9 steam cycle heat exchangers

STEAM CYCLE HEAT EXCHANGERS
Jay F. Nagori and Samuel Tarson

9.1

INTRODUCTION

Two classifications of heat exchangers are associated with
the steam cycle. These are the condenser located at the
exhaust of the steam turbine, and feedwater heaters located
in the condensate/boiler feedwater cycle that utilize extrac­
ted steam from the steam turbine. There are many other heat
exchangers in the power plant, even some associated with
the steam turbine or condensate cycle, but only the con­
denser and feedwater heaters are considered "steam cycle
heat exchangers" and discussed in this chapter.

9.2

CONDENSER

The function of the condenser is to condense the steam
leaving the turbine, collect the condensate, and lower the
turbine exhaust pressure. A turbine with inlet steam condi­
tions of 850 pounds per square inch (psi) [5,861 kiloPascals
(kPa)] at 900° F (482° C) discharging at atmospheric pres­
sure converts about 300 British Thermal Units (Btu) per
pound [698 kilojoules per kilogram (kJ/kg)] of steam to
electricity. The same unit operating with a steam condenser
at 2 in. of mercury absolute pressure (HgA) (50.8 mm HgA)
converts about 450 Btu per pound (1,047 kJ/kg) of steam to

electricity.
The condensing of the steam requires the condenser to
remove the heat of vaporization from the steam and reject it.
Condensers are designed to reject this energy directly into
cooling water or directly into the atmosphere. Condensers
using cooling water are the norm, except in a few locations
where water is very expensive, in that case air-cooled con­
densers may be used. This chapter addresses primarily the
water-cooled condenser, and only generally discusses the
air-cooled condenser.
The Heat Exchange Institute (HEI) has published the
Standards for Steam Surface Condensers, which covers rec­
ommended standards of definitions, performance, and con­
struction, and it is recommended that the designer become
familiar with this publication.
9.2.1

Condenser Arrangement

The condenser is a steam-to-water, tube and shell heat ex­
changer with the cooling water, normally called circulating
250

water, passing through the tubes and the steam in the shell.
Because of the large volume of steam to be condensed, the
specific volume of steam at 2 in. HgA (50.8 mm HgA) is 339
ft3/lb (21.2 m3/kg). The condenser is a relatively large piece
of equipment. In very small sizes, the condenser shell may be
round, but in most cases, the size of the condenser shell
requires it to be constructed of reinforced flat plate in a box

shape. Figure 9-1 shows a typical condenser with water
boxes for inlet and outlet circulating water, tubes, tube sup­
port plate, shell, and hotwell where the condensate collects.
Most condensers are mounted in the plant directly below the
turbine exhaust so that the steam passes down through the
condenser neck into the condenser. Other physical arrange­
ments may locate the condenser at the side or end of the
turbine.
On very large turbines that have two or three low-pressure
casings, a separate condenser shell is mounted under each
low-pressure casing. Each shell is in effect a separate con­
denser. Figure 9-2 represents various circulating water sys­
tems and condenser shell arrangements.
On Figure 9-2, the top row of condensers indicates the
circulating water system for one-pass condensers. In onepass condensers, the circulating water enters the inlet water
box at one end of the shell, passes through the straight tubes,
and exits from the outlet water box at the other end of the
shell.
The second row of condensers indicates the circulating
water system for two-pass condensers. In two-pass con­
densers, the circulating water enters the inlet water box
at one end of the shell and passes through the tube bundle
consisting of half the total number of tubes. It then exits the
tubes into the return water box, from that the water passes
through the other half of the tubes exiting from the outlet
water box at the same end of the shell as the inlet water box.
If there are two or three condenser shells, all are arranged
identically.
The third row of condensers indicates an arrangement that
utilizes two and three shell installations where the circulating

water passes through the shells in series. The water exiting
the first shell enters the second shell, and so on. As the water
entering the second shell is warmer than the water entering
the previous shell, the performance of the second shell is less
than that of the first shell, resulting in a higher pressure in the
second shell than in the first shell. The pressure in the third


Steam Cycle Heat Exchangers

251

ogy. Figure 9-4 indicates some of the terms used in calculat­
ing the size of a condenser, including the following:

NECK

* SHELL

Turbine exhaust steamflow,pounds per hour (lb/h) (Fls)
Turbine exhaust steam pressure, in. HgA (Ps)
Turbine exhaust steam enthalpy, Btu per pound (ВшДЬ) (hs)
Saturation temperature of exhaust steam, °F (Ts)
Circulating water inlet temperature, °F (Г,)
Circulating water outlet temperature, °F (Г2)
Circulating water flow, gallons per minute (gpm) (Flcw)
Effective tube length, ft (L).
Additional condenser sizing information includes the fol­
lowing:

SUPPORT PLATE

WATER
BOXES

Fig. 9-1.

HOT WELL

From these definitions the following are derived:

Typical condenser.

shell is even higher than in the second shell. This condition is
referred to as multipressure and is used where the lower
water flow and its economics exceed the loss in turbine cycle
efficiency resulting from higher condenser pressures.
Figure 9-3 gives standard nomenclatures for the various
parts of a condenser.
9.2.2

Condenser Sizing

Condensers are basically steam to water heat exchangers and
are sized according to normal heat exchanger design technolCIRCULATING WATER OUTLET (Typical)
1 SHELL

/

2 SHELLS

3 SHELLS

ONE-PASS

f i n
^ —

TWO-PASS

ТГ Я Я Я

CIRCULATING WATER INLET (Typical)■

я я я я я я
я тт тт тт тт я
SINGLE PRESSURE

2 SHELLS

ONE-PASS

3 SHELLS

я—я я
тт у—я
MULTI-PRESSURE

Fig. 9-2.

Number of passes (NP)
Tube diameter, in., and gauge, Birmingham wire gauge (BWG)
Condensate temperature, °F (Tc)
Condensate enthalpy, BtuVlb (Ac)

Circulating water flow alternatives.

Initial temperature difference = Ts — Tx (ITD)
Terminal temperature difference = Ts — T2 (TTD)
Temperature rise = T2 - Г, (TR).
9.2.2.1 Design Heat Load. The heat load to be used in
designing a condenser is that energy that must be removed
from the exhaust steam to cause it to condense. Turbine
manufacturers normally list the exhaust steam enthalphy on
their heat balance, as shown in Fig. 9-5. These are listed as
ELEP and TJEEP. ELEP stands for expansion line end point,
that would be the enthalpy of the exhaust steam if there were
no losses at the outlet of the turbine. However, some losses
do occur at the outlet of the turbine, that reheats the steam,
resulting in the UEEP, usable energy end point. This is the
enthalpy of the steam entering the condenser and should be
used in determining design heat load.
By subtracting the enthalpy of the condensate leaving the
condenser (hc) from the steam enthalpy (As) and multiplying
by the exhaust steam flow (Fls), the design heat load is
obtained. As Ah = (hs — hc) is normally approximately 950
Btu/lb (2,210 kJ/kg), this number is frequently used in the
computation of design heat load.
In addition to the heat load created by the turbine exhaust
steam, other flows into the condenser may add further heat

load. These flows may include steam flow from a boiler feed
pump drive turbine, other drive turbines, heater drains, tur­
bine leak-offs, and other sources. All these must be added to
the turbine exhaust steam load to determine total condenser
design heat load.
9.2.2.2 Condenser Operating Pressure Optimization. The
optimum condenser operating pressure must be determined
in conjunction with the steam turbine. Limitations imposed
by circulating water parameters of flow, inlet temperature,
and temperature rise establish lowest possible condenser
pressure. This lowest possible pressure may not be the most


252

Power Plant Engineering

О

о

- т ~ —г ■

.I

\
16

Iw V V 5
14

%

15 94

16

18
1

I.

ТУ

DIVIDED WATERBOX

NON-DIVIDED WATERBOX

SINGLE PASS CONDENSERS

NON-DIVIDED WATERBOX

SPRING SUPPORTED TYPE

DIVIDED WATERBOX

ТУ

NON-DIVIDED WATERBOX

ALT. DIVIDED WATERBOX

TWO-PASS CONDENSERS
1.
2.
3.
4.
5.

EXHAUST CONNECTION
EXHAUST NECK
VENTING OUTLET
CONDENSATE OUTLET
CIRCULATING WATER
INLET OR OUTLET
6. TUBES
7. INLET WATERBOX
8. OUTLET WATERBOX

9.
10.
11.
12.

