Numerical simulation of heat transfer performance of an air...

ORIGINAL

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

Xiufeng Gao Æ Chengwei Zhang Æ Jinjia Wei ÆBo Yu

Received: 10 April 2009 / Accepted: 28 July 2009 / Published online: 13 August 2009

� Springer-Verlag 2009

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 phys-

ical 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 plat-

form 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

Dh Hydraulic diameter

h Platform height (m)

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)

Umin 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/(m2K)]

b Wind direction angle (degree)

k Ground roughness coefficient

m Kinematic viscosity (m2/s)

q Density (kg/m3)

Subscripts

0 Reference situation

1 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 [1]. There is an

array of air-cooled units in the ACSC, each consisting of an

A-frame configuration of finned tube heat exchanger bun-

dles below which an axial flow fan is fixed. A stream of

X. Gao � C. Zhang � J. Wei (&)

State Key Laboratory of Multiphase Flow in Power Engineering,

Xi’an Jiaotong University, 710049 Xi’an, China

e-mail: [email protected]

B. Yu

Beijing Key Laboratory of Urban Oil and Gas Distribution

Technology, China University of Petroleum (Beijing),

Beijing, China

123

Heat Mass Transfer (2009) 45:1423–1433

DOI 10.1007/s00231-009-0521-x

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 [2, 3]. Secondly, the

surroundings of the ACSC, such as the main factory

building and terrain, also have influences [4, 5]. If build-

ings 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 [3].

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 [2, 6–13], 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 recir-

culation has a very large impact on the heat rejection rate

of the ACSC due to the existence of the main factory

building [10]. 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 dis-

tortion 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 tem-

perature 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 sur-

roundings, 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 air-

cooled 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 unfavor-

able 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.

2 Numerical model and methods

2.1 Geometric model

Direct air-cooled condensers from a 2 9 135 MW thermal

power plant are used for the simulation. To reduce the

length of steam pipes in a power plant, the platform of air-

cooled 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 28 m high steam turbine room and a 180 m high

chimney, together with the definition of the incident angles

of wind, are shown in Fig. 1. A steep hill with an elevation

angle of 80� and a maximum height of 18 m in the west of

the air-cooled platform is also shown in Fig. 1 for the

investigation of the terrain effect. The ACSC has 32 air-

cooled 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 turbine-

steam flows inside the steel elliptical tubes, and the cooling

air is drawn through the fan to take the heat from the

1424 Heat Mass Transfer (2009) 45:1423–1433

123

exhaust turbine-steam which converts to condensate. The

air-cooled units distributed in 8 columns 9 4 rows are

mounted on a 4 m high steel rectangular platform, which is

supported by 2 9 6 cylindrical pillars made of reinforced

concrete. To avoid the unfavorable effect of wind and

increase the efficiency of condensers, a windbreak is con-

structed around the platform with an equivalent height of

the condensers 10 m. Four different heights, 22, 25, 28 and

32 m, of the air cooled platform (pillars) are studied for

selecting an optimum designing height.

2.2 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 com-

putational effort. In order to guarantee the computational

efficiency without losing the correctness, the single air-

cooled unit in the present study is simplified to a

11.1 9 11.04 9 10 m3 (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 temper-

ature 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 prob-

lem 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 1:18 Re0:287Pr0:33 ð1Þ

where Nu is the Nusselt number, Nu = aDh/k; Re is the

Reynolds number, Re = UminDh/v. Here, a is the average

heat transfer coefficient, Dh the hydrodynamic diameter,

k the thermal conductivity, Umin 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 mea-

sured value in the air-cooled unit. The measured pressure

drop across the heat exchanger is correlated in a dimen-

sionless form as

f ¼ 2577:3 Re�0:61 ð2Þ

where f is the drag coefficient, f ¼ Dp=0:5qU2min: Here, Dp

is the pressure drop.

