Mechanical Design of Process System Volume-2(Shell and Tube Rotating Equipment)-Keith Escoe

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Mgchanica

Design

of

Process

Systems

Volume2

Shell-and-Tube

Heat Exchangers

Rotating Equipment

Bins,

Silos,

Stacks

A.Keith Escoe

Gulf

Publishing

Company

Book Division

Houston, London, Paris,

Tokyo



Design

Pmctss

Svsterns

2

Heat Exchangers

o

Equipnent

r

Bins,

Silos,

Stacks

right @

1986

by Gulf Publishing

Company,

Houston, Texas.

righrs

reserved.

Printed in the United

States

of America.

This

or

parts

thereof,

may not

be reproduced

in

any

form without

the

publisher.

ol

Congress Calaloging-in-Publicalion

Data

A. Keith.

design of

process

systems.

bibliographies

and indexes.

v. l.

Piping

and

pressure

vessels-v. 2. Shell-and-tube

exchangers;

rotating equipment;

bins, silos, stacks.

Ch€mical

plants

Design and construction.

1986 660.2

',

81

O.ATant

-562-9

(v

1)

(}ET2l)1-565-3

(v.

2)

85-22005

iv



Contents

Foreword ........vii

by John J. McKetta

Preface

..........ix

Chapter 5

The Engineering Mechanics

of

Bins,

Silos,

and

Stacks ........1

Silo

and

Bin Design,

I

Stack

Design,

8

Vortex

Shedding and Frequency

Responsc.

Ovaling. Helical

Vortex Breaker

Strakes.

Example

5-l: Granule

Bin

Design

for Roofing

Plant, 11

Bin Stiffener Design.

Vcssel

Supports.

Example

5-2: High-Pressure Flare

Stack

Design, 20

Effective

Diameters.

Section

Weights-Uncorroded weight.

Required

t

Thickness.

Anchor Bolt

Design. Cantilever

Vibration.

Static

Deflection.

Dynamic

Deflection. Anchor Bolt

Torque.

Design

Summary.

Example 5-3: Stack Vortex

Strake

Design, 27

Example 5-4: Natural Frequency

of

Ovaling

Ring Formula

(Michell

Formula), 28

Notation,29

References, 29

Chapter

6

Rotating

Equipment ......31

Pumps, 31

Centrifugal

Pumps.

Hydraulic

Requirements

of

Centrifugal Pumps. Positive Displacement

Pumps.

Pressure Protection

for

Positive

Displacement Pumps.

Compressors,43

Principles

of

Compression. Reversible

Adiabatic

(lsentropic)

Compression. Polytropic

Compression.

Isothermal

Compressron.

Dimensionless Reference

Numbers.

Centrifugal

Compressors. Reciprocating

Compressors.

\{ulriple

Staging

of

Reciprocating

Compressors.

Cas

Temperature for Reciprocating

Compressors.

Axial Flow

Compressors.

Specirying Compressor

Flow

Conditions. Mass

Flow.

Actual or

lnlet

Volumetric

Flow.

Standard

Volumetric

Flow. Properly

Specifying

Compressor

Flow

Conditions.

Piping Systems

for

Rotating Equipment, 60

Nozzle Loadings.

Pulsation Response

Spectra Induced by

Reciprocating Equipment, 62

Example 6-l: Horizontal Centrifugal

Pump

Sysrem

Design,

65

Suction

Line

Pressure

Drop.

K-Values.

Discharge

Line

Pressure

Drop.

The

Effects

of

Liquid Viscosity on Centritugal

Pumps.

Example

6-2: Positive Displacement Pump

Design,74

Suction

Line

Pressure

Drop.

K-Values.

A

word

About

Priming.

Example

6-3: Centrifugal

Compressor Selection,

Example 6-4: Installing a Compressor

at

Elevation,

34

Selecting the Reciprocating Compressor.

Example 6-5:

Naphtha

Pump System Design,

86

Flow from

Reservoir

to

Naphtha Storage Tank.

Naphtha Pump

Hydraulics. The

Maximum

Capacity

Condition.

Reevaluation

of

Reservoir

Line.

Notation,9T

References,

97

79



Chapter

7

The Mechanical

Design

of Shell-and-Tube Heat

Exchangers

......

99

Fundamentals

of

Shell-and-Tube

Heat

Exchangers,99

Design

Classifications

of

Heat Exchangers.

Fixed Tubesheet

Shell-and-Tube Heat

Exchangers.

U-Tube Shell-and-Tube

Heat

Exchangers.

Floating

Head

Shell-and-Tube Heat

Exchangers.

General

TEMA

Exchanger

Classes-R,

C,

and

B.

Basic

Components

of

Shell-and-Tube

Heat Exchangers.

TEMA

Formulations.

ASME

TUbe Joint

Load

Criteria.

Process

Evaluation

of Shell-and-Tirbe

Exchangers,

115

Tube Wall

Temperature

and

Caloric

Temperaturc.

Overall

Heat Transfer

Coefficient.

Fouling

of

Inside

and Ourside Tube

Surfaces. Tube

Film

Coefficients.

Tube Vibrations,

139

Plate-Fin Heat

Exchangers,

147

Example 7-1:

Regenerated

Gas Exchanger

Design, 148

Tube-Side

Film

Coefficient.

Shell-Side

Film

Coefficient.

Shell-Side

Pressure

Drop.

Example 7-2: Vibration

Check for

Regenerated

Gas

Exchanger,

153

Example 7-3:

Chlorine

Superheater

Design,

154

Tube-Side

Film

Coefficient.

Shell-Side

Film

Coefficient.

Shell-Sid€ Pressure

Drop. TUbe

Metal

Temperature.

Example 7-4: Asphalt

Coating Mix

Heater-A

Non-Newtonian

Fluid Application,

160

Tube-Side

Film

Coefficient.

Shell-Side

Film

Coefficient.

Shell-Side

Pressure Drop.

Example 7-5:

Zero

LMTD

Exchanger,

165

Notation,

165

References, 166

Chapter

8

External Loadings

on

Shell Structures

.... 169

Lifting Lug Design,

170

Example

8-1:

Lifting

Lug Design

and Location, 170

Notation,

175

References,

176

Appendix

A

Partial

Volumes

and Pressure

Vessel

Cafcufations

....,177

Partial Volume

ofa

Cylinder,

177

Partial

Volume

of

a Hemispherical

Head,

177

Partial Volumes

of

Spherically Dished

Heads, 178

Partial

Volumes

of Elliptical

Heads, 179

Partial

Torispherical

Heads, 181

Internal

Pressure

ASME Formulations

with

Outside Dimensions,

183

Internal

Pressure ASME

Formulations

with

Inside

Dimensions,

184

Appendix B

National Wind Design Standards

.........

187

Criteria for

Determining

Wind

Speed,

187

Wind

Speed

Relationships,

188

ANSI

A58.1-1982 Wind

Categories,

189

Appendix G

Properties

ot

Pipe

. . .....

193

Insulation Weight Factors,

200

Weights

of

Piping

Materials,

201

Appendix

D

Conversion

Factors

.

.....225

Alphabetical

Conversion

Factors, 226

Synchronous

Speeds, 233

Temperature

Conversion, 234

Altitude

and Atmospheric Pressures,

235

Pressure

Conversion

Chart,

236

vl



The engineer

who

understands

the impact

of

process

design decisions

on mechanical design details

is in a

po-

sition to save

his client

or

his company a

lot of

money.

That is because the test

of

any

process

design

is in how

cost-effectively

it

yields

the desired

product,

and how

"cost" generally

translates to

"equipment":

How much

will

the

process

require? How long will

it last? How

much energy

will

it

consume

per

unit of

product?

In

this two-volume

work

on Mechanical

Design

of

Process

Systems, A. K.

Escoe has

performed

a monu-

mental

service

for

mechanical design engineers

and

chemical

process

engineers

alike.

The

information is

presented

in

such a

manner that even the

neophyte

engi-

neer can

grasp its

full

value.

The author

has

produced

an

in-depth review of the way

in which

process

design spec-

ifications

are

interpreted into

precise

equipment designs.

Perhaps most

valuable of all

are the extensiv

e worked ex-

amples throvghout the text, of actual designs

that

have

been

successfully

executed in the field.

The

piping

system is the central nervous system

of

a

fluid

flow

process,

and

the

author has treated this with

proper

respect in two excellent chapters

on

fluid me-

t'oreword

chanics and

the

engineering

mechanics

of

piping

(Vol-

ume

1).

The chapter

on heat transfer

in

vessels and

piping

il-

lustrates

lucidly

the

interrelationship between

process

and

mechanical design. Every engineer

working with

in-

dustrial

process

systems

will

benefit from

reading this

chaDter.

Although the author

has made a herculean

effort

in

covering

the mechanical design of

pressure

vessels, heat

exchangers,

rotating equipment, and

bins,

silos

and

stacks

(Volume

2), it is true that there are omissions.

It

is

hoped

that,

as the author

hints

in

his

preface,

a future

volume might be added covering multiphase

flow, spe-

cific

cogeneration processes,

turbines,

and detailed

pip-

ing

dynamics.

Still, at this

writing

these

two

volumes

comprise

an

outstanding

practical

reference

for

chemical and

me-

chanical engineers

and a detailed instructional manual

for

students.

I recommend these volumes

highly

for each design

en-

gineer's professional

library.

John

J.

McKexa, Ph.D.

,

PE.

Joe C.

Waher Professor

of

Chemical

Engineering

Universitv of kxas,

Austin

vtl



Dedication

To

the memory

of

my

beloved

parents,

Aub-ri:y

tt.

Es-

coe and

Odessa

Davies

Escoe;

and

to the

dedicated

enei-

neer,

Dr. Judith

Arlene

Resnik,

U.S.

astronaut

aboid

the

ill-fated

space

shuttle

Challenger (Flight

51-L).

v||l



This book's

purpose

is

to show how to apply mechani-

cal

engineering concepts

to

process

system

design. Pro-

cess systems are common

to

a wide

variety of

industries

including

petrochemical processing,

food

processing

and

pharmaceuticals, power

generation

(including

cogenera-

tion),

ship building,

and

the aerospace

industry. The

book is based on

years

of

proven,

successful

practice,

and

almost

all

of the examples described are from

pro-

cess

systems

now

in

operation.

While

practicality

is

probably

its key

asset,

this

second

volume

contains a unique collection of valuable informa-

tion,

such as a

practical

approach to bin and silo

design

as

well

as

practical

methods

of

controlling wind vibra-

tions of stacks using vortex

strakes; new

information

on

nozzle loadings

on

compressors and turbines; compre-

hensive

discussions and examples

on

sizing

pumps

and

compressors for various

process

applications;

expanded

tube count tables

for

shell-andtube

heat exchangers;

a

practical

approach to design against tube bundle vibra-

tion;

and a comparative synopsis

of

the

various national

wind

codes.

Topics

included in

the text are considered to be those

typically

encountered

in

engineering

practice.

For

rea-

sons

of

time and space the

dynamic analyses of seismic

response spectra

and an extensive discussion

on

pulsa-

tion

response

spectra in

piping

induced by acoustic

pul-

sation

are

not

discussed.

However,

a short discussion is

given

on

pulsation

response

spectra induced by acoustic

pulsations.

Single-phase

flow is much more common in

mechanical

systems than two-phase

flow,

so

because

of

time and

space two-phase flow is

not discussed.

This

book is not intended

to be a substitute

or

a re-

placement

of

any

accepted code or slandard. The reader

is

strongly encouraged

to consult and

be knowledgeable

Preface

to

Volume

2

of any

accepted standard

or code that may

govern.

It is

felt

that this book

is

a valuable

supplement to any stan-

dard or

code used.