INLET-OUTLET WATERBOX
RETURN WATERBOX
BONNET TYPE RETURN WATERBOX
BONNET TYPE INLET - OUTLET
WATERBOX
13. WATERBOX COVERS

14. CONDENSER SHELL
15. HOTWELL
16. TUBE SHEETS

TUBE SUPPORT PLATES
HANDHOLES OR MANHOLES
SHELL EXPANSION JOINT
EXHAUST NECK
EXPANSION JOINT
2 1 . WATERBOX DIVIDING
PARTITION
22. WATERBOX PASS PARTITION
23. SPRING SUPPORTS
24. SUPPORT FEET
17.
18.
19.
20.

NOTE: MORE THAN ONE OF THE VARIOUS PARTS MAY BE INCLUDED IN A GIVEN CONDENSER. EXACT LOCATIONS
OF CONNECTIONS WILL VARY FROM ONE CONDENSER TO ANOTHER.
Fig. 9-3. Standard nomenclature for condenser parts. (From Standards for Steam Surface Condensers, Eighth Edition, Heat Exchange Institute, Inc.
Used with permission.)


Steam Cycle Heat Exchangers 253

STEAM INLET
Turbine
Turbine

Turbine
Turbine

Exhaust Flow, Ib/h (Fls)
Steam Enthalpy, Btu/lb (hs)
Exhaust Steam Pressure, in HgA (Ps)
Steam Saturation Temperature, F(TS)

EFFECTIVE TUBE LENGTH, FT (L)
<|fl
"^^
From Inside Face of Tube Sheet
to Inside Face of Tube Sheet

T

CONDENSATE
Temperature, F (T c )
Enthalpy, Btu/lb (h c )
INLET CIRCULATING WATER
Temperature, F (T-j)
Flow, gpm (Fl ow )

OUTLET CIRCULATING WATER
Temperature, F (T2)

TERMINAL TEMPERATURE
DIFFERENCE (TTD)

STEAM SATURATION

TEMPERATURE (Tg)

OUTLET CIRCULATING
WATER'TEMPERATURE (T 2 )

INITIAL TEMPERATURE
DIFFERENCE (ITD)INLET CIRCULATING
WATER TEMPERATURE (T-,)'

' CIRCULATING WATER
TEMPERATURE

CIRCULATING WATER
TEMPERATURE RISE (TR)

WATER FLOW

Fig. 9-4. Condenser sizing terminology.

economical from the viewpoint of first cost, operating cost,
and efficiency. An economic analysis is required to deter­
mine the optimum condenser operating design pressure.
Figures 9-6 and 9-7 from the HEI Condenser Standards
give recommended minimum operating design pressures for
varying cooling water inlet temperatures. It is recommended
that the operating pressure be limited by not allowing the
design terminal temperature difference to be less than 5 ° F
(2.8° C).
9.2.2.3 Condenser Sizing. The basic heat transfer equa­
tion used in sizing a condenser is listed below:

q = UAAT

(9-1)

ДГ:

ITD - TTD

Чтто)
A = tube surface area in square feet, and
U = Function of the heat transfer coefficient and the actual
construction and operation of the unit, and can be ex­
pressed as,

U=

Cx\/VxfxmxCl

where
С = heat transfer coefficient,
V = velocity of cooling water through the tubes,

where
q = design heat load in Btu/lb,
ДГ = log mean temperature difference defined as

(9-2)

/ = cooling water inlet temperature correction factor,
m = tube material and thickness correction factor, and

Cl = tube overall cleanliness factor.

(9-3)


254

Power Plant Engineering
LEGEND
H OR h - ENTHALPY, Btu/lb
» - FLOW, Ib/hr
P - PRESSURE, psia
F - TEMPERATURE, °F
Calculations based on
1967 ASME Steam Tables

4,444,386»

367.8F
394.4И

ЗБ1.3И

-rf\

BFP

3 4 5 . 0 h \ _ ^ 3164.7P
A h = 11.84

260.1h

147.9h

VALVE BEST POINT = 4,841,9610 11456.3-456.21 + 4,440,131 П 520.8 - 1299.01 = 8015 Btu kWh
727,072
NET HEAT RATE

142.9h

674,453 kW @ 1.5" HgA 0% MU
TC4F-30" LSB 3600 rpm
2400 psig 1000/1000°F
GEN: 806,500 kva @ 75 psig H2 PRESSURE & 0.90 (LIQ)

Fig. 9-5. Typical heat balance for a large turbine generator. (From General Electric Co. Used with permission.)

The Heat Exchange Institute has performed extensive
tests to establish reasonable heat transfer coefficients. These
are published in their Standards for Steam Surface Con­
densers. Table 9-1, taken from that source, gives values of С
for various tube diameters. These are based on admiralty
metal tubes of 18 BWG with inlet cooling water temperature
of 70° F. Figure 9-8, also from the HEI Standards, gives
correction factors for other inlet water temperatures, listed
above as Factor /. Table 9-2 gives correction factors for
several other tube materials and gauges, listed above as
Factor m. Table 9-3 lists tube characteristics. Cleanliness
factor provides for the loss of heat transfer with time attribu­
table to scale or other buildup and is established by the

designer based on the selected tube materials and cooling
water conditions.

In the design of a condenser, the majority of design pa­
rameters must be established by the designer. These nor­
mally include the following:
•
•
•
•
•
•
•
•

Turbine exhaust steam flow and enthalpy,
Turbine exhaust steam pressure,
Circulating water inlet temperature,
Tube diameter and gauge,
Tube material,
Number of passes,
Water velocity through tubes, and
Cleanliness factor.

In addition to the above parameters, the effective tube
length or the circulating water temperature rise is normally


Steam Cycle Heat Exchangers

255

ZA

CURVE A - CUTOFF
(Except Where the Absolute
Pressure is Limited by 5° F
Terminal Difference)

2.2

2.0

>-

cc
э
о
cc
Ш

-

/
/ i
/ /

2.0

>-

Ш

5

1.6

I

ш

У /д

ОС

1.2

(Я
(Л
ш

0_

1.8

cc

1.6

i

CC

-

CC
D

о 1.4
z
э

2.2

CURVE В - ZERO LOAD

1.8

7 ppb (CURVE A)
AND
14 ppb (CURVE B)

и
z

1.4

cc
w
и

1.2

y/d

Ш

ОС
Q.

1.0,

1.0

Ш

о

н
_1

(Л
03

СО

<

О 0.8

0.8

<
0.6

0.6

0.4

0.4
1

30

40

1
50

1
,60

L
70

'

•

80

90

100

TEMPERATURE OF COOLING WATER
TO CONDENSER, DEGREES F

Note: Performance is based on air removal equipment
having a capacity at one inch mercury absolute suction
pressure, of not less than that recommended by HEI and
the air and non-condensibles being removed from the
system not exceeding 50 percent of those values.
Fig. 9-6. Absolute pressure limit curves for steam turbine service.
(From Standards for Steam Surface Condensers, Eighth Edition, Heat
Exchange Institute, Inc. Used with permission.)

established. With this information, the other parameters of
the condenser, surface area, and circulating water flow
can be calculated. The following example illustrates this.
Example

9-1

Given:
Steam flow = Fl% = 800,000 lb/h
Steam turbme exhaust enthalpy = h3 = 1,038.6 BtuAb
Condenser pressure = Ps = 3.5 in. HgA
Circulating water inlet temperature = Tv = 85° F
Circulating water velocity through tubes = V = 7.0 fps

Temperature rise = TR = 27° F
Tube size = D = 1.0 in. OD; BWG = 22 gauge
Tube material = 304 SS
Cleanliness factor = 90%
Number of tube passes = 2

30

I
40

1
1
1
I
1
90
60
70
80
50
TEMPERATURE OF COOLING WATER
TO CONDENSER, DEGREES F

100

Fig. 9-7. Absolute pressure limit curves for steam turbine service for
oxygen contents. (From Standards for Steam Surface Condensers,
Eighth Edition, Heat Exchange Institute, Inc. Used with permission.)

Calculate:
Circulating water outlet temperature, T2 = T{ + TR = 85 + 27 =
112° F
Saturated steam temperature at 3.5 in. HgA, Ts - 120.56°F
Initial temperature difference, ITD = TS — Tv = 120.56 — 85 =
35.56° F
Terminal temperature difference, TTD = TS — T2 = 120.56 — 112
= 8.56°F

Table 9-1.

