N6 moIIIIIIIIIIII llllllllllllII uIuuuoilu · Superheater Outlet Steam Temperature 92 5. Water Inlet Temperature ----- 92 ... FIRST WATER-TUBE BOILER ... YARROW TYPE BOILER -----

A-A&07 454 NAVAL POSTGRADUATE SCHOOL. MONTEREY CA F/0 13/1DESIGN M DOEL FOR THE HEAT TRANSFER IN A SHORT STRAIGHT TUBE SOI-ETC(UJJUN31 81 L W VOLLMER

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cPOONAVAL POSTGRADUATE SCHOOLMonterey, California LEVEL-

DTICELECTE

NOV 19 1981

THESIS DDESIGN MODEL FOR THE HEAT TRANSFER

IN A SHORT STRAIGHT TUBE BOILER

by

Leo W. Vollmer, Jr.

June 1981

Thesis Advisor: P.F. Pucci

Approved for public release; distribution unlimited.

L 03

UnclassifiedsECUmITy CLASSIICATIOW OP Toins PAGE I1 000 Be

REPORT DOCUMENTATION PAGE RAD COMPTIUCNG NSlawn --Ft. GOVT AG3SO NO 4. jitsNT S CATALOG NUMERn

- ~i ?sters esis( Design Model for the Heat Transfer in a Short,, :

Straight Tube Boiler ... Me" i6. PERFORMING 000. REPORNT NUMR

j* ~6. CONTRACT OR GRANT MUMS901(sI

Leo W. (Vollmer, Jr/

S. PerPORullN ORGANIZATION NAME ANO AOOnU 10. PROGRAM ELE"MENT. PRJET. TAlE

AREA A WORK UNIT NUSERleS

Naval Postgraduate SchoolMbnterey, California 93940

I,. CONTROLLING OFFICE NAME ANO AoRSS ,.#-- T

Naval Postgraduate School is UM Junso lr1Monterey, California 93940 iS , /2ME P- ., -

I4. MONITONING AGENY NAME & ACORCSNSIl 01"0001 "M Ce"I"lld *OO) IS. S9CUrrV*CLAwZwr M5W)

Unclassified-"a OECL ASI PICATIONIOOWNGRAOING

SCNEDULE

S. DISTRIBUTION STATEMENT (oe roj RQPeM)


I?. OISTRIBUTION STATEMENT (of to 4a9~ed eMJod g* Wgeek ". Ii 4001 h RepU)

IS. SUPPLEMENTARY NOTES

19. KE9Y WORDS (CUM - ,,On * ee ,jr W ee it 11 old 1.4NM11110 6y 680041 011160)

.aste heat recoveryShort straight tube boilerStagnation boilerBoiling heat transfer

20. ABSTRACT (COM10 40 POY Sie It R0""W da iPAify a? moo&1 MINO

,.A design model for the Short Straight Tube Boiler with a segmented fin-tubearrangement was developed. This model was integrated for a single tube appliedin a computer program written in BASIC for the Hewlett Packard 9845 model Bdesk-top computer.

Water-side Reynolds numbers were varied in order to investigate the per-formance of this boiler. For an overall tube length of 39.4 inches, a Reynoldsnumber of 840 (29.65 lbm/hr) resulted in obtaining 50* superheat for anO "o" 1473 am'noN of, Iol o is o00OOL-6?

,A , Unclassified/ 0SECUNITY CLAIFICATIOI OP TRIS P -- ,I-M-. -awe

" Incasified

- __

"-operating pressure of 600 psig. With these conditions, saturated boilingbegins at 11.45 inches and superheating at 21.88 along the tube length.

Acces sion Fo"r/ACOS~lFor / --

NTIS GRA&I -

DTIC TAB [lUnannounced QJustificat ion

Distribution/Availability Codes

Avail and/or -

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Design Model for the Heat Transferin a Short Straight Tube Boiler

by

Leo W. Vollmer, Jr.Lieutenant, United States NavyB.S.E.E. Purdue University, 1974

Submitted in partial fulfillment of therequirements for the degree of

MASTER OF SCIENCE IN MECHANICAL ENGINEERING

from the

NAVAL POSTGRADUATE SCHOOL

June 1981

Author: & .Z

Approved by:

CO-Avisor/Second Re-a Xe-

Dean of Science and Engineering

.i3

ABSTRACT

A design model for the Short Straight Tube Boiler with a

segmented fin-tube arrangement was developed. This model

was integrated for a single tube and applied in a computer

program written in BASIC for the Hewlett Packard 9845 model

B desk-top computer.

Water-side Reynolds numbers were varied in order to in-

vestigate the performance of this boiler. For an overall

tube length of 39.4 inches, a Reynolds number of 840 (29.65

lbm/hr) resulted in obtaining 50 superheat for an operating

pressure of 600 psig. With these conditions, saturated boil-

ing begins at 11.45 inches and superheating at 21.88 along

the tube length.

4

TABLE OF CONTENTS

I. INTRODUCTION --------------------------------------15

A. BOILER -----------------------------------------1

1. An Overview ------------------------------- is

2. History ----------------------------------- 16

3. Boiler Types-------------------------------21

a. Fire-Tube Boilers----------------------22

b. Water-Tube Boilers---------------------31

4. Water Circulation ------------------------- 50

S. Heat Source--------------------------------527

B. RACER AND THE SST ----------------------------- 54

C. OBJECTIVE -------------------------------------- 8

I. MODEL DESCRIPTION ---------------------------------60

A. AN OVERVIEW -----------------------------------60

B. GEOMETRY --------------------------------------64

1. Segmented Fins-----------------------------65

2. Pin Fins-----------------------------------65

C. GAS-SIDE HEAT TRANSFER/PRESSURE DROP ---------- 70

D. WATER-SIDE HEAT TRANSFER -------------------- 7

E. OVER-ALL HEAT TRANSFER ------------------------ 84,

II.RESULTS AND CONCLUSIONS----------------------------90

A. BACKGROUND ------------------------------------90

B. DESIGN VARIABLES-------------------------------91

1. S.S.T. operating Pressure------------------92

S

2. Frontal Diminsions ---------------------- 92

3. Gas Flow Rate and Temperatures ---------- 92

4. Superheater Outlet Steam Temperature 92

5. Water Inlet Temperature ----------------- 92

6. Fin Geometry ----------------------------- 92

C. THE DATA SET -------------------------------- 93

D. CONCLUSION ---------------------------------- 108

IV. RECOMMENDATIONS FOR FURTHER RESEARCH ------------ 136

APPENDIX A: PROGRAM{ LISTING -------------------------- 138

LIST OF REFERENCES ----------------------------------- 178

INITIAL DISTRIBUTION LIST ------------------------------ 181

6

LIST OF FIGURES

FIGURE 1: BRONZE WATER-COOLED TUBES AND WATER-COOLEDFURNACE CHAMBER WALL ----------------------- 17

FIGURE 2: THOMAS SAVERY'S FIRST BOILER - 1698 -------- 22

FIGURE 3: THOMAS SAVERY'S IMPROVED BOILER - 1702 ----- 23

FIGURE 4: DR. DESAGULIER'S WINDING FLUE BOILER -1718 --------------------------------------- 24

FIGURE 5: HAYSTACK BOILER - 172S --------------------- 25

FIGURE 6: JAMES WATTS WAGON-TYPE BOILER - 1795 ------- 25

FIGURE 7: CYLINDRICAL BOILER - 1790 ------------------ 26

FIGURE 8: OLIVER EVAN'S RETURN-FLUE BOILER - 1800 27

FIGURE 9: RICHARD TREVITHICKS' CORNISH BOILER -------- 28

FIGURE 10: VERTICAL BOILER TYPES ---------------------- 29

FIGURE 11: FIRST WATER-TUBE BOILER - 1766 ------------- 32

FIGURE 12: BOX-SHAPE WATER-TUBE BOILER - 1793 --------- 33

FIGURE 13: JOHN STEVEN'S PORCUPINE BOILER - 1804 ------ 33

FIGURE 14: JOHN COX STEVENS' WATER-TUBE BOILER -

1805 - ---------------------------------------34

FIGURE 15: THE FIRST SECTIONAL WATER-TUBE BOILER -JOSEPH EVE - 1825 -------------------------- 35

FIGURE 16: GOLDWORTHY GURNEY'S CARRIAGE BOILER -

1826 --------------------------------------- 36

FIGURE 17: STEPHEN WILCOX'S BOILER WITH INCLINEDTUBES - 1856 ------------------------------- 37

FIGURE 18: FIRST SECTIONAL WATER-TUBE BOILER WITHINCLINED TUBES -TWBILL - 1865 -------------- 37

FIGURE 19: SIDE VIEW OF HEADER-TYPE BOILER ------------ 40

FIGURE 20: BASIC "A t TYPE BOILER ---------------------- 42

7

Adi-"-

FIGURE 21: YARROW TYPE BOILER ------------------------- 43

FIGURE 22: "A" TYPE BOILER WITH UNCONTROLLED INTEGRALSUPERHEAT ---------------------------------- 43

FIGURE 23: SECTIONAL EXPRESS TYPE BOILER - 1939 ------- 44

FIGURE 24: SEPARATELY FIRED SUPERHEATER BOILER -1930 --------------------------------------- 44

FIGURE 25: "A" TYPE BOILER FOR GENERATINGSUPERHEATED STEAM WITH CONTROLLED,INTEGRAL, INTERDECK SUPERHEAT -------------- 45