The book is slanted toward the

practices

of the ASME

vessel and

piping

codes and

the

TEMA

standard

for

shell-and-tube

heat

exchangers. The intent is not to be

heavily

prejudiced

toward any

standard, but

to

discuss

the issue-engineering. If one feels

that a certain

stan-

dard or code should be mentioned. olease remember that

lhere are olhe15

who

may be using

different

standards

and it is impossible

to

discuss

all of them.

The

reader's academic

level is

assumed

to

be a bache-

lor of

science

degree in mechanical engineering, but

en-

gineers

with bachelor of science

degrees in

civil,

chemi-

cal,

electrical,

or

other

engineering

disciplines

should

have little difficulty

with

the book,

provided,

of course,

that they have received

adequate

academic training

or

expenence.

Junior

or

senior undergraduate

engineering students

should find the book a useful introduction

to the applica-

tion

of

mechanical

engineering to

process

systems. Pro-

fessors

should

find

the book a helpful reference

(and

a

source

for

potential

exam

problems),

as well as a

practi-

cal

textbook

for

junior-,

senior-,

or

graduate-level

courses in the mechanical,

civil,

or

chemical engineering

fields. The book

can also be used

to supplement an intro-

ductory level textbook.

The French

philosopher

Voltaire

once

said,

"Common

sense is

not very

common," and

unfortunately, this

is

somelimes the case

in

engineering. Common sense is of-

ten the by-product of experience, and while both

are

es-

sential to sound

engineering

practice,

neither

can

be

Iearned from books alone. It is one

ofthis book's

soals

to

tx





The

engineering mechanics

of

bins

and silos

differ

from the

mechanics of

oressure

vessels

because solids

behave

differently from liquids and

gases,

both in stor-

age and

in

flow

conditions. The

mechanics

of

stacks are

almost identical

to

those

of

towers, but are somewhat

simpler.

An

engineer

has

more

fiexibility and

ap-

proaches

for solving vortex

shedding around stacks

than

around

towers,

because

stacks rarely have as many at-

tached structures.

SILO AND

BIN

DESIGN

The mechanics of

solid flow theory is a fairly compli-

cated

subject.

The

proper

design

of

silos and bins is

more

than meets the untrained

eye, and

involves

every

aspect of

engineering

mechanics.

This

chapter only

" sketches"

methods of approaching

this

complex

phe-

nomenon,

and refers

the

interested

reader to literature on

this specialty.

The field

of

solids

handling

has been augmented the

past

twenty

years

by two researchers-Jenike

and Johan-

son

[1].

The

methods

presented

in this chapter are

largely influenced

by their work.

Bins

and silos appear to be very

simple devices, but

what

goes

on inside is not

so

simple.

To design an

effi-

cient bin the

design engineer must understand why

solids

in bins

do

not

flow

(Figure

5-1):

1.

Development

of

a

rathole

or stable arch that

ceases

flow.

2. Erratic

flow-transient

arches

form

within the solid

resulting

in variance

of the

bulk

density such

that

flow becomes unstable.

3.

Fiushing-the

fluidization and

flushing

of

powders

creates

erratic flow.

The Engineering

Mechanics of

Bins,

Silos,

and

Stacks

4.

Dead

storage-residual build-up

of

solids

caused

by

the

inability

to exit bin.

5. Segregation-a

heterogenous

solid

of

varying spe-

cific

gravity

in which

the

lighter

particles

exit the

bin

first,

leaving

behind the heavier

particles.

6. Degradation-the

chemical change of solids caused

by remaining in storge

too long.

Spoilage,

caking,

and oxidation are

some examples.

Solids

behave

differently

from

gases

or

liquids

be-

cause they can transfer shear stresses without movement,

and

because of their cohesive strength, they can retain

their shape under

load. The

shear stress transferred

be-

tween the

solid

and

the channel

walls

is

a

function of

the

normal

pressure,

w. The relationship

between

the two is

as

follows:

S

1t

-

tdttrg

--

w

(5-l)

where

{'

:

kinematic

angle of friction between

the solid

and the bin wall

p

:

coefficient of friction between

the bulk solid

and the bin wall

Typical values of

@'

are

given

in

Table

5-1

for

various

solids

and

bin materials. This

table can be

used

in

appli-

cations where the

bulk

solid

properties

are

not

known

(as

is commonly

the

case). The

value

of

@'is

required by the

methods

presented

to

be

a constant value so that using

the table will

produce

a conservative

design.

There are

two

flow

conditions that can occur-mass

flow and

funnel

flow.

Mass

flow is

a flow Dattern in

which

all the material in the hopper

or bin

is

ln

motion

and the

flow occurs

along the bin

walls.

Funnel

flow

is

a

flow

pattern

in which the material

flows

primarily

in the

center resion

of

the bin.





The Engineering Mechanics of Bins.

Silos and Stacks

arch

lhickness,

T

Figure 5-2. Formatjon

of an

arch.

FR€E

SIJifACE

srREss

{q)

Mass

flow characteristics

1. Material

segregation

problems

are minimized

2.

Fine

Dowders

deaerate

3. Material

flows

unilormly

4.

Smooth steep

hopper

Figure

5-18.

Ideal

flow of

solids-mass flow.

The

strength of

the solid material

is the criterion

for

flow

behavior

in bins. Failure

conditions

ofthe solid oar-

ticles

can result in

arching. no flow. piping

(a

hole

formed

in

the solid formation),

or

limited flow Figure

5-2

illustrates

an arch formed

by

a

solid in

a

hopper.

The

failure

of the arch will

occur when

the major compres-

sive

stress,

R

equals

the unconfined yield

strength, fc. lii)

prevent

arching, the

critical

dimension,

B, ofthe

hopper

opemng

must

De

sTiEss

(L)

sti€ss

t

laLl)

CALCUIATEO

S-IRESS

I

IALL

)

Figure 5-3.

Stress

distributions along

hopper wall

[1].

per

wall.

When the hopper

angle is

less

than 30',

the

limits

of radial stresses will

occur in conical hoppers,

as

shown

in Figure

5-4.

Even though the hopper

opening is

large enough to

prevent

arching,

mass flow

piping

will

occur. The criti-

cal diameter at

which

the

pipe

is unstable is

given

by the

followine:

D

>

4\+

^l

(5-3)

_f-lJ>

'

7(1

+

m)

where

m

:

0 for slot opening

of width

B

m

:

1 for

circular opening

of

diameter B

?

:

bulk density

of the

solid,

lb/ft3

The calculated

stress and radial

stresses

are shown

in

Figure

5-3. When

the stresses

induced

between the

solid

particles

and

the hopper

wall

are

not

compatible

with ra-

dial

stress, a flow

pattern

will

not

develop along the hop-

Figure

5-5

shows

a

plot

ofthe

piping

factor, O, against

the angle

of

internal friction,

f.

The

limiting relations

for

arching and

piping

in Equations

5-2

and 5-3 are func-

tions

of

the material

yield

strength,

f".

This

parameter

can be

determined empirically

only

if

the consolidating

pressure

ol for steady flow

is known.

This

pressure

is

denoted bv

(5-2)

or

:

IBQ

(54)



Mechanical

Desisn

of

Process Svstems

z.^

E

=

-to

Figure 5-4.

The criteria

for mass

flow

when

0' < 30".

where

Q

=

2sin0

(s-5)

d

:

angle

of

hopper

slope

o

=

computed stress

function

along

the

wall

Combining

Equations

5-2

and

5-5 we obtain

1>

(r

+

-)e

(s-6)

t"

where

o1lf"

:

flow factor of solid

The critical flow factor for arching in channels

(ff)

is

o(1

+

sin

6)

represented by

n:

(?J".-*,

:

(1

+

m)Q

(s-'t)

'e_

F

I

o

z

igures

5-6-5-9

show the

values

of

ff

for

straight-

walled converging

bins with various material

properties

and

wall

slopes. These factors are

presented

as straight

lines in the

f" vs.

o1

graph

in Figure

5-10.

The consolidating

pr€SSUre

01

that the

flowing

solid

particles

exert in a vertical cylindrical channel is

ot

=

D"yG

(5-8)

300

40 50

60

70

ANGLE OF Ii{TERNAL

FRICTON IDEGREESI,Q

Figure

5-5.

Piping factor,

iD,

versus angle of internal friction,

6.



EFFECIIVE

AI{GLE

OF Ti|cNOfl IOEGf,EESI,

6

Figure

5-6.

Wall

friction

angle,

@',

versus

effective

angle

of

friction,6.

Figure

5-8. Wall friction

angle,

d',

versus

effective angle of

friction,6.

The

Engineering Mechanics of Bins, Silos and

Stacks

2O3.6070

E.rECrrE

^*GLE

OF

FitcT|Ox

roEci€Est,6

Figure 5-7.

Wall

friction angle,

{',

versus

effective

angle

friction,0.

EFFECTTVE

AXCTE Of FFICTION,6

Figure 5-9. Wall

friction

angle,

d',

versus effective angle

friction, d.

5

6ro



I

t-

(,

=

rl

E

F

(',I

ot

JI

lrJ

I

>l

o

trj

=

-

o

()

z,

=

Mechanical

Design

of Process Systems

--------)-

coNsoLroaTr

G

PRESSURE,

q

Figure 5-10.

Critical values

of or and f". Line

A

represents

strength

properties and

Line B the constant flow factor

[1].

where G

is a

function

of

the

effective angle of friction,

6,

and

the

internal

angle

of friction,

{.

This consolidating

pressure,

o1,

provides

the strength

of

the material

that

forms the

pipe

in

the bin. Combining Equation 5-3 with

5-8

we have

of

the

flow of

solid

particles.

This

pressure

is reduced

internally

somewhat

because

as the solid

particles

de-

scend through the

hopper,

a vacuum in the void between

particles

develops

and

produces

a

negative

gauge pres-

sure.

As

the particles

approach

the

outlet,

atmospheric

pressure

is obtained.

While

the

wall

pressure

is

maximum

at the

bin-hopper

tangent line in

mass

flow, it is only a fraction of a hydro-

static

pressure

for

a

liquid head equivalent to the height

ofthe

solid in the bin. Thus,

designing solid bins for

hy-

drostatic loads

results

in overdesign of

the bins. As

a

guideline,

the maximum hoop

pressure

at the bin-hopper

tangent

point

is

about seven times that of the

pressure

of

the solid induced by

gravity.

That is,

l6

P*:7{'y)*{H)ft

where

"y

:

bulk density of the solid, lb/ft3

H

:

height that solid is stored in bin, ft

(5-

10)

(+)

"

o*o

\r./

.,,,:(,1)",.""=o*o

The value of ff is

plotted

against

6 and

{

in Figure

)-l l.

Figure 5-12

shows flow

properties

of a

typical bulk

solid, which

are

quite

useful in

problem

solutions.

Thble

5-2

lists critical hopper dimensions for the material with

flow

properties given

in

Figure 5-12.

Once the

problems

of

arching and

piping

are solved

and the bin

is

designed to

handle

the solid

mixture,

the

next

step

is to examine

flow

pressures

induced by solid

particle

flow. As mentioned

previously,

solid

particles

suspended

in vertical

storage bins do

not

behave

linearly,

such as liquids.

To

a much

greater

extent than liquids,

solids

manifest shear forces between

particles

and on bin

walls. Figure 5-13 shows typical

pressure

distributions

for mass

flow

and funnel flow, and illustrates

how in

mass

flow

the

pressure

is maximum at the bin-hopper

junction

poilt.