Heat Transfer Coefficient for
Steam Surface Condensers

Tube Diameter, OD, inches
% and %
% and 1
Ш and VA
l%and Ш
W» and 1%
VA and 2

Coefficient (C)
267
263
259
255
251
247

Note: Based on new clean Admiralty metal tubes of
18 BWG gauge and with inlet cooling water temperature
of 70° F


256

Power Plant Engineering

Corrected heat transfer

e

1.2

z
о

1.0

gш

0.8

U-

g

(9-5)

= 263 x V 7 X 1.06 x 0.87 x 0.90

cc
cc
О
О 0.6

z
52

Схл/VxfxmxCl

= 577.5
_

_ 1.

30

40

L

50

_«

60

1

1

1

1

1

70

80

90

100

110

Condenser surface
120

A . U M - 69,410 ft*

TEMPERATURE OF INLET WATER, DEGREES F

577.5 x 18.96

Fig. 9-8. Inlet water temperature correction factor. (From Standards for
Steam Surface Condensers, Eighth Edition, Heat Exchange Institute, Inc.

Used with permission.)

Circulating water flow

LMTD, AT = (35.56 - 8.56) - In (35.56 -*- 8.56) = 18.96
Heat input, q = steam flow x Ah = 800,000 x (1,038.6 - 88.6) =
760 x 106 Btu/h
From the Heat Exchange Institute Standard, the following is
used:
Reference heat transfer coefficient, C, for 1.0-in. tubes = 263
(Table 9-1)
Inlet temperature correction,/, for 85°F = 1.06 (Fig. 9-8)
Tube material and thickness correction, m, for 304 SS, 22 BWG
= 0.87 (Table 9-2)
Tube characteristics for 1.0-in. 22 BWG
External surface, square feet per foot of length = 0.2618 ft2/ft
(Table 9-3)
Length for 1 ft2 of surface = 3.817 ft/ft2
gpm water flow at 1 ft/s (fps) velocity = 2.182 gpm/fps
(Table 9-3)
With this information, the condenser physical parameters can be
determined.
q = UAAT
(9-4)

Fl =-2cw

(9-6)

Tff

760 x 1Q6
= 281.5 x 10s lb/h
27
281.5 x 105
= 56,300 gpm
500
(llb/h = 1 gpm x [60 min/h/7.48 gal/ft3/0.016 ftVlb]
= 1 gpm x 500 lb/h/gpm)
Number of tubes per pass
Total gpm
gpm per tube per fps x velocity, fps
56,300
2.182 x 7.0
= 3,686 = tubes/pass

<7

UAT

Total number of tubes 2 x 3,686 = 7,372

Table 9-2. Tube Material Correction Factor of Heat Transfer Coefficient
Tube Wall Gauge--BWG
Tube Materials
Admiralty metal
Arsenical copper
Copper iron 194
Aluminum brass
Aluminum bronze

90-10 Cu-Ni
70-30 Cu-Ni
Cold-rolled low carbon steel
Stainless steelsa type 304/316
Titanium"

24

22

20

18

16

14

12

1.06
1.06
1.06
1.03
1.03
0.99
0.93
1.00
0.91
0.91

1.04
1.04
1.04
1.02
1.02
0.97
0.90
0.98
0.87
0.87

1.02
1.02
1.02
1.00
1.00
0.94
0.87
0.95
0.83
0.83

1.00
1.00
1.00
0.97
0.97
0.90
0.82

0.91
0.76
0.76

0.96
0.96
0.96
0.94
0.94
0.85
0.77
0.86
0.70
0.70

0.92
0.92
0.92
0.90
0.90
0.80
0.71
0.80
0.63
0.63

0.87
0.87
0.87
0.84

0.84
0.74
0.64
0.74
0.55
0.55

'The user is specifically cautioned that these noncopper bearing tube materials are more susceptible to biofouling than
tubes with high copper content. This may call for selection of design cleanliness factors that suitably reflect the
probable operating condition the tubes will be subject to in service.
Source: Standards for Steam Surface Condensers, Heat Exchange Institute, Inc. Used with permission.

(9-7)


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258

Power Plant Engineering

Table 9-3. Continued
ODof
Tubing Inches

Ш

1%

Ш

l 5 /8

BWG

Thickness
(inches)

Inside Diameter

(inches)

Surface External
(fWlinear ft)

Length in Feet
for 1 ft2 Surface

Water-gpm at
1 ft/s Velocity

13
14
15

0.095
0.083
0.072

0.935
0.959
0.981

0.2944
0.2944
0.2944

3.397
3.397
3.397

2.140
2.251
2.356

16
17
18
19

0.065
0.058
0.049
0.042

0.995
1.009
1.027
1.041

0.2944
0.2944
0.2944
0.2944

3.397
3.397
3.397
3.397

2.424
2.492
2.582
2.653

20
21
22
23
24

0.035
0.032
0.028
0.025
0.022

1.055
1.061
1.069
1.075
1.081

0.2944
0.2944
0.2944
0.2944
0.2944

3.397

3.397
3.397
3.397
3.397

2.725
2.756
2.797
2.829
2.861

12
13
14
15

0.109
0.095
0.083
0.072

1.032
1.060
1.084
1.106

0.3271
0.3271
0.3271
0.3271

3.057
3.057
3.057
3.057

2.607
2.751
2.877
2.994

16
17
18
19

0.065
0.058
0.049
0.042

1.120
1.134
1.152
1.166

0.3271
0.3271
0.3271
0.3271

3.057
3.057
3.057
3.057

3.071
3.148
3.249
3.328

20
21
22
23
24

0.035
0.032
0.028
0.025
0.022

1.180
1.186
1.194
1.200
1.206

0.3271

0.3271
0.3271
0.3271
0.3271

3.057
3.057
3.057
3.057
3.057

3.409
3.443
3.490
3.525
3.560

12
13
14
15

0.109
0.095
0.083
0.072

1.157
1.185
1.209

1.231

0.3600
0.3600
0.3600
0.3600

2.778
2.778
2.778
2.778

3.272
3.433
3.573
3.705

16
17
18
19

0.065
0.058
0.049
0.042

1.245
1.259
1.277

1.291

0.3600
0.3600
0.3600
0.3600

2.778
2.778
2.778
2.778

3.790
3.876
3.987
4.075

20
21
22
23
24

0.035
0.032
0.028
0.025
0.022

1.305

1.311
1.319
1.325
1.331

0.3600
0.3600
0.3600
0.3600
0.3600

2.778
2.778
2.778
2.778
2.778

4.164
4.202
4.254
2.292
4.331

12
13
14
15
16
17
18

19

0.109
0.095
0.083
0.072
0.065
0.058
0.049
0.042

1.282
1.310
1.334
1.356
1.370
1.384
1.402
1.416

0.3927
0.3927
0.3927
0.3927
0.3927
0.3927
0.3927
0.3927

2.546

2.546
2.546
2.546
2.546
2.546
2.546
2.546

4.018
4.196
4.351
4.496
4.589
4.683
4.806
4.902

20
21
22
23
24

0.035
0.032
0.028
0.025
0.022

1.430

1.436
1.444
1.450
1.456

0.3927
0.3927
0.3927
0.3927
0.3927

2.546
2.546
2.546
2.546
2.546

5.000
5.042
5.099
5.141
5.183

12
13

0.109
0.095

1.407

1.435

0.4254
0.4254

2.351
2.351

4.840
5.035
(Continued)


Steam Cycle Heat Exchangers

Table 9-3. Continued
ODof
Tubing Inches

l3/4

1%

2

BWG

Thickness
(inches)

Inside Diameter
(inches)

Surface External
(ft2/linear ft)

