Numerical simulation of heat transfer performance of an air-cooled steam condenser in a thermal power plant

Heat Mass Transfer (009) 5: DOI 0.007/s00-009-05-x ORIGINAL Numerical simulation of heat transfer performance of an air-cooled steam condenser in a thermal power plant Xiufeng Gao Æ Chengwei Zhang Æ Jinjia Wei Æ Bo Yu Received: 0 April 009 / Accepted: 8 July 009 / Published online: August 009 Ó Springer-Verlag 009 Abstract Numerical simulation of the thermal-flow characteristics and heat transfer performance is made of an air-cooled steam condenser (ACSC) in a thermal power plant by considering the effects of ambient wind speed and direction, air-cooled platform height, location of the main factory building and terrain condition. A simplified physical model of the ACSC combined with the measured data as input parameters is used in the simulation. The wind speed effects on the heat transfer performance and the corresponding steam turbine back pressure for different heights of the air-cooled platform are obtained. It is found that the turbine back pressure (absolute pressure) increases with the increase of wind speed and the decrease of platform height. This is because wind can not only reduce the flowrate in the axial fans, especially at the periphery of the air-cooled platform, due to cross-flow effects, but also cause an air temperature increase at the fan inlet due to hot air recirculation, resulting in the deterioration of the heat transfer performance. The hot air recirculation is found to be the dominant factor because the main factory building is situated on the windward side of the ACSC. List of symbols D h Hydraulic diameter h Platform height (m) X. Gao C. Zhang J. Wei (&) State Key Laboratory of Multiphase Flow in Power Engineering, Xi an Jiaotong University, 7009 Xi an, China e-mail: jjwei@mail.xjtu.edu.cn B. Yu Beijing Key Laboratory of Urban Oil and Gas Distribution Technology, China University of Petroleum (Beijing), Beijing, China k Thermal conductivity (W m/k) Nu Nusselt number p Pressure (Pa) Pr Prandtl number q Air mass flowrate (kg/s) Q Heat rejection rate (W) Re Reynolds number t Time (s) T Temperature ( C) U Wind speed (m/s) U min Air velocity at the minimum flow area (m/s) V Air velocity normal to the fan (m/s) x, y, z Coordinate directions (m) Greek symbols a Heat transfer coefficient [W/(m K)] b Wind direction angle (degree) k Ground roughness coefficient m Kinematic viscosity (m /s) q Density (kg/m ) Subscripts 0 Reference situation Introduction Although a direct air-cooled steam condenser (ACSC) has the shortcomings of the elevated turbine back pressure (absolute pressure) and thus the reduced cycle efficiency as compared to a water-cooled circuit, it is preferable to water-cooled ones in a thermal power plant for the areas rich in coal resources but poor in water []. There is an array of air-cooled units in the ACSC, each consisting of an A-frame configuration of finned tube heat exchanger bundles below which an axial flow fan is fixed. A stream of

Heat Mass Transfer (009) 5: ambient cooling air is forced to flow through the system and receives heat from the condensing steam in the finned tubes. Owing to the dynamic interaction between the steam turbines and the ACSC, a change in the heat rejection rate of the ACSC will directly influence the efficiency of the steam turbines. Therefore, it is very important to study the heat transfer performance of the ACSC. Many factors affect the running performance of the ACSC. Firstly, the environmental wind, especially strong wind, can generate flow distortions at the inlet of the axial fans to deteriorate the fan performance by reducing the cooling air flowrate and thus have an adverse influence on the heat rejection power of the ACSC [, ]. Secondly, the surroundings of the ACSC, such as the main factory building and terrain, also have influences [, 5]. If buildings are in the windward side of the fans, they will block air flowing to the inlet of fans. On the other hand, in the lee of the buildings a wake zone with low pressure is formed, and the hot buoyant outlet air plume from the ACSC is inhaled to this zone and then is drawn back into the ACSC inlet, resulting in an increase in the effective temperature of the cooling air with a corresponding reduction in heat rejection rate [6 8]. The heat transfer reduction caused by both the flow distortion in the fan inlet and the hot air recirculation leads to an increase of turbine back pressure, and occasional turbine trips occurs under extremely gusty conditions []. The heat transfer performance of ACSCs is closely related to the thermal-flow field about and through it. Some experimental and numerical studies have been conducted for investigating the thermal flow field [, 6 ], and it is found that computational fluid dynamics (CFD) is a very effective way to investigate the performance