The

geometric

discontinuity causes an

in-

crease

in flow

pressure

because

of

change

in momentum

Table 5-2

Critical

Hopper Dimensions

tor

Material With Flow

Properties

Shown in Figure 5-12

[11

Type

Critical

width

ot

a

slot

opening

lor

arching, ft

Freshly Stored

for

stored

24

hr

Flat

bottom or nonmass

flow

bins

Stainless

lined

hopper

(d,

=

30",

6"=

21.t

Mild

steel

hopper

(0'

:3o"

a'

:3s")

Critical diameter

of

a circular

opening

for arching, ft

Flat bottom or nonmass

flow

bins

Stainless

lined

conical

hopper

(0'

:

1s",0'

:27")

Mild

steel conical

hopper

(0'

:

15",

d'

:

35')

Critical dimensions

bins

5.6

7.7

+

Dictated only by

porticle

size

or

dynamic conditions.

+*

mese ralues are

the

same

as

the

flat

botrom bin

values because the

mid

steel

conical hopper

when

6' =

35" is too

rough to

proride

flor"'along

the

walls of the cone

when

0'

:

15"

(5-e)

0.2

0*

0,*

1.0

o.4

0.6

2.0

0.9

0.4

0*

0.4*

*

2.O**

CR

ITICAL

STREI{GTH

RoP(e$i{L

.

lrl

<=

ori

F

.I'

-t

aE

()C



The

internal pressure

in

Equation

5-10

can

be in-

crease.d

by

the use

of

pneumatic

air supplied

to the

bin.

In

the case

of

bins where

funnel

flow

exists

or for

small

bins

with

cohesive

solids,

supplying

forced

air

through

ducts

in

the

bin

is

desirable

to

prevent the

formation

of

arches and pipes

within

the

solid

itself.

To compensate

for

the additional

internal pressure,

Equation

5-9

be-

comes

The Engineering

Mechanics

of Bins,

Silos and Stacks

P.,":77H+Pu;

where

P";.:

air

pressure,

psig

60

e,

z

E

t40

=

o

o--

z

(s-1r

)

The

use of

pneumatic

air in

bins is

often

desirable

and

in

the

situations

where

air

cannot

be

used because

of

chemical

interaction

with

the

solids

in a closed

svstem.

nitrogen

is

commonly

used.

40

50

60

ANGLE

OF

FRICTION (OEGREES),6

Figure

5-11.

Critical flow

factor

for

piping.

Hlso

(

6'

?

1oo

3

Figure

5.12. Typical

bulk

solid

flow properties

used

to

deter-

mine

critical

dimensions

for piping

and

arching.

coNsoltDAT|NG

PRESSUAE,

q,

Lb/Fr2



Mechanical Desisn

of

Process Svstems

q,

PSI

+

0 Psl

bin fu

_

bin

haf tu

-

Figure

5-13.

(A)

Pressure

distribution

for

solid

flow

is

maximum

at

cylinder-cone intersection

primarily

because

of discontinuity

stresses;

@)

The

relationship

between

mass

flow

and funnel

flow

for conical sections.

The

angle of kinematic

friction,

d',

is a

function

of the

coefficient

of

friction

between the solid

and bin material and the compression

the solid is subjected to

in

storage.

F

I

STACK DESIGN

The analyses

of

stacks subjected

to wind and seismic

response

spectra are identical to those

methods used

for

process

towers

discussed

in

Chapter

4.

The differences

in

the

two

types

of

equipment

are

twofold:

(1)

stacks

have

different

values for logarithmic decrement

and dy-

namic magnification

factor, and

(2)

the solution

to

prob-

lems

induced

by vortex

shedding

are different.

Both of

these

factors are

a

result of stacks

having

simpler

geome-

trres.

The

simpler

geometry

of

the

stack

works

for

and

against

the engineer.

The

positive

aspect comes

as a

re-

sult

of the methods

used

to

break vortex shedding-vor-

tex

breakers

are much easier and more

practical

to install

on stacks

than on

process

towers.

The negative aspect

of

stacks

is

that

they do not have connected

piping

and

structures

to break

up

vortices and

to

damp

wind-in-

duced

vibrations. Thus, we

will

focus our

discussion

on

those aspects

of wind design that are

peculiar

to

stacks,

remembering

that the fundamental

basis of design

is

the

same

for stacks

and towers.

Vorter

Sheddlng

and

Frequency

Response

As explained

in Chapter

4,

only the fundamental

mode

of

vibration

is considered for

process

towers

and stacks.

Consequently,

the Rayleigh

method is applied

to obtain

the vibration

characteristics

of

the stack.

In

stacks,

lining

is

often

used

where high temperatures

are encountered

and carbon structural steel

is

the stack

material.

Lining must be used for

temperatures in excess

of

800

"

F because

of the danger of carbon

precipitation in

the steel.

To avoid this and

not

use

lining,

one

must use

hot-rolled,

high-strength

low-alloy

steels

that have good

elevated-temperature

properties.

Such steels are

not

gen-

erally

pressure vessel

quality

and require

heat treatment,

such

as

the

Cr-Mo steels described

in ASTM

specifica-

tions A-387

and A-542. These

low-alloy steels are

of

structural

quality,

contain

0.75-1.257o chromium,

and

are cheaper than

pressure-vessel-quality

alloys.

When common

carbon structural

steel

is

to be

used

with lining, the effect

of

gunite

lining

must be considered

with

the

mass and stiffness

to accurately determine

the

fundamental

frequency

of the stack.

An

approximate

value of the

modulus elasticity of

gunite

is 1.3

x

10opsi.

The effect of

lining in

a stack must also

be considered

with

the

flexibility

of

the base.

Table

5-3

is

a

list

of

con-

servative

values of the logarithmic

decrement

and dy-

namic magnification

factors for

various soil conditions

for lined and unlined

stacks.

For explanation

and

use of

these

values the reader

is referred to Chapter

4.

Ovaling

When slender

stacks, i.e.,

rings in which the

thickness

is

small

in comparison to

the radius, are subjected

to

vor-

tex shedding

caused

by

air

currents, the

elastic strain en-

FUNNEL FLOW



ergy of the cylinder

is distributed

in

such a manner

as to

induce flexural

and torsional

modes

of

vibration.

The

ring

is subjected to

the following

modes:

1. Extensional

(axial

elongation

and contraction about

the

ring's own

axis).

2.

Torsional

(twisting

of

the

ring

about

its own axis).

3.

In-plane

flexural

(inextensional

vibrations

in the

plane

of the

ring).

4.

Out-of-plane flexural

(inextensional

displacements

in

the

plane

of

the ring).

The

flexural

modes

are

generally

the

only

modes

of

practical

significance

since

the

fundamental natural

fre-

quencies

of

the torsional

and extensional

modes

are

much

greater

than

the fundamental

natural frequencies

of the flexural

modes.

Figure 5- l4

shows these various

modes.

The

flexural

modes,

in-plane and out-of-plane,

are

used in

determining

the resonance

frequency of

the stack

caused by ovaling.

Since

out-of-plane

flexural vibrations

are coupled

to torsional

vibrations,

it

is the out-of-plane

frequency

used ro describe

the

vibration

of the

siack;

however,

the natural frequencies

of

the

flexure

modes

in

and out of the

plane

of the ring

vary only

slightly for cir-

cular cross

sections. The natural

frequency

of

the

ring

is

siven as

,

_

I

I

Etn2(n2

-

l),

lo5

"

-

t

tpAr6t+

I

+

/t

(s-12)

The lowest

flexural mode

exists when

n

:

2 and

Eoua-

tion

5-12 reduces

to

The Engineering

Mechanics

of Bins,

Silos and Stacks

9

These relationships

were

formulated

by the

great

pio-

neers Michell

and Love

during

the nineteenth

century.

The reader is referred

to Example

5-4

for further clarifi-

cation

of

units.

In

practical

stack

design,

because

vortices

form alter-

nately on either

side

of

the stack, the

flexural

frequency

(ovaling

frequency)

given

in Equation

5-13 is

taken to be

twice that of

the

vortex

shedding frequency.

The vortex

shedding

frequency is

given

by

Equation 4-101

as

-

0.2v

'D

Now since f

:2f,

we solve

for

V

and obtain

,,

60f,D

(4-l0l)

(s-14)

"

4.4O9t

E

f'

n=l

n=2

Figure

5-14.

Stack

mode

shapes.

in which

s: the Strouhal

number

(is

equal to 0.2 for

a

wide

range of Reynolds

numbers).

The

value

of V.

is

the

critical

wind

velocity

in which

ovaling

occurs.

Both

the

vortex

shedding

and

flexural

frequencies

should

be

evaluated at

each elevation

if

ovaling rings

are

to

be used.

Norrnally,

rhe upper

third

of the

stack

is all

that is required

to be investigated,

based

on

various

wind

tunnel tests.

Now we

come to the

most

practical

aspect of stack

de-

sign-how

to alleviate

flexural

excitation.

This

can be

done

in two ways-ovaling

rings

or vortex

strakes.

Ovaling rings

are used

to increase

the mass

distributed

along

the tower to

dampen

flexural

vibrations.

When the

flexural

frequency

equals

twice

the

vortex

shedding

fre-

quency,

i.e.,

if

the

design

wind

speed

range includes

the

critical wind velocity,

V", stiffeners

are added

at those

sections where

f

=

2f.

The section

modulus

ofthe

stiff-

eners

is

given

by

-

(7

x

l0

)v:DrH,

(s-15)

where V"

:

"rnr"u,

,"r"0

velocity (Equation

5-14), fpm

D

=

internal

stack

diameter

at

elevation

under

investigation,

ft

H,

:

stiffening

ring

spacing,

ft

o,

:

allowable

tensile

stress

of stack

material.

DSi

Ovaling rings

provide

a redistribution

of

the mass

of

the

stack,

resulting

in

localized

stiffening

that tends

to

offset flexural

frequency

modes.

This

is

particularly

de-

sirable with

stacks

of

several

diameters.

However,

with

stacks

of

constant

or tapering

cross

section

the

use

of

vortex

strakes

is

becoming

increasingly popular.

(5-

l3)

s

l \

l):

j-r

i,

t



10

Mechanical Design of Process Systems

Helical Vortex

B?eaker

Strakes

For critical

wind velocities less

than 35

mph,

dynamic

stresses

should be investigated. One optimum solution

for

such stresses

in

stacks

has

been

found in wind tunnel

tests and in

practice

to be helical

vortex

strakes.

The application of helical vortex strakes to

vertical cy-

lindrical

towers

has

shown

remarkable results. The

strakes' function

is

to

break up

vortices

such that

flex-

ural frequency

modes are

quickly

dampened.

It is signif-

icant to

note that adding

the strakes

increases drag and

thus

wind loading.

These strakes

are

shown in Figure

)-l).

To

minimize the flow-induced drag and optimize the

vortex-breaking

effect,

the strake

height,

W(ft),

should

be

in the following range:

0.09D<w<0.10D

where D

:

OD

of

stack,

ft

Figure 5-15 shows a

helix

generated

on

a

cylinder

by

taking a

template z'D long

by

L high and wrapping

it

around

a cylinder. The length,

L,

of the

helix is

the

top

l/3

of the stack.

Wind tunnel

tests

have shown that

vortex

breaking

devices are most effective on the upper

third of

a stack.

The helix angle,

{,

should

fall

into the following

range:

54'<d<58"

There are always three strakes

per

stack to counter the

alternate

formation

of

vortices

on

either side

of the

stack.

Strakes can

be fabricated

from

a flat

piece

of metal,

normally

3/ro-in.

or 5 mm

thick.

Each strake is divided

up

into

a

certain number of strips, usually five to twenty

segments, depending

on

the

length of the stack. The

overall

length of the individual strakes that is divided

up

is determined by

S:[(?rD)2+L2]oj

(5-16)

where

D

=

OD

of

stack,

ft

L

:

height of tower

portion

straked

(V:

of

total stack

height),

ft

The number

"S"

is divided into individual strips that

are cut from

a

larger

piece

of

plate

shown

in

Figure

5-16.

