Estimation of Condensate Mass Flow Rate During Purging Time in HRSG of CC Power Plants

7/30/2019 Estimation of Condensate Mass Flow Rate During Purging Time in HRSG of CC Power Plants

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condensate is formed in the superheater . If the condensate is not completely drained from

all superheater tubes before steam flow is established during the start up, two adverse

consequences are result [5]:

First, condensate that often is still sub-cooled is ejected in large quantity into the

outlet header and pipe manifold where it quench-cools hotter material. On hot starts

after trips, the outlet header and manifold can be more than 200 oC above saturation

temperature. During warm starts, the outlet header will be close to saturation

temperature but the outlet manifold temperature will still be substantially higher.

Second, tubes have different average temperatures at different times because they do

not clear of condensate simultaneously. Some tubes will be clear of condensate and

therefore will be cooled by steam flow, while others will have part of the tube filled

with condensate at saturation or lower temperature. Still other tubes will be stagnant

and therefore at the purge air or exhaust gas temperature. Designs that have stiff tube

arrangements connected to headers at each end develop. Thus, the unit will develop

large stresses at tube-attachment welds when other tubes in the same row are at

different temperature.

To eliminate both of these effects, the condensate needs to be removed from the lower

headers at the peak rate at which it forms. But that is not easily accomplished. Many

researchers have been conducted about modeling of HRSG [6-11]. Dechamps described a

method used to compute the transient performances of assisted circulation heat recovery

steam generators. In his mathematical modeling bundle of tubes in each part of HRSG

replaced with ‘equivalent’ linear heat exchanger [6]. Valdes et.al proposed a methodology to

identify the most relevant design parameters that impact on the thermal efficiency and the

economic results of combined cycle gas turbine power plants in full and part loads [7].

Mohagheghi et.al developed a new method for modeling a steam cycle in advanced combinedcycles by organizing non-linear equations and their simultaneous solutions by use of the

hybrid Newton methods [8]. Godoy et.al proposed Optimal designs of a CCGT power plant

characterized by maximum second law efficiency values. These thermodynamic optimal

solutions were found within a feasible operation region by means of a non-linear

mathematical programming (NLP) model, where decision variables (i.e. transfer areas, power

production, mass flow rates, temperatures and pressures) could vary freely [9] . Woudstra

et.al investigated internal exergy efficiency for combined cycles used the same gas turbine

but have different steam bottoming cycles. Differences did originate from the number of



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pressure levels at which steam is generated in the HRSG (Heat Recovery Steam Generator).

The evaluation includes respectively a single pressure, double pressures and triple pressures

HRSG. The steam pressures were optimized with regard to overall plant efficiency using a

multi-parameter optimization procedure. They showed that internal exergy evaluation was

useful method to cycle performance promotion [10]. Bahadori et.al provided a simple model

to estimation of the percent of blowdown that is flashed to steam as a function of flash drum

pressure and operating boiler drum pressure followed by the calculation of the amount of heat

recoverable from the condensate [11].

Due to literature survey, no research has been conducted the HRSG during purging time. In

this paper, transient modeling of HRSG to estimate of condensate mass flow rate during

purging time was conducted. For this purpose, a transient analytical modeling of HRSG was prepared and the effect of cooling gas temperature and mass flow rate on condensate was

investigated. In addition effect of purge time on condensate was also considered. In

summary, the followings are the specific contribution of this study in the subject matter area:

Transient modeling of heat recovery steam generator (HRSG), especially during

purge time

Comparing the numerical output of the model with the actual measurements at a real

power plant,

A computer program code was developed in the MATLAB which could be applied

for simulation of any types of HRSG,

2- System description

The typical power and steam generation plant, which we considered in this research consists

of the followings items:

Gas turbine and generator: Four 4 V94.2 gas turbine generator and auxiliaries,

HRSG : Two 2 Single pressure level waste heat recovery steam generators,

Common system: Two 2 spray tray type deaerators, Seven 7 boiler feed water pumps

(six 6 electric motor driven and one 1 steam turbine driven), Two 2 recirculation

pumps, Four 4 Booster pumps, Two 2 fuel oil pump, blow down system, sampling

system and dosing system

Waste heat from gas turbine exhaust heat passes through the individual HRSG will generate

steam. Each HRSG generates HP steam to HP steam common header. Also LP steam is



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produced through let down station which its source is HP steam extracted from HP common

header. Demine water is supplied to external deaerators through preheater which its

temperature shall be increased before enters to preheater by recirculation pumps. Feed water

to the HRSG and Auxiliary Boiler economizer shall be fed by a common feed pumps, which

shall take their suction from the feed water storage tank.

This HRSG has the following condition:

Each HRSG maximum capacity : 340000 kg/h (gross),

Installation: Outdoor,

Type : Fired, Natural circulation,

Operating mode : Constant pressure,

Duct & casing will be insulated for surface temperature below 60 °C at ambient

temperature 37 °C,

Gas pressure drop: 300.00 mmH2O based on static head,

HRSG configuration: Length (From Diverter center to Outlet Stack Center): 34.32 m

And Inlet Duct Angle: 51 Deg,

Figure (1) shows the schematic diagram of system. For mathematical modeling, three control

volumes are considered. These control volumes are two superheaters 1, 2 and evaporator,

which are shown in Figure (1). During purging time, inlet HP steam drum and attemperator

are closed. Purging time is about 300 (s), which condensate is formed in superheaters 1, 2 and

evaporator are drained through a drain which is located below superheaters.

Specification of HRSG heating surface is shown in Table (1). Steam conditions in evaporator

and superheaters 1 and 2 are shown in Table (2). The purge gas specification is shown in

Table (3).



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Figure (1): Schematic diagram of system.

Table (1): Specification of HRSG heating surface

Component

Geometry Data Fin Data Weight

SL ST STN SLN HS

No

Tube

s

OD Thick Finned

Length

No

FinsThk Height Type Tube Fin Total

(mm) (mm) (-) (-) (m^2) (-) (mm) (mm) (m) (/m) (mm) (mm) (-) (Ton) (Ton) (Ton)

SH-HP2 101.6 114.3 64 3 1647 192 50.8 2.7 17.0 100 1.2 9.0 SOLID 10.5 5.2 15.7

SH-HP1 101.6 114.3 64 2 2206 128 50.8 2.7 17.0 142 1.0 11.0 SOLID 7.0 5.2 12.1

EV-HP2 90.0 96.0 78 6 13797 468 38.1 2.4 17.0 260 1.0 19.0 SERRATED 16.8 43.2 60.0

EV-HP1 90.0 96.0 78 12 27594 936 38.1 2.4 17.0 260 1.0 19.0 SERRATED 33.6 86.3 119.9

Table (2): Steam conditions in evaporator and superheaters 1 and 2

ComponentPressure

(bar)

Temperature

(oC)Steam condition

SH-HP2 52.00 502.00 Superheat

SH-HP1 52.00 392.00 Superheat

EV-HP2 52.00 266.00 Saturate

EV-HP1 52.00 266.00 Saturate



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Table (3): Purge gas specification

Gas component Temperature (oC) Pressure (Bar) mP (kg/s) Time (s)

N2 (76%)

100 3 386.86 300O2 (21%)

CO2 (1%)

Others (1%)

3- Mathematical modeling

The heat balance for each control volume can be written as following:

)TT(Cmhmdt

dUm

dt

dTCm

dt

dTCm goutginggfgw

s

sf

f f P

P p (1)

In which, p

m is a total pipe mass in each control volume (kg), p

C is a specific heat of pipe

(kJ/kgoC), pT is a pipe temperature (oC), t is time (s),

f m is a total fin mass in each control

volume (kg),f

C is a specific heat of fin (kJ/kgoC),f

T is a fin temperature (oC),s

m is a

total steam mass inside pipes in each control volume (kg),sU is an internal energy of steam