FIGURE 26: "M" TYPE BOILER WITH CONTROLLED,

INTEGRAL SUPERHEATER ----------------------- 46

FIGURE 27: MODIFIED "A" OR GUEST BOILER --------------- 47

FIGURE 28: "D" TYPE BOILER ---------------------------- 48

FIGURE 29: NEWER 1200 PSIG SINGLE-FURNACE BOILERFOR POST WORLD WAR II DESTROYERS ----------- 49

FIGURE 30: THE "BENSON" BOILER ------------------------ 51

FIGURE 31: SCHEMATIC DIAGRAM OF CONTROLLED

CIRCULATION BOILER ------------------------- 53

FIGURE 32: SHORT STRAIGHT-TUBE BOILER UNIT ------------ 57

FIGURE 33: SEGMENTED FIN PROFILE ---------------------- 66

FIGURE 34: PIN FIN PROFILE ---------------------------- 67

FIGURE 35: MODEL TUBE LAYOUT -------------------------- 69

FIGURE 36: HEAT TRANSFER CONSTANTS ---------------------- 71

FIGURE 37: SURFACE HEAT FLUX VS. INLET SUBCOOLING ----- 75

FIGURE 38: METHOD OF ROHSENOW ------------------------- 77

FIGURE 39: TWO-PHASE FLOW DEVELOPMENT ----------------- 79

FIGURE 40: SUPPRESSION FACTOR, S ---------------------- 91

FIGURE 41: WATER MASS FLOW VS. REYNOLDS NO. ------------- 94

FIGURE 42a: SUPERHEAT VS. REYNOLDS NO. ------------------ 96

FIGURE 42b: STEAM MASS FLOW VS. SUPERHEAT -------------- 97

8

FIGURE 43: REYNOLDS NO. = 310--------------------------- 98

FIGURE 44; REYNOLDS NO. - 300--------------------------- 99

FIGURE 45: REYNOLDS NO. - 310---------------------------- 100

FIGURE 46: REYNOLDS NO. = 380---------------------------- 101



FIGURE 49: LOCATION OF ZERO% QUALITY-------------------- 104

FIGURE 50" LOCATION OF 100% QUALITY--------------------- 105

FIGURE Sia: REYNOLDS NO. = 840---------------------------- 109

FIGURE Sib: REYNOLDS NO.= 840---------------------------- 110

FIGURE Sic: REYNOLDS NO. = 840---------------------------- 111

FIGURE 52a: REYNOLDS NO. = 840---------------------------- 112

FIGURE 52b: REYNOLDS NO. - 840---------------------------- 113

FIGURE 52c: REYNOLDS NO. = 840---------------------------- 114


FIGURE 53b: REYNOLDS NO. = 840---------------------------- 116



FIGURE S4b: REYNOLDS NO. = 840---------------------------- 119

FIl;URE 54c: REYNOLDS NO. = 840---------------------------- 120

FIGURE 55a: REYNOLDS NO. - 840---------------------------- 121

FIGURE SSb: REYNOLDS NO. = 840---------------------------- 122


FIGURE 55d: REYNOLDS NO. = 840---------------------------- 124

FIGURE 56a: REYNOLDS NO. - 840---------------------------- 125

FIGURE 56b: REYNOLDS NO. - 840---------------------------- 126

9

F IGU E S c : R YNO DS N .= 840 --- ---- --- ---- --- 1 2FIGURE 56d: REYNOLDS NO. - 840-------------------------- 127

FIGURE 57a: REYNOLDS NO. = 840-------------------------- 129

FIGURE S7a: REYNOLDS NO. - 840-------------------------- 129

FIGURE 57b: REYNOLDS NO. = 840-------------------------- 130

FIGURE 58a: REYNOLDS NO. = 840-------------------------- 131

FIGURE 58b: REYNOLDS NO. - 840-------------------------- 132

FIGURE 58b: REYNOLDS NO.= 840-------------------------- 133

FIGURE 59b: REYNOLDS NO.- 840-------------------------- 134

10

LIST OF TABLES

Table I. HEAT BALANCE RESULTS ----------------------- 106

Table 2. TUBE/FIN DIMENSIONS AND AREAS -------------- 107

11

- _ _ _ __- - 7

NOMENCLATURE

English Letter Symbols

A - Area (FT2)

Ab - Frontal Area Blocked by Tubes and Fins (FT2)

Abt - Bare Tube Area (FT2)

Af - Heat Exchanger Frontal Area (FT2)

Afin - Fin Area (FT2)

Aff - Cross-Sectional Area for Fluid Flow (FT2)

Amin - Minimum Cross-Sectional Area for Gas Flow (FT2)

As- Inside Heat Transfer Area (FT2 )

A - Outside Heat Transfer Area

Cmax Maximum Heat Capacity (BTU/Hr-F)

Cmin - Minimum Heat Capacity (BTU/Hr-F)

Cpf - Specific Heat of Water/Steam (BTU/lbm-F)

C - Specific Heat of Gas (BTU/lbm-F)pg

df - Fin Outside Diameter (FT)d - Inside Tube Diameter (FT)

d

d - Outside Tube Diameter (FT)

D - Pin Outside Diameter (FT)p

- Friction Factor

F - Reynolds Number Factor

Gmax - Maximum Gas Flow Rate Per Square Foot (lbm/hr-FT 2)

h - Heat Transfer Coefficient (BTU/Hr-FT 20 F)

Hps - Horizontal Pin Spacing (FT)

12

- Enthalpy (BTU/lbm)

If - Enthalpy of Saturated Water (BTU/lbm)

Ifs - Enthalpy of Vaporization (BTU/lbm)

j - Heat Transfer Colburn j-Factor

k - Thermal Conductivity of Gas (BTU-Hr-FT-F)

X - Fin Height (FT)

Zc - Length of Cut from Fin Tip (FT)

Lp - Average Pin Length (FT)

LT - Tube Length (FT)

N, NTU - Number of Transfer Units

Nf - Number of Fins Per Inch

Ns - Number of Segments in One Fin

OONB - Heat Flux to Initiate Boiling (BTU/Hr-FT2)

P - Pressure (psia)

Q - Heat Transfer Rate (BTU/Hr)

Ri - Heat Exchanger Inside Resistance (Hr-FT20 F/BTU)1

R0 - Heat Exchanger Outside Resistance (Hr-FT20F/BTU)

Rth - Thermal Resistance (Hr-FT2 -F/BTU)

T f - Fin Thickness (FT)

T - Gas Temperature (OF)

T sa - Temperature of Saturated Water (F)

TTO - Outside Tube Wall Temperature (OF)

TTi - Inside Tube Wall Temperature (*F)

Uoi, U0 - Overall Heat Transfer Coefficient (BTU/Hr-FT2 -F)

V - Vertical Pin Spacing (FT)

13

x - Steam Quality

Xe - Equilibrium Quality

XT - True Quality

XTT - Martinelli Parameter

Dimensionless Groups

Nu - Nusselt Number

Pr - Prandtl Number

Re - Reynolds Number

ReTp - Two-Phase Reynolds Number

St - Stanton Number

Greek Letter Symbols

a - Water Surface Tension (lbf/FT)

- Void Fraction

AP - Pressure Change (psia)

AT - Temperature Change (F)

C - Effectiveness

1b - Viscosity at Bulk Temperature (Ibm/FT-Hr)

Uw - Viscosity at Tube Wall Temperature (ibm/FT-Hr)

1 - Fin Efficiency

ns - Surface Efficiency

P - Density (lbm/FT 3)

- Density of Saturated Water (lbm/FT3)

- Density of Saturated Vapor (lbm/FT 3)

14

I. INTRODUCTION

A. BOILER

1. An Overview

A boiler is an apparatus in the form of a closed

vessel constructed for the continuous generation of vapor

under pressure through the transference of heat to the liquid

which is contained in it. The energy for most boilers is pro-

vided by the combustion of the fossil fuels - coal, oil, coke,

or gas. Coal is the major fuel, but most boilers are conver-

tible from one fuel to another. Other fuels such as wood,

waste gases from industrial processes, and solid wastes such

as bagasse (from sugarcane), sawdust, and even trash and gar-

bage serve as energy sources. An increasing number of large

steam plants built since 1960 for generating electricity are

designed to use nuclear fuel, which provides heat from nuclear

fission.

The simple term "boiler" ordinarily refers to the

steam boiler which utilizes available heat energy to convert

water into steam. Because a boiler produces saturated steam

only, it must be distinguished from the steam generator which

may include a superheater, economizer, and air preheater as

integral and necessary parts of the equipment. In their

simplest form, boilers are closed caldrons of water placed

over an open fire. Boilers range in size and function from

i5

the compact units in domestic heating systems to 20-story

complexes that drive giant steam turbines for electrical power

production. They are bu'ilt in many forms with rated steaming

capacities from 40 lbm/hr to 1,000,000 lbm/hr steam production

and operating pressures ranging from 2 psia to the critical

pressure of water, 3208 psia.

2. History

The development of the boiler was initiated and has

been sustained by the development of the steam engine and

steam turbine. Early versions of both, more toys than serious

inventions, are referred to in the writings of Hero of Alex-

andria during the third century B.C. Contrary to popular

belief, the principles used in boiler design are not recent

developments but are of ancient origin. As early as 200 B.C.,

boilers were employed for warming water, heating, and house-

hold services. An example of this type of boiler is shown in

Fig. 1. In this boiler, found in the ruins of Pompeii, the

grate consisted of bronze water-cooled tubes and a water-cooled

furnace chamber wall. The boiler was internally fired, a

feature which did not appear again until the early 18th century.

Water-cooled furnaces and water tubes also were innovations

not incorporated in boiler design until recent.

However, basic principles for boiler construction,

comparable to those set down in 1769 by James Watt for the

steam engine, were not widely known until the 19th century.

Beginning with Hero, the caldron form of boiler was used.

16

HINGED wOB

TWATER

fooooo

WATER

SECTION ToRDX-X

FIGURE 1: BRONZE WATER-COOLED TUBES AND WATER-COOLED FURNACECHAMBER WALL

17

It was not until the latter part of the 17th century that

boilers in a real sense appeared on the scene. During this

period, the greatest potential use for power seemed to be

pumping water out of mines, and a number of engines were in-

vented for this purpose.

These pumping engines, from those first successfully

pioneered by Thomas Savery in 1699, to the engine patented

in 1769 by James Watt, were not true steam engines, but atmos-

pheric engines. Steam was used to fill a cylinder in which a

piston was mounted. When the steam was suddenly condensed by

a spray of water, a partial vacuum was created, and the piston

was forced down by atmospheric pressure. The steam pressure

that raised the piston, with the help of a counterweight on

each cycle of operation, was only slightly above atmospheric

pressure. The steam pressure was so low that without the

counterweight the piston would not have moved. Consequently,

the boilers for these engines were of relatively simple design.

The wagon boiler, designed by James Watt, and in use

from about 1774, was hardly more than a long, riveted, metal

water tank placed over a brick setting or narrow brick oven.

The furnace was at one end and the stack of chimney at the

other. The combustion gases passed along the bottom of the

boiler to the other end where they were channeled back to

heat the sides before entering the chimney at the hearth end.

Although many odd forms were tried between 1780 and 1820, the

wagon boiler was most common until the invention of the

18

internal flue by Oliver Evans in the United States and, inde-

pendently, between 1800-1805 by Richard Trevithick in England.

In true steam engines, where steam does mechanical work,

it is advantageous to impart the highest pressure possible to

the steam before it is released into the pressure cylinder.

Both Evans and Trevithick saw the necessity of high pressure

engines for marine and locomotive applications and were chiefly

responsible for the adoption of higher steam pressures. Built

in two forms, the boiler was a long cylindrical shell with a

large internal cylindrical tube. In one form the flue contained

a furnace at one end and opened to a stack at the other. In

the second a brick setting was used. Hot gases from the fur-

naces were drawn through the setting, heating the lower surface

of the shell, and then returned through the internal flue to

the stack above the furnace.

With the addition of a number of small internal flues,

called fire tubes, operating pressures increased rapidly. The

usual pressure in the time of Watt was only S-7 psi above

atmospheric. Evans, Trevithick, and John Stevens of New York

early in the 19th century used pressures from 50-75 psia.

Jacob Perkins of Massachusetts, in work done between 1823 and

1827, actually obtained 1400 psia. But in general, fire-tube

boilers are limited to relatively low operating pressures

because of the difficulty of constructing a shell strong enough

to withstand the high pressure required for efficient operation.

As pressure demands of steam engines and turbines increased,

19

the fire-tube boiler was supplemented by the water-tube boiler.

During the 1820's, steam pressures on the Mississippi steam-

boats and elsewhere in the United States had been raised to

100 and 150 psia, but explosions were frequent.

It was early recognized that the strength requirements

for increased pressure could be met if the tubes contained the

water rather than the hot gases. Two problems impeded the

development of the water-tube boiler. First, adequate circu-

lation of the water within the tubes had to be maintained in

order to absorb the heat transferred through the tube wall.

Without this circulation the tube-wall temperature and the tube

would be burnt! out. A second difficulty was the explosion

hazard of higit pressure operation where relatively large

amounts of wr-ir are in contact with heating surfaces liable

to fail bicause of overheating. In the event of such failure,

liquid water, at the saturation temperature corresponding to

a high pressure, flashes to steam with explosive force. If,

however, the total amount of water in a boiler flows in a

sufficient number of parallel circuits, as in a sectioned

boiler, a failure in the heating surface will not be serious.

Although these problems were recognized by Joseph Eve and

Goldsworth Gurney in 1825 and 1826, the only successful water-

tube boiler prior to 1870 was that patented by Babcock and

Wilcox in 1867.

The period from 1870 to 1900 was one of improvements.