The strips must be cut to

a

radius of curvatue,

r,

that

is

determined as

follows:

a2a2 + 8

(5-17)

. D-

wherea:

--,

lt

z

,L

2rw

r,r

:

number

of

revolutions around

stack

cylinder

made

by

helical strake

(usually

<o

:

1)

An alternative formula, developed

by Dr. Frank Mor-

gan,

and

two

to three

percent

in error of Equation 5-17,

IS

XW

1-)\

S, interior

arc Iensth

of helix

\rhefe A

=

_

:

------:--------:

S" exterior

arc length of helix

aa2

0.090s

W<0.1D

d

=

Helix angle

54o

<C358'

(s-18)

(5-le)

T

L

I

|-,D

Figure 5-15. Cylindrical

strake helix

geometry.

The value Si

is

determined

by using the outside diame-

ter of the

stack

in Equation 5-15, and S"

is obtained by

using D

* 2W in

place

of D

in

the same equation.

For

the

most accurate

results,

Equation

5-16 should be used,

as it

is

the

exact radius of curvature

of

a

helix

projected

on a

cylinder

[3].



Strips

are laid

out,

as shown in

Figure 5-16, with

an

inner

radius

of

curvature

determined

bv Eouation

5-17

and

outer

radius

of

curvature

of r

:

r

+ W.

it is

desired

that

the helix be

perpendicular

to the centerline

of the

cylinder

along

the entire

length

of

the helical

strake

shown

in

Figure 5-15.

To

obtain this

each metal

strip is

placed

in

a rig

shown in

Figure

5-17. The

rig

is com-

posed

of two

clamps,

each 45'

from

the

plane

perpendic-

ular

to the

table,

or 90" offset

from

each othe;.

O;ce the

metal

strip is

clamped-in,

a

hot

torch

is

run up and

down

the

length

of the metal

strip hot-forming

it

to

the shape

formed

by the clamps.

The

strip

should not

be heated

any

longer

than necessary

to hot-form.

The

metal strips

should

be the

same material

as

the

stack.

The effectiveness

ofthe

system

is

not

impaired

by

a

gap

of

0.005D

between

the

helical

strake

and cylinder.

This

method

leads

to ease and

quickness

in

fabricating

helical vortex

strakes.

EXAMPLE

5.1:

GRANULE

BtN DEStcN

FOR

ROOFING

PLANT

Twelve granule

bins

are

to be

designed

to

provide

granules

for the

manufacture

of

roofing

shingles

of Ex-

ample

3-6. Each

bin is

to contain

10.02

tons

of

sranules.

yielding

120.24

lons rolal

capacity

for

all

twe'ive

bins.

The client

desires

to

use an

existins

steel

frame

that lim-

its

the bin

to a rectangular

shapJwith

an off-centered

opening

as

shown

in

Figure 5-18. From this

figure

we

consider

the

first

criterion

in

bin

design-to

satisfu flow

conditions

such

that the

granules

wili

move.

The Engineering

Mechanics

of Bins,

Silos and

Stacks 1t

As seen

in Figure

5-13b,

the minimum

hopper

angle

for

mass flow

is

0

:37.74'1"

From Figure

5-4,

6'

:

l0

From Figure

5-5,

<D

=

0, which

implies

that

we

will

not

have

piping

forrning

in the bin

6=70'

For

a circular

opening,

m

=

1

^

s'(l +

sin

6)

zslnd

From

Figure

5-6,

ff

:

1.6

ff=(l+m)Q

Q:

=

=

0.80

).

or

:

7BQ

1

=

90 lb/ft3

B

:

0.667

ft

o1

:

(90)(0.667)(0.80)

:

43.6, tbrtU

From Figure

5-12,

f"

:

s0 lb/fC

(5-5)

(s-7)

Figure

5-17.

Clamping

each

strip

on 45

degree

offsets

and

hot

forming

with

torch

obtains

desired

geometry.

igure

5-16.

Strake fabrication

detail.



12


l--j*---l

t;;-

lr\l

tl \ I

/\

,.T

;l

1

E

Since 0.278

ft

<

0.667

ft

=

8

in.

archins

will

not

form in

the bin

After

flow criteria have

been met, we

proceed

to the

structural

design

of

the

bin.

The

allowable

stress used

in

the case

of

bin design is

the

ASME allowable,

since the

granule

weight forms a

pressure

distribution, thus mak-

ing the bin

walls

pressurized

components.

For simplicity

and

ease

in calculations, the solid

pres-

sure distribution exerted on the bin walls is taken to be a

simple

hydrostatic load. The

bin

walls

are

fixed on three

ends and free on the top edge. The solution for the maxi-

mum

stress

is

given

by

Thus, the critical arching

dimension is

f{o

B=

'

-

-'

:0.278

ft

r(l

+ m)

(90X2)

:

v{bt

uno

F

:

orPb

at x

=

0, z

:

0

=

*1 o'

unoF

=

orPbatx

=

+a.z

=

b

'

ifu > borz:0.4bif a'( b

Figure 5-18. Granule bin

silo.

(5-2)

(5-20)

(5-21)

In this

problem,

a

:

12.625

ft

and b

=

4.00

ft.

The

pressure

at the bottom

of

the

plate

is

P

:

eo*

(n.6zs)ttffi

:

z,ur

p,r

a 4.000

b

12.625

From Figure 5-19

we obtain

the

following:

*r

:0.030

Vz

=

0.032

The maximum

stress

occurs

at

the

bottom side at x

=

0 and

z:0

_

_

vrPb2

/<.)n\

b

:

12.625

ft:

151.50 in.

For 5.4-516 Gr. 55,

o4

:

13,700

psi.

Solving for t

in

Equation 5-20

we have

/v,pu'\o'

r:

l__-l

here Vr,

V2, 01, and 02 are

in

Figure 5-19

F

:

reaction

force exerted on

the

plate

edge normal to the

plate surface,

lb/in.

P

:

load

per

unit area,

psi

t

:

plate thickness, in.

ko.o:o)(z.ssr)

.\

(rsr.sofin.'lo,

1:l

tn'

|

:0.627

in.

I

r:,eoo

--lb-

I




of Bins,

Silos and Stacks

Deflections of bin

plate"

b

12.625

a 4.O

At

x

=

0,

z

:b

=

12.625

ft

(0.00020)(7.891)

',Jb-

1+t.oy

in.,

D

:

flexural rigidity

of

plate

Et3

12(1

-

v'?)

o

16

13

The

stress at mid-plane

is

v,Pb2

""

t2

z

:

0.4b,

P

:

4.734 psi

,

_

lro.orzlr+.rl+rr

rs

r.

sor'lo

'

=

0.502 in.

Selecting

SA-516 Gr.

70, oat: 17,500

.

_

[ro.o:otr

z. ae1)fl51.50),lo

5

I

--

0.557 in. at bottom

edge

,

_lto.ozzx+.tt+xrsr.sor,lo'

_

=

0.446 in. ar z

=

5.050 fr

Dlb

ln.

I

Figure

5-19.

Rectangular

flat

plate

solutions.



14

Mechanical Design of

Process

Systems

D

(30.0

>

106x0.562)r

:48:

.b49.25J

12(1

-

0.311

in

which,

w

:7.4565

x

10-o in.

-b

Atx

=

u,z

=

t

w

_

(0.004x7.891)(48.0),

481

,649.253

w

=

1.4913

x

l0-1 in.

For a

e/ro-in.

plate

deflections are negligible and

no

stiff-

eners are

required

for

this

plate

thickness.

Bin Stilfener

Design

To

reduce

bin

plate

thickness, stiffeners

can

be

used

with thinner

plate.

A thinner bin

plate

makes fabrication

simpler

because

a

thinner

plate

is

easier

to

weld

and is

cheaper.

With

stiffeners,

each enclosed area is

analyzed

as

a

flat

plate

with

three

edges

fixed and one edge simply

supported.

The

stress

in

the

plate

is

given

by the follow-

ing:

't,

Ph2

ob

=

'l:-:

and

F

-

QrPb at x

=

0.

z

=

0

\5-22)

t-

Itr

^Ph2

"

:

*,5o'

and

F

-

02Pb ur *

:

tJ.

z'0.4b

(5-23)

where V1,

V2,

01,

and 02 are

shown

in

Figure 5-20

.09

.o8

.o7

.05

.o4

Figure 5-20.

Rectangular

flat

plate

solutions.




of Bins,

Silos and Stacks 15

F

:

reaction force

exerted on the

plate

edge normal to the

plate

surface,

lb/in.

P

:

load

per

unit area,

psi

t:

plate

thickness, in.

First

Stiffener

Consider b

:8.0

ft,

a/b

:

0.50. From Figure

5-20

we

obtain Vr

=

0.064.

Thus,

from

Equation

5-22

we have

_

_

(0.064x

7.891)(96)2

_

11 r

.",

:

(0140

=

JJ.009.228

psi

>

al)owable

Consider

b

:

4.0 ft,

a/b

=

1.0. From Figure 5-20 we

have

i{''

:

0.192

and

from

Equation

5-22,

o^,

:

24,756.921

psi

>

allowable

Similarly,

considering

b :

2.0

ft,

o-"-

:

11,475.865

psi

<

allowable

=

17,500 psi

By

a

process

of iteration

we

obtain

a

value

ofb

:

2 ft

8

in.,

in which

o^,

:

17,364.2?9

psi

< 17,500

psi

allowable

Thus, we

place

the first stiffener 2

ft

8

in. above

the

bot-

tom

seam,

Second Stiffener

At 2

ft 8 in. the maximum

pressure

exerted on the bin

wall is

rh / rf':

\

P

-

e0; 2.62s

-

2.667\

ft

ln;-l

=

6.224

psi

Consider b

=

4.0

ft.

a/b

:

1.0 in which

Vr

:

0.192 from

Figure

5-19. Thus,

o-",

='o'n',)lu:?.', '08)2

:

19.s26.e

psi

>

17.500

psi

(0.

141)

By a

process

of

iteration we

arrive at b

=

3

ft

I

in.

in

which

o-""

:

17,502

psi

Third

Stiffener

At the new

elevation, 6.167 ft

above the bottom

seam,

we obtain

the maximum

pressure

exerted

on the

wall.

rhI

P

=

90--l

f

2.625

-

(2.667

-

3.5O)

fcl

:

4.036

psi

-(rq)

The top

portion

ofthe

bin

is

now

a

plate

with

three sides

fixed and

the

top edge free.

Thus, Equations 5-20

and

5-21

hold,

using Figure 5-18. By iteration we

obtain

b

:

6.458 ft,

P

-

4.036

psi,

a/b

=

0.619,

Vr

:0.091

and

o^":15,643

psi

o

17,500 psi

Since the maximum

stress

is less

than the

allowable

for

the top

portion,

no

third

stiffener is required.

First

Stiffener

Design

a

=

4 ft-o

in.: b

:

2

ft-8 in.

a/b:#:t.roo

/'-TR

I

lt I

Pt l.

| | tco

fi l'

,H1 |

/r

)ll

\q--7891

psi

'Yr

=

0.383

R

:

.yrpb

:

(.383)(7.891X32.0)

:96.712lbhn.

:

w

With

plate pushing

uniformly

on stiffener,

the

latter

will

be analyzed

as

a fixed

end beam

with

uniform loading.

v

u,

ffi

UV

I

STIFFENER

96.712

lb/in



16

Mechanical

Design

of

Process Systems

w/

M.*:

^-;

W=wf

24

w

:

(96.712X48)

:

4,642.18

tb

M-^.

