3 ft) [3]. However, it is typical to assume the pressure drop from the ST exhaust flange to that location is not significant. The key features are shown in Table 3 below. 3 Copyright 2014 by ASME

4 The dry section includes the air-cooled condenser (ACC) and the ACC duct. The condenser pressure is the same as the ST exhaust pressure, because the ACC duct is typically in the vendor s scope and PTC 30.1 requires a test pressure measurement as close as possible to the ST exhaust flange [2]. The key features are shown in Table 4. Table 3: SSC and Cooling SSC Steam flow rate t/h (539.3 x 10 3 lb/hr) SSC Condenser duty (Q SSC ) MW (526.6 x 10 6 Btu/hr) Configuration of condenser 2-pass CW flow rate in condenser t/h (2700 x 10 3 lb/hr) Velocity in condenser tubes 2.6 m/s (8.5 ft/s) Overall heat transfer coefficient, U 3075 W-m 2 - C Btu/hr-ft 2 - F Effective area, A m ft 2 HEI Cleanliness factor [9] 85% CW inlet temperature (T COLD ) 13.6 C (56.5 F) CW outlet temperature(t HOT ) 24.4 C (76.0 F) Auxiliary cooling flow rate t/h (4.0 x 10 6 lb/hr) Auxiliary cooling duty MW (44.2 x 10 6 Btu/hr) Cooling tower duty MW (574.2 x 10 6 Btu/hr) Cooling tower configuration Mechanical draft 8 cells, in-line Cooling tower plan 13.6 m x m 44.5 ft x ft Table 4: ACC and ACC Duct ACC Steam flow rate t/h ( x 10 3 lb/hr) ACC Condenser duty (Q ACC ) MW ( x 10 6 Btu/hr) ACC Configuration Mechanical draft 6 rows of 6 cells ACC plan 71.3 m x 71.3 m 234 ft x 234 ft ACC Overall heat transfer W/hr-m 2 - C coefficient, U Btu/hr-ft 2 - F ACC Effective area, A m ft 2 Pressure drop in ACC duct bar (0.045 psi) Velocity in ACC duct m/s (331.9 ft/sec) A power plant also has auxiliary cooling requirements, such as lube oil coolers and air compressors, which are often served by the CW system and add duty in the cooling tower. The auxiliary cooling duty was included in the CW system. EFFECT ON CONDENSER PRESSURE AND DDF The heat balance model was run with inputs for off-design conditions that might be expected during a performance test. The SSC duty and ACC duty were recorded and the DDF was calculated. The condenser pressure P COND was also recorded, as this is a typical objective in a performance test [2][3]. The following conditions were modeled; the data are tabulated in the Annex : 1. Entering dry-bulb temperature T DB. This has a significant effect on the DDF and P COND, as shown in Figure 3 below. 2. Entering wet-bulb temperature T WB, input as relative humidity at constant T DB. Higher humidity somewhat increases the DDF and P COND. 3. Pressure drop in the ACC duct. Higher pressure drop decreases the DDF, but increases the P COND. These effects are slight. 4. Number of operating fans in the ACC. The performance test would normally include all fans operating. However, in unfavorable wind conditions, the upwind row of fans (6 fans, in this example) may stall so there is no cooling air flow [12]. This reduces the capacity of the ACC but has no effect on the cooling tower and SSC. The DDF drops from 70% to 67% and the P COND rises. 5. Fouling in the ACC. This includes not just dust or debris on the exterior surface of the finned tubes, but also air blanketing or condensate flooding of the interior surfaces. Greater fouling in the ACC decreases the DDF and increases the P COND. 6. Number of operating fans in the cooling tower. The performance test would normally include all fans operating. If one or two fans are shut off for any reason, the CW continues to be cooled by natural convection in those cells and there is very little effect on the cooling system. However, if several fans are shut off, the DDF and P COND are increased. 7. Fouling in the SSC. This is evaluated with the HEI-style Cleanliness Factor [9], so it includes not just insulating deposits on the interior of the condenser tubes, but also obstruction of the CW flow through some of the tubes and air-blanketing on the shell side due to a poorly-performing air removal system. Fouling in the SSC increases the DDF and the P COND. 8. Flow rate of the circulating water. This may be more or less than its design value, as discussed below. Higher CW flow rate decreases the DDF and the P COND. 9. Degradation of the steam turbine. If the ST is less efficient as-tested than as-designed, the enthalpy of the ST exhaust steam is higher and the overall duty of the cooling system is higher than design. This has no effect on the DDF but increases the P COND, as shown in Figure 4 below. 4 Copyright 2014 by ASME