The multiple drum bent-tube (Stirling) boiler, successfully

20

introduced in 1893 is actually the most recent form of boiler

construction. Subsequent improvements in materials and methods

of construction increased the steaming capacity of the largest

central power plant boilers from 30,000 lbm/hr in 1905 to

1,000,000 lbm/hr by 1935. In 1926, Iving Moultrop was respon-

sible for the increase in operating pressures from 350 to 1200

psia. By 1935 this pressure had gone to 1600 psia. Modern

plants are built to operate above critical temperature at

4500 to 5000 psia. The steam generator superheat temperatures

have likewise increased from a maximum of 550*F in 1905 to

750*F in 1925 and 950*F in 1938. Moder'n plants are designed

to operate at 1050OF to 1100 0 F.

3. Boiler Types

There are basically two types of boilers: fire-tube

and water-tube boilers. All boilers that restrict the passage

of the hot gases through flues or tubes where they transfer

their heat to the surrounding liquid are called fire-tube

boilers. In the water-tube boiler the flue gases pass over

the exterior surface of the tubes while the water being heated

passes within the tube. Another distinguishing feature between

these two types of boilers is that the fire-tube boiler is

supported by the setting sidewalls or the fire box and the

water-tube boiler is usually suspended from the overhead steel

work and columns.

21

- z

a. Fire-Tube Boilers

Figures 2 to 10 show a historical development

of the fire-tube boiler from the first boiler used by Savery

in 1698 to the famous Cornish boiler designed by Trevithick.

FIGURE 2: THOMAS SAVERY'S FIRST BOILER - 1698

22

FIGURE 3: THOMAS SAVERY'S IMPROVED BOILER - 1702

Modern engineering practice limits the use of

fire-tube boilers to about 500 horsepower and a working pres-

sure of 150 psig. Designers of pressure vessels proportion

the shell thickness to the diameter and working pressure to

be encountered. Upon reaching the above limits the shell

thickness requirement becomes such that it is difficult to

fabricate. These limitations are also placed for safety

reasons; for an explosion of one of the larger fire-tube

23

AI

,~ ~ ~ ~ ~ ~ ~ ~ ~~ ~&R , -. .i .i .-- o o' .. -'''-

./

FIGURE 4: DR. DESAGULIER'S WINDING FLUE BOILER - 1718

boilers can be quite disastrous due to the instant release

of large volumes of steam having tremendous expansive power.

The Watt wagon-type boiler, Fig. 6, and the

cylindrical or flueless boiler, Fig. 7, are made up of flat

plates rolled to shape and riveted or welded longitudinally

and also circumferentially when several plate lengths are

required to obtain the desired capacity provided with a

24

__L

SAH'ITy VALVE

FIGURE 5: HAYSTACK BOILER - 1725

SFET YAL OOS _

FIGURE 6: JAMES WATTS WAGON-TYPE BOILER - 1785

25

water-inlet, a condensate return, a steam outlet, a gauge

glass for ascertaining the water level, and a safety valve

for the relief of excess pressure makes the simplest kind of

boiler when mounted over a suitable refractory combustion

box or chamber and entirely insulated against heat loss.

This elementary design is rather awkward and requires much

space since the tank or drum is limited in its pressure with-

standing capabilities. The use of this type of boiler is

restricted to a moderate working pressure and stationary

applications only.

FIGURE 7: CYLINDRICAL BOILER - 1790

26

Cylindrical boilers function by burning fuel in

a refractory combustion chamber constructed under the cylin-

drical drum. The combustion products flow along the underside

of the drum from front to back and then exit through an uptake

to the chimney or stack. The heat in the flue gases must be

transmitted through the metal drum into the liquid in order

to preheat and vaporize it. Since the time element is short,

with the gases passing from front to back only once, then

exiting, the overall efficiency of this arrangement is low.

The boiler is a notorious slow steamer; however, once the

steaming action has started, it is a steady steamer. These

boilers are limited to 'small capacities, usually to 350 boiler

horsepower, because of their inherent tendency to explode.

FIGURE 8: OLIVER EVAN'S RETURN-FLUE BOILER - 1800

27

.. . . .

- *T

FIGURE 9: RICHARD TREVITHICKS CORNISH BOILER

The flue boiler is an improvement over the cylin-

drical boiler, see Figures 8 and 9, in that one or more inner

flues are incorporated for the purpose of conducting the flue

gases through the center of the shell, cylinder or drum.

There are many variations of design with some using these flues

as actual combustion chambers. The use of flues increases

the ratio of heating surface to the volume of water to be

evaporated; therefore, the flue boiler generates steam more

quickly than the ordinary cylindrical boiler. When the flue

boiler is arranged so that combustion takes place within the

flue greater savings in weight and space can be effected

rendering the design highly applicable to locomotion and

propulsion applications because of its portability.

28

The vertical upright fire-tube boiler, Figure 10,

is avialibe in ranges from S to 300 horsepower, and consists

of a cylindrical steel shell completely filled with tubes

vertically mounted over a fire box. The flue gases pass in

single flow directly through the tubes and into the stack at

the top. While the vertical boiler is not highly efficient

it has other virtues such as portability, a small floor re-

quirement, a large heat surface in a small volumetric space,

and an extreme ruggedness which has won it the name of "The

Donkey Boiler". The popular fire engine "pumper" of several

I I

! ] I

1N41

- -

WRM 3U QR A Wff N DASO FM930Z WfH B4A*0 FOX=

FIGURE 10: VERTICAL BOILER TYPES

29

decades ago used this type of boiler to generate high pressure

steam in order to motivate a reciprocating piston plunger

pump. It also found wide use in the inaccessible, mountainous

terrain of the Pacific Northwest during the logging boom of

the late 1880's.

Fire-tube boilers become limited as capacity and

pressure requirements increase. Larger shell diameters re-

quire thicker plates to withstand the temperature and pres-

sure stresses. Temperature differentials in the boiler create

high stresses. These stresses combined with the effects of

precipitates and other deposits have caused many boiler

explosions (Ref. 2]. Other disadvantages include long warm-up

time requiring up to 15 hours to produce steam for propulsion,

unequal expansion of metal parts at an increased firing rate,

faster expanding fire tubes tend to pull loose from the boiler

shell, heavy weight, up to 15 lbs of boiler and water to pro-

duce 1 lb of saturated steam per hour and inefficiency

requiring large amounts of fuel to produce a relatively small

amount of steam.

Fire-tube boilers tend to be dangerous at high

steam pressures. This is partly due to inherent structural

problems and partly due to the relatively large quantity of

water and steam contained in the boiler. In the event of a

rupture in the boiler, the large amount of water and steam

releases great destructive energy. With this ever-increasing

demand for more and more steam at higher and higher pressure

30

. . . .. . , . . . . . , i, ,< I

the water-tube boiler was developed. When steam and water

at elevated pressure are confined within a tube of small or

moderate diameter the thickness requirement is reasonable and

practical.

b. Water-Tube Boilers

The water-tube boiler is composed of drums and

tubes which are external to the drums and serve as intercon-

nections for them. The drums store water and steam and con-

tain no tubular heating surface, thereby allowing for these

vessels to be much smaller in diameter than a fire-tube

boiler shell and resulting in resisting higher pressures.

This boiler may be a straight or bent-tube type, and in both

cases, the tubes possess the entire heating surface.

The principle of using water tubes is an old one

as shown by the use of the Pompeiian household boiler. Modern

development of the horizontal, straight water-tube boiler

began in 1766 by William Blakely who patented an improvement

in the Savery engine which included a novel steam generator.

The arrangement of heat transfer surfaces shown in Figure 11,

probably was the first step toward the development of

water-tube boilers, which has resulted in the modern high

pressure water-tube boiler.

Three men, James Rumsey, John Fitch, and James

Marlow, worked independently upon a box-shaped water-tube

boiler with horizontal tubes, as shown in Fig. 12. The dis-

advantages of this design were the weak box-shaped pressure

31

FIGURE 11: FIRST WATER-TUBE BOILER 1766

vessels and the restricted water circulation in the horizon-

tal tubes. This was not critical since steam pressures were

between 3 to 7 psig.

In 1804, John Stevens designed a boiler for a

steamboat operating on the Hudson River. This boiler, Fig.

13, had slightly inclined tubes connected at one end with a

32

reservoir. The Steven's Porcupine boiler, as it is known,

marked a definite advance in design of high pressure steam-

generating apparatus, as its working pressure was 50 psig.

Again, lack of proper water circulation in the tubes was a

________ * ooooooooI_________ -ooooooooo0(

_________ 000000000000oI_________ -000000000000I.

- ooooooo0Cooo________ - oooooooooo

_________ - o000ooo0oo1_________ 00oo1o00oo000I__________ -ooo0oo0oo 11__________ - oooooooo0oo

FIGURE 12 BOX-SHAP WATER-TUE BOILER-19

FIGURE 13: JOX-HNATEENS'APRCTUPIE BOILER -1804

serious defect. Stevens' son, John Cox Stevens, designed a

water-tube boiler in 1805. This boiler, shown in Fig. 14,

consisted of 20 vertical tubes arranged in a circle, connecting

a water space at the bottom and a steam space at the top.

The steam and water chambers were annular spaces of small

cross section and small volume.

4i

FIGURE 14: JOHN COX STEVENS' WATER-TUBE BOILER 1805

Figures 15, 16, 1' and 18 show the development of

the water-tube boiler in the next sixty years. The first

sectional water-tube boiler with well defined circulation was

built by Joseph Eve in 1825. Sectional boilers have the

34

R4

FIGURE 15: THE FIRST SECTIONAL WATER-TUBE BOILER - JOSEPHEVE - 1825

water-steam spaces divided into small sections, none of which

is subject to disastrous explosion. Basically all water-tube

boilers are sectional in character. Figure 15 shows the

sectional composed of small tubes, practically vertical but

with a slight double curve. The tubes were fixed in larger

horizontal tubes called headers, which, in turn, were connected

to a steam space above and a water space below. The steam and

water spaces were joined by outside pipes (down-comers) to

35

...... . t

secure a circulation of water up through the sections and down

through the external pipes. In 1826, Goldworthy Gurney built

a number of boilers for use on his steam carriages, one of

which is shown below. This boiler consisted of a number of

FIGURE 16: GOLDWORTHY GURNEY'S CARRIAGE BOILER - 1826

small "U" tubes, laid sidewise, with the ends connected to

larger horizontal pipes. These, in turn, were connected by

vertical pipes to permit circulation and also were connected

to vertical cylinders serving as steam and water reservoirs.

In 1856, Stephen Wilcox first used inclined water-tubes to

connect water spaces at the front and rear with a steam space

above, Fig. 17. The first to use inclined tubes in a sectional

form was George Twbill in 1865, Fig. 18. He used wrought iron

36

r

FIGURE 17'. STEPHEN WILCOX'S BOILER WITH INCLINED TUBES -

FIGURE 18: FIRST SECTIONAL WATER-TUBE BOILER WITH INCLINED

TUBES - TWILL - 1865

37

tubes connected at the front and rear to standpipes. These

standpipes carried the steam to horizontal cross-drums at the

top, the entrained water being separated and led to the rear

standpipe. In the following years, development of water-tube

boilers continued with great vigor. Steam drums soon were

substituted for the nest of cast iron tubes and pipes origi-

nally used for steam and water storage. Steel replaced cast

iron for the generating tubes. Large tubes were replaced by

small tubes to increase heat transfer areas. The transition

to water-tube boiler design didn't arrive until the end of the

19th century when the triple expansion reciprocating steam

engine was introduced to the navies of the world.

The head type boiler was the first form of water-

tube boiler to displace the fire-tube type for use in steam

vessels. A header type boiler was installed in the yacht

Reverie in 1889 with a designed pressure of 225 psig. The

success of this and other similar installations led to a

decision by the Admiralty to use this boiler in the British

Navy. In 1896, the U. S. Navy installed this type in three

vessels, Marietta, Annapolis, and Chicago.