of ACSCs [9]. Rooyen and Kroger studied numerically the performance of ACSCs under windy conditions, and they found that the effect of flow distortion at the fan inlet was much larger than that of the hot air recirculation [9]. In the simulation, they did not consider the effects of the main factory building s location which is usually situated adjacent to the ACSC. Wang et al., however, found that the hot air recirculation has a very large impact on the heat rejection rate of the ACSC due to the existence of the main factory building [0]. They found that the factory building located at the windward side of the ACSC could generate large eddies with low pressure between the factory building and the ACSC, resulting in a strong hot air plume recirculation. In their simulation, however, the wind-induced flow distortion at the fan inlet was not considered, and the air mass flowrate at the inlet of the fan was assumed to be constant for simplification. The outlet air temperature from the ACSC was also assumed to be a prescribed uniform value. Since both the air mass flowrate and the outlet air temperature were the given input computational conditions for the simulation, this method can not be a good way to obtain accurate quantitative results. Therefore, it seems that a reasonable numerical model should give full consideration to both the surrounding effects on the hot air recirculation and the flow distortion at the fan inlet. In this work, a simplified physical model of the ACSC is developed. A heat exchanger model is used for simulating the flow and heat transfer in the ACSC, in which the heat exchanger is simplified as a porous media and all flow losses are taken into account by a viscous coefficient and an inertial loss coefficient. In addition, a fan model is used to get the flow condition at the heat exchanger inlet by giving the actual performance curves of the fan. The surroundings, including chimney, main factory building and terrain, are considered in the simulation to investigate their effects. At first, we look into the effects of wind direction to find the most unfavorable wind direction. Then, at the most unfavorable wind direction, the effect of wind speed on the ACSC performance for different heights of aircooled platform is investigated to determine an optimum height. Finally, at the designed height of platform, how the ambient wind speed and direction affect the total heat transfer capacity of the ACSC and the back pressure of the running turbine are investigated under the most unfavorable summer condition. At the same time, the contributions of all the factors that affect the ACSC performance are analyzed. The results of our study provide a reference for both the design and running of the ACSCs in thermal power plants. Numerical model and methods. Geometric model Direct air-cooled condensers from a 9 5 MW thermal power plant are used for the simulation. To reduce the length of steam pipes in a power plant, the platform of aircooled condensers is usually sited right behind the steam turbine room. The configurations of the proposed power plant, including the air-cooled condensers platform and buildings which comprises two 56 m high joint boiler rooms, a 8 m high steam turbine room and a 80 m high chimney, together with the definition of the incident angles of wind, are shown in Fig.. A steep hill with an elevation angle of 80 and a maximum height of 8 m in the west of the air-cooled platform is also shown in Fig. for the investigation of the terrain effect. The ACSC has aircooled units in total, and each consists of an A-frame configuration of steel finned elliptical tube bundles and an axial flow fan is mounted underneath. Exhaust turbinesteam flows inside the steel elliptical tubes, and the cooling air is drawn through the fan to take the heat from the

Heat Mass Transfer (009) 5: 5 Fig. Geometric model Chimney S E β W β Boiler rooms Turbine room N Air-cooled platform Wind wall Hill Wind direction angle Pillar exhaust turbine-steam which converts to condensate. The air-cooled units distributed in 8 columns 9 rows are mounted on a m high steel rectangular platform, which is supported by 9 6 cylindrical pillars made of reinforced concrete. To avoid the unfavorable effect of wind and increase the efficiency of condensers, a windbreak is constructed around the platform with an equivalent height of the condensers 0 m. Four different heights,, 5, 8 and m, of the air cooled platform (pillars) are studied for selecting an optimum designing height.. Simplified physical model The air thermal-flow around the direct air-cooled platform is assumed as incompressible and steady. The governing equations are Reynolds averaged Navier Stokes equations. The effect of the buoyancy force on the air is simulated via the Boussinesq variable density model. In a single air-cooled unit, the heat exchanger has many