Effects of diffuser orifice plate on the performance of air-cooled steam condenser

Accepted Manuscript Title: Effects of diffuser orifice plate on the performance of air-cooled steam condenser Author: Xuelei Zhang, Tingting Wu PII: DOI: Reference:

S1359-4311(15)01399-X http://dx.doi.org/doi: 10.1016/j.applthermaleng.2015.12.018 ATE 7432

To appear in:

Applied Thermal Engineering

Received date: Accepted date:

18-6-2015 8-12-2015

Please cite this article as: Xuelei Zhang, Tingting Wu, Effects of diffuser orifice plate on the performance of air-cooled steam condenser, Applied Thermal Engineering (2015), http://dx.doi.org/doi: 10.1016/j.applthermaleng.2015.12.018. This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.

Effects of diffuser orifice plate on the performance of air-cooled steam condenser Xuelei Zhanga, * Tingting Wua a

School of Energy, Power and Mechanical Engineering, North China Electric Power University,

Baoding 071003, China

Highlights    

A diffuser orifice plate is considered to improve ACSC performance. The kinetic energy of ambient wind can be utilized by adding diffuser orifice plate. The mechanisms how diffuser orifice plate gets into reaction on ACSC are illustrated. A more uniform flow field distribution can form via diffuser orifice plate.

Abstract: Ambient wind has adverse impacts on the thermal-flow performances of ACSC. In this paper, a new measure to improve ACSC performance is proposed by installing diffuser orifice plate under the ACSC platform. As ambient wind flowing through diffuser hole in the plate, the velocity head converts to the pressure head, which contributes to shrink the negative pressure region under the ACSC platform and further improve ACSC performance. The numerical model of Column 4 section in a 600 MW ACSC is investigated, and the results show that under windy conditions, the diffuser orifice plate not only increases the volume flows across almost all fans due to the shrinkage of the negative pressure zone, but decreases fans inlet air temperatures, especially for upstream fans. At a wind speed of 9m/s, the volumetric effectiveness of windward Fan(1,4) is increased by 19.9%, and the temperature drop at windward Fan(1,4) inlet is 1.1K, as considering diffuser orifice plate. Moreover, a more uniform flow field distribution such as pressure, temperature as well as air flow rate can form via diffuser orifice plate, which may also improve heat transfer of condenser cell. At a wind speed of 9m/s, the heat transfer effectiveness is * 

Corresponding author: School of Energy, Power and Mechanical Engineering, North China Electric Power University, Baoding 071003, China. Tel.:+86 312 7522541; Fax: +86 312 7522440. E-mail address: [email protected] (Xuelei Zhang).

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increased by 6.6 percent points. The numerical results indicate that diffuser orifice plate is an effective way to restrain the adverse impacts of ambient wind, and it can be beneficial to operation improvement of the direct dry cooling systems in power plants. Key words: air-cooled steam condenser; ambient wind; diffuser orifice plate; numerical simulation. Nomenclature ACSC

air-cooled steam condenser

b

exponent of the wind speed in power-law equation

cpa

isobaric specific heat of air (J kg-1K-1)

C1ε, C2ε, C3ε

k-ε model constants

Ci

inertial loss coefficient

d1

inlet diameter of diffuser orifice (m)

d2

outlet diameter of diffuser orifice (m)

e

specific energy (J kg-1)

fn

the polynomial coefficient

g

gravitational acceleration (m s-2)

Gb

turbulence kinetic energy generation due to buoyancy (m2 s-2)

Gk

turbulence kinetic energy generation due to mean velocity gradients (m2 s-2)

h

specific enthalpy (J kg-1)

H

height of diffuser orifice plate (m)

k

turbulence kinetic energy (m2 s-2)

L

thickness of diffuser orifice (m)

lh

heat exchanger thickness (m)

ma,i4

air mass flow rate through Fan (i,4) (kg s-1)

NTU

number of transfer units

p

pressure (Pa)