Although the Reverie design was an outstanding

success and proved to be far more satisfactory in service than

other boilers previously used by the Navy, combustion was not

entirely satisfactory. The shape of the furnace made it

necessary to slope the grates down toward the rear to gain

furnace volume. At high ratings, this furnace was not suitable

38

for good combustion. Also, the fittings and water gages were

not readily accessible from the firing aisle because of the

steam drum location.

To overcome these difficulties, the design was

modified to permit firing from the opposite end. Combustion

conditions improved since the furnace was enlarged in the

direction in which combustion took place. The grates were

relocated and the accessibility of the tubes for cleaning from

the fireroom floor was improved greatly. The steam drum was

now at the firing end of the boiler, where the water gage could

be seen easily and all fittings were more accessible. These

modifications, seemingly very simple, revolutionized marine

water-tube boiler design. The many improvements throughout

the years have resulted in the sectional header type boiler,

shown in Fig. 19, as we know it today.

The necessity for more steam brought on experi-

mentation with forced draft to accelerate combustion of marine

boilers. Although Robert Stevens tried in 1827 to supply air

by forced draft, it was not until 1880 before a successful

forced draft system was designed by James Howden. Interest-

ingly, his design also included means by which waste heat in

flue gases was used to heat the incoming air for combustion.

The increased steam pressures and rapidity of steam generation

led to serious trouble arising from overheating of tube ends

and tube-sheets, causing many breakdowns and explosions from

endeavors to work cylindrical fire-tube boilers under forced

39

-- - - - -- - - - - - - - --i- - - - ~" .. .... ' .. .. ... .... .. . II : " - ' -

WATER WALL CCUAO .......... ECONOMIZER TUBESRISER TUBE TU....

UPAE STEAM DRUM

~ / OOWP4TAKE 141PPLEI ~ / GENERATING TUBES

* * OOWNTAKE HEADER

AIRINTAE.~0:0./ / SUPERHEATER TUBES

WATER SCREEN TUBES

JNCTION HEADER

AIR SPACE

- OWN TAKE HEADER

WATER WALL UPTAKE HEADER WTRWLTBSARPC

FIGURE 19: SIDE VIEW OF HEADER-TYPE BOILER

40

draft. The proponents of this boiler type, satisfied with

the partial solution of installing tube inserts in the tube

ends, argued that the generated pressures were sufficient for

their needs.

Then in 1894, Charles A. Parsons developed and

proved the superiority of the steam turbine for marine propul-

sion. The advantages of the turbine were immediately apparent

and within a relatively short time (1894-1914) almost all

important ships of the world navies were driven by turbines.

Higher pressures, greater steam capacity and utilization of

propulsion turbines concurrently accelerated the development

of the water-tube boiler. A major contribution to higher

pressure and greater steam capacity was the evolution of

treating boiler water to prevent scale formation and boiler

corrosion.

From the days of the caldron boiler until well

into the turn of the 20th century there was very little im-

provement in steam generation except in the larger power-

generating stations operated by the greater public utilities.

At the end of the 19th century the new fast torpedo boats and

destroyers, demanding boilers of light and compact design,

brought forth the drum-type express boiler. The name "express"

is used in its sense "dispatched with special speed" and is

applied to this type boiler because of the "speed" with which

steam can be raised and with which this boiler can answer

steam demand changes.

41

- - - l'.r,

UPTAKCES

t _STEAM DRUMA

DOWNCOMERSGENERATINGAND CIRCU-

LATING TUBE BANK_/ , \_ FURNACE WAE DR oUMWTER DRUM

FIGURE 20: BASIC "A" TYPE BOILER

The development of drum-type boilers as applicable

to naval vessels to present time is shown in Figs. 20-29.

Figure 20 illustrates the basic design of A-type boilers (so

named because of the shape of the generating elements resem-

bling the letter "A") which evolved from such early designs as

the Yarrow express boiler shown in Fig. 21. The A-type boiler

was originally designed as a coal-fired boiler (converted from

coal to oil during the decade between 1905-1915) and was widely

used for many years with only minor changes and additions.

42

FIGURE 21: YARROW TYPE BOILER

/ PTAM

PROTECTION STEAM CCU'

IA1U DINM

FIGURE 22: "All TYPE BOILER WITH UNCONTROLLED INTERGRALSUPERHEAT

43

HEAI CIRCUIATDNI PIPES

REAR

~** SUFEMOETER

VERTI

SUPWIEATERBUNR DM

WATER SCREEN OEIG

FIGURE 23: SECTIONAL EXPRESS TYPE BOILER -1939

,/000 0 ol IN

00nmM aa 10

m. 0 a a

ma f

WL PN A LL OV41* 0 A~SP I* IWA1U T IV.PAM " M W ALYW

FIGUREHWM 24: SEPARATEL FIE UERETRBILR 13

44rm yuI"v

World War II created a huge demand for powerful,

packaged, compact steam generators simple to operate and easy

to maintain for use by the armed forces ashore and afloat.

Under this stimulus, designs were developed that are capable

of delivering two to three times the output of earlier boilers

containing the same heating surface. In some cases this can

be accomplished in one-quarter the previous volumetric space

requirement at great savings in weight and with efficiencies

running on the average of 80% [Ref. 3]. This was extremely

important to offset the increasingly congested space conditions

aboard destroyers.

...... .. O 0

TUBE GAS-TIGHT

SLIPrM0lEATED

SUPEREAT STEAMS SA T SID ETEMPERATURECONTROL FURNACE S PRESSURE

DIVISION WALL HEAOER FEOEM SION WAL HEADER CONTROL FUM E

FIGURE 25: "A" TYPE BOILER FOR GENERATING SUPERHEATED STEAMWITH CONTROLLED, INTEGRAL, INTERDECK SUPERHEAT

4S

* - ... ... .......- ,-.--------....----,., -

The general design most widely used for combat

vessels built during WW II is shown in Fig. 26. The "M"-type

boiler, as this design is commonly called because of the shape

of its generating elements, was a two furnace, single-uptake,

controlled superheat type boiler generally designed to operate

at 600 psig-8500 F. The modified "A" or Guest-type, Fig. 27,

had a separately fired superheater of the radiant-convection

type temperature control. However, this type superheater did

not prove well adapted to naval requirements for rapid load

changes and steaming at low rates for considerab,. :riod of

@000000

/ v/

um %\.! SATUATO SEIX/

FIGURE 26: "M" TYPE BOILER WITH CONTROLLED, INTEGRALSUPERHEATER

46

\oooo0\ man

. - ..... 0000%'

31ATSItEATE31Ult WAR

0 00

FIGURE 27: MODIFIED. "A ORGETBIE

il Ku a

t " W IE / "E -- -

FIGURE 27: MODIFIED "All OR GUEST BOTLER

time. Another boiler used mainly in WW II destroyer escorts

is the single furnace boiler shown in Fig. 28. The "D" boiler

is the progenitor of boilers used in post-World War II ship

construction. Figure 29 shows the 1200 psig boiler which has

become the major naval fossil-fired steam generator.

47

UPTAKES

TEI STAV DRUM

\S'JPERNEATER

T\PROTECTION

'*'MAIN STEAM

WATER WALL HIEADER WATER 611M?

FIGURE 28: I'D" TYPE BOILER

48

PERTMSgM? DATA

FIUR 2: EWR 20 PIGSIGL-F RAE BOI$UL ...OR 2 PSTWORLD WARA OUPU~~ DESTROYERSMl

To49TNG$*AL _.$ 0P

-~~~UMC V__________________ 4nCU P

4. Water Circulation

The majority of the previously discussed boilers have

depended upon natural convection for the water-side circula-

tion. That is, the difference in density between the water

near the combustion chamber being heated and the cooler water

in more remote parts of the boiler causes a natural flow away

from the combustion chamber which in many cases is adequate.

As boiler pressure increases, however, there is less difference

between the densities of water and steam. At pressures over

1000 psig the density of steam differs so little from the

density of water that natural circulation is harder to achieve

than it is at lower pressures [Ref. 61 . At high pressures

positive circulation boilers have a distinct advantage because

their circulation is controlled by pumps and is independent of

differences in density. The utilization of pumps to circulate

and agitate boiler water increases heat transference, thereby

increasing the output of a boiler without increasing the heat-

ing surface.

There are two main kinds of positive circulation

boilers; one type is known as a controlled circulation or

forced recirculation boiler and the other is known as a once-

through or forced-flow boiler. In both types of boilers

external pumps are used to force the water through the boiler

circuits; the essential difference between the two lies in

the amount of water supplied to the boiler.

so

In a once-through forced circulation boiler all (or

very nearly all) of the feedwater pumped to the boiler is con-

verted to steam the first time through without any recirculation.

In its course along the full length of the tube, the water is

heated, evaporated and superheated as additional enthalpy is

applied. The steam generation rate is numerically equal to

the water content supplied. Conventionally this type of boiler

does not require a steam and water drum. Because no excess

water is forced through the boiler, pumping other than that

provided by feedwater pumps is not required. The very small

amount of water that is not converted to steam in the generating

circuit is separated from the steam and discharged to the feed

pump suction. The once-through type of positive circulation

boilers include the Benson and Besler boilers.

FIGURE 30: THE "BENSON" BOILER

51

The Benson Boiler, Fig. 30, is characterized by the

complete absence of steam separating drums. The unit responds

sensitively to combustion rate changes and is therefore well

adapted for use with a special turbine design in which variable

pressure is used tc accomodate a variable load. This boiler

was first built in England (1923) by Mark Benson, a Czechos-

lovakian, but was developed by Siemens, Inc., in Germany. It

has been built with capacities as large as 220,000 ibm/hr at

3200 psig.

Figure 31 shows the boiler circuit of a controlled

circulation (or forced recirculation) boiler. In this boiler

the water in the tubes is not evaporated to complete dryness

but only to the point at which dissolved salts and solids are

retained in solution. The mixture of water and steam passes

to the steam-and-water drum where the steam is separated. The

steam is passed through a separator where excess water is re-

moved, then passed to a superheater. The separated water,

along with feedwater, is returned to (or recirculated through)

the heating circuit through the downcomers. The quantity of

water passing through the boiler (circulation rate) is from

3 to 20 times the amount evaporated. This ratio requires

constant speed recirculating pumps in addition ta the boiler

feedwater pumps.

5. Heat Source

The heat may be derived from (1) the combustion of

fuel (solid, liquid, or gaseous); (2) the hot waste gases of

52

P

_ [JKZ Z

53

other chemical reactions; (3) the application of electrical

energy; or (4) the utilization of nuclear energy.

Items 1, 3, and 4 are the most prevelant methods of

supplying a heat source, but the increasing need for energy

conservation has necessitated increased utilization of waste

heat recovery. Wherever a process waste product or gas is

continuously discharged at a temperature of 1,000'F or higher,

heat recovery should be considered. In addition to producing

useful steam, the lowering of the flue gas temperature reduces

maintenance of flues, fans, and stacks. One proposal for

"by-product heat recovery" utilizes the hot exhaust gases from

a shipboard gas-turbine propulsion plant as the heat source

for a conventional Rankine cycle. This provides the means by

which useful work can be extracted from a once wasted energy

source.

B. RACER AND THE SST

The realization of finite petroleum resources has prompted

researchers to accelerate the development of alternative fuels

and energy sources, the employment energy conservation methods

for short term relief, and the investigation of long term con-

servation techniques. The Navy, having made a major commitment

to gas turbine propulsion systems, is currently investigating

a means of fuel conservation without a reduction in steaming

1Waste heat is a misnomer; a more appropriate term is "by-pro-duct heat."

54

time. The RAnkine Cycle Energy Recovery (RACER) system is being

developed to recover some of the heat contained in the gas tur-

bine exhaust gases.