Length in Feet
for 1 ft2 Surface

Water-gpm at
1 ft/s Velocity

14
15
16
17
18
19

0.083
0.072
0.065
0.058
0.049
0.042

1.459
1.481
1.495
1.509

1.527
1.541

0.4254
0.4254
0.4254
0.4254
0.4254
0.4254

2.351
2.351
2.351
2.351
2.351
2.351

5.205
5.363
5.465
5.567
5.701
5.806

20
21
22
23
24

0.035
0.032
0.028
0.025
0.022

1.555
1.561
1.569
1.575
1.581

0.4254
0.4254
0.4254
0.4254
0.4254

2.351
2.351
2.351
2.351
2.351

5.912
5.958
6.019
6.065
6.111

12
13
14
15

0.109
0.095
0.083
0.072

1.532
1.560
1.584
1.606

0.4581
0.4581
0.4581
0.4581

2.183
2.183
2.183
2.183

5.738
5.950
6.135
6.306

16
17
18
19

0.065
0.058
0.049
0.042

1.620
1.634
1.652
1.666

0.4581
0.4581
0.4581
0.4581

2.183
2.183
2.183
2.183

6.417
6.528
6.673
6.786

20
21
22
23
24

0.035
0.032
0.028
0.025
0.022

1.680
1.686
1.694
1.700
1.706

0.4581
0.4581
0.4581
0.4581
0.4581

2.183
2.183
2.183
2.183
2.183

6.901
6.950
7.016
7.066
7.116

12
13
14
15

0.109
0.095
0.083
0.072

1.657
1.685
1.709
1.731

0.4909
0.4909
0.4909
0.4909

2.037
2.037
2.037
2.037

6.713
6.942
7.141
7.326

16
17
18
19

0.065
0.058
0.049
0.042

1.745
1.759
1.777
1.791

0.4909
0.4909
0.4909
0.4909

2.037
2.037
2.037
2.037

7.445
7.565
7.721
7.843

20
21
22
23
24

0.035
0.032
0.028
0.025
0.022

1.805
1.811
1.819
1.825
1.831

0.4909
0.4909
0.4909
0.4909
0.4909

2.037
2.037
2.037
2.037
2.037

7.966
8.019
8.090
8.143
8.197

12
13
14
15

0.109
0.095
0.083
0.072

1.782
1.810
1.834
1.856

0.5236
0.5236
0.5236

0.5236

1.910
1.910
1.910
1.910

7.764
8.010
8.224
8.422

16
17
18
19

0.065
0.058
0.049
0.042

1.870
1.884
1.902
1.916

0.5236
0.5236
0.5236

0.5236

1.910
1.910
1.910
1.910

8.550
8.679
8.845
8.976

20
21
22
23
24

0.035
0.032
0.028
0.025
0.022

1.930
1.936
1.944
1.950
1.956

0.5236
0.5236
0.5236
0.5236
0.5236

1.910
1.910
1.910
1.910
1.910

9.107
9.164
9.240
9.298
9.354

Source: Standards for Steam Surface Condensers, Heat Exchange Institute, Inc. Used with permission.

259


260

Power Plant Engineering

Surface per tube

1.0

лг
,
7/

.80

_ Total surface
No of tubes

(9-8)
Ш

.60
.50

■D
{Li-

.40

CO

69,410
~ 7,372

//,

/j

ГА

СЗ
I- .30

= 9.42 ft2 per tube

n
оu.

Length of tube
= Surface per tube x length for 1 square foot

(9-9)

= 9.42 x 3.817

//'

a: .20
ш
t<
Š
u.
О
ь .10
ш
ш .08

TUBE

O.D. IN
с/о

LiC/l

= 36.0 ft

(Л

It should be noted that the quantity of noncondensible gases
leaking into a condenser affects its performance. Normally, the
effect of this leakage is negligible, but prediction of performance is
less precise when the terminal difference, TTD, is less than 5° F. For

о—1
2
О

'/

7/8

.06
.05

1

1-1/8
1-1/4
1-3/8

1-1/2
1-5/8
1-3/4
1-7/8

.04

(О
ее 03
.02

I

I

.015

3

4

5 6

8 10

15 20

30

WATER VELOCITY, FT/SEC

GAUGE CORRECTION FACTOR FOR FRICTION LOSS
Tube
O.D. In.
0.625

12
BWG

14
BWG

1.38

0.750

1.28
1.25

1.21
1.16

0.875
1.000

18
BWG
1.00

20
BWG

1.00
1.00

0.95

1.06
1.05
1.04

1.00

0.94

22
BWG
0.91
0.93
0.94
0.94

24
BWG
0.89
0.90
0.92
0.93
0.94

1.00

0.96
0.96
0.97

1.04

1.00

0.97

0.96

0.94

1.13

1.08
1.07

1.03

1.00

0.97

0.96

0.94

1.12
1.10

1.06
1.05

1.03
1.02

1.00

0.97

1.00

0.97

0.96
0.96

0.95
0.95

1.125

1.19
1.16

1.250

1.14

1.375
1.500
1.625

1.13
1.11

16
BWG
1.10
1.06

1.09

0.95

1.750

1.10

1.05

1.02

1.00

0.98

0.97

0.96

1.875

1.09

1.02

1.00

1.08

1.02

1.00

0.98
0.98

0.97

2.000

1.05
1.04

0.96
0.96

0.97

Fig. 9-10a. Friction loss for waterflowingin 18 BWG tubes. (From
Standards for Steam Surface Condensers, Eighth Edition, Heat
Exchange Institute, Inc. Used with permission.)

1

2

3

4

5

6

7

8

9

VELOCITY THROUGH TUBES OR NOZZLE, FT/SEC

Fig. 9-9. Waterbox and tube end losses two pass condensers. (From
Standards for Steam Surface Condensers, Eighth Edition, Heat
Exchange Institute, Inc. Used with permission.)

10

this reason, condensers should not be designed with TTD less than
5°F.
To determine circulating water pressure drop, use Figs. 9-9,
9-10a, and 9-10b, with the following:
AF = friction loss per foot for water flowing in 1-in. 18 BWG
tubes at 7 fps (Fig. 9-10a) x correction factor for 22 BWG
tubes (Fig. 9-10a) X correction factor for average water tem­
perature (85 + 27/2 = 98.5° F) (Fig. 9-10b) x tube length x
number of passes + combined inlet and outlet tube end loss
(Fig. 9-9) + water box inlet loss (Fig. 9-9) + water box outlet
loss (Fig. 9-9).
Using Figs. 9-9, 9-10a, and 9-10b, results in the following:
№ = 0.22 x 0.94 x 0.949 x 36.0 x 2 + 1.74 + 0.76 + 0.38 =
17.0 ft of water


Steam Cycle Heat Exchangers

1.16 |

I

1.14

r

L12

I

1

1

1

1

r—

9.2.4

X

a. 1.06
о

\
\
V

u.

\
1.02

>r

\

<->

\

£ 1.00

\

.98

9.2.5

.V-

.96
.94

^Ц4^-

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^ ^

.90 ■

.88 I—J
30 40

1

1

50 60

1
70

1
80

1——i

1

1

90 100 110 120 130

AVERAGE WATER TEMPERATURE - DEGREES F
Fig. 9-10b. Temperature correction for friction loss in tubes. (From
Standards for Steam Surface Condensers, Eighth Edition, Heat
Exchange Institute, Inc. Used with permission.)

Calculations of condensers and their performance, with a
knowledge of other parameters, can be preformed in a simi­
lar manner. Most condenser manufacturers publish an empir­
ical method of calculating condenser size that is based on the
preceding approach.

9.2.3

Corrosion Protection

To protect the waterboxes and tube sheets from corrosion, a
combination of cathodic protection and protective coating is
used. Cathodic protection is provided by fastening sacrificial
anodes to the waterboxes. Protective coatings must be pro­
vided with cathodic protection. These coatings vary with the
circulating water analysis and may range from a coal tar
enamel to an epoxy.

X

1.04

u

Waterboxes normally are made of fabricated steel and are
protected with a coating and cathodic protection.
As different materials normally have different life spans,
an economic analysis should be performed to evaluate the
various materials.

V

1.08

О
H

1

V

1.10

H
^

1

261

Materials of Construction

Condensers are normally fabricated with a welded steel shell
and tube support plates. The tube sheets, tubes, and water
boxes are constructed of material compatible with the circu­
lating water. Material specifications and details of construc­
tion are recommended in the HEI Standards for Steam Sur­
face Condensers.
When fresh water is used as circulating water, a typical
condenser uses admiralty tubes, muntz metal tube sheet, and
cast iron or carbon steel waterboxes. The air-cooling section
of the tube bundle uses 90-10 copper-nickel (CuNi) tubes.
The selection of tube, tube sheet, and water box material
for brackish water or seawater must carefully consider the
type and amount of impurities in the water, including the
amount of sand and dirt being carried by the water.