∆p

pressure drop (Pa)

∆pf

fan pressure rise (Pa)

Qa,4

heat transfer rate of Column 4 (W)

Qa,i4

heat transfer rate of No.(i,4) condenser cell (W)

Si

momentum source (N m-3)

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s1

horizontal pitch (m)

s2

vertical pitch (m)

ta,i4

temperature at the inlet of Fan (i,4) (℃)

T

temperature (K)

v

air velocity (m s-1)

v10

wind speed at 10m height (m s-1)

vf

axial velocity across a fan surface (m s-1)

vw

wind speed (m s-1)

V

volumetric flow rate (m3 s-1)

z

height above the ground (m)

Greek symbols 1/αi

viscous loss coefficient

ρ

density (kg m-3)

μ

dynamic viscosity (kg m-1 s-1)

ε

turbulence kinetic energy dissipation rate (m2 s-3)

τ

stress tensor (J m-3)

λ

thermal conductivity (W m-1 K-1)

σ

turbulent Prandtl number

εi4

heat transfer effectiveness of No.(i,4) exchanger

ηht

heat transfer effectiveness

Subscripts a

air

ftb

finned tube bundles

id

ideal

eff

effective

t

turbulent

n

normal

v

vapor

1. Introduction Air-cooled steam condensers (ACSCs) are widely applied in power plants in the regions where water source is of shortage. In ACSCs, heat is transferred from the exhaust steam to a cooling air stream via finned tubes. Because ambient air replaces the water as the cooling medium, the

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thermo-flow performance of ACSC is then susceptible to influence from the ambient conditions, such as wind, temperature and atmospheric instabilities [1]. Windy weather occurs frequently in north China. Ambient wind will lead to rapid increase of the turbine back pressure that results in the reduction of cooled performance. Furthermore, the steam turbine trip could occur when turbine back pressure becomes higher than previously regarded to be the safe limit under intensive and unfavorable winds, especially in summer. Therefore, it is very important to estimate and improve ACSC performance under windy condition. The effects of ambient wind on the performance of ACSC have been the focus of extensive study in literatures. Duvenhage and Kröger [2] showed that wind blowing across a bank of ACSC reduces the flow rate of air delivered by the fans and increases the recirculation of hot exhaust air. It indicated that there exists a negative pressure region under the ACSC platform, especially at the inlet of windward fans, which decreases fan performance greatly [3, 4]. In addition, hot air recirculation also occurs under windy conditions. Owen and Krӧger [5], in their numerical study of an ACSC, showed that hot air recirculation increases with wind speed. Gu [6] found that hot air recirculation is very sensitive to the wind direction, wind speed and the height of the ACSC platform. Bustamante [7] indicated that wind has the potential to reduce ACC heat transfer rates by reducing fan performance or causing hot plume recirculation. Very high wind speeds can even lead to air backflow through the fans. Rooyen [8] and Owen [9] found that reduced fan performance is the primary contributor to the ACSC performance deterioration while recirculation of hot air only reduces performance by a small amount. To weaken the adverse impacts of ambient wind on the thermal-flow performances of ACSC, various measures had been proposed. He [10] studied the effect of varying the fan blade angle in