WE

N

S

β

β

Wind direction angle

Chimney

Boiler rooms

Turbine room Air-cooled platform

Hill

Pillar

Wind wall

Fig. 1 Geometric model

Heat Mass Transfer (2009) 45:1423–1433 1425

123

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 = 217 - 3.26 V - 8.46 V2 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.

2.3 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 2,000 9 2,000 9 1,000 m3, 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. 2. The

total grid number is about 3,000,000. 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 2,5000,000. In Fig. 2, 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.

2.4 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 (1) and (2) 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

U10

¼ z

10

� �kð3Þ

where U10 is the standard wind speed at the height of 10 m;

z is the height above the ground; U is wind speed at z, and

k is the ground roughness coefficient ranging from 0.125

to 0.25.

2.5 Simulated conditions

To determine a reasonable platform height, four different

platform heights, 22, 25, 28 and 32 m, are simulated under

the most unfavorable operating conditions of the ACSC in

summer when the ambient air temperature is 28�C. The

wind speed, U, ranges from 0 to 4 m/s, and the wind

direction angle, b, is in the range of 0–180�.

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.

Fig. 2 Grids of geometry model


123

3 Results and discussion

3.1 Effect of wind direction

Figure 3a–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, 45 and 180�, respectively. The

height of the air-cooled platform (pillar) is 22 m and the

wind speed is 2.5 m/s. For b = 0�, a wake zone is gener-

ated 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. 3a; whereas for a reversed wind

direction of b = 180�, 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 gener-

ation 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. 3c. 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. 3a, 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 4 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. 3.

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.

x-directional profile y-directional profile z-directional profile



(a)

(b)

(c)

Fig. 3 Velocity and

temperature fields. a b = 0�,

b b = 45�, c b = 180�


123

3.2 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 = 22

and 32 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 [ 22 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 = 2 m/s at the

platform height h = 22 and 32 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 gen-

erate hot air recirculation, which induce an increase of air

temperature at the fan inlet below the ACSC [14]. 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 temperature-

increased area at the fan inlet becomes less as h increases

from 22 to 32 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 22, 25,

28 and 32 m. As the wind speed increases up to 2 m/s, the

total heat rejection rate decreases sharply, and then

decreases slowly with the further increase of wind speed up

to 4 m/s, indicating that there is no remarkable influence on

the total heat rejection rate for the wind speed ranges from

2 to 4 m/s. As the platform is lift up, the heat rejection rate

increases due to the less effect of the hot air recirculation

0 45 90 135 180350

400

450

500

β (o)

Q (

MW

)

h = 22 mU = 2.5 m/s

429.64 MW

Fig. 4 Effect of wind directions on the heat rejection rate

(a)

(b)



Fig. 5 Flow pattern and

temperature distribution at

U = 0 m/s. a h = 22 m,

b h = 32 m


123

shown in Fig. 6(b). The heat dissipation can always reach

above the design value of 429.64 MW for h = 32 m.

Figure 8 shows the effect of wind speed on steam tur-

bine back pressure at four different platform heights of 22,

25, 28 and 32 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 air-

cooled platform at h = 22, 25 and 32 m. The air mass

flowrate q is normalized by that in the ideal no wind case, q0.

In ideal case of no wind, a fan is assumed to work

independently without interference from the others. q/q0

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. 10 displays the wind speed effect on the increment of

the air temperature at the fan inlet area compared with the

ambient temperature at h = 22, 25, 28 and 32 m. It can be

(a)

(b)



Fig. 6 Flow pattern and

temperature distribution at

U = 2 m/s. a h = 22 m,

b h = 32 m

0 1 2 3 4 5350

400

450

500

550

U(m/s)

Q(M

W)

β = 0o

h = 22 mh = 25 mh = 28 mh = 32 m

429.64 MW

Fig. 7 Effect of wind speed and platform height on total heat

rejection rate

0 1 2 3 4 520

30

40

50

U(m/s)

Pb

(kPa

)

β = 0o

h = 22 mh = 25 m h = 28 mh = 32 m

35 kPa

Fig. 8 Effect of wind speeds on steam turbine back pressure


123

seen that the inhaled air temperature increases as the nat-

ural wind speed increases, suggesting an increasing hot air

recirculation. The inhaled air temperature decreases as the

platform height increases, indicating that elevating plat-

form 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 investi-

gates the effects of the steep slope to the west of the air-

cooled platforms and the 180 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 air-

cooled 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

hardly affect the thermal-flow field and thus the perfor-

mance of the ACSC.