-

(4'92

181x48

0)

=

9.284.36r

in.-rb

24

:

773.697

ft-Ib

Mc

I

For design

purposes

select

a design

stress

of o

:

17,000

psi.

With

a factor

of safety of 2. This

would

give

a

yield

stress

of

34,000

psi,

which

is conservative.

I:M"

g

Select

a

3-in.

x2-h.

a

t/+-in.

thick,

,

'

(9,284.361)

inlb(.49)

in.

:

0.268 in.a

17,000

lb/in.'?

I of .zr

:0.39

in.a

Therefore,

3-in.

x

2-in.

><

tla-in.

4

is sufficient

Second

Stiffener

Design

P

=

6.224

psi

a:

4ft-}in.;b

:

3 tt-6

in.; a/b

= :

t.t+l

3.5

By

linear interpolation,

1t

:9.349

R:

(0.340)(6.224)

lbl\n.2(42.0)

in.

:

88.879

lb/in.

pr-".

=

{:

w

=

(88.879)

lb/in.(4E.0)

rn.

w

:

4

,266

.192 lb

M

-

@

'266

'1921(48

'ol

=

Rs1?.3E4in.-rb

A-","".

I:M"

q

Select a

2rl2-in.

x

z-in.

x

tl4-in.

4

,

_

(8,532.384)

in.Jb

(0.54)

in.

_

.,

".,,

,-

o

rtun

-

l?soo

rbfinj

I

:

0.37

in.a

Therefore,

Ztlz-in.

x

2-in.

x

t/+-in.

4

is sufficient

Stiffener at

Junction

Point

ot

Bin

P

Hoop

Force

=

-

yD

From

data

provided

by the

client,

P

=

400 lb/ft'z at

junc-

tion

point.

Using a

factor of 7 we have

P

=

7(a00)

=

2,800

lb/ft2

P

:

2,800 rb/rt

(r-lq)

=

re.zt44

psi

UseP

=

20

psi

For bottom

plate,

a:4

ft-0 in.: b:2 ft-8

in., a/b

=

1.500

rr

:

0'383

R

:

(0.383X20.0X32.0)

=

245.

r20 lbl in.

w,

M.*:

=-:

w

=

(245.t20X48.0)

=

11.765.760

Ib

1.+

M

_

(

I 1.765.760X48)

:

1"t

slj.520

injb

A_.-'--

Select

a 3rl2 in.

x

3

in.

x

tla

in.

4

I.in

:

(23,531.520X0.79)

_

1.094

in.a

17,000

I

=

1.3 in.a

for

section

Therefore,

3rlz-in.

x

3-in.

x

r/+-in.

r

is sufficient

with

long side

facing

bin



:

55.92% of minimum yield

:

30.36% of ultimate

yield

The Engineering

Mechanics of Bins,

Silos and

Stacks

0. | 825(6.31 3X50.928)'

-

(0.438)2

'-'----'''

Therefore, use 716

in.

f,

for bottom

plates

Bending Stress in Bottom Portion

From

previous

information,

P

=

6.313

psi

on triangular plate

A

:

area

of

triangte

=

Ia'20'lro.z*>

=

t0.40

ftj

-

\21

ot A

:

1,497

.589 in.2

=

F

:

(6.3 13)

lb/in.,(I,497.589)

in.2

:

9,454.279

lb

at3

:

(4.244)(12)13

:

16.916

in.

M,

:

F(a/3)

:

160,495.84

in.-lb

Mc

I

thJ r/{O Otl\3

r:-=-:[,007.49Er

l,/.

Iz

17

Bottom Portion

of Bin

Bottom

portion

of bin will be approximated with

four

tdangular

plates

welded together,

as shown

in

Figure

5-18.

Pr

=

7.891 lb/in.2

pz

:

e0 lb/n3(16.50 ft)

[-]q144

:

10.313

psi

At an

angle

ot90o-0:37.7474,

P

:

10.313

sin 37 .747"

=

6.313

psi

By linear interpolation,

B'

:9.3659

o:41

=o.rszs

2

0.1825Pa'z

o=

,,

,

qan

=

l/,JWPSl

m.l8r5x6.rl3x5o.%y

=

u.4rJ

rn.

\l

17.s00

with

t:3/E

in.,

,'

_

0.1825(6.113)(50.928),

=

,t ,4q s?? nci

" - (0it5,- - '''-"

-J-

YJ'

For

SA-516 Gr. 70, minimum

yield

:

38,000

psi

11

*ll

->l

I

Ptt

ll

-tl

ll

--'l'

Y

CROSS

SECTION CUT AT

MIDPLANE

OF

TEIANGLE

--tt

t

La-

t-ll

It-

It;

-il

1$

rJ

-_tt

%

yreld

:

%

yield,

:

21249.532

38,000

21249.532

70,000

with

rhe

in.

f,

,



18 Mechanical Design

of

Process Systems

,^

50.928

atJ:_=lD.y/orn.

3

(1r,007.498X

,

_

(160.495.84)

in.-lb

(16.976r

in.

_

.,.,,,,

'

-

(r

1,00?/98xr?J00)

i"rlb/i"r

-

"

"'-'

Therefore,

tlrc

in.

t_

is

sufficient.

Vessel

Supports

Consider all trusses as

pin

connected.

Side

Truss

End Truss

For three horizontal plates,

(r2.62s

itx8.0 rt1

=

1 30-1f

-

z,3ts.22JIb

'2

or

for

three

plates,

wt

:

6,945.669 lb

For simplicity and

to

keep things conservative, let us

analyze the

internal

plate

to determine if we need any

supports on

inside

of

structure.

For two outside

plates,

t

:

3/8

in.;

wt

:

(12.625)(8.0)(0.375)(1,14)(.283)

:

1,543.482 Ib

wto'.r

:

3,086.964 lb

For

two

side plates,

Wtt"d

:

2(3,086.964)

=

6,173.9r,

tO

Under

Bins-4

Triangular Plates

For each bin,

/a

qor

\

A

-

4 l-

'"'l

A.244\tt44\

=

5.990.355

in.1 of metal

\21

wt of each bin

-

(5.990.355)(.283)

=

1.695.270 lb

(

160,495.84)(16.976).

o"u

:

17,500

psi

m.

weighr of internal load

:

(t20.24)

lz'z+o

v\

'on'I

,on

/

:

269,337 .60 tb

Weight

of

steel

(Wt):

(12.625

ftx16.0 ftXt)

:

ro*r

:

0.283 lb/in.3

wt:

(16,362.0)(0.283)

(1s

1.s0)(192)(0.s63)

16,362.0 in.3

:

4,630.446 tb

i:\\:-j

w rblfr



Number

of Bins

:

8

Therefore,

Wtrorur

:

13,562.164

lb

Empty weight

of structure

:

4,630.446

lb

+

6,945.669 lb

+

3,086.964

lb

+

6,173.928

lb

+

13,562.164 tb

:

34

,399

.r7

|

Ib

Wt of

granules

=

269,337

.60 lt:

Total

wt

loaded

:

303,'736.7711b

Total

number

of internal plates

:

4

Total

length

:

4.0 ft

w

-

303'739 771

..

75.934.r93

lb/rt

4.0

:

911,210.313

lb/in.

Considering

the

plate

in

Figure

5-18,

rur

Y,

w

.

(9

.210.313r

lb

92.1 ,n.

E

in.

:

174,952,380.1

lb

M -

(174952'380

1)(192)

=

4

rqx

x\7

r)l,n

-rh

8"

Therefore,

bin

must have

internal

supports

under bot-

aom.

Number

of vertical

supports

=

9

=

R

:

303

'73-6'771

9

=

33,748.530

tb

Number

of

ioint

suDDorts

:

9

tl

tol 716 ?71

F

:

--"'

_-:j____:

:

20,249.118 lb

IJ

The

Engineering Mechanics

of

Bins.

Silos and

Sacks

19

The frame structure

shown in Fieure

5-18 is analvzed

as continuous beams in

the longitudinal

and lateral direc-

tions.

FoR EACH

spAN

wL:

lzsss.+rglli

[+.olrt

:

so,g73.ozo

ro

RA

:

0.393 wt

=

0.393(30,373.676)

=

11,936.3tt

,O

RB:

Ll43

wf:

1. 143(30,373

.676)

=

34,117.rt

b

Rc

:

0.928

wf

:

0.928(30,373.676)

:

28,186.77t

tb

Ro

:

1.143

wf:

1.143(30,373.676)

:34,717.rt

rO

Solr ing

for

reacrion\ in

lateral

plate

FOR

EACH SPAN

WL= 30.373.676

lb

v.*

:

0.607(30,373.676)

tb

V-*

:

18,436.821

lb

RB

=;

(10.373.676X2)

=

37,967.0q5

lb

6

Ra

=

ft.

=

11,390.129

lb

Design each

support

column

for

37,967.095Ib

=

38,000 lb

srde

saructure

The

bin

structural

detail is

shown

in Figure

5-21.



20 Mechanical

Design

of

Process

Systems

BIN

JUNCTURE

DEIAIL

STIFFENER DETAIL

EXAIIPLE

5-2:

HIGH.PBESSURE

FLARE

STACK

DESIGN

A

high-pressure

flare

stack shown

in

Figure

5-22 is

to

be

designed and construcred

to the

following

specifica-

trons:

Base diameter

:

l0 ft

Height from bottom

of steel

base

to tip of

flare stack

:

200

ft

Gas

pressure

in stack

=

2

psig

Gas

temperature

=

100oF

Design

wind velocity

=

100 mph

Maximum

gas

flow

rate

:

300 MMscfd

Earthquake

design

:

World

Mercali

6-7

Figuie 5-21. Bin

struclural

frame detail.

Effectlve

Diameters

Add

12

in. for

platforms

and

12

in. for

ladders.

Add

4-2-in. d

lines.

2-in.

g

dia. line

:

2.3'75

in.-Add

t/z

in.

insulation

D

:

(3.375X4)

:

13.50 in.

D"^"".,

=

2(12)

+

13.50

:

37.50 in.

De

:

42

+

37.50

:

79.50

in.

DB

:

90

+

37.50

=

127.50 in.

Dc

:

120

+

37.50

:

157.50

in.



Height Wind Pressure

(fD

P,

(rb/ftr)

w

=

B

x De x

Pz

lb/tt

The

Engineering Mechanics

of

Bins,

Silos and

Stack

Wind

Load

Moment

(5,270.98X110.5

+

2.5)

+

(2,862.0)

x

(90.0

+

2.5)

+

(10,404.0)(65.5

+

2.5)

+

(13,604.25)(24.2s

+

2.5)

0-30

26

:

to6)( f)tz6):20415

:

ro.olffit:3):25e.88

:

toor(lle)o

:2ee.2s

:

too(l#J(44):34650

:

,0.u,(]?Za)r*)

:

28o.so

/r

r:

so\

=

t0.6tl'-'""1t48t =

306.00

\

12

/

:

ro.orfifJt+t):

reo.8o

:

<o.orit#)or)

=

202.'73

30-40 33

866.25

40-74 38

74-76.5

44

'16.5-125

44

125

159

48

159-t74

48

174-200

51

Wind Load

s.270.98

2,862.00

r0,404.00

Moment

(s,270.98)(13.0

+ .0) +

-0)

+-

o,404

(5,270.98)(28.0

+

x

(7.5

+

34.0)

+

llrl

\21

44 ft-lb

76 ft-lb

(5,270.98)(62.0

+

48.5)

+

(2,862.0)

x

(41.5

+ 48.s)

+

(10,404.0x17.0

+

48.5)

/an s\

+

(13.604.25)

|

-'l

=

1.851.388.35 fr-lb

\2l

l5

34

(1

(2,

(')

.0)

862.00)

169,052.