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the presence of ambient wind and suggested the use of regulable blade fans to improve ACSC performance. Subsequently, He [11] found that the promotions of rotating speeds in windward fans improve the performance of fan array and the heat transfer characteristic of ACSC. Yang [12] numerically investigated a representative 2×600 MW direct dry cooling power plant with three different configurations of the wind-break wall and found the extensions of the inner and outer walkways and elevation of the wind-break wall improve the thermal-flow performances of ACSC. Gao [13] and Yang [14] proposed a measure to remove the strong wind effect by adding deflecting plates under the air-cooled platform, which contributes to forming a uniform air mass flow rate in the axial fans by leading enough cooling air to the fans in the upwind region. Similarly, Zhang [15] proposed the arc-shape flow guiding devices installing in A-frame condenser cell and compared thermo-flow characteristics in the A-frame condenser cell and through the finned tube bundles. Owen [9] investigated the effect of porous wind screen installed in a cross-type arrangement below the fan platform, and the simulation results showed that the wind screen configuration can improve ACSC performance. Gu [16] numerically simulated the effect of line-screen windbreak structure on the ACSC performance in a 2×350 MW power station, and the optimal structure size was achieved in the further analysis. As mentioned above, attempts at improving ACSC performance under windy conditions should focus on improving fan performance and avoiding hot air recirculation. In this paper, a new measure to improve ACSC performance is proposed by adding diffuser orifice plate under the ACSC platform. Different from other measures, the diffuser orifice plate distributes a lot of diffuser hole, and the velocity head of ambient wind can convert to the pressure head as ambient wind flows through diffuser hole, which may improve fans performance. Consequently, the

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adverse effects of ambient wind can be restrained. The impacts of diffuser orifice plate upon the thermal-flow characteristics of Column 4 section in a 600 MW ACSC are numerically investigated. Numerical simulation is not only useful to predict the impact of diffuser orifice plate and help the design of diffuser orifice plate, but saves time and cost. The mechanisms how the diffuser orifice plate gets into reaction on the ACSC performance are illustrated. It can be beneficial to operation improvement of the direct dry cooling systems in power plants. 2. Methodology Ambient wind has adverse impacts on the thermal-flow performances of ACSC. It had been found that the negative pressure region under the ACSC platform results in volumetric effectiveness reduction of the windward fans under windy conditions [17]. Moreover, the negative pressure region plays a leading role in the heat transfer effectiveness reduction. To shrink the negative pressure region under the ACSC platform and sequentially alleviate the impacts of ambient wind, a diffuser orifice plate is installed under the ACSC platform in one side, as shown in Fig.1. The diffuser orifice plate distributes a lot of diffuser orifice, as shown in Fig.2. As ambient wind flowing through diffuser hole in the plate, the wind velocity decreases gradually with the orifice cross sectional area increases, and the velocity head of ambient wind converts to the pressure head according to Bernoulli equation. This velocity change results in a pressure rise at the orifices outlet, which may shrink the negative pressure region under the ACSC platform and sequentially improve the ACSC performance. The dimensions of diffuser orifice plate are shown in Table 1. The plate thickness should be thin as considering investment and installation, therefore a thickness of 1.0m is selected. The vertical diffuser orifice plate locates below ACSC platform and outside the steel structure as considering

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easy installation and fixation. The dimensions of diffuser orifice plate such as d2/d1 and height have considerable impact on the performance of ACSC. The dimensions of diffuser orifice plate will affect the expansion pressure ratio, flow loss as well as heat transfer rate of ACSC. For example, if the value of d2/d1 is too large, a greater expansion pressure ratio will be obtained, but it may result in flow loss increase. The optimal dimensions can be obtained as the ACSC heat transfer rate reaches a maximum. In this paper, the diffuser orifice plate with a height of 10m is considered, and the ratio of outlet diameter d2 to inlet diameter d1 is 1.5. 3. Numerical models 3.1 Physical model and computational grids The 600MW ACSC, shown in Fig.3, consists of eight long adjacent fan rows. The Column 4 section is representative of the fan units located the centre of the ACSC. The effects of diffuser orifice plate are investigated based on the numerical model of Column 4 section under windy conditions. The structure of condenser cell is shown in Fig.4. It mainly consists of an axial flow fan, 9.15m in diameter and a series of finned tube heat exchanger bundles. The cooling air is accelerated towards the A-frame plenum chamber by the fan, and then it is heated as flowing across the finned tube heat exchanger bundles. As a result, the heat rejection of the exhaust is carried away by cooling air. Obstacles or appurtenances such as supports, beams, inlet screens, and electrical fan drives are not individually modeled to avoid the additional complications and computational expense. The size of computational domain is 500m×160m×300m (x×y×z). The characteristic wind direction of +X is considered as schematically indicated in Fig.3. The computational meshes are generated with commercial software Gambit, using tetrahedral unstructured grid approach. For the