cooling fins on the tube, which will entail an unreasonably elevated number of elements and thus an excessive computational effort. In order to guarantee the computational efficiency without losing the correctness, the single aircooled unit in the present study is simplified to a. 9.0 9 0 m (length 9 width 9 height) cubic box fulfilled with porous media, under which an axial fan is arranged. The tube surface is set at the saturation temperature of the exhaust turbine-steam and the turbine back pressure is the saturation pressure. In our work, the HEAT EXCHANGER model in the FLUENT software is used to solve the heat transfer problem in the air-cooled unit, in which the heat exchanger core is treated as a fluid zone with momentum and heat transfer. Heat transfer is modeled as a heat source in the energy equation and the fluid zone representing the heat exchanger core is subdivided into macroscopic cells along the cooling air path. The steam is in a state of constant saturation temperature. The air inlet temperature to each macro can be computed and is then used subsequently to compute the heat rejection from each macro. This approach provides a realistic heat rejection distribution over the heat exchanger core, and the total heat rejection from the heat exchanger core is computed as the sum of the heat rejection from all the macros. Experimental heat transfer coefficient can be used for the computation of the total heat rejection in the heat exchanger core, which is correlated in a dimensionless form as Nu ¼ :8 Re 0:87 Pr 0: ðþ where Nu is the Nusselt number, Nu = ad h /k; Re is the Reynolds number, Re = U min D h /v. Here, a is the average heat transfer coefficient, D h the hydrodynamic diameter, k the thermal conductivity, U min the air velocity at the minimum area, m the kinematic viscosity. The porous media model is used in the present study to simulate flow resistance in the air-cooled unit. In the analysis, the pressure loss is added to the standard momentum equation of fluid flow as a momentum sink, contributing to the pressure gradient in the porous cell. The momentum sink term is composed of two parts, a viscous loss term defined by Darcy s law and a conventional inertial loss term. The constants in the two terms are determined by creating a pressure drop equal to the measured value in the air-cooled unit. The measured pressure drop across the heat exchanger is correlated in a dimensionless form as f ¼ 577: Re 0:6 ðþ where f is the drag coefficient, f ¼ Dp=0:5qUmin : Here, Dp is the pressure drop.

6 Heat Mass Transfer (009) 5: The main function of the axial flow fan is to boost the ambient cooling air to the heat exchanger. Therefore, the FAN model in the FLUENT software is selected in the present study because it can increase the pressure of the fluid flow across the fan to fulfil this purpose. Here, the fan is simplified as a infinitely thin layer for discontinuous pressure rise to overcome the flow resistance, thus the flow across the fan is considered to be one-dimensional without consideration of the swirl component. The pressure rise is specified as a function of the velocity through the fan, Dp = 7 -.6 V - 8.6 V Pa, which is determined from the actual performance curve of the fan. Here, V is the magnitude of the local fluid velocity normal to the fan. By employing the HEAT EXCHANGER model and FAN model in the FLUENT software and giving the experimental data of heat transfer coefficient, pressure drop and the fan performance curve, the simplified model can simulate the flow and heat transfer characteristics of a single air-cooled unit reasonably.. Computing grids In our work, Gambit is used to model geometry shape and generate computational grids. To ensure accuracy and save time, structured grids are used for the air-cooled units and solid surfaces, and unstructured grids are used for other computational zones. To study the heat transfer of the ACSC in an infinite space, a solution domain which is large enough to avoid the domain size effect should be selected without affecting the accuracy of the results. A cuboid form of the computational zone is adopted with the dimensions of,000 9,000 9,000 m, almost twenty times larger than the size of the ACSC in each direction of x, y and z coordinates, resulting in a volume almost 8,000 times the volume of the air-cooled platform. Numerical results show that the velocity and temperature distribution at the inlet of the computational zone are the same as those at the outlet, indicating that the selected computational zone is large enough to carry out the proposed simulation. The grids of the air-cooled platform, boiler and turbine rooms, the chimney and the hill are shown in Fig.. The total grid number is about,000,000. The results of this set of grids are considered to be the solutions independent of the grid number since this grid number yields a