The DD-963, FF6-7, and the DDG-47 class of ships are powered

by the General Electric LM 2500 Marine gas turbine engine. The

RACER system, an unfired waste heat recovery system, will func-

tion as a bottoming cycle for the LM 2500. Program requirements

specify the development of a high efficiency cruise engine that

utilizes hot exhaust gases to heat the working fluid (water)

and convert it to useful energy through a steam turbine. It

is intended that the design be compatible with the basic LM

2500 Marine gas turbine providing one shafts worth of power to

the DD-963 and FFG-7, U.S. Navy surface cambatants. Further,

specifications require provisions that will permit satisfactory

inspection and cleaning of watersides/steamsides. It is im-

portant that internal surfaces can be properly observed for

corrosion, scaling, defects, repairs, etc. Cleaning methods

and techniques should be simple, easy to use, and practical for

use with the proposed design.

The boiler, placed in the exhaust ducting, would by necessity

have to be compact and lightweight, have a high steam generation

rate and excellent response characteristics as well as simplicity

in design, installation, and maintenance. LCDR Combs [Ref. 71,

in his thesis conducted a feasibility study for propulsion

assistance on waste heat recovery as applied to the Spruance

class destroyers. Results of the study indicate that this

55

* -- ---- --- -.-- ---- - t

proposal provides for substantial fuel savings based upon

operational considerations. The proposed Rankine cycle in-

cluded a once-through (O.T.) boiler having an optimal operating

pressure of 600-800 psig.

For this application, the O.T. boiler is particularly

attractive due to its small size. A special class of once-

through boilers, which provides a possible design alternative,

is the Short Straight Tube (S.S.T.) boiler. Also known as

the "stagnation" (non-flow) boiler, it has excellent perfor-

mance as a steam generation unit. This boiler has the same

components as the O.T. boiler; however, in contrast, its

construction consists of short (1-2 meters) and straight tubes

with toroidal headers, Fig. 32. Due to boiler water level

control and water chemistry problems, the S.S.T. boiler all

but disappeared since the early 1900's. With the advancements

made in water chemistry procedures and with the development

of automatic boiler control systems, these problems should be

alleviated, allowing present day application.

When compared to the normal forced or natural convection

boilers, the once-through boiler is more desirable for waste

heat recovery on the Spruance class destroyer. The O.T.

boiler is lighter and safer, since no steam drum is required

and the total mass of water is substantially less. However,

a reveiw of its construction reveals, Fig. 33, a complicated

arrangement of headers and long "serpentine" tubes. This

leads one to question its applicability, especially in terms

56

STE.'AM OUT

TOPHEADER

TUBES

I

BOTTOM LHEADER

f JWATER IN

FIGURE 32: SHORT STRAIGHT-TUBE BOILER UNIT

S7

of simplicity. In contrast, the S.S.T. boiler has a charac-

teristic which makes it very attractive for this purpose. As

it is the simplest of boiler designs, it lends itself to lower

fabrication cost and simplified maintenance.

C. OBJECTIVE

The S.S.T. boiler is undergoing a resurgence in Japan for

industrial usage due to its performance and construction

(Ref. 81. These industrial applications are for steam supplied

up to 150 psig. In order to employ such a design at higher

pressures, it is necessary to obtain some essential information.

In general, this need for information arises from three

sources. First, the designerreauires detailed information in

order to provide for an optimal design. This entails a de-

tailed quantitative study, for example, of heat transfer co-

efficients and two phase pressure drop. Secondly, operational

conditions must be evaluated to ascertain the optimal steady

state and transient operating conditions. Inclusive is

casualty control data to diagnose faults due to the departure

from the optimal conditions. Finally, and most paramount, is

the requirement for the precise maximum safe operational limit.

For these reasons, experimental investigation to ascertain

design, operation and safety criteria is necessary. Firm

establishment of the characteristics unique to the Short-Straight

Tube Boiler is required to provide the means by which performance

evaluation and comparison with other evaporators can be made.

However, a lack of design and operational experience doesn't

58

allow the engineer to anticipate S.S.T. characteristics a

priori. Therefore, the logical beginning to experimentation

is having a model from which predictions can be made.

The objective of this thesis therefore, is the development

of a suitable model and computer program to provide the means

for analytic study of the S.S.T. principle.

59

II. MODEL DESCRIPTION

A. AN OVERVIEW

The Short Straight Tube Boiler is modeled as a cross-flow

heat exchanger with one fluid mixed; the other unmixed. The

model is applied in a computer program for use in heat exchang-

er analysis for a phase change, or boiling situation. The

same program could be easily adapted for general cross-flow

heat exchanger performance study.

The two basic tools required to perform any heat transfer

analysis are the first law of thermodynamics and the applicable

rate equation.

A first law consideration of each fluid with mass flow

rate i, for an incremental area AA of the exchanger, yields

AQ = (MCpAT)gas = (MCpAT)water (la)

Another expression for AQ is achieved using the rate equation

and writing an expression for the conditions at AA; the

resulting equation is

AQ = UAA (Tg-Tw) (lb)

where Tg is the gas temperature in OF, Tw is the water tempera-

ture in oF, and U is the overall heat transfer coefficient in

BTU/(Hr FT 2 OF). For this analysis, U is assumed to remain

constant throughout the incremental heat exchanger. Equations

la and lb can also be written as

AQ = (ACP) min ATmax - UAA(Tg-Tw) (2)

60

which can be written as

ATmax UAA

(Tg-T = - (c~m (3)(;Cp )min

For the incremental heat exchanger area, AA, the number of

transfer units, ANTU is defined by the following equation:

NTU - AT max UAA (4)T g- Tw) (mc p)min

The number of transfer units is a measure of a heat ex-

changer's ability to transfer heat. For a number of heat

exchangers in series, the number of transfer units for the

system can be written as follows:

let Cmin = (;cp)min

NTU -UiAAi + U2VAz .. U n A j

Cmin, C min2 .. C minj

or N U. AAU (NTU = U , U5)j = Cminj

where N represents the total number of incremental heat

exchangers; AAj's are the incremental heat transfer areas,

and U.'s are the overall heat transfer coefficients for theh

jth elemental heat exchanger [Ref. 9].

For a particular type of heat exchanger, a mathematical

relationship can be calculated between NTU, the number of

transfer units; e, the effectiveness of heat transfer; and

C, the heat capacity rate ratio. If two of these quantities

61

-j -A.- 4-.

are known, this mathematical relationship can be used to cal-

culate the unknown. Reference [10] lists this relationship

for cross-flow as follows:

C min N=NTU UACmax min

Cmax mixed, Cmin unmixed

= (1/C) I-EXP i-C(1-e'N) (6a)

Cmax unmixed, Cmin mixed

= -EXP [I(1/C 11-EXP(-NC)(J (6b)

If the hotter fluid is a condensing vapor or the cooler fluid

is a boiling liquid, the heat capacity ratio C = 0. For this

case, Kays and London [Ref. 11] show that equation 6 reduce

to

e = I-EXP(-NTU) (7)

The proposed model utilizes this procedure to determine the

overall heat transfer coefficient for the straight-vertical-

tube cross-flow heat exchanger used as a boiler.

In the development of the model, the following assumptions

were made:

1) A tube could be modeled as a group of N heat exchangers

connected in series. The value of N is determined from fin

geometry as the product of tube length times the number of

fins per inch.

62

2) Each "elemental" heat exchanger has an overall heat

transfer coefficient which can be considered constant over

the segment area. With this assumption, equation 5 can now

be applied to find the tube NTU.

3) The "elemental" applied heat flux (Q s/A si) is con-

stant for each individual tube segment with Qs given byQs = E(;c p) min (T -T W) (8)

where Qs is the segment heat transfer rate in BTU/hr, and

e is the segment effectiveness.

4) The water side pressure drop is negligible.

The SST model is divided into three principal sections:

heating, boiling and superheating. The specified initial

conditions for the model are: (1) gas temperature into the

heat exchanger (T gin); (2) minimum average gas exit tempera-

ture from the heat exchanger (T ); (3) gas flow rate ( g);goug

(4) water inlet temperature (T w); (5) water/steam pressurein

(Pw); (6) amount of superheat at the tube exit (Tsh).

Since each tube is separated into N segments, it will be

possible to determine axial temperature distribution of both

the gas and water. At this point, a simplification should be

noted; that is, what happens for any particular segment of

any given tube is taken to be representative of the entire

tube row. This varying gas-to-water temperature difference

across each row of tubes is then utilized to determine a

possible profile for the non-uniform gas temperature distri-

bution across the entire heat exchanger.

63

B. GEOMETRY

Two different fin-tube configurations were selected for

this model. The fins can be either segmented or pin fins.

Tube banks can be arranged with either in-line or staggered

tube rows. From heat transfer and pressure drop measurements

made on in-line and staggered banks, Weierman [Ref. 12] stated

"it was seen that even tube rows in an in-line layout produced

about the same heat duty as four tube rows in a staggered

layout for the same pressure drop." He further stated "in-line

layouts should be reserved for those cases when a serious

justification exists. When cleaning lanes are required, a

staggered 450 layout should be considered." From this infor-

mation, the advantage of staggered layout is obvious, and it

is for these reasons that in-line banks are not considered.

The tube length, LT-' and the number of tubes per row, N tr

are chosen by the designer. For this model the tube length

and heat exchanger height are constrained to that associated

with the S.S.T. boiler, that is, 3.2808 to 6.5616 feet (1-2

meters) in length. Additionally, it should be noted that the

maximum width is limited by shipboard space to 12 feet.

Since this model dissects a tube into N "elemental" heat

exchangers, all heat transfer areas are those associated with

one tube segment. These areas are essentially entire tube

areas divided by N; therefore, the following equations repre-

sent "elemental" heat exchanger areas.

1. Segmented Fins

This fin profile is shown in Fig. 33 and the descrip-

tion of the segmented finned tubes is as follows:

d. = tube inside diameter =1.86 in.I

d = tube outside diameter =2.00 in.

N f =fins per inch -5.94

Z= fin height = 1.015 in.

i= length of cut from fin tip = 0.82 in.

d f = fin outside diameter = 4.03 in.

t f = fin thickness -0.048 in.

w= fin segment width = 0.17 in.

N s= number of segments in 360 degrees = 38

In order to establish the minimum gas flow cross-sectional

area the total "blocked" frontal area, A b9 of the "elemental"

heat exchanger must be calculated from.

L Tdo + 2LTN f it f

Ab N()

The fin surface area is

Afi [N (21 w + 2tf9X + w tf + (df2 }2-d 2)] (10)

and the bare tube area is

7rd oLTAbt N tffi(1

2. Pin Fins

This fin profile is shown in Fig. 34 and the descrip-

tion of the pin finned tubes is as follows:

d. =tube inside diameter -1.86 in.

65

ws

, -±

.<zc - I !D -

... -. Do

FIGURE 33: SEGMENTED FIN PROFILE

66

Vp 5

4- D0 DH

H PS

FIGURE 34: PIN FIN PROFILE

67

d = tube outside diameter = 2.00 in.0

d = fin root diameter - 2.00 in.r

Nf = pin groups per inch 2.00

Z. = average pin length = 1.015 in.P

d = pin diameter = .125 in.p

Ns = number of pins per group = 12

Vps = vertical pin spacing = .375 in.

hps = horizontal pin spacing = .20 in.

The "blocked" frontal area is

L Td 2LT N fd 1Ab = N (12)

The fin surface area is found from

Afi= (1 + dP/4) (13)

and the bare tube area is

doLT d- __Nf) . (14)

bt N ( Ns

The previously determined areas are those which are particu-

lar to the given fin profile. With this information, the

minimum gas flow area is

Ami n = Af -Ab (15)

where Af is the frontal area. The inside area available for

heat transfer per segment heat exchanger is

7wd LTA.i= iT'N- (16)si N

68

- -. . , . . .~- - --- ..- -, .

and the outside heat transfer area is

Aso Afin Ab (17)

Finally, the cross-sectional fluid flow area is calculated

from

Aff - 2[di (18)

For both segmented and pin fin-tube configurations,

the center-to-center tube spacing in the transverse direction

is 4.5 inches. Following Weierman's suggestion (450 staggered

layout) leads to a spacing normal to the gas flow, Sn, of 4.5

inches and a spacing parallel to the gas flow, Sp, of 2.25

inches. This tube layout is shown in Fig. 35.