Tube materials for these applications normally vary from
90-10 CuNi to stainless steel to titanium. Tube sheet mate­
rials are selected to be compatible with the tube material.

Air Extraction

Because the condenser operates at below atmospheric pres­
sure, air leaks into the system. This must be continuously
removed to maintain the design vacuum.
The air removal equipment for small condensers usually
consists of a steam jet air ejector. This ejector uses steam
from the main steam header as the operating medium. A twostage dual unit is normally provided. The steam is condensed
out of the exhaust stream by a small after-condenser, and the
remaining gasses are discharged to the atmosphere.
On larger condensers, mechanical evacuators are used.
These are basically air compressors that raise the pressure of
the air from condenser vacuum to atmospheric. They are
specially designed for this service.
The HEI Standards for Steam Surface Condensers dis­
cusses air removal equipment and its capacity. Table 9-4 lists
recommended capacities from that source (HEI 1984). To use
Table 9-4:
1. Determine the total steam flow of the unit by adding the
main turbine exhaust flow and the auxiliary turbine exhaust
flow entering all shells of the condenser.
2. Determine the total number of MAIN turbine exhaust open­
ings for all shells. Do not include auxiliary turbine exhaust
openings.
3. Divideflowobtained in Step 1 by exhaust opening number
obtained in Step 2. The resultant number is the EFFECTIVE

STEAM FLOW FOR EACH MAIN EXHAUST OPENING.
4. Enter the appropriate section of Table 9-4 based on whether
the unit is a single-shell, twin-shell, or triple-shell con­
denser and locate the flow obtained in Step 3 in the left
vertical column.
5. Determine TOTAL NUMBER OF EXHAUST OPENINGS
for all shells by adding the total number of main turbine
exhaust openings to the total number of auxiliary turbines
exhausting into the condenser. Split auxiliary turbine ex­
haust ducts coming from one auxiliary turbine count as one
auxiliary turbine exhaust.
6. Locate the appropriate column and capacity using the num­
ber obtained in Step 5.


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ТГ 5> й


Steam Cycle Heat Exchangers

7. If independent venting systems are utilized for each shell
of a multishell condenser, the capacity of each system is
determined by dividing the total capacity obtained from the
appropriate table by the number of independent venting
systems.
When starting a turbine, it is desirable to rapidly evacuate
the air from the condenser and reduce its pressure to some
lower value, after that the main ejector takes over. To do this,
a starting air (hogging) ejector is installed in addition to the

regular steam air ejector. It has a much higher capacity but
normally will not produce the low vacuum desired during
normal operation. The mechanical evacuators also have a
method to permit rapid evacuation of the condenser during
startup.
Table 9-5 lists recommended capacities.
9.2.6

Performance

Condensers are very large heat exchangers, with thousands
of tubes handling very large volumes of steam. In addition to
providing the required tube surface area, the manufacturer
must ensure that the design provides for the proper distribu­
tion of the steam throughout the shell so that all of the surface
is effective. Manufacturers employ various tube location
patterns to permit steam access to all the tubes while main­
taining minimum pressure drop. If properly designed, the
condenser will perform as intended, provided that air leak­
age is within limits and that the tubes are maintained within
the cleanliness factor allocated.
9.2.6.1 Leak Detection. Air leakage into the condenser
can be detected by a measuring device provided with the air
removal equipment that measures the quantity of air dis-

Table 9-5.

Hogger Capacities

Total Steam Condensed

(lb/h)
Up to 100,000
100,001-250,000
250,001-500,000
500,001-1,000,000
1,000,001-2,000,000
2,000,001-3,000,000
3,000,001-4,000,000
4,000,001-5,000,000
5,000,001-6,000,000
6,000,001-7,000,000
7,000,001-8,000,000
8,00,0001-9,000,000
9,000,001-10,000,000

SCFMa—Dry Air at 10 in.
Hg abs Design Suction Pressure
50
100
200
350
700
1,050
1,400
1,750
2,100
2,450
2,800
3,150
3,500

"SCFM—14.7 psia at 70° F
Note: In the range of 500,000 lb/h steam condensed and above, the above
table provides evacuation of the air in the condenser and LP turbine from
atmospheric pressure to 10 in. Hg abs in about 30 min if the volume of
condenser and LP turbine is assumed to be 26 ft3/l,000 lb/h of steam
condensed.
Source: Standards for Steam Surface Condensers, Heat Exchange Insti­
tute, Inc. Used with permission.

265

charged. Air leakage into the condenser during plant startup
can come from incomplete welds, connections not plugged,
and other similar construction errors. Other sources that may
contribute at startup or later during operation include leaking
valves or valve packing on lines leading to the condenser,
leaking or broken gauge glasses, weld cracks, and similar
sources.
Large leaks, such as incomplete welds, are usually located
by physical inspection or by flooding the steam space prior to
startup. There are several methods of detecting a small leak­
age. One of the most common is by installing a freon detec­
tor on the air removal equipment outlet. A small stream of
freon directed at the suspected leak point will show up on the
leak detector if a leak actually exists.
A defective condenser tube results in leakage of circulat­
ing water into the shell where it combines with the conden­
sate. This is caused by lower pressure in the shell than inside
the tubes. This leakage, unless very large, does not affect

condenser performance. The circulating water, because of its
dissolved chemicals, has a much higher conductivity than
the condensate. Leaking circulating water mixing with the
condensate raises the conductivity of the condensate and is
detected by a conductivity meter in the condensate outlet.
9.2.6.2 On-Line Cleaning. Lowering of the performance
of a condenser may be a result of scaling of the inside surface
of the tubes, accumulation of sediment in the tubes, or an
accumulation of leaves or other debris on the tube sheet,
shutting off the inlet to the tubes. Several methods are avail­
able for on-line cleaning of the condenser.
To remove an accumulation of material from the face of
the tube sheet, the condenser can be "back flushed" by
reversing the flow of the circulating water through the con­
denser. This is performed by adding additional valves and
piping to the circulating water system. It is relatively simple
for a two-pass condenser, but more complicated for a singlepass condenser. Figure 9-11 shows the pipe and valve ar­
rangement required to permit backwash. To switch flow on­
line requires that the flow be restricted for a very short time.
To accomplish this, the circulating water valves are all mo­
torized and programmed to function in the proper order on
start of the program.
Backwashing removes accumulation of material from the
tube sheet and to a lesser extent may remove some sediment
in the bottom of the tubes. Two proprietary systems have
been developed to help remove sediment settling in the tubes
and may help in removing scaling. One system uses small
plastic shuttles; the other uses sponge balls.
In the shuttle system, a small plastic shuttle is placed in
each tube and a cage is placed on the end of each tube. The

shuttle normally rests in the cage at the outlet of the tube.
When the condenser is back-washed, the reverse flow causes
the shuttle to move down the tube, pushing the sediment
ahead of it, and to a lesser extent, scraping the scale from the
tube. The shuttle is captured in the cage on the inlet of the
tube where it stays until the direction of the water is reversed


266

Power Plant Engineering

TWO-PASS

SINGLE-PASS

NORMAL FLOW
CIRCULATING
WATER IN

CIRCULATING
WATER OUT

CIRCULATING
WATER IN

CIRCULATING
WATER OUT

CIRCULATING

WATER IN

CIRCULATING
WATER OUT

CIRCULATING
WATER IN

CIRCULATING
WATER OUT

BACKWASH

Fig. 9-11. Backwash piping arrangements.

to normal, when it again travels through the tube to be
captured by the outlet cage.
The other system employed to clean the condenser on­
line uses sponge balls of the same diameter as the tube's
inside diameter. These balls are injected into the circulating
water upstream of the condenser, pass through the tubes, and
are collected downstream of the condenser. A pump returns
the balls to the upstream injection point. In passing through
the tubes, the balls push the sediment out of the tube. Balls
can be provided with an abrasive coating that helps remove
scale from the tube.
9.2.6.3 Performance Testing. Performance testing of a
condenser requires an accurate measurement of the various
parameters. These measurements can best be taken if arrange­
ments for them are incorporated into the original design.

ANSI/ASME Performance Test Code, PTC 12.2, Code on
Steam Condensing Apparatus outlines the recommended
method for measuring condenser performance.