total heat rejection rate of the ACSC only 0.07% larger than the grid number of,5000,000. In Fig., x coordinate indicates the horizontal direction in which the turbine room points to the air-cooled platform; y-direction shows the horizontal direction perpendicular to the x-direction; and z-direction is the vertical direction from the ground to the sky. The x, y and z coordinates meet the right-hand rule. Fig. Grids of geometry model. Boundary conditions The boundary of the wind-inlet and wind-outlet surfaces is set as velocity inlet and pressure outlet. Ground and windbreak wall are set as wall. The chimney, the turbine room, the boiler room, the support pillars and the hill are set as solid zones. The HEAT EXCHANGER model combined with FAN model in the FLUENT software is used for the air-cooled unit. Equations () and () are used for the average heat transfer coefficients and pressure drop, respectively. The input parameters of the fan are chosen based on the actual performance curves of the fan supplied by the manufacturer. The wind speed distribution across the ground is used for the inlet velocity profile of the overall computational zone and is usually expressed as follows U U 0 ¼ z k ðþ 0 where U 0 is the standard wind speed at the height of 0 m; z is the height above the ground; U is wind speed at z, and k is the ground roughness coefficient ranging from 0.5 to 0.5..5 Simulated conditions To determine a reasonable platform height, four different platform heights,, 5, 8 and m, are simulated under the most unfavorable operating conditions of the ACSC in summer when the ambient air temperature is 8 C. The wind speed, U, ranges from 0 to m/s, and the wind direction angle, b, is in the range of 0 80. After a final platform height is determined, the effects of different ambient temperatures, wind speeds and wind directions on both the heat rejection rate of the ACSC and the steam turbine back pressure are investigated under the most unfavorable summer condition.

Heat Mass Transfer (009) 5: 7 Results and discussion. Effect of wind direction Figure a c shows the velocity and temperature fields of the air-cooled platform in x, y and z directions for the wind direction angle b = 0, 5 and 80, respectively. The height of the air-cooled platform (pillar) is m and the wind speed is.5 m/s. For b = 0, a wake zone is generated behind the boiler and turbine rooms, taking some hot air flowing out from the ACSC to the fan inlet and developing a high temperature region below the air-cooled platform, as shown in Fig. a; whereas for a reversed wind direction of b = 80, the natural wind blows back toward the boiler and turbine rooms, making the hot air plume turn to the negative x-direction. There is no wake zone generation and thus there is no obvious hot air recirculation phenomenon behind the boiler and turbine rooms, resulting in no large increase of cooling air temperature at the fan inlet, as shown in Fig. c. Therefore, the location of the main factory building has a large effect on the hot air recirculation for b = 0 and thus can not be neglected in the simulation of the ACSC performance. The comparison of Fig. a, b and c shows that the wind direction deviated from the x-direction leads to the excursions of the buoyant outlet hot air plume from the air-cooled platform in both x and y directions. The flow direction of hot air accords with the natural wind direction. Figure shows the effect of wind directions on the heat rejection rate of the ACSC. The total heat rejection rate increases with the increase of the wind direction angle, and drops to the lowest when b = 0 which accords with the flow and temperature distributions shown in Fig.. Therefore, b = 0 is the most unfavorable wind direction, and the main factory building should be arranged on the windward side of the ACSC with the statistical average wind speed and frequency on this direction being the smallest in recent years. Fig. Velocity and temperature fields. a b = 0, b b = 5, c b = 80 (a) (b) (c)

8 Heat Mass Transfer (009) 5: Q (MW) 500 50 00 h = m U =.5 m/s 9.6 MW 50 0 5 90 5 80 β ( o ) Fig. Effect of wind directions on the heat rejection rate. Effects of wind speed and platform height Since b = 0 is the most unfavorable wind direction, we investigate the effects of wind speed and platform height by fixing the wind direction at b = 0. Figure 5a and b show the thermal-flow map around the ACSC for the no wind case at the platform height h = and m, respectively. The ambient cooling air inhaled by the axial flow fans flows into the ACSC to take heat from the steam flow inside the finned tubes. The heated air flow forms a plume rising upward after leaving from the ACSC. We can see when h [ m, there is no obvious effect of h on the thermal-flow pattern. Figure 6a and b show the thermal-flow map around the ACSC for the natural wind case with U = m/s at the platform