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central domain with Column 4 section, a finer mesh is chosen. For other zones, a coarse mesh is generated with size function. Grid independence test is achieved based on a very small deviation of the heat transfer rate from the three set grids of the numerical model. The final grid number is about 1,523,000. 3.2 Numerical method The phenomenon of fluid flow is dominated by the conservation laws including mass, momentum, energy and extra turbulence transportation. The governing equations for the mass, momentum and energy conservations are as follows [18]:  (  vi ) xi

 (  vi v j ) x j



    eff x j  

 v v j i    x xi j   xi

0

 2 vk  p   g i  S i (i, j, k=1,2,3, and i≠j)     eff  3 x k  xi  

( v i (  e  p )) 

   T  v j ( ij ) eff    eff xi  xi 

(1)

(2)

(3)

where vi is the velocity in the xi direction, T is the temperature, p the pressure, and gi is the gravitational acceleration in the xi direction. In this model, gi only exists in the -z direction. e  h  p /   v i2 / 2 is the specific energy, h is the specific enthalpy, μeff=μ+μt is the effective dynamic viscosity, μ and μt are the dynamic viscosity and turbulent dynamic viscosity respectively, τeff is the stress tensor, λeff is the effective thermal conductivity, Si is the momentum sink and equals to the pressure drop per flow passage length through the tube bundles, and ρ is air density. The effect of turbulence on the flow field is included in the application of the standard k-ε turbulent model. Robustness, economy, and reasonable accuracy for a wide range of turbulent flows explain the popularity in industrial flow and heat transfer simulations of the standard k-ε

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turbulent model [19].  xi  xi

(  kv i ) 

(  v i ) 

 t k    ) (     G k  G b    x j   k  x j 

 t    ) (    x j    x j

2      C 1 ( G k  C 3  G b )  C 2   k k 

(4)

(5)

where k and ε are the turbulence kinetic energy and its rate of dissipation, σk and σε are the turbulent Prandtl number for k and ε, Gk represents the generation of turbulence kinetic energy due to the mean velocity gradients, and Gb is the generation of turbulence kinetic energy due to the buoyancy. The model constants have the following default values, C1ε=1.44, C2ε=1.92, C3ε=1, σk =1.0, σε=1.3. 3.3 Setup and boundary conditions As air flows through the ACSC it experiences mechanical energy losses due to the presence of flow obstructions, such as screens, support beams, and the heat exchanger bundles. A porous media model in FLUENT is used to deal with the finned tube bundles. The pressure drops resulting from the screens, support beams, plenum chamber and outlet losses are taken into account in the porous media model. The pressure drop across the finned tube bundles is determined from the equation given by the manufacturer.  p ftb  16.1 v ftb  17.4 v ftb 2

(6)

where ∆pftb and vftb are pressure drop and air velocity across the finned tube bundles, respectively. For implementation of the boundary condition in the model, the pressure drop across the finned tube bundles is simulated by the momentum sink term, which consists of two parts, namely the viscous resistance and the inertial resistance term as Si  

 p ftb lh

 (

 i

vi  C i

1 2

 a | v | vi )

(7)