Circulating water flow is normally determined by taking
pitot tube readings across the diameter of the circulating
water pipe. These readings should be taken in a straight run
of pipe. As in any flow measurement, the pitot tube should be
located properly from any upstream or downstream pipe
fitting.
Circulating water inlet and outlet temperatures must be
measured accurately. With a condenser designed for a tem­
perature rise of 10° F (5.6° C), a one-degree error in reading
inlet or outlet temperature results in a 10% error in the
calculated temperature rise. The circulating water does not
leave each condenser tube at exactly the same temperature.
To obtain an average outlet temperature, the measurement
point should be located some distance down the pipe, away
from the condenser to allow time for uniform mixing of the
water.
Steam flow is almost impossible to measure directly.
Indirectly, the steam flow can be determined by performing a
complete heat balance on the turbine and feedwater system at


Steam Cycle Heat Exchangers

the same time as the condenser test is being performed. The
other method is to measure condensate flow leaving the
condenser and correct this flow for other flows entering the

condenser, such as heater drains.
Condenser vacuum should be determined from connec­
tions on the turbine exhaust casing or on the condenser shell
close to the connecting point with the turbine. The measuring
instrument should be accurate to 0.01 in. Hg (0.254 mm Hg).
With these readings and the design parameters of the
condenser, calculations can be performed, similar to those
for the original design, to determine expected vacuum, that is
then compared with actual. Alternatively, by using actual
vacuum the cleanliness factor can be calculated to determine
tube status.
9.2.7

Air-Cooled Condensers

The previous paragraphs have discussed water-cooled steam
condensers. The great majority of the condensers installed in
the United States are of this type. The other type of con­
denser is the air-cooled condenser.
The air-cooled condenser has both advantages and dis­
advantages. Among the advantages are that it minimizes
water make-up requirements and eliminates cooling tower
blowdown disposal problems, cooling tower freeze-up,
tower vapor plume, and circulating water pollution restric­
tions. The disadvantages of the air-cooled condenser include
higher condenser operating pressure (lower cycle efficiency),
higher first cost, larger site, higher noise levels, and higher
operating cost.
There are two basic types of air-cooled condenser sys­
tems, as shown in Fig. 9-12. These are the jet condenser with

dry cooling tower arrangement, and the direct air-cooled
condenser system. In the jet condenser with dry cooling
tower, part of the steam condensate is cooled in a dry cooling
tower. It is then returned to the condenser where it is sprayed
into the steam flow, causing the steam to condense and
collect in the bottom of the condenser.
In the direct air-cooled condenser, the steam is piped from
the turbine exhaust direct to air-cooled steam coils. The
steam condenses in the coils, the condensate draining to the
bottom collection tank.
The jet condenser/dry cooling tower has several advan­
tages. The most prominent advantage is that the cooling
tower may be located farther from the plant than the aircooled condenser. The large size of the steam duct and the
need to maintain a low-pressure drop in this duct dictates that
the air-cooled condenser be located as close to the turbine as
possible. It also eliminates the cost of the large steam duct
from turbine to condenser.
The advantages of the air-cooled condenser are that it
eliminates the cost of the jet condenser and the cost of the
circulating water pumps and piping. It normally also pro­
duces slightly better condenser vacuum. In the air-cooled
condenser, there is only one approach temperature involved,
from air to steam. In the jet condenser/dry cooling tower, two

II

TO CONDENSATE
SYSTEM

CONDENSATE

PUMP

DRY
COOLING
TOWER

267

1
T

CIRCULATING
WATER PUMP

JET CONDENSER WITH DRY COOLING TOWER

STEAM FLOW
DIRECT
AIR-COOLED
CONDENSER

TO CONDENSATE SYSTEM-

CONDENSATE
PUMP
DIRECT AIR-COOLED CONDENSER
Fig. 9-12. Air-cooled condensers.

approach temperatures are involved—from air to cooling
water and from cooling water to steam, although the cooling

water to steam approach is very small (less than Г F).
Although the use of air-cooled condensers in this country
has been very limited, the scarcity of water, the need for
zero water discharge, and the better modern designs of this
equipment indicate that more of these types of installations
will be used in the future, particularly in the northeast United
States.


268

Power Plant Engineering

9.3

pressure turbine generator producing 727,072 kW (727.072
MW). This turbine operates with inlet steam at 2,520 psig
(17,375 kPa), 1,000° F (538° C) with 1,000° F (538° C) re­
heat, and with a condenser back pressure of 3.5 in. HgA
(88.9 mm HgA). It uses seven feedwater heaters and has a
heat rate of 8,015 Btu/kWh (8,448 kJ/kWh).
Obviously, with the unit depicted in Fig. 9-5, the higher
initial steam pressure and temperature, along with the steam
reheat, contribute to the lower cycle heat rate, as compared to
the unit shown in Fig. 9-13. However, the use of seven
feedwater heaters instead of four also contributes to this
lower heat rate.
The number of feedwater heaters used in any specific
cycle depends mainly on the size of the turbine, the inlet and
exhaust steam conditions, to some extent the overall plant

cycle, and certainly the economic considerations.
The overall plant cycle may help determine the total
number of feedwater heaters. In a stoker-fired unit, only a
minimum of combustion air heating is normally desired. It

FEEDWATER HEATERS

9.3.1

Regenerative Cycle

In the regenerative cycle, steam is extracted from the steam
turbine at various stages and used to heat the feedwater. This
results in a higher cycle efficiency by increasing the tempera­
ture of the feedwater and by reducing the amount of energy
lost in the condenser.
Heat balances for two representative turbine generators
will illustrate the method of using these heaters in the cycle.
The first, shown in Fig. 9-13, is a heat balance for a small,
low-pressure turbine generator producing 26,860 kW (26.86
MW). This turbine operates with inlet steam at 850 psig
(5,861 kPa), 900° F (482° C), with a condenser back pressure
of 1.5 in. HgA (38.1 kPA Hg). It utilizes four feedwater
heaters and has a heat rate of 9,605 Btu/kWh (10,124 kJ/
kWh).
Figure 9-5 (p. 254) is a heat balance for a large, high-

STEAM FROM
STEAM GENERATOR
LEGEND

H OR h - ENTHALPY, Btu/lb
# - FLOW, Ib/hr
P - PRESSURE, psia
F - TEMPERATURE, °F

13,704*
1307.5H

FEEDWATER
TO STEAM
GENERATOR

213,600#
365.2F
339.1h

HEATER 4

HEATER 3

173.7P

71.7P
304.6F
274.3h

5FTD

u

Fig. 9-13.

91.9F
60.2h

5FTD

10FDC

Gross Heat Rate =

, 205,703*.

OF TD

10FDC

13,704*

231,600*

12,944*

317.0F
287.3И

307.0F
278.7h

191.OF

159.0И

Ф

10FDC

1

28,830*
104.3F
72.3h

231,600(1453.1-339.1)
„ „ „ , ,.,,.
■
■
= 9605 Btu/kWh
26,860

Typical heat balance for a small turbine generator. (From General Electric Co. Used with permission.)


Steam Cycle Heat Exchangers

then can be advantageous to eliminate the highest pressure
feedwater heater, thus reducing the temperature of the feedwater to the steam generator economizer and allowing it to
recover more of the energy from the steam generator exhaust
gasses. In cogeneration plants, the entire cycle must be an­
alyzed carefully to establish the optimum plant cycle and
number of feedwater heaters.