86rn)

I34.01

\)

I

622,44r.

Figure

5-22.

High-pressure

flare stack;

unless otherwise

indi-

cated,

all dimensions in feet,

design wind

speed

:

100 mph.

51 PSF

48

PSF

__

159

_t

-l

1.

44

PSF

,rO-l

30_

38

PSF

33

PSF

26 PSr

t3,6U.25



Wind Load

22

Mechanical

Design

of

Process Systems

(5,270.98)(113.0

+

34.0)

+

(2,862.0)

x

(92.5

+

34.0)

+

(10,404.0X68.0

+

34.0)

+

(13,604.25)(26.'75

+

34.0)

+

(866.25)

-

.-.

-

/:+.0\

x

.25 +

34)

+

00.174.5)

l:-jj:

I

\'2 I

:

3,228,045.06

ft-lb

2,598.80

(5,270.98X147.0

+

10.0)

+

(2,862.0)

x

(126.50

+

10.0) +

(10,404.0)(

102.0

+

r0.0)

+

(r3

,6O4.2s)(60.7

5

+

10.0) +

(866.25)

x

(35.25

+

10.0) +

(10,174.5)

/,,r.0\

x

(r7.0

+ t0.0) r

(2.s98.80)

l+-l

\2

t

:3,672,858.86

6,142.50

(5,270.98X157.0

+ 30.0)

+

(2,862.0)

x

(136.50

+

30.0)

+

(10,404.0X112.0

+

30.0)

+

(13,6M.25)(70.7

5

+ 30.0)

+

(866.25)

x

(45.25

+

30.0)

+

(10,174.5)(27.0

+

30.0)

+

(2,598.80)(s.0

+ 30.0)

+

(6,142.50)

For

Section

D

*

(0'875

-

0 125)

:

o.oo6

>

0.00425

120

_

(0.56)(0.006x29.0

x

109

""

tl

+

(0.004x29.0

x

109(30,000)]

:

20,02i.918

psi

For Section C

:

d

(0750

-

o l25)

:

o.oo5

>

o.oo425

120

(0.56X0.005)(29.0

x

109

li

+ 0.004(29.0

x

109/(30,000)l

:

16,684.932

psi

For Section

B

13

d

o,

t"_(0.625-0.125)

d

90.00

o.

=

20,021.918

psi

For Section

A

:0.006

:

d

(0.500

-

0.12s)

=

0.009

o.

=

30.032.877

psi

Section

Weights-Uncorroded

Weight

Section A

i3o'oJ

:

s,

r:t,+rr.zo rt-ru

Allowable Shell

Buckling

Stress

0.56 t"

E

'.E:29

x

'

d

(1

+ 0.004

E/y)

'

''

,[l/€)'-litt\'l',

r

=

(0.2833)

j:

(37.0)(

12)

'n.

'

[\,

/ \2 I

)

-

8,199.69

lb

Section B

wr

-

(02813)

{

rzoo,rz,'

"

[(T)'

- (*,

)']'"'

y

:

30,000

psi

106

psi;

=

45,340.61

lb



Section

C


of

Bins,

Silos and Sacks

Section

A

wt

=

(0.2833)

--ll(30.0)(r2)

in.

"

[(9'

- (r94,)]'"

:

33,397

.r9

tb

Total

wt

:

128,966.580 lb

Required

t

Thickness

16

D" Mr

r(D"'?+Dr'?)@"+D)oE

r(D. + D)oE

Section D

(16)(120.OXs,

138,419.76)(12)

'n

n

[('r),

(ry

ro]l_.,

wr

-

(0.2833);

(44.0X12)'n.

[\, | \

2

l)

:

42,029.09

lb

Section

D

(16)(42

.0)(169

,0s2

.44)(12)

rl

+

(42

+

@D2l(1.2.0

+

41.0X30,032,877)(1.0)

8,199.69

r(42.0

+ 41.0X30

,o32.877)(l

.0)

t.

:

0.052 in.

=

1/z

in.

[ ,

OK

for

buckling

Anchol

Bolt

Design

Try 24-11+-rn.

d

anchor

bolts

=;

dec

:

l2o

+

2(2.50): 125.00

Total tension in each bolt

:

Wn

*,

=

ottl

''] ;01?,tu'

-

l? 'e 6

58

-

76,84r.ros lb

-

24(125.00)

24

oe

:

40,000 psi

A.

-

76'841 109

=

|.921

in.2

<

1.980

in.']

"

40.0(n

:

l3/+ in. dia, 8-thread series

Check

[/av\

1

:

t-wl

^

t\d/

I

AR:-

-

No,

r[(120)'?

+

(1

18.25F](120.0 +

I 1

8.25X14,

182.

19X1.0)

128,966.580

r(120.0

+

118.25X14,182.19)(1.0)

t,

:

0.381 in.

+

7r in.

[

,

OK

for

buckling

Section

C

(4X12X5,138,419.76)

(12s.50)

-

128,e66.58]

r(120.0

+

,

=

0.245

in.

-

I

18.5)(16,684.932X1.0)

t/q

iI^.

'll_

,

OK for

buckling

(16)(120.0X3,672,858.

86)(12)

r1r20)'?

+

(1

18.5F1(120.0

+ I I

8.5)(16,684.932)(1.0)

95,569.39

(24) (40,000)

Ar

=

1.913 in.'?

<

1.980

in.'?

Bearins

pressure

=

P-

=

48Y

+

W

:i- 7rl:in.

nDu'

j

r

Drj

"

^

48(s, r38.419.76)

t28.966.58

'

r( 125.00)'/(7.50)

7r( 125.00X7.50)

Pt

:7\3.734

psi

< Fb

:

1.33(900)

:

1,197

psi

Tr

:

Base

fl

thickness,

T1

:

compression

I

thickness

Section B

(16)(90.0X1.8s1.388.35)(l2l

rl

+

(90),

+

(88.75f1(90.0

+

88.7s)(20,021.918)(1.0)

53,540.300

r(90.0

+

88.75)(20,021.918)(1.0)

t,

=

0.183

in.

.r

:/s-in.

[

,

OK for

buckling

t"

:

"

(;oiltJ

;

e

=

B *

C

:

Z3tqin.

+

Zttcin.

:

5.5o



24

Mechanical

Design

of

Process Systems

Il,lr r

l

rarl

"t

Te

=

(5.50r

l:j;;:;=l

=

1.800

in.

I

zu.uuj

I

After

K:

:0.151

one iteration,

1

-

[

:twu)

o

l'''

[:1zo.r+r.roenorl''

''

[4(20.000)el

[

4(20.000X5.5)

I

1+

(61,789.8ss)

(10x1,096.373)

1.268.836

psi

B.ownell and

Young Base

f,

Method

Bolt

circle

d

:

125.00 in.

Base

P

:

4

-

125.00

-

212.50\:

130.00

in.

After six

iterations,

K:0.178

f"

=

n

E'

--

lo(1.096.373)

:

t0,963.73

-Eq

t",^

,"8)(t25.0)

+ 7.001

fc,-o,.area,

=

(1.0e6.373)

[46.,rr'l,,rr.*,

I

di

:

130.00

-

2(7

130.00

-

1

(7.00)

:

116.s0

1t.

16.00

:

7.00

in.

-

,. ,..,,

[:r

t,26t.

sto)1"'

-

''

"'

t

,o"ooo

I

:

2.181

in.

(without

gussets)

Using 24

gusset

f,'s,

gusset

spacing

is

-r

lr(

O\

h

=

""-"

"'

=

32.725

in..

|

=

A

=

5.00

in.

t2

n

5.00

b

32.'125

From

tble

4-8,

using linear interpolations,

My:

-

O.467fcrt2

My=

-

0.467

(1,268.836)(5

'00f

:14'813.660

in.-lb

r,

_

l(oJ{l+.6rr.oou)l

=

2.10g in.

-

t

20.000

I

t

=

use

2rls in

base

Brownell

and

Young

External Chair

Design

r

5.00

22

b

:

gusset spacing

=

32.725 in.

t

15

For

|ta-in.O

bolts,

e

-:"

:

t.375

in.

2

with fc,".,

:

1,000;

K

=

0.333

fc(Bc)

:

(1,200)

[

:

559,723.403

A

1.980

in.2

(12)

r,=-

=

-

U.UOI ln.

'

z'd

?r( 125.00)

I

^"

\r/2

I JI. I

L4

-

^1^-l

2(0.333X12s.00)

2(0.333x125.00)

+

7.00

:

1,106.925

psi

For

K

=

0.333;

c"=

1.588;

C,

=

2.316

z:0431l.

j

=

0.782

:

6l,789.855

Fc

=

559,'723.403

+

128,966.58

:

688,689

983

t:

:

7.00

-

0.061

:

6.939

in.

(5.138.419.76)

-

(128.966.58){0.r'''

ll25

00l

.,"\

12

/

688,689.983

r6.e3e

-

(10)(0.06rI

($Q)<r.sasr

rc-

=

1,096.373



PB

:

max. bolt load on upwind side

:

fsAB

=

(10,963.73)(1.980)

:

21,708.185

lb

\r.

-

2r.708.185[

(t

+

0.30),n Izrs.ool

I

* ,l

4r

t

[z'(L375)J I

=

3,612.549

ir..-lb

.

_

[{6x3612.549)lr" _,

"^.

t

15,000

I

=

lr/+

in.

in.

f,

for

compression

ring

Calculation

of Gusset

f,

Thickness

for

Compression Rings

r,2

[ =

4qr2-r2

=].

withk=%(1.250)=0.469

t2

,

=

[ereL,

]

:0

r35 in.

h:G+H

I21h n.

:9

+

lt/+ in.

i

2t/c in-

=

l2tlz

in.

h

12.500

P

Bolr

Load

r

0.135

18,oo0rtr-Ptt-

htP

:0

I,500

18,000(5.00)t63

-

(21,708.

185)k,

_

(12.500),(21.708.185)

_

0

1,500

qt

-

0.24ltS

-

0.025:0

q

=

0.40

in.

=

r/z-in.

f,

is OK

Skirt-to-Base

Ring

Weld

,:

(#ft)*

(";)'0u."

The Engineering Mechanics of Bins,

Silos

and

Stacks

fw

:

1.33'yn(0.55), for wind or earthquake

fw:

i.33(20,000X0.55)

:

14,639.99

5 154 1)1

Weld size

=

-.'-

---

=

0.396

14,630.00

or

0.396

=

0.198+

Va

in.

minimum

weld each

side

'2

Cantilevel

Vibration

'.

:

(,aJo

o

.

(,$n',

=

5 860

rt

Corroded Stack

weisht

lttt

*,^

:

6nluurf(,sl

-

(91

:6,16'".ze4tb

*,"

:

<arr

oezrl(r1)'

- (r91

:36,323217

tb

*,:,ounn

rl(?l-

(rtl]I|l:

35,o6oe6o,b

23,905.217 tb

101,457.688

lb

Lc

:

8.00 +

5.00: 13.0 ft

r^ rlnn

=

U.UtJ

< U.5

L

200

trl

=

4 4:1688=

=

14.773

<

2r)

LD,2

(200X5.8601

Therefore, vibration

analysis r,?as,

be

performed.

Wa

:

101,457.688

lb,

L"

=

200

-

lr.O *

ff

:

193.50 ft

-

t.648

L?

1.648(

193.501

'

:

5r(ET,

=

o.gaOx:sJt-lItc

:

r'v)) seconds

I

t:

,uB

:

0.511 cps

vc

:3fDrr:3(0.s11X5.860)

:

8.983

mph

,

_

[r+xs.

r:a.+

ro.76x

r2)]

r28.966.58

r-[

--;6

20"0) -l

t

"(t20-00)

=

)'tv+

tzl



26

Mechanical

Design of

Process

Systems

V:o

:

100 mph

v*

:

(roo)

(*J'"'

:

13r.165

mph

Maximum

gust

velocity

:

13 1.