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where Ci and 1/αi are the inertial and viscous loss coefficients, respectively, the footnote i represents the individual Cartesian directions, ρa air density, and lh is heat exchanger thickness, 0.219m. By Equations (6) and (7), the inertial and viscous loss coefficients in the normal direction to the finned tube bundles Cn and 1/αn are 139.6 and 4,465,730 respectively. The inertial resistance coefficients in the other two orthogonal directions are specified to be 103 times higher than in the normal direction, to restrict the flow in these directions. Heat transfer is modeled to capture any buoyancy effects or hot air recirculation. Due to the excellent thermal conductivity of finned tube bundles, the temperature difference between exhaust steam and heat exchanger bundles is negligible. As a result, the porous zone is set the saturated temperature of steam condensing of 46.6℃. The fan is simplified as a pressure jump surface. In the Fan boundary, the pressure rise ∆pf is expressed as the polynomial form of the axial velocity vf across a fan surface. N

pf 



n -1

f nvf

(8)

n 1

where fn is the polynomial coefficient and the term number N is 3. By fitting the performance curve of the typical fan adopted in air cooled condensers, the polynomial coefficients are obtained. f1=144.8, f2=-5.86, f3=-0.81. The effect of buoyancy force on the air is modeled via the incompressible variable density model. The atmospheric wind velocity profile vw is appointed on the windward surface of the computational domain as v w  v10 (

z 10

)

b

(9)

where v10 is the wind speed at 10m height, usually measured by meteorological stations, and z is

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the height above ground. There are four categories for ground surface in China, and the air-cooled power plant is located in category C. Considering the roughness of the ground and atmospheric stability, the exponent b equals to 0.2 in this paper. Three wind speeds of 3m/s, 6m/s and 9m/s are assigned v10 to investigate the effects of wind speeds and diffuser orifice plate. On the windward surface of the computational domain, the velocity-inlet boundary condition is applied. On the downstream surface of the computational domain, the outflow boundary condition is appointed due to the unchanging flow velocity profile, namely the zero diffusion flux in the direction normal to the exit plane. On the other surfaces of the computational domain, symmetry boundaries are designated. The ground and wind break wall are assumed to be wall surfaces. The surfaces of the steam duct are given a constant temperature that is equal to the saturated temperature of the exhaust steam approximately. The THA condition is taken into consideration. The atmospheric temperature is 14℃, and the exhaust flow is 349.4kg/s. 3.4 Solution and data reduction The governing equations for the momentum, energy, turbulent kinetic energy and specific dissipation rate are discretized using the first-order upwind differencing scheme. The SIMPLE algorithm is employed for the pressure-velocity coupling. The solution is iterated until sufficient convergence is achieved. The criteria for convergence of the scaled residuals are set as 10-4. It is also important to confirm the solution does not change by monitoring variables on specified locations at different calculation steps. For this study, the values of volume flow across fans are monitored at each iteration. To access the ACSC performances under windy conditions, the effects on fan volumetric

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effectiveness and heat transfer effectiveness are investigated. For Column 4 section consisting of 8 cells, the fan flow rate of each cell is inevitably different from each other. Volumetric effectiveness, which is just the actual fan volume flow rate V divided by the ideal fan volume flow rate Vid is used to characterize the influence of the environmental wind on the fan performance. Here, the design volume flow rate of fan, 428 m3/s, is assigned to the ideal flow rate. The summation of individual condenser cell performances would consequently give the performance of the entire Column 4, as Q a ，4 

8

8

i 1

i 1

 Q a,i4 =  m a,i4 c pa  i4 ( t v  i4  1  e

 N T U i4

 t a,i4 )

(10)

(11)

where Qa,4 is the heat transfer rate of Column 4, Qa,i4 is the heat transfer rate of No.(i,4) condenser cell, ma,i4 is numerically predicted air mass flow rate through Fan (i,4), ta,i4 is temperature at the inlet to Fan (i,4), tv represents the steam turbine exhaust temperature, cpa is the isobaric specific heat of dry air, εi4 is the heat transfer effectiveness through the No.(i,4) finned tube heat exchanger, and NTUi4 is the number of transfer units. The overall heat transfer effectiveness ηht is defined as the ratio of heat transfer rate under the windy operating condition, calculated from the numerical results by Equation (10), to the heat transfer rate under the windless operating condition. w indy

 ht =

Qa

w indless

(12)