Overall plant economics always affect plant design. Each
feedwater heater added to the plant results initially in higher
costs. These higher costs result from the first cost of the
heater; piping and valving systems for extraction steam,
feedwater, and drains; controls and instrumentation; and
space requirements. In addition, there are costs for mainte­
nance and upkeep as well as additional pumping cost due to
pressure drop across the heater. On the other hand, fuel costs
are lower because of the improved cycle efficiency due to the
additional feedwater heater.
An economic analysis should be performed as part of the
design to determine the optimum number of heaters to be
used in any particular plant. Obviously, fuel cost plays a
major role in this analysis. High fuel cost normally justifies
additional heaters.
9.3.2

Heater Construction

As part of the cycle analysis, consideration must be given to
the type of feedwater heaters to be used, their design parame­
ters, and the method for disposal of the condensed steam
from the heater.
Feedwater heaters are classified as "open" or "closed"
heat exchangers. The open heater directly mixes the extrac­
tion steam with the feedwater to be heated. The extraction
steam is condensed and becomes part of the feedwater leav­
ing the heater.
The closed heater maintains a separation of the extraction
steam and the feedwater to be heated. The extraction steam is

condensed in its chamber of the heater and leaves the heater
separately from the feedwater.
The open heater is designed to deaerate the incoming
condensate.1 This action liberates the dissolved, noncondensible gasses consisting mainly of oxygen, nitrogen, ammonia,
and carbon dioxide, from the condensate. They are present as
a result of leaks and chemical reactions. From this action, the
open heater has received the name of "deaerator." Deaeration of the feedwater is essential to the proper operation of
the plant and is discussed in more detail in Chapter 15, Water
Treatment. In addition to deaeration, this heater provides
proper suction conditions for the boiler feed pump and a
natural break between high-pressure and low-pressure
closed heater design. The disadvantages to open heaters are
that they are large and heavy, and require a pump on the
feedwater outlet to move the feedwater forward in the cycle.
The closed heater is normally referred to as the feedwater
heater. It is physically smaller than the open type of heater

269

and is easier to control. One pump can be used to move the
feedwater through a series of heaters, thereby saving the cost
and complications of a series of pumps.
A plant could be designed using all open heaters, but then
a pump would be required with each open heater.
A plant could also be designed to use all closed heaters.
Some plants in the past have been designed in this manner,
employing the condenser to perform the required deaeration.
However, deaeration was found to be not as complete, and
control of the boiler feed pump was more difficult. Most
modern cycles using high-pressure steam generators are de­

signed with one deaerator and with the remaining heaters of
the closed variety.
9.3.2.1 Open Heater Construction. The open heater (de­
aerator) is constructed of three sections: the heater section,
the vent condenser, and the storage section. Figure 9-14
shows a cross section of a typical deaerator. Although this
configuration is typical, other configurations such as those
shown in Fig. 9-15 are also used.
In the heater section, the incoming condensate comes into
contact with the extraction steam, heating the condensate,
and condensing the steam. Deaeration of the condensate is
based upon Dalton's and Henry's laws. These laws combine
to state that the quantity of a gas that dissolves in a liquid
decreases as the temperature of the liquid rises, and if the
liquid is raised to the boiling point all the dissolved gases
will be liberated. To heat the incoming condensate to the
boiling point and release the dissolved gases, the incoming
water must be broken down into a fine spray or thin sheets of
water and permitted to come into contact with the extraction
steam. Deaerators are made using both principles, some
using a spray system, but most using a stack of small trays
that allow the condensate to cascade over the edges of the
trays, falling from one layer of trays to the next. The steam
flows upward through the falling condensate, heating it and
causing liberation of the dissolved gases, and, of course,
condensing the extraction steam. A tray-type deaerator is
shown in Fig. 9-14, and a spray-type unit in Fig. 9-16.
An opening in the top of the heater section allows the
released gasses and some steam to be vented from the heater
section. To prevent loss of this venting steam, a "vent con­

denser" uses the incoming condensate to condense the steam
being vented with the gases. The vent condenser can be built
as a simple tube and shell heat exchanger mounted on top of
the heater section, or as is done in most modern large units,
can be constructed as a spray section located in the top of the
shell of the heater section. Figure 9-17 shows a typical deaerating heater section using a tray distribution system and
an internal spray vent condenser. Figure 9-15 shows the shell
and tube-type vent condenser as a dashed circle at the top of
the heater section.
The heated and deaerated condensate, along with the
condensed steam, falls to the bottom of the heater section and

'The condensed steam leaving the condenser and passing through the closed heaters to the deaerator is normally called "condensate." The mixture of
"condensate" and extraction steam leaving the deaerator is called "feedwater."


270

Power Plant Engineering

STAINLESS STEEL
WATER DISTRIBUTORS

-SPRAY BOX INSPECTION
- CONDENSATE INLET

SPRAY VALVE ACCESS

-ATMOSPHERIC VENT

STAINLESS STEEL
HIGH PRESSURE
HEATER DRAIN
TRAY ACCESS
MANWAY
STAINLESS STEEL
INTERNAL VENT
CONDENSER
STAINLESS STEEL
VALVE SUPPORT PLATE

EQUALIZER

STAINLESS
STEEL HOOD
GRID SUPPORT

TRAY STACK
SUPPORT GRID
SECTION A-A

OUTLET TO
BOILER FEED
PUMP SUCTION
Fig. 9-14.

DRAIN

Typical tray-type deaerator.

s

4

HORIZONTAL
SINGLE SHELL

^

ь
VERTICAL
SINGLE SHELL

N

fT

-\.

VERTICAL SHELL ON
VERTICAL STORAGE TANK

then passes into the storage section. For small heaters, the
storage section can be located within the same shell as the
heater section. However, for most large heaters, the storage
section is a separate vessel with the heater section mounted
on top of it to permit gravity discharge of the condensate into
the storage section.
The deaerating capability of a unit is measured by the
amount of dissolved oxygen in the feedwater leaving the

storage section. Modern plants normally specify a dissolved
oxygen guaranteed maximum of 0.005 cc/L. In specifying

ffl

ATMOSPHERIC VENT
SPRAY VALVE ACCESS
RELIEF VALVE

CONDENSATE INLET
STAINLESS STEEL SPRAY VALVES
HOT CONDENSATE INLET

HORIZONTAL SHELL
ON HORIZONTAL STORAGE TANK

TRAP RETURNS
VACUUM BREAKER
STEAM INLET

-ID

A<

еэ- -Е э
VERTICAL SHELL ON, HORIZONTAL STORAGE TANK

(1)

EXTERNAL VENT CONDENSER, IF USED.

Fig. 9-15. Typical deaerator configurations. (From Standards for Steam
Surface Condensers, Eighth Edition, Heat Exchange Institute, Inc. Used
with permission.)

OVERFLOW
CONTROLLER

Fig. 9-16.

PUMP SUCTION'

Typical spray-type deaerator.


Steam Cycle Heat Exchangers

ATMOSPHERIC VENT
INLET
STAINLESS STEEL
SPRAY VALVES

well storage capacity, provides for sudden changes in con­
densate or feedwater flow in the system. It also assists in
flash protection of the boiler feed pump, as discussed later.
The deaerator is also used for several other purposes.
Drains from the high-pressure heaters are normally cascaded
into the deaerator. If an air preheater is installed on the steam
generator, it normally uses extraction steam with the conden­
sate returned to the deaerator. As an alternative, the air

preheater may use hot water from the deaerator, with the
cooled water being returned to the deaerator for reheating.
Discharges from high-pressure traps throughout the plant
are also normally piped to the deaerator. Because these
drains come from sources of higher pressure than the deaera­
tor shell design pressure, a safety valve is provided on the
shell to relieve a possible overpressure by these drains. If an
auxiliary steam supply is connected to the heater that can
provide a large capacity of high-pressure steam, the unit
must be adequately protected from this additional source.
The unit must also be protected from a loss of steam supply,
that can create a subatmospheric pressure in the unit. A
"vacuum breaker" valve is installed on the unit for this
protection.
The unit should be constructed in accordance with the
ASME Boiler and Pressure Vessel Code, Code for Unfired
Pressure Vessels, Section VIII.

STAINLESS STEEL INTERNAL
VENT CONDENSER
STAINLESS STEEL
VENT HOOD
DISTRIBUTING PANS
FLASHING RETURNS

STEAM BAFFLE
TRAY ACCESS
MANWAY

STEAM INLET

EQUALIZER

STAINLESS STEEL
DEAERATING TRAYS

OUTLET TO
STORAGE TANK

271

OUTLET SCREEN
AND BAFFLE PLATE

Fig. 9-17. Typical deaerating heater section with internal spray vent
condenser.

the capacity of the unit, the reference is to the quantity of
feedwater leaving the storage section. The unit can be de­
signed to operate at any pressure, although subatmospheric
operation requires additional facilities to provide for re­
moval of the released gasses from the shell. Most units are
designed for positive pressure only, and the shells are spe­
cified for a pressure exceeding the maximum pressure of the
turbine extraction steam. Water storage capacity of the stor­
age section is measured as that volume below the overflow
level. It is normally specified as the equivalent to 10 min of
rated capacity flow, but on very large units this may be
reduced somewhat because of limits in physical size. This
storage capacity, working together with the condenser hot-

9.3.2.2 Closed Heater Construction. The closed heaters
(feedwater heaters) are of the tube and shell design, with the
condensate or feedwater in the tube side and the extraction
steam and resulting condensed steam (heater drains) in the
shell side.
Figure 9-18 shows a cross-section through a straight tube
feedwater heater. In this type of construction, the feedwater
enters the divided channel, passes through the tubes to the
reverse channel, reverses direction, passes through return
tubes to the divided channel, and exits the heater. A partition

STEAM INLET

FLOATING HEAD COVER
FLOATING HEAD TUBE SHELL

/

SHELL

/

UJ

IMPINGEMENT BAFFLE

I

■ TUBE BUNDLE

rh

REVERSE CHANNEL

Fig. 9-18.

yt

DRAIN
OUTLET

Straight tube feedwater heater with floating reverse channel.