3.3 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 dif-

ferent 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 28 m is finally

determined to be the designed value. Based on this plat-

form 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 11 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 180�.

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 rap-

idly 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

0 1 2 3 4 50.9

0.92

0.94

0.96

0.98

1

U (m/s)

q/q 0

β = 0o

h = 22 mh = 25 mh = 32 m

Fig. 9 Effect of wind speeds and platform height on cooling air mass

flowrate

0 1 2 3 4 50

2

4

6

8

10

U(m/s)

∆Tr (

k)

β = 0o

h = 22 mh = 25 mh = 28 mh = 32 m

Fig. 10 Effect of wind speed and platform height on fan inlet air

temperature rise

0 45 90 135 180350

400

450

500

β (o)

Q (

MW

)

TRL

U = 2 m/s U = 3 m/s U = 4 m/s U = 5 m/s

429.64 MW

Fig. 11 Effects of wind speed and direction on total heat rejection

rate


123

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. 11 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 10 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 12 shows the relationship between the steam

turbine back pressure Pb 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. 11, Pb decreases rapidly at first

as b rises for b\ 90�, and then levels off. While Pb can be

always kept at the normal value of 35 kPa for U = 2 m/s,

it becomes higher than 35 kPa for b\ 30� 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 13a and b show the distribution graph of the heat

rejection rate of the ACSC in summer condition for U = 0

and 3 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 air-

cooled platforms is 426 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

350 MW.

Figure 14a and b show the air mass flowrate inhaled by

the axial fans of different air-cooled units in TRL condition

for U = 0 and 3 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 12,225 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 air-

cooled platform. For a fan working independently without

interference from other fans, the inhaled air mass flowrate

is 390 kg/s in the present study. Therefore, the total air

mass flowrate will be 12,480 kg/s for 32 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 426 9 12,480/12,225 = 434 MW. So

the heat transfer capacity is about 2% 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

11,756 kg/s and a corresponding heat transfer reduction of

0 45 90 135 18025

30

35

40

β (o)

Pb

(kPa

)

TRL U = 2 m/s U = 3 m/s U = 4 m/s U = 5 m/s

35 Kpa

Fig. 12 Relationship of turbine back pressure with wind speed and

direction

1 2 3 45

67

8 0

1

23

45

1.00

1.25

1.50

1.75

q(10

7W

)

q(10

7W

)

Row

Column

1 2 3 45

67

8 0

1

2

3

4

1.00

1.25

1.50

1.75

RowColumn

(a) (b)Fig. 13 Heat rejection rate

distribution of air-cooled units

on summer condition. a No

wind (U = 0 m/s), b U = 3 m/s


123

426 - 426 9 11,756/12,225 = 16.3 MW. As shown in

Fig. 13, the total heat transfer reduction is 426 - 350 =

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 perfor-

mance, leading to a corresponding heat transfer reduction

of 76 - 16.3 = 59.7 MW. It is obvious that the hot air

recirculation is the main factor responsible for the reduc-

tion 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 [15].

4 Conclusions

Numerical simulation of the thermal-flow characteristics and

the heat transfer performance is made of an ACSC in a

2 9 135 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:

1. 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.

2. The ACSC performance increases and the correspond-

ing 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.

3. 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 perfor-

mance of the ACSC. This effect should not be

neglected in the simulation.

4. 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 sup-

port from the NSFC Fund (No. 10602043, 50821604, 50876114).

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300

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Q(k

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RowColumn

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