Qa

3.5 Validation of the numerical model The performances of condenser cells in Column 4 are numerically investigated under windless condition to validate the models. The average volume flow rate of the entire 8 fans in Column 4 is

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426.6 m3/s, and relative error is only 0.5% compared with the design value. Furthermore, the numerical result of average heat transfer rate of condenser cells in Column 4 is 14.1MW, and there is a less than 0.3% difference between the simulation results and the design values. The consistency of fan volume flow rate as well as the heat transfer rate shows that the simplifications associated with the air-cooled condensers and the computational methods are reliable enough for the purpose of this investigation. 4. Results and discussion Under the windy conditions, the ACSC performance will be affected both by fan inlet temperature increase and by fan performance degradation. In the paper, +X wind conditions are designated to investigate the effects of diffuser orifice plate on fan volumetric flow, average air temperature at fan inlet as well as heat transfer of the finned tube exchangers. 4.1 Fan volumetric effectiveness Fig.5 shows the volumetric effectiveness of Column 4 including 8 axial fans. It is found that the volume flows across fans, especially across upstream edge fan, are quite sensitive to the wind speed. Under windless condition, the fan volume flows are close to the design value. However, the air flows of upstream edge fans greatly decrease during windy conditions. The stronger the wind is, the lower the air flows across upstream edge fans are. Fan(1,4) has a volume flow reduction of 25% of design value at the wind speed of 6m/s, and then the volume flow reduction increases to 42% at the wind speed of 9m/s. The static pressure contour for the Column 4 at the profile of y=6.6m under a wind speed of 6m/s is presented in Fig.6 (a). It is evident that there is a negative pressure region under the ACSC platform, especially at the inlet of windward fans, which results in volumetric effectiveness reduction of the fans. Fig.7 (a) denotes the static pressure contour for

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Fan(1,4) shroud inlet plane under a wind speed of 6m/s, and an uneven pressure distribution can be found at Fan(1,4) shroud inlet. Fig.5 also indicates that the volume flows across almost all fans increase as considering diffuser orifice plate, especially for upstream fans. The improvement amplitude is enlarged with increasing wind speed. The volumetric effectiveness of Fan(1,4) is increased by 19.9% at the wind speed of 9m/s as considering diffuser orifice plate. As can be seen from Fig.6, there is an obvious negative pressure region which pressure is lower than -80Pa, while this region disappears by adding diffuser orifice plate. As ambient wind flowing through diffuser hole in the plate, the velocity head converts to the pressure head, which shrinks the negative pressure region under the ACSC platform and sequentially improves the fans performances. On the other hand, a more uniform flow field can be obtained via diffuser orifice plate, as denoted in Fig.7(b) and Fig.8(b). It is evident that the static pressure distribution at Fan(1,4) shroud inlet is more even, which contribute to forming a uniform air flow across finned tube bundles and further improving ACSC performance [14]. The diffuser orifice plate also makes vortex flow at the inlet of upstream condenser cell disappear, as shown in Fig.8(b). 4.2 Fan inlet air temperature The average inlet temperature of 8 fans in Column 4 is calculated and presented in Fig.9. It is found that the average inlet temperature is obviously higher than the ambient as wind speed exceeds 6m/s. It has a temperature rise of 0.05 K at the wind speed of 6m/s, and then increases to 0.3 K at the wind speed of 9m/s. Not all fans inlet temperatures are sensitive to the wind speed. Fig.10 denotes each fan inlet air temperature at the +X wind speed of 9m/s. It is evident that the inlet air temperatures of most of the upstream and interior cells are higher than that of fresh