(kJ/kg), p

gm is a purge gas mass flow rate (kg/s),

gC is a constant pressure specific heat of

purge gas (kJ/kgoC),ing

T is an inlet purge gas temperature (oC),ou tg

T is a outlet purge gas

temperature (oC), pw

m is a condensate mass flow rate (kg/s),wC is a specific heat of

condensate (kJ/kgoC) and p

fgh is a condensate enthalpy (kJ/kg). Above equation can be

written for three control volumes which are shown in figure (1). In general, this equation

shows energy conservation in control volumes. First to fourth terms in left hand of above

equation, show pipes, fins, steam, condensation energy variation with cool gas flowing. Right

side of above equation shows energy variation of purge air which it flows control volumes

showed in Figure 1.

Also energy gas reduction rate is equal to heat transfer between gas and HRSG control

volume:

)TT(CmTAU goutgingglntt (2)

In which,t

U is an overall heat transfer conductivity (W/m2oC),t

A is an overall heat

transfer surface (m2) andln

T is a log-mean temperature (oC).

Fin temperature can be calculated by the following equation [12]:



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ttii

goutttsii

f AUAU

TAUTAUT

(3)

In whichi

U is an inside tubes overall heat transfer coefficient (W/m2oC) andi

A is an

internal surface (m2

).

Overall heat transfer coefficient can be calculated by the following equation [12]:

ii

t

i

o

m

t

i

t

i

o

o

thA

A)

d

d(ln

k 2

A

A

Aff

h

1ff

U

1

(4)

In which,oi ff ,ff are fouling factors inside and outside of tubes (m2oC/W),

mk is a thermal

conductivity of tube wall (W/moC),io d,d are outlet and inlet pipe diameters (m),

io h,h

are outlet and inlet heat convection coefficient (W/m2oC) and is a fin effectiveness. Inlet

and outlet overall heat transfer coefficients can be calculated by the following equations [12]:

)d

d(ln

k 2

Aff

U

1

i

o

m

i

i

i

(5)

o

o

oh

1ff

U

1

(6)

Outlet overall heat transfer coefficient can be calculated by the following equation [12]:

67.0

pm

f

gm

o

o

531oPr CG)

15.273T

15.273T()

d

h2d(CCCh

(7)

In which, 531 C,C,C are constants, do is outlet pipe diameter (m), h is a fin height (m), Tgm

is an average purge gas temperature (oC), G is a gas mass velocity (kg/m2s), Cpm is a mixture

gas heat capacity (kJ/kgoC) and Pr is a prandtl number. G gas mass velocity can be calculated

by the following equation:

nbh2dSL NmG

oTW

g

(8)

Nw is a number of pipes wide, L is a effectiveness length of pipe (m), ST is a transverse pitch

(m), n is a number of fins per meter (NO/m), b is a fin thickness (m) and h is a fin height (m). C1, C2 and C3 are obtained from Table (4). In this table ST and SL are transverse and

longitudinal pitches (m) and s is a fin clearance (m), which can be calculated by the following

equation:

bn1s (9)



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And Reynolds number:

Gd

Re (10)

Table (4): Factors C1,C3 and C5 for solid and serrated fins in in-line and staggered arrangements Solid fins: 35.0

1Re25.0C

In-line: )s/h25.0exp(65.02.0C3

,

)S/S2exp()) Nd7.0exp(5.175.0(1.1C TL5

Staggered: )s/h25.0exp(65.035.0C3

)S/Sexp()) Nd7.0exp(8.07.0(7.0CTL

2

5

Serrated fins:

35.0

1Re25.0C

In-line: )s/h35.0exp(5.035.0C3

,

)S/S2exp()) Nd7.0exp(5.175.0(1.1CTL5

Staggered: )s/h35.0exp(45.055.0C3

)S/Sexp()) Nd15.0exp(8.07.0(7.0C TL

2

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For average inside pipe heat transfer coefficient, we use the following equation [13]:

2/1

i

2/1

f

f

2/1

f f

i Pr k

065.0h

(11)

In which k f is a thermal conductivity of liquid phase inside pipe (W/moC),f

is a liquid

phase density inside pipe (kg/m3),f

is liquid phase viscosity (NS/m2), Pr f is a prandtl

number of liquid phase andi

is an interfacial shear stress (N/m2).i

is an interfacial shear

stress which is calculated by the following equation [14]:

g2

Gf

2

g

ii(12)

In which f i is an interfacial friction factor evaluated as for single-phase pipe flow at a mean

vapor mass velocity Gg. For both solid and serrated fins, effectiveness is [13]:

t

f

A

A)E1(1 (13)

For solid fins [46]:

bh2 bdhd2h2nA oo

2

f (14)



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)nb1(dAAof t

(15)

E is fin efficiency can be calculated by the following equation [13]:

5.0

o

o22

dh2dhm330034.01

1E

(16)

m is a factor which can be calculated by the following equation [14]:

5.0

o

kb

h2m

(17)

In which k is a fin thermal conductivity (W/moC). For serrated fins [13]:

ws

bh2h b5.0ndA of

(18)

In which, ws is a fin width (m).

Factor m can be calculated by the following equation [14]:

5.0

o

kbws

)ws b(h2m

(19)

The calculation process is as following:

1-

Equation (1) which shows energy balance between cool gas (purge air) and controlvolumes is coupled with equation (2) which shows heat transfer between purge air

and HRSG.

2- Parameters included in equations (1) and (2) are calculated by equations (3) to (19)

3- One second is considered as step calculation.

4- Results and discussion

For validation the results, the comparison between model and experimental data is applied.

Figure (2) shows comparison of system pressure variation during purging time between

model and measured data. A good agreement between model and measured data is observed.



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Figure (2): Comparison of system pressure variation during purging time between model and measured data

Figure (3) shows variation of outlet purge gas temperature from superheaters 1&2 and

evaporator during purge time. At first, purge gas entered the first control volume

(supeheater2), and it is warmed in this control volume, after it, purge gas entered the second

and third control volumes (superheater 1 and evaporator) and also it is warmed in these

control volumes. It is observed that, outlet purge gas temperature from evaporator is higher

than superheater 1 and also outlet purge gas temperature from superheater 1 is higher than

superheater 2 during purge time. Figure (4) shows the steam temperature variation in super

heaters 1&2 and evaporator during purge time. Since, purge cool gas has a minimum

temperature in this control volume, and has not been warmed yet, steam temperature in

superheater 2 has the most temperature decrease rate during purge time, so steam temperature

is decreased from 502 (oC) to 130 (oC). Minimum decrease steam temperature rate is

occurred in evaporator. It can be seen from Figure (4), this reduction is about 47 ( oC) during

purge time. This variation has two reasons; First reason is related to purge gas which it has

been warmed in superheaters 1&2 before it entered to evaporator, so the effect of purge gas

cooling is reduced; Second reason is related to mass of steam, pipes and fins which are morethan superheaters 1&2.

Figure (5) shows the pipe temperature variation in superheaters 1&2 and evaporator during

purge time. This variation is similar to Figure (4).