E (9) (. I

FIGURE 35: MODEL TUBE LAYOUT

69

-E - M UM

With the mass flow rates, terminal temperatures and

heat exchanger geometry established, the remainder of the

model may be solved for the number of tube rows or passes,

actual iterim temperatures, and gas side pressure drop.

C. GAS-SIDE HEAT TRANSFER/PRESSURE DROP

For both the segmented and pinned-tube profiles, the gas-

side Reynolds number is calculated initially using the gas

bulk temperature to find the gas properties. With this Reynolds

number, the segmented fin, j-factor, is obtained from a poly-

nomial fit to the data for tube layout number 5 in Ref. [12].

The j-factor is related to the heat transfer coefficient hg

by the following relationship.

j = StPr2

By introducing

= NuRePr

the previous expression can be written asNu 2/3 Nu

Reg Pr Re Pr 1/3

and

Nu = j Re Pr /3

gg

wherehd

Nu = KoKg

Therefore, a relationship may be written for the heat transfer

coefficient as follows:K 1/3

hg -j K RegPrg (19)

70

70i

For this equation all gas properties are evaluated at the film

temperature.

For the pin fin, the gas-side heat transfer coefficient

has contributions from the pins and bare tube, Equation 6-14

of Ref. [10], supplies the pin heat transfer coefficient_kf 1/3 (O

h -f- C Re Pr (20)

where the constants C and n are listed in Fig. 36.

Re C ng

0.4-4 0.989 0.3304-40 0.911 0.385

40-4000 0.683 0.4664000-40,000 0.193 0.618

40,000-400,000 0.0266 0.805

FIGURE 36: HEAT TRANSFER CONSTANTS

Properties for use with the above equation are evaluated at

the film temperature.

Eckert and Drake [Ref. 13] recommend the following rela-

tions for heat transfer from tubes in cross-flow:

o 5 0 . 38 Prf '2

Nu (0.43 + 0.50 Re )Pr (21)

for l<Re<103

71

-r- -.. .. . ..... . ... . .. .. ..

0.6 038Pr 0.25

Nu 0 .2SRe Pr P 1 (22)

for 10 3<Re<2 x 10s

For gases the Prandtl number ratio may be dropped, and fluid

properties are evaluated at the film temperature. The presure

drop for flow of gases over a bank of tubes may be calculated

from1 4

333.4 f'(G max) 2 N T UwApb (23)

Ap = pressure drop = inches H20

Gma x = mass velocity - lbm/(FT 2-s)

p = free-stream density = lbm/FT3

NT = number of transverse rows

The emperical friction factor f' is given by Jakob [Ref. 14]

as 0.18 0 *1f' = 0.25 + 0.18 ]Remax (24a)

[(Sn - d)/d jmx2

for staggered tube arrangements and

S0.08S /d o _0.15

fl = 0.044 + pRe (24b)I (.43 + 1. 1d/Sp I max

[(n - d)/d] dS

for in-line tube arrangements. For pin fins both friction

factor expressions are required. Although the tubes are

arranged in the staggered tube arrangements, the pins are

essentially an in-line configuration.

The same equation is used for the segmented fin profile

with the friction factor, f', obtained from a polynomial fit

to the data for tube layout Number 5 in Ref. [12], (Fig. 5).

72

" --

D. WATER-SIDE HEAT TRANSFER

Water-side heat transfer involved the heating, boiling and

superheating of water or steam. A multitude of correlations

are currently available to describe each individual mode.

Although not prohibitive, a literature review resulted in

several being chosen to represent the S.S.T. model.

With little known about this boiler type, four heat transfer

correlations were candidates to represent the heating mode.

One could reasonably surmise that the laminar situation is the

most probable; turbulent correlations however, are also con-

sidered. An empirical correlation based on experimental data

that takes into account the effect of varying physical proper-

ties and free convection is listed below [Ref. 15]:

kf 33 4 /Prf, ji~ 3 2go,~T.

hH 0 = .17 k4/( " 2g-- B AT (25)

This relationship is valid for heating in vertical upflow for

Re<2000. Water properties are evaluated at the film tempera-

ture; Prw, at the wall temperature.

The work done by Sieder and Tate [Ref. 16], resulted in

two simple empirical relationships which take into account

fluid property variations. For laminar heat transfer in tubes,

use the following equation

hH0 - 1.86 k (RePr) () ( (26)

ford.

RePr --I >10T

73

and for turbulent heat transfer

hH20 M .027 k Re Pr (27)

for

7<Pr<16.700

Re>10,000

For both situations, all fluid properties are evaluated at the

mean bulk temperature of the fluid, except w , which is evalua-

ted at the wall temperature. Recommended by Dittus and Poelter

[Ref. 17] for turbulent flow is

3k Re Pr (28)hH20 02 - Re Pr

The properties in this equation are evaluated at the fluid bulk

temperature. This variety of expressions allows the designer

to study a greater range of possibilities.

Consider now the conditions under which boiling will be

initiated in the vertical heated tube. No boiling can occur

while the temperature of the heating surface remains below the

saturation temperature of the fluid at that particular location.

It should be realized that fully developed subcooled boiling

is not initiated whenever the heated surface first exceeds

the saturation temperature. As shown in Fig. 37, a region of

"partial boiling" exists between the subsaturation zone, A,

and fully developed subcooled boiling, B. This "partial boil-

ing" zone consists of comparatively few nucleation sites and

in this zone, a proportion of the heat would be transferred

by normal single-phase convection between patcher or bubbles.

74

F"' do w - ,Od

%% (Rgion )

%

F

A////

Partial iing

FIGURE 37: SURFACE HEAT FLUX VS. INLET SUBCOOLING

For forced convective flow, Bergles and Rohsenow (Ref. 18]

developed a criterion based on analysis by Hsu [Ref. 19] and

Han [Ref. 20] for the incipience of boiling. In this work,

Bergles proposed a convenient numerical expression to determine

the wall superheat required to initiate boiling. This expres-

sion is for the steam-water system in the 15-2000 psia pres-

Oonb - 15.60 p (T wall - Tsat) P ) (29)

75

where 0 onb (Btu/hr.ft2 ) is the heat flux to cause nucleation

at a wall superheat T wall Tsa t and at a system pressure p

(psia).

As the surface temperature is increased further, the whole

surface is covered by bubble sites, boiling is "fully developed"

and the single-phase component reduces to zero. In the "fully

developed" boiling region, velocity and subcooling have little

or no effect on the surface temperature as observed experimen-

tally. Throughout the partial boiling region Rohsenow [Ref.

21] suggests that

0Total = OSPL + 0SCB

where 0Total is the total average surface heat flux, 0 SPL is

the average surface heat flux transferred by single-phase

convection, and 0SCB is the average surface heat flux trans-

ferred by boiling. This method of superposition, shown in

Fig. 38, utilizes the single-phase component given by

OSPL ' hc (Twall - Tbulk)

where h is found by the familiar Dittus-Poelter equation.

The boiling contribution was successfully correlated with

the experimental data using the equation suggested by Rohsenow

[Ref. 22] for staturated nucleate pool boiling

0SCB I g(f - )1Tall Tsatd] 3 (30)

SfgPr Csf Ifg

where Csf is a constant described by the liquid surface com-

bination. The data reduced and correlation used a value of

Csf - .006. The surface-tension is calculated from (Ref. 23].

76

Single-phase I SubcooledLiquid I boiling

(Region 8)

Non-boboiinging

I- - Oc

forcedc tm iraueT

FIUEC8OMTOMOMOHEO

OW77

a = 58 x 10 (1 .000142T) lbf/FT

with T in *F

In their introduction to Chapter 8, Hsu and Graham [Ref.

24] delineate the metamorphosis of two-phase flow typically

shown in Fig. 39.

A fluid enters the heating section as a subcooledliguid. Some distance downstream, subcooled boilingbegins on the wall, but the bubbles do not leave thewall. Farther downstream, the bubbles begin to growand depart into the main stream (low subcooled boilingregion). As the fluid is near the saturation tempera-ture, enough bubbles are generated to populate thewhole cross section (bubbly flow) and finally coalesceinto slug flow. However, slug flow exists only overa narrow range of valid fractions. As the void frac-tion is increased, the flow pattern quickly undergoesa transition into annular flow with a dispersed mistin the core (dispersed annular flow). The annularfilm is gradually depleted and thinned out. In thefinal stage, the flow is a mist flow with dropletscarried by the superheated vapor. The wall regioncan be dry.

This excerpt from Hsu and Graham should raise the question

of when do these individual changes occur. Reference [24]

states that Haberstroh and Griffith found that the transition

between slug and annular two-phase flow occurred at the void

fraction between .8 to .9. The void fraction, a, can be

determined from

with the vapor velocity found from

u Gx G mTOTALv v AREA

and the fluid velocity isuf G(l-x)

f78

78

....... ......

Dry-wallmist

Annularflow

Increasedcz 3 quality• ,and

Slug voidflow

o d08Incipientboiling 0

Saturatedinlet

FIGURE 39: TWO-PHASE FLOW DEVELOPMENT

79

4 _ . . . _ l m . .

where x is the mass quality, and is defined asm vv mfxl-xmTotal mTotal

Now provided both liquid and vapor are in thermodynamic equi-

librium, i.e., they exist at the saturation pressure and tempera-

ture, then an alternative definition of mass quality x can be

given on the basis of thermodynamic properties.

I - fe I

fg

Only when the thermodynamic equilibrium exists are the values

of x and x e identical.

However, this result was obtained for the adiabatic condi-

tion. Some important features of flow patterns in boiling

two-phase flow and differences between this and the adiabatic

case are discussed in Hsu and Graham. Furthermore, it has been

suggested that the Haberstroh-Griffith criteria for this

transition overpredict the flow quality. According to The

Handbook of Heat Transfer, [Ref. 25], Rohsenow observed that

the slug-to-annular flow transition occurs for qualities

between 5-10%. In this thesis a local true quality of 10% was

used as the accepted transition.

Slug flow occurs when the flow equilibrium enthalpy is

near saturation, the actual quality when xe = 0 can be

determined by using the analysis for the subcooled boiling

regime. Based upon Hsu's [Ref. 19] postulate, Levy proposed

a criterion for bubble detachment, and therefore, a condition

80

for fully developed subcooled boilipg. He showed that the

true quality can be related to the equilibrium quality through

the expression [Ref. 26]

X T = X e - xd EXP (x e/Xd - 1) (31)

where

xd = -c P (Tsat - Tbulk)Ifg

Using Levy's approach, the fraction of heat flux used for

evaporation 0ev can be determined from

0 = Total [1 - EXP (xT/Xd - 1)1 (32)

At the present, a workable analysis of annular two-phase

heat transfer is not available; therefore, empirical correla-

tions must be employed. References 24, 25 and 27, contain

expressions derived from the work of Dengler and Addoms,

Schrock and Grossman, and Collier and Pulling, for example.

Chen [Ref. 28] compared the existing correlations and found

while the expression obtained by a particular research group

represented their own data fairly well, 4t failed to represent

any other groups' data with any reasonable accuracy. In

general, none of the existing correlations appeared satisfactory

for general use. As did Rohsenow, Chen postulated that two

basic mechanisms take part in the heat transfer process for

boiling of saturated fluids with flow.

The macroconvective mechanism, or single-phase convection,

is described by a modified form of the Dittus-Roelter relationship.