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ambient air at a wind speed of 9m/s, while a few of the downstream cells are at or near the fresh air temperature. Fig.11 (a) and Fig.12 (a) are temperature contours for the Column 4 at the profile of y=6.6m and for Fan(1,4) inlet plane under a wind speed of 9m/s, respectively. It is obvious that the greatest amount of temperature increase normally occurs at the upstream edge of Column 4 where the low pressure region under the ACSC platform draws a portion of the hot air down and into the inlet air stream. The diffuser orifice plate can decrease effectively fan inlet air temperature under windy conditions. As shown in Fig.9, the average inlet temperature of 8 fans in Column 4 is almost unchanged with increasing wind speed for diffuser orifice plate case, even at high wind speed. Fig.10 compares the inlet temperature of each fan at the +X wind speed of 9m/s. It is found that almost all fans inlet air temperatures decrease to ambient temperature as considering diffuser orifice plate. Especially, the temperature drop with a maximum value of 1.1K is found for windward Fan(1,4) at the wind speed of 9m/s. Fig.12 also compares the temperature contour for Fan(1,4) inlet plane under the wind speed of 9m/s. For without diffuser orifice plate case, a high temperature region is found, while it disappears for diffuser orifice plate case. Moreover, a more uniform temperature distribution at fan inlet plane can be obtained via diffuser orifice plate, which is helpful to improve heat transfer of condenser cell. 4.3 Heat transfer effectiveness As mentioned above, the fan performance deterioration and air temperature increase at fan inlet caused by ambient wind lead to a decrease of heat transfer performance of condenser cell. Fig.13 displays the heat transfer effectiveness at Column 4 against wind speed. The heat rejection rate drops sharply as the wind speed exceeds 3m/s, and the heat transfer effectiveness decreases to 0.92

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with the further increase of wind speed up to 9m/s. Almost no difference in heat transfer effectiveness is found between without and with diffuser orifice plate case under windless condition. However, the diffuser orifice plate obviously improves the heat transfer performance of Column 4 section under windy conditions. The stronger the wind is, the greater the improvement amplitude is. At the wind speed of 9m/s, the heat transfer effectiveness is increased by 6.6 percent points. The heat transfer rate improvement indicates that diffuser orifice plate is an effective way to restrain the adverse impacts of ambient wind. For the Column 4 section with diffuser orifice plate, the heat transfer effectiveness no longer decreases monotonically with the increase of wind speed. Enhancing wind speed makes the fan performance degradation, but on the other hand, the pressure head that is converted from wind velocity head via diffuser orifice plate increases simultaneously. Due to the cooperation action of these two factors, the heat transfer effectiveness increases at first, and then it decreases. As a result, the Column 4 section with diffuser orifice plate shows the best performance under the wind speed of 6m/s. 5. Conclusions Wind can have an important deleterious effect on ACSC performance due both to recirculation of hot exit plume air into the ACSC inlet and to degradation of fan performance resulting from negative pressure region under the fan platform. A new measure to improve ACSC performance is proposed by installing the diffuser orifice plate under the ACSC platform. The impacts of diffuser orifice plate upon the thermal-flow characteristics of Column 4 section in a 600 MW ACSC are numerically investigated. As ambient wind flowing through diffuser hole in the plate, the velocity head converts to the pressure head, which contributes to shrink the negative pressure region under the ACSC platform.

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Under windy conditions, the diffuser orifice plate not only increases the volume flows across almost all fans, but decreases fans inlet air temperatures, especially for upstream fans. At a wind speed of 9m/s, the volumetric effectiveness of windward Fan(1,4) is increased by 19.9%, and the temperature drop at windward Fan(1,4) inlet is 1.1K, as considering diffuser orifice plate. Moreover, a more uniform flow field distribution such as pressure, temperature as well as air flow rate can form via diffuser orifice plate, which may improve heat transfer of condenser cell. As a result, the diffuser orifice plate obviously improves the heat transfer performance of Column 4 section under windy conditions, which indicates that it is an effective way to restrain the adverse impacts of ambient wind. Due to the cooperation action of diffuser and ambient wind, the Column 4 section with diffuser orifice plate at the wind speed of 6m/s shows the best performance. Acknowledgments This research is supported by “the Fundamental Research Funds for the Central Universities No. 2015MS116”. References [1] C. Butler, R. Grimes, The effect of wind on the optimal design and performance of a modular air-cooled condenser for a concentrated solar power plant, Energy, 68 (2014) 886-895. [2] K. Duvenhage, D.G. Kröger, The influence of wind on the performance of forced draught air-cooled heat exchangers, J. Wind Eng. Ind. Aerod., 62(2-3) (1996) 259-277. [3] X.L. Zhang, J.P. Wang, H.P. Chen, Calculation model of exhaust pressure of direct air-cooled unit considering the effect of environmental wind, Proceedings of the CSEE, 32(23) (2012) 40-47. [4] W.F. He, Y.P. Dai, J.F. Wang, Performance prediction of an air-cooled steam condenser using UDF method, Appl. Therm. Eng., 50 (2013) 1339-1350.