0 50 100 150 200 250 30020

25

30

35

40

45

50

55

Time(sec)

P ( k P a )

Model

Measured data



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Figure (3): Variation of outlet purge gas temperature from superheaters 1&2 and evaporator during purge time

Figure (4): Steam temperature variation in superheaters 1&2 and evaporator during purge time

Figure (5): Pipe temperature variation in superheaters 1&2 and evaporator during purge time

0 50 100 150 200 250 300100

120

140

160

180

200

220

240

260

280

time(sec)

T g

( C )

Outlet gas temp SHI

Outlet gas temp SHII

Outlet gas temp EV

0 50 100 150 200 250 300100

150

200

250

300

350

400

450

500

550

time(sec)

T s

( C )

Steam temp SH.I

Steam temp SH.2

Steam temp EV



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Figure (6) shows condensate mass formation during purge time. It can be seen, condensate

formation starts after 80 (s) of purge start. At end of purge time, condensate formation is

about 238 (kg) or 238 (l) which it should be drained from superheaters. Figure (7) shows

variation of condensate mass formation through purging time with purge gas temperature

(Model and measured output data). Average model error is about 5%.

From this figure, it can be concluded that, if purge gas temperature decreases from 125(oC) to

25(oC) with same purge gas mass flow rate (386.86 (kg/s)), condensate mass formation

increases from 238(kg) to 258(kg).

Figure (6): Condensate mass formation during purge time

Figure (7): Variation of condensate mass formation through purging time with purge gas temperature



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5- Conclusions

Boiler and combustion safety codes require that HRSG’s are purged prior to start-up to

ensure that any remaining combustible gases from the gas turbine are removed and do not

present a risk of explosion. This purging process requires the movement of large volumes

of cold ambient air through the HRSG which cools the steam within the hot tubes and

results in condensation. Large volumes of cool condensate can be formed which then fall

onto bottom headers which, because of their thick walls and mass, remain hot. This

causes local quenching and possible cracking at stress concentrations such as tube

penetrations. It is important that the high volumes of condensate which may be generated

during the purge are drained prior to re-start. Otherwise, this condensate would remain at

the lower saturation temperature (337oC at 140 bar) until evaporated, cooling areas of the

superheater and reheater even though adjacent sections may have reached the superheater

temperature (560oC). Again this can give rise to large thermal stresses and possible

permanent deformation of tubes known as tube bowing. Tube bowing, once it has

occurred, can lead to by-passing of hot gases which can then impinge directly on cooler

downstream components and result in vibration issues, causing further fatigue failures.In

this paper investigated the mathematical modeling of transient HRSG during purge time

in order to drain condensate formation in superheaters. For this purpose, a computer

program code was written in MATLAB to simulate HRSG during purge time. For

validation the model, numerical output of program was compared with measured data in

real power plant which average error is about 5%. During purge time, steam temperature

decreases from 502 (oC), 392 (oC) and 266 (oC) to 130 (oC), 130 (oC) and 220 (oC),

respectively and also steam pressure decreases from 52 (bar) to 23(bar) during purge

time. At end of purge time, condensate formation is about 220 (l) when inlet gas

temperature is equal to 100 (

o

C) and purge gas mass flow rate is equal to 386.86 (kg/s).

Nomenclature:

Ai Internal surface (m2)mp

Total pipe mass in each control

volume (kg)

At Overall heat transfer surface (m2)ms

Total steam mass inside pipe in each

control volume (kg)

bFin thickness (m)

mw

p Condensate mass flow rate (kg/s)



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C1,C2,C3,C5 Constantsn Number of fins per meter (NO/m)

Cf Specific heat of fin (kJ/kgoC) Nw Number of pipes wide

Cg Specific heat of purge gas (kJ/kgoC)Pr Prandtl number

Cp Specific heat of pipe (kJ/kgo

C)Pr f Prandtl number of liquid phase

Cs Specific heat of steam (kJ/kgoC)Re Reynolds number

Cw Specific heat of condensate (kJ/kgoC)s Fin clearance (m)

Cpm

The mixture gas heat capacity

(kJ/kgoC)SL Longitudinal pitch (m)

di Inlet pipe diameters (m)ST Transverse pitch (m)

do Outlet pipe diameters (m)t Time (s)