height h = and m, respectively. The hot air will be blown by the wind and flows aligned with the wind direction. At the same time, a low-pressured wake zone develops in the lee area of the turbine and boiler rooms because the natural wind blows straightly on the turbine and boiler rooms. Therefore, some hot air from the vertical plume is drawn into this low-pressure area again to generate hot air recirculation, which induce an increase of air temperature at the fan inlet below the ACSC []. The hot air mainly concentrates in the central zone near the turbine rooms and two corner zones in the lee areas. We can see that due to the hot air recirculation, the temperatureincreased area at the fan inlet becomes less as h increases from to m, therefore, a higher platform can improve the heat transfer performance of the ACSC. Comparison of the y-direction velocity and temperature profiles shows that the hot air flow leaving the ACSC turns into a little divergent plume for the case of no wind, whereas it becomes a convergent taper for the case of natural wind. Figure 7 shows the effect of wind speed on total heat rejection rate at four different platform heights of, 5, 8 and m. As the wind speed increases up to m/s, the total heat rejection rate decreases sharply, and then decreases slowly with the further increase of wind speed up to m/s, indicating that there is no remarkable influence on the total heat rejection rate for the wind speed ranges from to m/s. As the platform is lift up, the heat rejection rate increases due to the less effect of the hot air recirculation Fig. 5 Flow pattern and temperature distribution at U = 0 m/s. a h = m, b h = m (a) (b)

Heat Mass Transfer (009) 5: 9 Fig. 6 Flow pattern and temperature distribution at U = m/s. a h = m, b h = m (a) (b) Q(MW) 550 500 50 00 9.6 MW β = 0 o h = m h = 5 m h = 8 m h = m 50 0 5 U(m/s) Fig. 7 Effect of wind speed and platform height on total heat rejection rate P b (kpa) 50 0 0 β = 0 o h = m 5 kpa h = 5 m h = 8 m h = m 0 0 5 U(m/s) Fig. 8 Effect of wind speeds on steam turbine back pressure shown in Fig. 6(b). The heat dissipation can always reach above the design value of 9.6 MW for h = m. Figure 8 shows the effect of wind speed on steam turbine back pressure at four different platform heights of, 5, 8 and m. The variation trend corresponds to the heat rejection rate as shown in Fig. 7. The less the heat rejection rate, the higher the turbine back pressure is. The turbine back pressure (absolute pressure) increases as the wind speed increases and decreases as the platform height increases. Figure 9 shows the effect of natural wind speed on the cooling air mass flowrate inhaled by the fans under the aircooled platform at h =, 5 and m. The air mass flowrate q is normalized by that in the ideal no wind case, q 0. In ideal case of no wind, a fan is assumed to work independently without interference from the others. q/q 0 decreases almost linearly with the increase of wind speed, and the effect of platform height is not so significant. The comparison between Figs. 7 and 9 shows that the decrease of q is far less than that of the total heat rejection rate for a given wind speed, indicating that the decrease of air mass flowrate is not the dominant factor for the reduction of the total heat rejection rate, and implying that the main factor is probably the hot air recirculation. To present the effect of hot air circulation quantitatively, Fig. 0 displays the wind speed effect on the increment of the air temperature at the fan inlet area compared with the ambient temperature at h =, 5, 8 and m. It can be

0 Heat Mass Transfer (009) 5: q/q 0 0.98 0.96 0.9 0.9 β = 0 o h = m h = 5 m h = m 0.9 0 5 U (m/s) Fig. 9 Effect of wind speeds and platform height on cooling air mass flowrate T r (k) 0 8 6 β = 0 o h = m h = 5 m h = 8 m h = m 0 0 5 U(m/s) Fig. 0 Effect of wind speed and platform height on fan inlet air temperature rise hardly affect the thermal-flow field and thus the performance of the ACSC.. Performance of the ACSC and turbine back pressure under summer condition According to the above numerical simulation of the ACSC performance and steam turbine back pressure at four different platform heights, it is found that the higher the platform is, the better the ACSC performs. However, the construction cost will increase greatly with the increase of the platform height. Considering the performance and the cost, a compromised platform height of 8 m is finally determined to be the designed value. Based on this platform height, the effects of ambient wind speed and wind direction on the performance of the ACSC and the back pressure of the steam turbine are investigated under the most unfavourable summer condition. Figure shows the relationship between the total heat rejection Q and the natural wind direction angle b with the wind speed U as a parameter on summer