165(1.3)

:

170.5 l5

mph

k

_

0.0077D,5E

_

(0.0077X5.860)5(29.0

x

106)

L..Ws

(

193.50)1t

0 t.457.688.)

:0.002

<

1

15

Therefore,

the

stack

is

free

from

cantilever vibration.

Static Deflection

^ ^

tt.467V

"f

.-{.,o-

2

_

(

1.0X0.00238X

1.467P(8.983,,

:

0.107

psf

2

,

:

(,$,o.to

-

0.r25)

+

(,$,o.uro

:

0.523 in.

,"

=

(_..)(?

* 0,,,)

*

(,z,|(?..,")

:35.285

in.

i+r

\/

+r

r

o.zs

\ / ss

\/ss.zs

+ o.zs\

'=

\'oo/\

,

/.

\-/l ,

/

:

34."111 ir.

r

:

it{zs.zts),

-

(34.71

lfl

in.a

:

77,307.326 in]

-

P"D,(LF(12)j

D.=

-

-

.

8EI

_

(0.207X5.

86oX l93.so)4(t2f

8(29.0

x

106)(77

,307

.326)

Dynamic Deflection

Using a

magnification lactor

of

30.

6

:

0.164

(30)

:

4.915 in., which

is

permissible

Ovaling Vibration

Natural

frequency of free

ring

:

t

^

7.58r.(E)o

5

''

6oD2

At

42-in.

dia

:

3.50 ft,

r

-

7 58(0'3zs-iq9

0-l-]06t

:

20.826

cps

60(3.50F

fu

=

vortex

shedding

frequency

0.2v

0.2\66,

:D=(35;=r'l/rcPS

2f"=7.54t

.

t,

At

90-in. dia

:

7.5 ft,

-

7.58(0.5x29.0

x

106)0

5

t.:

'

t

:6047

cps

"

60(7

5)2

-'

f

_

0.2(66)

_

|

%n

/.f

2f,:2

(1.760)

=

3.520

<

6.047

cps

At

120-in. dia

:

10

ft,

f

_

7.58(0.625X29.0

x

lfff5

=

a)\).^<

'"

-

6o(to0l--

where

the

uplift load

on each

bolt,

F, is

-

4(5,

r38,4r9.76X

r2)

r0r,457.688

t,:

,2a1us.0ot

a

:

tt't6t

'tztD

_

0.2(66)

10

:

1.320

cps

2t,:2.640

<

4.252

At

bottom

section,

f,

_

7.58(0.751(2?.0

x

t06)05

=

5.102

cps

'

60(10.01

i,:o2t66t=1.320cos

'10

2f,:2.640

. t.rO,

Therefore,

stack is free

from ovaling

vibration.

AIICHOR

BOLT TOFOUE

Anchor bolt torque

on stack bolts is handled exactly

like

tower anchor bolts

as discussed in

Chapter

4.

Using

Equation

4-66 and considering lubricated

bolts we

have

T:CDFi

(4-66)

=

0.164 in.



.\hich

results

in a required bolt torque of

r

:

(0.

rs)

(r.75)(77

,987

.312)

=

20.471.67

in.lb

=

1.706 ft-lb

Use 1,706

ft-lb

torque

with

lubricant

grease

Fel-Pro

C-

5,A,,

or equivalent.

The

skirt base and anchor bolt

detail for the stack

is

hown

in Figure

5-23.

Design

Summary

Static

wind

shear

at base

=

22,355.110 \b

Static wind

moment at

base

=

1,299,115.509

ft-lb

Dynamic

wind shear

at base

=

22,844.841 lb

Dynamic

wind moment at

base

=

1,308,916.974

ft-lb

Total

deflection

at top of original

tower

:

4.418

in.

Total

deflection

at

top of modified

tower

:

5.898

in.

Base

plate

thickness:2lle-in.

plate

Compression

plate:

1l/4-in.

plate

,

16) l:/+-in.

anchor bolts

Required

anchor

bolt torque:

1,710

ft-lb

Total

operating weight

=

128,966.580 lb

EXAMPLE

5.3:

STAGK VORTEX STRAKE

DESIGN

An exhaust

stack 126 ft tall is

to be Drovided

with

heli-

.'al vortex

strakes. The length of the

stack

to be

straked

is

the

top

portion

31

ft

6

in.

long. Cornpute the radius

of

iurvature

of

the

strake

to

be

cut

from

flat

olate. Refer-

ring

to Figure 5-15 we have

the following:-

D:ODofstack:7ft4in.

L:31 ft

6

in.

D

7.333

.i

=

_

=

_

:

J.DO/

L

31.5

_

:

.t

t

{

2ro

2rtl)

\ou,

_ _

a2cu2

+b2

-

--;F-

_

_

(3.66'7)2(r)2

+

(5.013F

-

(356?X1t-

:

=

10.521

ft

(5-17)


of Bins,

Silos

and Stacks

ALL

MATERIAL

TO

BE

SA-285

_C

ALL WELD

SIZES IN

INCHES

zl-tci

I

a-tgg-'et.

a THBEAo

sERtEs

BoLTS

BOLTS

TO

STRADDLE

CEI.ITERLINES

trL_u--l

ffi

Figure

5-23. High-pressure

flare

stack base

support detail.



28

Mechanical

Design of Process Systems

Check

Using the approximate

Morgan equation

we have,

Si

:

interior

arc length

:

[(rDJ'?

+

L2]0

5

:

39.025

ft

52:exterior

arc length

=

[[?r(8.667)]'?

+

(31.5)'?105

:

41.637

ft

x:9:t?'o,T:r.nt

(s-1e)

s.

41.637

\w

r

:

._________

(5-lg)

r

-

(0

937)(0

667)

-

9.966

ft

=

9

ft

i

t.594

in.

|

-

0.937

_

10.521

-

9.966

va

e:,rof

=

ff

=

5.276Eo

errol

The final

product

is

shown

in Figure 5-24.

EXAMPLE

5.4:

NATURAL

FREOUENCY OF

OVALING

HING

FORIIULA

IMICHELL

FORUULA}

To use the

Michell

equation

(5-12)

dimensional

analy-

sis

must

be

applied

to

obtain Equation

5-13.

The

original

Michell

equation

is

as

follows:

, I rtJrJ

-

'J

f.

=

-r/

.--.-

.

-.

]-

(5-12)

''

2"Y PAf

(n'

+l+/)

where

p

:0.283

lb/in.3 for steel

A:

(t)

in.

x

(1)

in.

f

:

in.a

E

:

lb/in.2

T-

I

-

;

.

per

unit lenpth

ofring.

in.'

t2

z

:

l/r for

steel

I

z7f

BASE PLA?E

- 3/16r

STRIPS

CUT

FRO}I

BASE PLATE

t

0.5ft

+

0,66?ft

=

11.-2.

-

4.409r

E

Ir

:

----

Vt1

386 lb.-in.

rgl

'i-c'

1(36)r(in4)

(0.283)

-.l l

1

in.2 1 in.a

(5.333)

(5-13)

Figure 5-24.

Manufactured strake elements.



NOTATION

A

:

cross-sectional

area of stack,

in.2

AB

:

anchor

bolt

area,

in.2

a

:

stack

radius

=

D/2,

ft

B

:

critical

arching

parameter, dimensionless

D

:

critical diameter

at which

piping

is unstable,

di-

mensionless;

internal stack diameter

(Equation

5-15), ft;

outside

diameter

of

stack

(Equation

5-16),

ft;

dynamic magnification

factor

(Thble

5-

JI

E

=

modulus of elasticity.

psi

f.

:

material

yield

strength,

psi

ff: critical flow

factor for arching

in channels,

di-

mensionless

f,

:

natural

frequency

of

a

ring,

Hz

f"

=

stack

vortex

shedding

frequency,

Hz

G

:

consolidation

particle

parameter

(Equation

5-8),

dimensionless

H

:

height

that solid is stored

in bin, ft

H,

:

stiffening

ring spacing,

ft

I

=

moment

of

inertia of stack

cross section,

in.a

L

:

height of tower

portion

straked,

ft

m

:

geometric

parameter

for arching

(Equation

5-2),

dimensionless

n

:

flexural

mode

(Equation

5-12), dimensionless

P"1.

=

air

pressure

(Equation

5-11),

psig

Pn**

:

maximum

hoop

pressure

at bin-hopper

tangent

point,

psi

r

:

outside radius

of

stack

(Equation

5-12),

ft;

n-

dius

of

curvature

of vortex strake

(Equation

5-

17),

ft

S

=

over-all

length of vortex strake

(Equation

5-16),

ft

Si

:

interior arc

length of helix

(Equation

5-18), ft

So

:

exterior arc length

of helix

(Equation

5-18),

ft

-

S.

=

section

modulus of stiffeners

(Equation

5-15),

ft'

t

:

shell thickness of stack,

in.

V

=

wind

velocity, ft/min

V"

:

critical wind

velocity

in

which ovaling

occurs

(Equation

5-14),

fum

w

:

width

of

strake,

ft;

normal

pressure

applied

on

bin walls by solid

(Equation

5-1),

psi

XI

] )

mode

shapes relating

translational

displace-

7.1

ments about

the

x,

y

and z axes,

respectively

Greek

St/mbols

7

:

bulk density of solid.

lb/ftl

6

=

logarithmic decrement, dimensionless

The Engineering

Mechanics

of

Bins,

Silos

and

Stacks

29

Table 5-3

Conservative

Values

for Logarithmic Decrement

and

Dynamic

Magnification Factor tor

Various

Stacks

Low

Oamping

6D

Average

High

Damping

Damping

6D6D

Unlined

Stacks

0.035

90

0.052

60

0.105

30

Lined

Stacks

2"

gunite lining 0.070

4"

gunite

lining

0.117

10

9

31

25

45

27

0.100

0.r25

0.300

0.360

Inw

Danping

=

rocky,

very

stiff soil;

Iow-stressed

pile

suppon, or struc'

tural

Itame

support.

Average Dampin?

=

modetutelt stiff soil; aormol spreadfooting

or

pile

sup-

port

HiBh Damping

:

soft soil;

foundation

on highlJ stressed

Iriction

piles

6

:

piping

factor, dimensionless

0_:

ungle

of

hopper slope, degrees

0

:

modal

shape

relating

to

rotation about an axis

perpendicular

to stack centerline

(Figure

5-14),

dimensionless

p

:

coefficient of friction between the bulk solid

and

the

bin

wall

(Equation

5-1),

dimensionless

d' :

kinematic

angle

of

friction

between the

solid

and

the

bin wall,

degrees

dr

:

consolidating

pressure

for

steady flow

(Equation

5-4\,

tbflft2

ot

:

allowable tensile

stress

of stack material,

psi

or

:

number

of

revolutions around

stack

made

by

a

helical

strake, dimensionless

REFERENCES

1. Jenike, A.

W.,

Johanson,

J.

R.,

and Carson,

J. W,

Storage

and

Flow

of

Solids, American Institute

of

Chemical Engineers, New York, New

York,

1981.

2

.

Blevins

,

R.

D

.

,

Formulas For Natural Frequency and

Mode Shape, Van Nostrand

Reinhold

Company,

New

York. NY.

1979

3.

Thomas, G.

B.,

Calculus and

Analytic

Geometry,

Addison-Wesley Publishing Co., Inc., Third Edition,

1960.





Fluid movers

and

their

use are vital to the

process

in-

dustries.

This

chapter focuses

on two

basic

types-

pumps

and compressors.