81

- -

h a 023 k Re'Pr" F (33)1

In this expression the factor F is the ratio of the two-phase

Reynolds number (ReTP) to the liquid Reynolds number (Re),

where ReTP is defined by ReTP = ReF '25 . Chen reasoned and

verified that this ratio is a function of the Martinelli para-

meter, X TV given by. )T . .9 S.I

xTT = T

The work of Foster and Zuber was expanded to include the

previously neglected effective and wall superheat difference

in order to obtain an expression for the microconvective, or

bubble nucleation and growth (boiling) term.. A surpression

factor given by

AT e

AT

where

AT = Twall T sat

ATe - effective uperheat with flow1 . 25

was defined and later verified to be a function of ReTP(ReF

Manipulating S through the Clausius and Clapeyron equation and

combining with the modified Foster and Zuber expression, Chen

obtained an expression for microconvection in terms of the

suppresion factor and total superheat, AT.

kmc . 0 2 p " g .2' , .75

= 00122 k '.4 P('.g 2 5 AT AP S (34)i auf 29Ifg 2 4O v.2 4

82

tl~ u. . . -'- k, -' - i - ' J . .. . . ... : .. . . .. ' - l - l . . .. V

where a is the surface tension of water previously defined,

and

AP pressure difference between the saturationpressure correspondint to T wall minus the

bulk saturation pressure and is in lbf/ft2.

The total heat transfer coefficient is then obtained as the

h Total h ha i

This correlation was tested with available data for water

and organic fluids. The average deviation between calculated

and measured boiling coefficients for all (over 600) data points

from ten experimental cases was + 12%. They include those for

water in the pressure range of 1 to 35 atm with liquid flow

velocities up to 14.7 ft/sec, heat flux up to 760,000 Btu/hr.

ft2 , and quality up to 71%. However, no substantiating evi-

dence was found to refute using this correlation up to a quality

of 100%.

Computation of the water-side heat transfer coefficient

from a true quality of 10-100% is accomplished by the utiliza-

tion of the Chen correlation. Both the Reynolds number factor,

F, and the Suppression factor, S, are obtained from polynomials

fit to the data of Reference 27, Figs. 6 and 7 respectively.

From the point of 100 percent quality to the tube exit, the

evaluation of the water-side heat transfer coefficient becomes

a choice of the previously discussed forced convection heat

transfer correlations which are obviously dependent upon the

Reynolds number.

83

E. OVER-ALL HEAT TRANSFER

Newton's law of cooling may be written as

Q =hAAT

It is often convenient to express the heat transfer rate for

a combined conductive-convective problem in this form with h

replaced by an overall heat transfer coefficient U.

The primary objective in the thermal design of heat ex-

changers is to determine the necessary surface area required

to transfer heat at a given rate for given fluid temperatures

and flow rates. The utilization of U facilitates this procedure

in the above fundamental heat transfer relation

Q = UAATLM

where ATLM is the average effective temperature difference

known as the log-mean temperature difference.

It is well known that the primary purpose of fins is to

increase the effective heat transfer surface area exposed to

a fluid in a heat exchanger. The performance of fins is often

expressed in terms of the fin efficiency, nf

[actual heat transfer]heat transfer if entire fin wereat the base temperature

Analytical expressions for nf are readily obtained for several

common fin configurations.

For this model, both segmented and pin fins are considered

to have ends that are at the free-stream temperature. Using

a corrected length L (as was suggested by Harper and Brown)c

the insulated tip fin equation can be utilized, as follows

84

tanh (mLC) (36)qf = mL c 6

where L = L + t/2. However, for the fin dimensions utilized

in this thesis L>>t which allows Equation 36 to be approximated

by:

- tanh (mL) (37)- mL

h Pwhere m =

h = gas-side heat transfer coefficientg

P = "wetted" fin perimeter

k = fin thermal conductivity

A = fin cross-sectional area

L = fin length

Now in terms of ff, the heat transfer rate is given by the

simple expression

Q - hi (Abt +lf Afi)AT

remembering that Abt is essentially

Abt = Aso - A fin

allows the heat transfer equation to be rewritten in terms of

an "elemental" heat exchanger efficiency as

Q = hns A soAT

where ns M 1 - (1 - nf) fin

so

Using the tube wall resistance, together with the gas-side and

water-side heat transfer coefficients and the "segment" effi-

ciency, the overall heat transfer coefficient can be found as

follows:

85

-7=277.

(a) in terms of the inside area

Uo. A 1 (38)1 +Asi 01n(d/d si+ +

hH20 27ktube Ls so sh g

(b) in terms of the outside area

Uo 1 (39)o Aso A ln(do/d.)+ +

sihH20 2ktube s

Using the effectiveness - NTU method, an average "segment"

effectiveness is calculated from equations 7a or 7b; if a

phase change is occurring, equation 8. The expression for e

is now used in an iterative fashion. Each time an effective-

ness is calculated it is compared to the previously found £

until a difference of less than .001 is obtained. This differ-

ence represents the convergence of the "elemental" heat exchanger

effectiveness. During each iteration, the gas and water

temperatures are determined in the following way. The total

heat transfer rate is, per segment,

Qs 0c ATmin max

then the outlet gas temperature is

Tgou t 0 Tgin - Qs/Cgas (40a)

and for the water/steam temperature

Twouy a TWin + Qs/CH20 (40b)

or if boiling is present, the outlet enthalpy of water is

found from

out in + QS/H 2 0 (40c)

86

where I * enthalpy in BTU/lbm

With these segment terminal temperatures, inside and outside

tube wall temperatures may be calculated as follows:

Tgou t + Tg in .Tg

TW out ' TWinTWbulk 2 b TWb

Now performing a heat balance upon the "elemental" heat ex-

changerTgb, Twb (1IQs gbRthw (41)

Rth

iwith Rt Ai~ AsoUo

or Tgb TT = TTi - TwbR0

i

here TTO outside tube wall temperature

TTi - inside tube wall temperature

R 0 nsh 1Ag so

R1Ri A Asi hH2 0

Combining to find the outside tube wall temperature in terms

of the overall temperature difference asRo

TTOa Tgb - t (Tgb Twb) (42)

and for the inside wall

R.T Ti = w Tb + r)h ( T g b -TVb (43)

th

these wall temperatures are then utilized to find the gas-side

and water-side film temperatures

87

- - ---- " . ... . --- - -. ----1

Tgb + TToTgf 2

Twb + TTiand Twf = b2T

which are used in calculations requiring fluid property evalua-

tion of the film temperature.

This completes the required calculations for the "elemental"

heat exchanger. The final values of the water-side variables

are used as the inlet conditions for the next segment, and the

whole of the aforementioned procedure is repeated N (total

number of elemental heat exchangers per tube) times. This re-

presents the end of the tube.

At this point an overall tube heat balance is written based

upon enthalpy considerations. The total heat transfer rate is

found fromN

QTotal j*i Qs.

and QTotal = mH20 (exit - inlet

solving for the water mass flow rate yields

QTotal ibm

HO I exit Iinlet r-

Obviously as the gas transverses the rows of tubes in the

heat exchanger, its temperature decreases, producing a corres-

ponding decrease in the gas-to-water temperature difference.

The heat transfer rate being a function of this temperature

difference also becomes less, however, the inlet and exit

conditions are specified design parameters which remain con-

stant. Since one of the assumed parameters was a water-side

88

- - ~- ----- ---- - - .-

Reynolds number, this step allows for the actual mass flow

rate, and Reynolds number to be compared with the original.

If the relative error defined here as

H20 calc -1 20 INITIALE

r "~H20 calc

is greater than 5%, the entire set of calculations is repeated

using the calculated mass flow rate as the initial m H20 In

this manner, a profile of the water mass flow rates is found

on a row basis for this type of heat exchanger.

A matrix of the segment outlet gas temperatures is utilized

at the inlet of the next row. The total number of rows is

obtained when the following condition is satisfied

T0 * ~ ou Tg i Design outlet gas temperature

89

III. RESULTS AND CONCLUSIONS

A. BACKGROUND

The foregoing model was applied in a computer simulation

program witten in BASIC for the Hewlett Packard 9845 model B

desktop computer. Inclusive are a set of supporting subpro-

grams for the determination of water, steam and air properties.

Finally, one subprogram was included to provide a choice for

the thermal conductivity of both the selected fin and tube

metal. With the establishment of the design variables to test

the model and to develop an understanding of the behavior of

the S.S.T., the program was activated to produce a heat ex-

changer design.

A review of the initial results revealed that the proposed

model required modification in order to produce more realistic

simulated data. The problem arose from the attempt to match

Rohsenow's and Chens correlation at 10% steam quality. It is

well known that boiling heat transfer is associated with rather

high heat transfer -coefficients, and these results reflected

this fact up to the 10% quality point. The heat transfer co-

efficients were of the order of 20,000 BTU/HR FT2 OF. However,

upon program transfer to the Chen correlation for the

determination of the water-side heat transfer coefficient, these

values were of the order of 6-20 BTU/HR FT2 OF. Obviously,

this tremendous diference in magnitudes is definitely not a

90

real simulation of the probable actual situation. This mis-

match in heat transfer coefficients no doubt arises from the

suppression factor, to utilize in the Chen microconvective

term (Eq. 35), way below the lower limit of approximately

18,000 for the Reynolds number as shown in Figure 40.

1.0 1 T I m I I 'SA-

.6

2-

REGIO OFt FDATA

0

.5

FIGURE 40: SUPPRESSION FACTOR, S

Therefore, the proposed model was changed to utilize the

Rohsenow method for the entire range of 0-100% steam quality.

The single phase contribution was calculated by using a value

for the heat transfer coefficient, h c , being determined by the

local Reynolds number. Remember that initially h c was cal-

culated from the Dittus-Boelter correlation. A listing of

the final program appears in Appendix A.

B. DESIGN VARIABLES

The following set of design variables was selected to

produce the basic set of data.

91

1. S.S.T. Operating Pressure

The previous work of Combs [Ref. 7] indicated that the

optimal operating pressure ranged from 600-800 psig. For this

simulation, 600 psig was selected.

2. Frontal Diminsions

Space constraints for a DD-963- type engineroom limits

the maximum heat exchanger width to 12 feet. To be consistent

with this constraint, the frontal dimension of 3.658 m x lm

(12'x3.2808') was selected for the design.

3. Gas Flow Rate and Temperatures

The outlet from the gas turbine was taken to be 145

lbm/sec (522,000 lbm/hr) at 900'F under steady state and steady

flow conditions.

4. Superheater Outlet Steam Temperature

The target superheat at the outlet was specified to be

S0OF. At 600 psig this corresponds to 536*F. This amount of

superheat will vary with water-side mass flow rate.

S. Water Inlet Temperature

The water inlet temperature was fixed at 50'F subcooling

or 4360F.

6. Fin Geometry

The segmented fin profile was selected for the gas-side

heat transfer.

92

C. THE DATA SET

A set of data was produced for the combination of design

variables just discussed. This data was generated by varying

the inlet Reynolds number, or if one prefers, the inlet mass

flow rate, and was the result from the investigation of 60

different such numbers. The following set of output variables

was considered.

1. Gas bulk temperature variation

2. Gas-side heat transfer coefficient

3. Overall heat transfer coefficient based upon the out-

side area

4. Tube wall temperature variation

S. Water/steam mass flow rate

6. Water-side heat transfer coefficient

7. Overall heat transfer coefficient based upon the inside

area

8. Inside surface heat flux

9. Water-bulk temperature variation

10. Thermodynamic quality variation

This investigation included a Reynolds number range of

75 to 1650. Figure 41, shows graphically the water/steam mass

flow rate corresponding to each Reynolds number in order to

provide information concerning the range of mass flow rates

involved. It should be re-emphasized that these Reynolds

numbers are those associated with the water properties at the

tube inlet temperature of 436 0F. The outlet Reynolds number

93

0zM

z

00 z

U) U

0 -

LAI

U) -

949

becomes simply Reout = Re. in INLET

in"EXIT

A logical expectation is the decreas~e in outlet superheat

with an increase in mass flow rate. This trend is displayed

in figure 42a. Point A corresponds to a Reynolds number of

310 and a superheat of 610; d.,hereas, point B, 380 and 91.