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[5] M.T.F. Owen, D.G. Krӧger, An investigation of air-cooled steam condenser performance under windy conditions using computational fluid dynamics, J. Eng. for Gas Turbines and Power, 133(2011)604502. [6] Z.F. Gu, X.R. Chen, W. Lubitz, et al., Wind tunnel simulation of exhaust recirculation in an air-cooling system at a large power plant, Int. J. Therm. Sci., 46(3)(2007)308-317. [7] J.G. Bustamante, A. S. Rattner, S. Garimella, Achieving near-water-cooled power plant performance

with

air-cooled

condensers,

Appl.

Therm.

Eng.

(2015),

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condensers by using flow guiding device, J. Enhanc. Heat Transf., 19(2012) 63-74. [15] W.X. Zhang, L.J. Yang, X.Z. Du, et al., Thermo-flow characteristics and air flow field leading of

the

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Fig.1 Location of the diffuser orifice plate.

(a)

(b)

(c)

Fig.2 Diffuser orifice plate: (a) 3D, (b) 2D, and (c) Diffuser hole.

Fig.3 Column 4 section in a 600MW ACSC.

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Fig.4 Numerical model of A-frame condenser cell.

(a)

(b)

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(c)

(d) Fig.5 Volumetric effectiveness of Column 4 including 8 axial fans at wind speeds of (a) 0m/s, (b) 3m/s, (c) 6m/s, and (d) 9m/s.

(a)

(b)

Fig.6 Static pressure plots for the Column 4 at the profile of y=6.6m under a wind speed of 6m/s (a) without diffuser orifice plate, and (b) with diffuser orifice plate.

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(a)

(b)

Fig.7 Static pressure plots for Fan(1,4) shroud inlet plane under a wind speed of 6m/s (a) without diffuser orifice plate, and (b) with diffuser orifice plate.

(a)

(b)

Fig.8 Streamline plots for the Column 4 at the profile of y=6.6m under a wind speed of 6m/s (a) without diffuser orifice plate, and (b) with diffuser orifice plate.

Fig.9 The average inlet temperature of 8 fans in Column 4.

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Fig.10 Fan inlet air temperature at the +X wind speed of 9m/s.

(a)

(b)

Fig.11 Temperature plots for the Column 4 at the profile of y=6.6m under the wind speed of 9m/s (a) without diffuser orifice plate, and (b) with diffuser orifice plate.

(a)

(b)

Fig.12 Temperature plots for Fan(1,4) inlet plane under the wind speed of 9m/s (a) without diffuser orifice plate, and (b) with diffuser orifice plate.

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Fig.13 Heat transfer effectiveness at Column 4 under windy conditions.

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Table.1 Dimensions of the diffuser orifice plate. Items

Values

Inlet diameter of diffuser orifice (m)

d1

0.6

Outlet diameter of diffuser orifice (m)

d2

0.9

Thickness of diffuser orifice plate (m)

L

1.0

Horizontal pitch (m)

s1

1.0

Vertical pitch (m)

s2

1.0

Height of diffuser orifice plate (m)

H

10

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