working condition (TRL). Considering the x-axis symmetry of the air-cooled platform, the wind direction angle b ranges from 0 to 80. We can see that as b increases, Q increases rapidly at first, and then increases slowly for b [ 90. The x-direction component of the natural wind speed is always positive for b \ 90, leading to the formation of a low-pressure wake zone in the backside of the turbine room and thus bringing about the hot air recirculation, which reduces the total heat rejection rate of the air-cooled platforms. For b \ 90, the x-direction component of the natural wind decreases rapidly with the increase of b, reducing the wake zone and thus weakening the strength of the hot air recirculation, which increases the total heat rejection rate rapidly. For b [ 90, the x-direction component of wind is always seen that the inhaled air temperature increases as the natural wind speed increases, suggesting an increasing hot air recirculation. The inhaled air temperature decreases as the platform height increases, indicating that elevating platform height is an effective way to eliminate the effect of hot air recirculation. These conclusions are in agreement with the previous analyses and van Rooyen and Kroger s study [9]. In addition, the present numerical simulation investigates the effects of the steep slope to the west of the aircooled platforms and the 80 m high chimney to the east of the main factory building. The chimney is a very tall and slender building, and there is no obvious effect on the aircooled platform. The steep slope is much lower than the main factory building and there is a long distance between them. Therefore, the chimney and the steep slope can Q (MW) 500 50 00 U = m/s U = m/s U = m/s U = 5 m/s 9.6 MW TRL 50 0 5 90 5 80 β ( o ) Fig. Effects of wind speed and direction on total heat rejection rate

Heat Mass Transfer (009) 5: P b (kpa) 0 5 0 U = m/s U = m/s U = m/s U = 5 m/s TRL 5 Kpa 5 0 5 90 5 80 β ( o ) Fig. Relationship of turbine back pressure with wind speed and direction negative, and there is no obvious wake zone in the space between the turbine room and the air cooled platform, resulting in no remarkable hot air recirculation. It can also be seen from Fig. that the total heat rejection rate of the air-cooled platforms decreases with increasing wind speed. It should be noted that for the wind speed range studied here, the decrease of the total heat rejection rate is not so remarkable with the increase of wind speed. However, when the wind speed is greater than 0 m/s, the natural wind will severely reduce the air mass flowrate at the fan inlet, which will reduce the heat rejection rate significantly. Figure shows the relationship between the steam turbine back pressure P b and the natural wind direction angle b with the wind speed U as a parameter on summer working condition (TRL). Corresponding to the heat rejection rate shown in Fig., P b decreases rapidly at first as b rises for b \ 90, and then levels off. While P b can be always kept at the normal value of 5 kpa for U = m/s, it becomes higher than 5 kpa for b \ 0 when U increases up to 5 m/s. Therefore, it should be careful to adjust the turbine back pressure accordingly under these critical conditions to avoid turbine trips. Figure a and b show the distribution graph of the heat rejection rate of the ACSC in summer condition for U = 0 and m/s, respectively. The square column represents the heat rejection rate of each air-cooled unit. The higher the square column, the greater the heat rejection rate is. In the case of no wind, the heat rejection rate of the whole aircooled platforms is 6 MW, and the heat rejection rate of outer rings air-cooled units is lower than that of inner ones. In the case of natural wind, the heat rejection rate of windward air-cooled units is much lower than that of the others, resulting in a lower total heat rejection rate of 50 MW. Figure a and b show the air mass flowrate inhaled by the axial fans of different air-cooled units in TRL condition for U = 0 and m/s, respectively. Each column represents the air mass flowrate of an air-cooled unit. In the case of no wind, the total air mass flowrate is,5 kg/s, and the air flowrates of the fans in the outermost ring of the air-cooled units are lower than those of the inner fans. This is because the fans in the outermost ring inhale the static air from the outside surroundings, resulting in a heavier burden than those in the inner units. In the case of no wind, there is no obvious hot air circulation, so the reduction of the total heat rejection rate mostly results from the decrease of air mass flowrate inhaled by the fans in the outermost ring of the aircooled platform. For a fan working independently without interference from other fans, the inhaled air mass flowrate is 90 kg/s in the present study. Therefore, the total air mass flowrate will be,80 