Figures 43 and 44 show that there is very little difference in

the water-side heat transfer coefficient for these two values

of Reynolds numbers during the preheating of the water toward

the saturation temperature. However, the value of h represen-

tative during superheating for Re = 380 is approximately twice

the value of h for Re 310 (Figures 45 and 46).

The logical question to be raised at this point is why

this difference e.,ists. Outwardly, there definitely appears

no logical explanation. Although not a specific part of the

data set, the Reynolds number variation with axial position

should have been plotted. It was observed on the CRT of the

computer that in this "transition" region the Reynolds number

gradient steadily increased along the tube for the inlet

values of 310 and 380. This observation revealed the greatest

magnitude of the Reynolds number was obtained at the point of

100% steam quality inside the tube.

For a Reynolds number of 310 the maximum value obtained

was approximately 2100, whereas a Reynolds number of 380 the

maximum value obtained was approximately 3300. The effect on

heat transfer is observed as the water-side tube flow trans-

verses from the laminar to turbulent regime, with the steam

95

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105

bulk temperature variation reflecting this fact. In figure

47, the change in steam temperature is approximately 24 .4 OP FT

whereas in Figure 48, it is 38.60 F /FT.

Figures 49 and 50 show the axial location of initiation of

boiling (zero percent quality) and that of complete evapora-

tion (100 percent quality) as a function of the inlet Reynolds

number.

The target of 50 degrees superheat was obtained with an

inlet Reynolds number of 840.

Table I. HEAT BALANCE RESULTS

Inlet water-side Reynolds number 840

Outlet water-side Reynolds number 5180.19

Overall water mass flow rate 29.648202 lbm/hr

Total overall heat addition rate 24621.249 BTU/hr

The inlet water temperature 436.21 OF

The saturation temperature 486.21 O

The outlet steam temperature 536.22 O

The amount of superheat 50.01 OF

The inlet enthalpy 414.867 BTU/lbm

The outlet enthalpy 1245.314 BTU/lbm

Figures 51 through 59 graphically represent the data

obtained from the model simulation for Re =840.

106

These figures show the axial distribution of the associated

variables under the prescribed design conditions for the seg-

mented fin profile.

Table 2: TUBE/FIN DIMENSIONS AND AREAS

The selected fin profile is the segmented-fin

The outside tube diameter 2 Inches

The inside tube diameter 1.86 Inches

The tube length 39.37 Inches

The total outside tube area 1.71889 Square Feet

The total inside tube area 1.59857 Square Feet

The fin root diameter 2 Inches

The number of 'fins' per inch 5.94

The number of 'segments' per fin 38

The fin height 1.015 Inches

The fin thickness .048 Inches

The transverse tube pitch 4.5 Inches

The longitudinal tube pitch 2.25 Inches

The fin outside diameter 4.03 Inches

The fin segment width .17 Inches

The length of cut from fin tip .82 Inches

The total integer number of fins 234 Fins

Actual tube length will be 39.394 Inches

The fin thermal conductivity 144 BTU/Hr.Ft./F

The tube thermal conductivity 20 BTU/Hr.Ft./F

107*.

Calculated heat transfer areas in square feet:

Total heat exchanger frontal area 39.394

Total fin area 26.9509S1

Total bare tube area 1.228802

Total outside tube area 28.179753

Single fin area .115175

Outside 'elemental' area .120426

Inside 'elemental' area .006831

'Elemental' frontal area .005261

'Elemental' blocked area .003015

'Elemental' minimum gas-flow area .002246

This data shows what would reasonably be expected. That

is, relatively low values for heat transfer coefficients and

heat flux for the preheat and superheating sections and high

values for the boiling section.

D. CONCLUSION

The short straight tube model and the accompanying graphi-

cal data provide the basic framework for additional study of

the stagnation, or S.S.T boiler. This model, as formulated,

provides a fairly reasonable estimate of the possible perfor-

mance for the fin-tube configuration considered. Performance

estimates are encouraging even considering the assumptions

involved.

108

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135

IV. RECOMMENDATIONS FOR FURTHER RESEARCH

Investigation of the Short Straight Tube Boiler model in

this thesis was essentially the adjustments of the water/

steam mass flow rate to determine some possible operating

conditions and heat transfer characteristics. This investi-

gation was conducted upon only one tube under steady state

steady flow operation. Additionally, the model consisted of

preheat, boiling, and superheat with accompanying single

phase flow.

In order to increase the precision of the model and lead

to greater insight into the possible performance of the S.S.T.,

consideration must be given to the inclusion of two phase

heat transfer; in particular to "frothy" flow correlations.

Possible sources of this information may exist in research

and experimentation conducted in "flash-type"' water distilla-

tion systems.

Secondly, consideration should be given to apply enhanced

heat transfer configurations to the inside af the tubes;

helical wire inserts or helical vane inserts. This enhance-

ment would be required to increase the heat t-ansfer in the

preheat and superheat sections and would be very important

especially as gas temperatures decreased.

FFinally, the program should be transferred to the IBM

3033 computer in order to more efficiently process an expanded

136

program to determine actual heat exchanger size. Although a

powerful tool the HP 9845-B lacks the storage capability,

especially, disk storage, to handle a complete heat exchanger

design analysis that would be required in order to make a

reasonable comparison with other heat exchangers.

With the aforementioned analytic changes implemented, the

calculated heat transfer coefficients should be compared with

those coefficients obtained experimentally. The single tube

experiment should also include the capability for the obser-

vation of the boiling regime in order to gain possible greater

qualitative insight into the most probable flow patterns in

existence in the boiling regions within the tube.

137

APPENDIX A: PROGRAM LISTING

i-=l

(0whi C4 N

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V' w. LL. - U -T Q'') Ix- = w W) L -W - T 4

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LIST OF REFERENCES

1. Babcock and Wilcox Co., Steam, Its Generation and Use,Babcock a.zd Wilcox Co., New York, 1963.

2. Shields, Carl D., Boilers: Types, Characteristics, andFunctions, F.W. Dodge Corporation, New York, 1961.

3. Latham, Robert F., Introduction to Marine Engineering,U.S. Naval Institute, Anapolis, Maryland, 1955.

4. Croft, Terrel, Steam Boilers, McGraw-Hill Book Co., Inc.New York, 1937.

5. Fryling, Glenn R., Combustion Engineering, CombustionEngineering Co., Inc. New York, 1966.

6. Principles of Naval Engineering, 3rd ed., NAVPERS10788-B, p. 255, Training Publications Division, NavalPersonnel Program Support Activity, Washinton, DC, 1970.

7. Combs, R.M., LCDR, U.S.N., Waste Heat Recovery UnitDesign for Gas Turbine Propulsion Systems, M.S. Thesis,Naval Postgraduate School, Monterey, California, 1979.

8. Egusa, Tatsuo, Short Straight Tube Boiler - S.S.T.,Lectures presented at Naval Postgraduate School, Monte-rey, California, 23-27 June 1980.

9. Holland, F.A., Moores, R.M., Watson, F.A., and Wilkinson,J.K., Heat Transfer, American Elsevier Publishing Co.,Inc., New York, 1970.

10. Holman, J.P., Heat Transfer, 4th ed., McGraw-Hill BookCo., Inc., New York, 1976.

11. Kays, W.M. and London, A.L., Compact Heat Exchangers,McGraw-Hill Book Co., Inc., New York, 1964.

12. Weierman, C., Taborek, J., and Marner, W.J., Comparisonof In-Line and Staggered Banks of Tubes with SegmentedFins, paper presented at the 15th National Heat Trans-fer Conference, AIChE-ASME, San Francisco, California,i9"5.

Skert. E.R.G., and Drake, R.M., AnalZsis of Heat and,.: rjnqfer, VzGraw-Hill Book Co., New York, 1972

A

14. Jakob, M., Heat Transfer and Flow Resistance in CrossFlow of Gases over Tube Banks, Trans. ASMR, vol. 60,p. 384, 1938.

15. Collier, J.G., Convective Boiling and Condensation,McGraw-Hill Book Co., New York, 1972.

16. Sieder, E.N., and Tate, C.E., Heat Transfer and PressureDrop of Liquids in Tubes, Inc. Eng. Chem., vol. 28,p. 1429, 1936.

17. Dittus, F.W., and Boelter, L.N.K., Heat Transfer in Auto-mobile Radiators of Tubular Type, Publications inEngineering, University of California, p. 443, 1930.

18. Bergles, A.E., and Rohsenow, W.M., The Determination ofForced Convection Surface-Boiling Heat Transfer, J. HeatTrans. vol. 86, pp. 365-372, 1964.

19. Hsu, Y.Y., On the Size Range of Active Nucleation Cavitieson a Heating Surface, J. Heat Trans., 843(3), pp. 207-216,1962.

20. Han, C.Y., and Griffith, P., The Mechanism of Heat Trans-fer in Nucleate Pool Boiling, Part I, Bubble InitiationGrowth and Departure, Int. J. Heat Mass Trans. 8(6),pp. 887-904, 1965.

21. Rohsenow, W.M., A Method of Correlating Heat TransferData for Surface Boiling of Liquids, Trans., ASME, 74,p. 969, 1952.

22. Rohsenow, W.M., Heat Transfer with Evaporation, Universityof Michigan Press, p. 101-150, 1953.

23. Sucec, J., Heat Transfer, Simon and Schuster TechnicalOutline, Simon and Schuster, Inc., New York, 1975.

24. Hsu, Y.Y., and Graham, R.W., Transport Processes inBoiling and Two-Phase Systems, McGraw-Hill Book Co.,New York, 1976.

25. Rohsenow, W.M., and Harnett, J/P., Editors, Handbook ofHeat Transfer, McGraw-Hill Book Co., New York, 1973.

26. Levy, S., Forced Convection Sbucooled Boiling-Predictionof Vapor Volumetric Fraction, Int. J. Heat Mass Trans.,10, pp. 951-965, 1967.

27. Tong, L.S., Boiling Heat Transfer and Two-Phase Flow,John Wiley and Sons, Inc., New York, 1965.

179

28. Chen, J. A Correlation for Boiling Heat Transfer toSaturated Fluid's in Convective Flow, ASME, 63-HT-34, 1963.

29. Afimiwala, K.A., Interactive Computer Methods for DesignOptimization, Ph.D. Thesis. State University of New York,BuiititalNew York, 1976.

180

INITIAL DISTRIBUTION LIST

No. Copies

1. Defense Technical Information Center 2Cameron StationAlexandria, Virginia 22314

2. Library, Code 0142 2Naval Postgraduate SchoolMonterey, California 93940

3. Department Chairman, Code 69 2Department of Mechanical EngineeringNaval Postgraduate SchoolMonterey, California 39340

4. Professor Paul F. Pucci, Code 69Pc 3Department of Mechanical Engineering iNaval Postgraduate School

Monterey, California 93940

S. Professor Jack Stodghill 1IIDept. of MathematicsDickenson CollegeCarlisle, Pennsylvania 17013

6. Lt. Leo W. Vollmer, Jr.1Rural Route $4Vincennes, Indiana 47591

181

Documents

N6 moIIIIIIIIIIII llllllllllllII uIuuuoilu · Superheater Outlet Steam Temperature 92 5. Water Inlet Temperature ----- 92 ... FIRST WATER-TUBE BOILER ... YARROW TYPE BOILER -----