EFins efficiencyTf Fin temperature (oC)

f i Interfacial friction factor Tp Pipe temperature (oC)

ff i

Fouling factors inside of tubes

(m2oC/W)Ts Steam temperature (oC)

ff o

Fouling factors outside of tubes

(m2oC/W)Tg in Inlet purge gas temperature (oC)

GGas mass velocity (kg/m2s)

Tg

out Outlet purge gas temperature (oC)

Gg mean vapor mass velocityTgm Average purge gas temperature (oC)

hFin height (m)

∆T

Ln log-mean temperature (oC)

hi

Inlet heat convection coefficient

(W/m2oC)Ut

Overall heat transfer conductivity

(W/m2oC)

ho Outlet heat convection coefficient(W/m2oC)Ui

Inside tubes overall heat transfer

Steam internal energy (kJ/kg) Us coefficient (W/m2oC)

hfg p Condensate enthalpy (kJ/kg)wsFin width (m).

k Fin thermal conductivity (W/moC)Greek symbols

k f

Thermal conductivity of liquid phase

inside pipe (W/moC)ρf

Liquid phase density inside pipe

(kg/m3)

km

Thermal conductivity of tube wall

(W/mo

C)ɳ Fin effectiveness



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L Length of pipe (m)Liquid phase viscosity (NS/m2)

mf

Total fin mass in each control volume

(kg)τi Interfacial shear stress (N/m2)

mg p Purge gas mass flow rate (kg/s)

References:

[1] J.H.Horlock. "Combined power plants — past, present, and future". ASME Journal of Engineering for Gas

Turbines and Power 117 (1995) 608 – 16.

[2] S.Sanaye,S. and M.Rezazadeh, " Transient thermal modeling of heat recovery steam generators in combined cycle

power plants", International journal of energy research 122 (2006) 547-556.

[3] M.Pearson. and R.W.Anderson " Reliability and Durability from Large Heat Recovery Steam Generators".

Proceeding Institute Mechanical Engineering 1999: 213 Part A.

[4] R.Krowech, R.; L.Stanley; LLC Deltak Avoid damage from HRSG cycling; Power Magazine, March/April 1998

[5] "HRSG Users Handbook, "Design Operation and Maintenance", R.C.Swankekamp , HRSG Users Group,2006

[6] P.J.Dechamps. "Modeling the transient behavior of heat recovery steam generators". Journal of Power and Energy

209 (1995) 265 – 273.

[7]: M. Valdes, A. Rovira, M.D. Duran, Influence of the heat recovery steam generator design parameters on the

thermoeconomic performances of combined cycle gas turbine power plants, Int, Journal of Energy Research 28

(2004) 1243-1254.

[8]: M. Mohagheghi, J. Shayegan, Thermodynamic optimization of design variable and heat exchangers layout in

HRSGs for CCGT using genetic algorithm, Applied Thermal Engineering 29 (2009) 290-299.

[9]: E. Godoy, N.J. Scenna, S.J. Benz, Families of optimal thermodynamic solutions for combined cycle gas turbine

(CCGT) power plants, Applied Thermal Engineering 30 (2010) 569-576.

[10]: N. Woudstra, T. Woudstra, A. Pirone, T. van der Stelt, Thermodynamic evaluation of combined cycle plants,

Energy Conversion and Management 51 (2010) 1099-1110.

[11]: A. Bahadori*, H.B.Vuthaluru. A method for estimation of recoverable heat from blow down systems during

steam generation. Energy 35 (2010) 3501-3507.

[12]: ESCOA Corp. ESCOA Fintube Manual. Tulsa, OK: ESCOA, 1979.

[13]: Carpenter, E.F. and Colburn , A.P., 1951." The effect of velocity on condensation inside tubes", Proc. Of

General Discussion on Heat Transfer,20-26, IMechE/ASME.

[14]: Collier, J.G., "Convective boiling and condensation", Second edition, 1981.

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Estimation of Condensate Mass Flow Rate During Purging Time in HRSG of CC Power Plants

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