5 The most significant test condition is the dry-bulb temperature T DB, both because the effect is dramatic and because the test T DB is very likely to differ from the design T DB. PTC 30.1 allows the ACC to be tested for T DB ±10 C (± 18 F) from the design value [2], and PTC 23 allows the cooling tower to be tested for T DB ±14 C (± 25 F) [4]. Over this range of entering T DB, the DDF changes from 64% to 75%. This represents a significant change of duty to either the wet section or the dry section of the hybrid cooling system. The results are shown in Figure 3. Figure 3: Effect of Entering T DB tests of the wet cooling tower and the condenser. CW flow is an essential input for these tests. PTC 12.2 [3] and ATC-105 [5] both include instructions for determining the CW flow from a pitot traverse of the CW pipe. PTC 23 also includes detailed instructions for the tracer dilution method and use of ultrasonic flowmeters [4]. Accurate measurement of CW flow is important in testing the hybrid cooling system. CW flow is used to compute the Q SSC and the DDF with equations (3) and (1). The effect of off-design CW flow in the hybrid cooling system was analyzed by modeling several different flow rates. As expected, the condenser pressure improves with higher CW flow rate, and some duty shifts to the wet tower. However, the effect is small (See Table A8). EVAPORATION If there is a guarantee on evaporation, it is necessary both to establish it and to test for it. The evaporation rate can be estimated from the CW flow rate and temperatures by the usual Marley formula [7]: E G T HOT T COLD (4) (for E & G in gpm, T HOT and T COLD in F) Another significant test condition is degradation of the steam turbine, if any. The test and the analysis must compute the increased duty of the cooling system, as shown in Figure 4. Otherwise, the increased P COND may give the incorrect impression that the cooling system is not adequate. Figure 4: Effect of ST Degradation Evaporation rate can be computed more accurately by a heat balance model (such as Thermoflex [8], which displays water evaporated as an explicit output of the cooling tower icon). However, the best way to establish the guarantee is to request a preliminary curve from the prospective vendor, and if the evaporation rate meets the project requirements, use the curve as the basis. The evaporation rate during the test is significantly affected by T DB over the PTC-allowable range [2][4], as shown in Figure 5. When the T DB is high, the ACC releases a smaller fraction of duty by dry convection. The SSC and cooling tower take on a larger fraction of the duty, with concomitant consumption of water. Figure 5: Effect of Entering T DB on Evaporation CIRCULATING WATER FLOW The CW pumps are often specified with margin on the head or the flow. This may improve the thermal performance of the cooling system, but it also introduces uncertainty into the 5 Copyright 2014 by ASME