kg/s for fans. In the ideal case of no wind without obvious hot air recirculation, the heat rejection rate is mainly affected by the cooling air mass flowrate and is 6 9,80/,5 = MW. So the heat transfer capacity is about % lower than that in the ideal case at U = 0 m/s due to the reduction of air mass flowrate of the fans in the outermost ring of the air-cooled units. In the case of natural wind, the air mass flowrate of the fans in the windward air cooled units is much lower than that of the others, resulting in a much lower value of,756 kg/s and a corresponding heat transfer reduction of Fig. Heat rejection rate distribution of air-cooled units on summer condition. a No wind (U = 0 m/s), b U = m/s (a) q(0 7 W).75.50 (b) q(0 7 W).75.50.5.5.00 Column 5 6 7 8 0 5 Row.00 Column 5 6 7 8 0 Row

Heat Mass Transfer (009) 5: Fig. Air mass flowrate distribution of air-cooled units on summer condition. a No wind (U = 0 m/s), b U = m/s (a) 00 (b) 00 75 Q(kg/s) 50 Q(kg/s) 50 5 00 Column 5 6 7 8 0 Row 00 Column 5 6 7 8 0 Row 6-6 9,756/,5 = 6. MW. As shown in Fig., the total heat transfer reduction is 6-50 = 76 MW. Therefore, besides the reduction of the cooling air mass flowrate, the hot air recirculation also plays an important role in decreasing heat transfer performance, leading to a corresponding heat transfer reduction of 76-6. = 59.7 MW. It is obvious that the hot air recirculation is the main factor responsible for the reduction of heat rejection rate. For both no wind and natural wind cases, the influences on edge fans of the air-cooled platforms are greater than inner fans. If the power of outermost ring fans is increased and the other fans are kept to be constant, the cooling air mass flowrate will be increased effectively and better heat exchange efficiency will be obtained [5]. Conclusions Numerical simulation of the thermal-flow characteristics and the heat transfer performance is made of an ACSC in a 9 5 MW thermal power plant by considering the effects of ambient wind speed and direction, the air-cooled platform height, and the location of the main factory building and terrain condition. The main conclusions are as follows:. The performance of the ACSC and the corresponding steam turbine back pressure decreases with the increase of wind speed and increases as the platform height is elevated.. The ACSC performance increases and the corresponding steam turbine back pressure decreases rapidly with the increase of wind direction angle up to a critical value, and then both levels off. The critical wind direction angle is dependent on the platform height. The lower the platform, the larger the critical wind direction angle is. The relationship of the turbine back pressure with the wind speed and direction can be used for adjusting the running back pressure of the steam turbine to prevent turbine trips.. The impact of wind speed on the windward fringe fans of the air-cooled platform is larger than that on the other fans, resulting in a reduced inhaled air mass flowrate and thus a decreased heat transfer performance of the ACSC. This effect should not be neglected in the simulation.. The direct factors affecting the ACSC performance are hot air recirculation and the reduction of cooling air flowrate through the axial flow fan, and the former shows a dominant effect due to the existence of the main factory building on the windward side of the ACSC. The effect of main factory building should be considered in the simulation. Acknowledgments We gratefully acknowledge the financial support from the NSFC Fund (No. 0600, 50860, 50876). References. Adibfar A, Refan M (006) The air cooled condenser for dry countries. In: Proceedings of the ASME power conference, pp 0. Duvenhage K, Vermeulen JA, Meyer CJ, Kroger DG (996) Flow distortions at the fan inlet of forced-draught air-cooled heat exchangers. Appl Therm Eng 6(8/9):7 75. Duvenhage K, Kroger DG (996) The influence of wind on the performance of forced draught air-cooled heat exchanger. J Wind Eng Ind Aerodyn 6:59 77. Goldschagg HB (99) Lessons learned from the world s largest air cooled condenser, EPRI proceedings, international symposium on improved technology fro fossil power plants-new and retrofit applications, Washington, DC 5. Liu HP, Zhang BY, Sang JG, Cheng Andrew YS (00) A laboratory simulation of plume dispersion in stratified atmospheres over complex terrain. J Wind Eng Ind Aerodyn 89: 5 6. Krus HW, Haanstra JO, van der Ham R, Wichers Schreur B (00) Numerical simulations of wind measurements at Amsterdam Airport Schipol. J Wind Eng Ind Aerodyn 9:5 7. Zhao WS, Lei M, Wang SL, Cui N, Liu Y (008) Numerical simulation and analysis of the hot air recirculation phenomenon observed in direct air-cooled system. In: Proceedings of the rd IEEE conference on industrial electronics and applications, vol, pp 6 68

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