WATER LEVEL
CHANNEL
FEEDWATER


272

Power Plant Engineering

plate in the divided channel separates the incoming feedwater from the outgoing. To accommodate the differential
expansion between the tubes and the heater shell, the reverse
channel is "floated" in the heater shell. This type of con­
struction permits access to both ends of the tubes and allows
for mechanical cleaning of scale from the inside of the tubes.
With the advent of better water chemistry and chemical
cleaning, however, the floating head design has given way to

the "U" tube construction shown in Fig. 9-19.
In a typical, straight condensing feedwater heater, such as
that shown in Fig. 9-19, the extraction steam condenses on
the tubes, drains to the bottom of the heater, and exits
through the drain outlet. Since the steam condenses on the
outside of the tubes, the outside tube temperature is the steam
saturation temperature at the prevailing pressure. Therefore,
the feedwater temperature can only approach the saturation
temperature even when the extraction steam entering the
heater is superheated. An economical design is one in which
the unit is designed for a 5°F (2.8° C) terminal difference
between the saturated steam temperature and the outlet feedwater temperatures.
In the straight condensing heater, the drains falling to the
bottom of the shell pass through the steam atmosphere,
resulting in their leaving the heater at approximately satura­
tion temperature.
Thermodynamically, it is often advantageous if the feedwater outlet temperature can be raised above the steam satu­

ration temperature or the drains cooled below the steam
saturation temperature.
Figure 9-20 illustrates a typical two-zone feedwater heater
constructed with both condensing and subcooling zones. In
this design, the tubes containing the inlet and coldest feedwater are enclosed in such a manner that the drains must pass
over these tubes before exiting. This zone now becomes a
water-to-water heat exchanger, and the outlet drain tempera­
ture can approach the inlet feedwater temperature. An eco­
nomical design for this zone is one in that the drain to
feedwater approach temperature is 10° F (5.6° C).
Figure 9-21 illustrates a typical two-zone feedwater heater
constructed with both desuperheating and condensing zones.

In this design, the tubes containing the outlet and hottest
feedwater are enclosed in such a manner that the inlet steam
must first pass across these tubes before entering the main
portion of the shell. In this zone, the temperatures and veloc­
ities are such that the steam is cooled but not condensed. The
zone becomes a gas-to-liquid heat exchanger, and the outlet
feedwater temperature can be raised above the steam satura­
tion temperature. Obviously, this zone is used only when the
extraction steam contains a considerable amount of super­
heat. An economical design for this zone is one in that the
final feedwater temperature is 2 to 3° F (1 to 1.7° C) above the
steam saturation temperature. As approach temperature is
normally referring to approach to the steam saturation tem­
perature, a temperature of 2° F (1° C) above saturation, is
OPTIONAL

FEEDWATER
OUTLET
U-TUBES

PROTECTIVE
SHIELD

/
' CHANNEL
FEEDWATER
INLET

Fig. 9-19. Typical straight-condensing "U" tube feedwater heater. (From Standards for Steam Surface Condensers, Eighth Edition, Heat Exchange
Institute, Inc. Used with permission.)

Steam Cycle Heat Exchangers

273

OPTIONAL

TIE RODS
AND SPACERS

FEEDWATER
OUTLET

U-TUBES

OPTIONAL
' DRAINS SUBCOOLING
ZONE BYPASS

HEATER
SUPPORT

TUBE
SUPPORTS

CHANNEL

DRAINS SUBCOOLING
ZONE ENCLOSURE

Fig. 9-20. Typical two-zone feedwater heater (subcooling and condensing zones). (From Standards for Steam Surface Condensers, Eighth Edition, Heat
Exchange Institute, Inc. Used with permission.)

DESUPERHEATING
ZONE SHROUD

U-TUBES

DRAIN
INLET

DESUPERHEATING
ZONE BAFFLES

FEEDWATER
OUTLET

CHANNEL

Fig. 9-21. Typical two-zone feedwater heater (desuperheating and condensing zones). (From Standards for Steam Surface Condensers, Eighth Edition,
Heat Exchange Institute, Inc. Used with permission.)


274

Power Plant Engineering

DESUPERHEATING
ZONE SHROUD

U-TUBES

FEEDWATER
OUTLET

DESUPERHEATING
ZONE BAFFLES

/
HEATER
SUPPORT

TIE RODS
AND SPACERS

DRAINS SUBCOOLING
ZONE ENCLOSURE

CHANNEL

Fig. 9-22. Typical three-zone feedwater heater (desuperheating, condensing, and subcooling zones). (From Standards for Steam Surface Condensers,
Eighth Edition, Heat Exchange Institute, Inc. Used with permission.)

referred to as a minus two degree (—2° F) (—1° C) approach.
Figure 9-22 illustrates a typical three-zone feedwater heater
constructed with desuperheating, condensing, and subcool­
ing zones.
To perform maintenance on the tube bundle, it must be
removed from the heater shell. Figures 9-19 and 9-20 show
the two types of construction available, flanged and bolted or

all welded. In the all welded design, the shell must be cut to
separate it from the tube bundle. A protective shield protects
the tubes during the cutting process. Most larger heaters are
constructed all welded, as it is less expensive than flanges
and eliminates the problems associated with leakage on a
large-diameter, high-pressure flanged joint.
To expose the tube bundle, the designer can arrange to
"shell pull," in that the shell is pulled off the tube bundle, or
"bundle pull," in that the bundle is pulled out of the shell.
With the present use of all-welded piping systems, it is
necessary to cut the piping connections on the shell or on the
channel before the two parts can be pulled apart. Most de­
signers feel that it is easier for the maintenance personnel to
cut and reweld the thinner walled piping associated with the
shell than the heavy walled piping associated with the feedwater. Therefore, in most installations, the units are designed
for shell pull. When specified, the manufacturer can provide
wheeled shell supports to facilitate pulling of the shell. In
either design, the designer should provide adequate space for
disassembly of the unit.
Figure 9-23 illustrates construction of the channel for a
low-pressure feedwater heater. The channel is that portion of
the heater in that the feedwater enters and leaves. It consists
of the tubesheet, barrel with inlet and outlet connections,
cover, and pass partition to separate the inlet water from the

outlet water and force it to pass through the tubes. This type
of construction is common for low-pressure heaters with
diameters of less than 48 in. (121.9 cm). To eliminate the
gasketing of the pass partition in the main channel cover, an
internal pass partition cover such as that shown in Fig.

9-23(b), is used. The pressure differential across the pass
partition and pass partition cover is only the pressure drop
through the heater tubes, normally only a few pounds per
square inch. The partition and cover are relatively light­
weight and the gasketing is relatively simple.
Figure 9-24 illustrates various types of construction of the
channel for a high-pressure feedwater heater. Because of the
high pressure of the feedwater, a full-size flanged cover, as
shown in Fig. 9-23, is impractical. The design utilizing a
hemispherical head is usually the most efficient but presents
difficulty in fitting nozzles on the hemisphere's periphery.
The larger size low-pressure heaters are often constructed in
the same manner.
The feedwater tubes are "rolled" into the tubesheet. Roll­
ing is a process in that a tool is inserted into the tube and it
expands the tube beyond its elastic limit, forcing it out tight
against the tube sheet. On some high-pressure heaters, the
hole in the tube shell may be serrated to provide further
resistance to pull out of the tube. On some high-pressure
heaters, the tubes are rolled into place and then a small weld
bead is placed around the tube end that welds the tube to the
tube sheet, providing further protection against leaks be­
tween the tube and tube sheet.
Feedwater heaters can be constructed for mounting in the
horizontal or vertical position as illustrated in Fig. 9-25.
While most heaters are installed in the horizontal, some
smaller plants have the heaters mounted vertically. The verti-