6 The evaporation rate can be tested by isolating the makeup and blowdown to the CW system and measuring the drop in water level in the cooling tower basin during the test run. However, the actual difference may be small so it may be necessary to take the measurement in a stilling well. (For example, at the design evaporation rate, 196 t/h, the drop in water level is about 13 cm/h). The submergence requirements for the CW pumps must be reviewed to prevent problems. Accurate measurement of the water level during the test will also allow the CW flow to be estimated from the CW pump curve, suction and discharge pressures. This is not as accurate as the pitot traverse, but it can confirm that the CW flow is the same during the test as on a previous day when the pitot traverse was conducted. OVERALL PLANT PERFORMANCE These variations in performance of the hybrid cooling system have little impact on the overall plant performance test. In the example cooling system, the greatest impact on ST generation was about 5 MW or 1.5%; that is, about 0.7% of the 645 MW overall net generation. However, the overall plant performance test generally includes a correction factor for T DB, such as 1 [6]. T DB affects plant performance by way of the gas turbine or boiler and also by way of the ACC or cooling tower. In addition to these effects, the hybrid cooling system responds to T DB by shifting the duty between the wet and dry sections changing the DDF. This moderates the impact of T DB on overall plant performance; that is, it reduces the required test correction factor. Therefore, the correction curve for T DB must be prepared with a heat balance model that allows the DDF to change in response to the performance of the ACC and SSC. If possible, the performance test for the cooling system should be simultaneous with the performance test for the overall plant or the ST. This will allow the quality of the exhaust steam to be calculated by heat balance [2]. The quality, together with the condenser pressure and condensate flow rate, will allow the overall duty of the cooling system to be calculated with minimum uncertainty. CONCLUSION A hybrid cooling system can be tested with as much rigor as either a cooling tower or an ACC. However, it is essential to compute the dry duty fraction DDF, which is not necessarily at its design value. The expected performance of the SSC or ACC depends on the actual steam flow to the device. But if the actual performance of the SSC or ACC falls short of the expected, the hybrid system will divert some duty to the other device. The test analysis must include this effect. The guarantee for the hybrid cooling system should be based on a specific overall duty for the ST exhaust steam, and this duty should be determined in the test. Thus, if by chance the ST is less efficient than expected, pushing the condenser pressure higher than its guarantee, the poor performance will not be incorrectly blamed on the cooling system. ACKNOWLEDGMENTS The heat balance model used in this study was prepared by Frank Senkel and Nick Zervos, also of CB&I, in the course of other work. Their detailed and complete product made the author s work much easier. NOMENCLATURE DDF Dry duty fraction, defined in equation (1) E gpm Evaporation rate of circulating water G gpm Circulating water flow rate h cond kj/kg [Btu/lb] Specific enthalpy of condensate in SSC or ACC h exh kj/kg [Btu/lb] Specific enthalpy of ST exhaust steam ṁ exh kg/hr [lb/hr] Mass flow rate of ST exhaust steam P COND bar a [psia] Condenser pressure Q MW [Btu/hr] Overall heat transfer duty in the hybrid cooling system Q ACC MW [Btu/hr] Heat transfer duty in the air-cooled condenser Q SSC MW [Btu/hr] Heat transfer duty in the steam surface condenser T DB C [ F] Dry-bulb temperature of air entering the cooling tower or ACC T WB C [ F] Wet-bulb temperature of air entering the cooling tower or ACC T COLD C [ F] Temperature of circulating water leaving the cooling tower or entering the surface condenser T HOT C [ F] Temperature of circulating water leaving the condenser or entering the cooling tower x % Quality of exhaust steam REFERENCES [1] Card, R.W. Economic Design of Hybrid Wet-Dry Cooling Systems, Power , ASME 2013 Power Conference [2] ASME PTC Air-Cooled Steam Condensers [3] ASME PTC Steam Surface Condensers [4] ASME PTC Atmospheric Water Cooling Equipment [5] Cooling Technology Institute, CTI Code ATC-105 (00), Acceptance Test Code for Water Cooling s [6] ASME PTC Overall Plant Performance [7] Marley Cooling Technologies, Cooling Fundamentals, 2nd edition, 1998, p 31 [8] Thermoflow, Inc, Thermoflex, Version 23.0, Revision of March 30, 2013 [9] HEI, Standards for Steam Surface Condensers, 10th edition, 2006 [10] Groves, J. et al, Afton Combined Cycle with Hybrid Cooling, Power Engineering, February Copyright 2014 by ASME

7 [11] Maultbesch, John S.; and DiFilippo, Michael N.; Hybrid Cooling System Overview, presented at ACC Users Group Meeting, Gillette, Wyoming, September 25 & 26, /11/07-Hybrid-Cooling-System-Overview.pdf accessed on 1/27/2014 [12] Maultbesch, J.S and DiFilippo, M.N. Effect of the Wind on the Performance of Air-Cooled Condenser, California Energy Commission Report CEC , June 2010 [13] GEA Heat Exchangers, PAC SYSTEM Installations, current 7/9/2013 ANNEX A DATA FROM HEAT BALANCE MODEL The heat balance model displays results in SI units. Data for the design condition are shown in italics. Table A1: Ambient/Entering Dry Bulb Temperature Tdb, C Twb, C DDF Pcond, bar Tdb, C gen, Table A2: Ambient/Entering Wet Bulb Temperature RH Twb, C DDF Pcond, bar RH gen, Table A3: ACC Duct Pressure Drop % loss DP, bar DDF Pcond, bar % loss gen, Table A4: ACC Fans ACC Fans DDF Pcond, bar ACC Fans gen, Table A5: ACC Fouling Fouling DDF Pcond, bar Fouling gen, Copyright 2014 by ASME

8 Table A6: Cooling Fans fans Tcold, C DDF Pcond, bar fans gen, Table A7: SSC Fouling Fouling DDF Pcond, bar Fouling gen, Table A8: CW Flow Rate Margin To Cond, t/h DDF Pcond, bar Margin gen, Tcold, C Thot, CTrise, C Table A9: ST Degradation Degrade Dry Step Eff, % DDF Pcond, bar Degrade ST Exh ST Exh Overall gen, Enth kj/kg quality Duty, Copyright 2014 by ASME