The sizing of these

units and

their interaction with the other components

of

a

process

system

are discussed.

This chapter does

not

address

the

detailed mechanical design

of

sophisticated equipment,

such as

turbine

blade design

and

gas

dynamics

in a

tur-

bine. That type of material is a separate

field

of study

and lies outside this text's

objective of examining

how to

select and apply rotary bquipment to

process

systems.

For

further

reading,

see

the

bibliography

at

the

end

of

the

book.

PUIIPS

As the

primary

movers

of liquids,

pumps

come in

many

types

and an understanding of the

various kinds is

essential

in

successfully applying them

to

process

sys-

tems.

Pumps

are used to transfer

liquids

from

one

point

to

another.

They basically

fall

under two categories-cen-

trifugal

and

positive-displacement.

The centrifugal

pump

gets

its name from the fact that the

pump's

impeller im-

parts

kinetic energy to the liquid with centrifugal

force

acquired

by the impeller's

rotation.

This simple

mecha-

nism

allows

the

centrifugal

pump

to

be

practical

for

high

capacity,

at low to medium

heads.

The

aspect

of low to

medium heads will

be discussed shortly.

Typical centrif-

ugal

pumps

include mixed

flow,

propeller,

peripheral,

and turbine.

Positive-displacement

(PD) pumps

trap a

quantity

of

liquid

and force it out

of

the

cavity

against the

pressure

of the

discharge by

means

of

rotary

or

reciprocating ac-

tion.

Ideally, a PD

pump

will

produce

whatever

head is

impressed on

it

by

the

system

restrictions to the flow.

Rotating Equipment

Not all

PD

pumps

are

purely

rotary or reciprocating, but

we

will

focus

our

attention

on

these

types. PD pumps,

by

definition,

deliver

fluids at a

rate

proportional

to

the

speed

of the

pump

action

and this rate

is

independent of

the

pressure

differential

across the

pump.

For this reason

means must be

provided

to

limit

the discharge

pressure

and this

will be discussed under

the

section

of

positive-

displacement

pumps.

Typical rotary

positive-displace-

ment

pumps

include screw,

gear,

vane, cam, and lobe.

Reciprocating

positive-displacement

pumps

include

pis-

ton,

plunger,

and diaphragm.

Selecting

the type of

pump

to

use

is

a

function of the

service to be

handled.

Sometimes, the selection

is obvi-

ous;

for

example,

if

you

wanted

to

pump

molasses,

you

would

choose

a

positive-displacement pump.

In the situ-

ation

where neither a standard type

of

pump

is used

for

the

service, nor

is

it

obvious

what

type to use, a

centrifu-

gal

pump

is always considered

first.

The

reason

for

con-

sidering a centrifugal

pump

initially

is because

of its low

initial cost, economical cost

of maintenance, wide range

of

materials of

construction,

and

relatively

large clear-

ances. Factors to be considered in selecting a

pump

are

as follows:

1.

Efficiency

2.

Net

positive

suction

head

(NPSH)

required

by

pump

3.

Operating

costs

4. Shaft speed

5.

Magnitude of

clearances

6.

Materials

of

construction

7.

Fluid

service

to be handled

8.

Availability

and

delivery

time

of

pump

The type of

pump

to

be used

for

a

specified

service

or

duty

can be selected from Figure 6-1. This figure clearly

indicates

how

different

pumps

have

overlapping

charac-

31



Mechanical

Design

of

Process Systems

10

ro-

F

o

I

J

I

-l

5

teristics.

Depending on the

relative importance

of the

previously cited criteria, a certain

type of

pump

will be

selected.

Figure 6-1

will

help the reader determine

from

a

quick

glance

what type(s) of

pump(s)

will

be required.

Gentrifugal

Pumps

Centrifugal

pumps

are the

most widely

used because

of their

wide

operating

range and the

reasons

previously

cited. These

pumps

come in

a

vadety of

types,

depend-

ing

on the type

of

impeller, casing, stuffing

box,

and

bearings.

These components

are shown

in

Figure 6-2.

The

radial type

impeller is by far the

most common

centrifugal

pump

in the

process

industries.

The

flow

is

directed

by

the

impeller imparting

motion on the

fluid,

driving

the

fluid

to the

periphery of the impeller.

This

allows

the

velocity head to be converted

mostly to

pres-

sure head

in

the volute.

The

mixed flow

pump

impeller

consists

of

vanes

dou-

bly

curved

or

screw-shaped so

that

the impeller

moves

the

fluid

by both centrifugal

and

pushing

action. The

re-

sult is a discharge

of

axial and

radial flows.

The

axial

flow

pump

impeller develops

head by a

lifr

ing or

pushing hydrodynamic action that

results

in totally

axial

flow on discharse.

234

Figure 6-1.

Pump selection

guide.

The impeller

is

hydrodynamically

balanced to ensure

minimal

vibration. The casings can

come in a

variety

of

designs,

but

are

either

vertically

or

horizontally split.

A

vertical-split

casing implies that

the casing is bolted to-

gether

along a

vertical

plane.

Similarly, a

horizontally

split casing

is

bolted

or connected

along a

horizontal

plane.

The advantage of the

vertical split casing

is

that

the

pump

is

supported along

the shaft allowing

for ther-

mal

movements

without

causing

shaft

misalignment.

Packing

and seals

on the shaft are the

most common

source

of

failure

for a

pump.

In low-pressure

applica-

tions,

soft or

metallic

packing

will

suffice in a stuffing

box.

In most low-pressure

applications, a single

seal

will

usually suffice.

When

pressures

exceed about

50

psig

and there

can be

no tolerance

for

leakage, a

double

seal

is utilized.

These seals come

in various configurations-

tandem.

bellows.

and face-to-face.

When

process

conditions

get

severe enough,

a

double

inside-outside seal,

where

part

of

the seal is outside

the

stuffing

box, is

used. The disadvantage

of

this type

of

seal

is that

not

all

stuffing

box

arrangements

allow such

a configuration.

For

proper

cooling and

lubrication the

seal must be

supplied

with a

fluid,

called

a seal flush.

Figure

6-3

shows such

a system.



Group

I

Standard

Pump

'Pafls

10rtra'y

sl0ck.d by cLsrome.lor e4erqenc/

'Ppd

rs

"Trrd.name

ol lnternanonal Nrrel Coooanv

(A)Nor

avarable In Recessed h0eller

pumps

(BlNor

avr'abre

In Seri

Pnmno

oumoe

(Cr

\or

rva ubre on

4x3

LS.loii

4d

US

I3

o'

614 US

l3A

rcast

sleel suotntuledr

{01Jackeled cover oral€s are carhon sre€l

(E)

Used

n

Packed PonPs

only

{t)

Trtanrum

Dumos

havs

GraJor

rmpell€.

oaskels

Cdro,r

b a

reo'9ercd

lraoe

name

or

un'on Carbrde Coro0 anon

lGr Allov

rs

B7 Sio. Duclilp

lron

rnd Crlbon Sleel

oumos

{H)

Icd€name ol E

I

Duponl deNamoors

&

ComDafiy

Inc

G:oup

ll

and

lll

Standard Pumps

Materials

Common

io all

Alloys Unless

otherwise

Noted

Parl No. Parl

Malerial

104 lmoeller Gasket'

107 Rear Cover

Plate Gasket*

Durabla

108 Bearing

Housing

Adapler

Casl lron

109 Bearinq flousrno

Fool

Casl lron

111

Gland Studs or

F

anqe

Studs

with

Hex

Nuts

3M

S.S./303 S.S

112

Sealcaqe'(E)

PTFE

113 Molded

Rino Packinq'rE)

Kevay'il

114

Inboard 0ellector

PTFE

115 Casino Studs/Hex

Nuls

304 S.S./316

S.S.10

118

Inboard

0ilSeal'

TFSB

119

Bearina Housing

Cast

lron

120 Inboard Eearinq'

Sleel

121

0utboard

Bearino'

Steel

122

0ilSlinoer

Steel

123 Bearino Cover

Cast

lron

124 Bearing LockNut

Steel

125 Bearin0

Lockwash€r

Steel

126 Beaino Cover Gasket

Cork

127 Bearino

Shim'

Steel

129 outboard

oilseal'

TFSR

130

Shall Couolino

Kev

Steel

131

Beanno

Housrng Adapler' 0"

Binq SBR

132

Soherical

Washer

lor Foot Steel

133

Trico 0iler

(nol

shown)

Steel-Plaslic

134 Bearinq

Housino

Venled Drarn Plu0 Plastic

136

Cao

Screw

for Foot Steel

138

Cap

Screws

lor Eearinq

cover Steel

139 Machine Eolts lor

Bearing Housrng Steel

140 CaD Screws

iorAdarterto Cover Sleel

Figure

6-2. Centdfugal

pump

components.

(Courtesy

of the

Duriron

Company.)



Mechanical Design oI Process Syslems

The various types

of seals

are shown

in

Figure 6-4.

The

pump

manufacturer should be relied upon for the

choice of seals. Sealing technology

is

a subject vast

enough

to encompass this book and the reader is referred

to Buchter

[1]

for

additional

sources.

Bearings, like seals, are for the most

part

the main

re-

sponsibility of the

pump

manufacturer. In all situations,

the

bearings should be

of

the outboard type

(not

sub-

jected

to

the process

fluid),

unless

situations prevent

this

type

of

arrangement.

Hydraulic

Bequirements

of Centrifugal

Pumps

In this section the reader

will

find

it

advantageous to

refer to Chapter

1

. The

most important hydraulic

param-

eter

in

pump

selection is the net

positive

suction head

(NPSH),

which

is the

total

pressure

at the

pump

suction

point

minus

the

vapor

pressure

of

the

liquid

at the

pump-

ing temperature.

NPSH

is

the energy that

forces the

liq-

uid into the

pump,

and

is

expressed

in foot-pounds of

en-

ergy

per pound

of

mass

(normally

referred to as

feet

of

head) or

pounds per

square inch

of

absolute

pressure.

When

values of

pressure

are

expressed in feet of liquid,

the theoretical

height

to

which a

liquid

can

be lifted

at

any

temperatnre

is

the difference between the

atmo-

spheric

pressure

and the

vapor

pressure

of

the

liquid

at

that temperature.

Figure 6-5 helps simplify the calcula-

tion of

the NPSH.

Figure 6-3.

A

seal flush configuration.

(Courtesy

of the

Durametallic

CorDoration.)

In

selecting a

pump

the engineer must

refer to

the

per-

formance curves

the

pump

manufacturer

prepares

for

each

model ofpump. Most

performance

curves

are

plots

of flow

capacity

(gpm)

of

water versus break horse-

power

or

total dynamic

head in

feet.

Such

a curve

is

shown

in

the examples that follow.

As

seen,

the

effi-

ciency curves are

plotted

with

various lines indicating

impeller

size and the

NPSH required

at

various

points.

In

reading

the

performance curves,

it

is

emphasized

that the

extreme

right side

of

the curve should be avoided, be-

cause the capacity

and head change abruptly. Pumps are

normally

selected

to

operate in the area of high effi-

ciency.

The danger in selecting a

pump

on the extreme

left

is

that

at low flows

the

pump

horsepower

overheats

the

liquid.

If

low rates carmot be avoided, a

by-pass may

be

required to

prevent

vaporization and subsequent

pump

damage.

Thus, vaporization of the

pumped

liquid

can occur two

ways:

(1)

the

NPSH required is not being

met

and

cavitation

occurs

in

the

liquid

causing

vapor

bubbles that can severely damage

the

impeller or

(2)

the

pump

horsepower overheats

the

pumped

liquid, forming

vapor

bubbles

that can

(and

normally

will)

damage the

pump.

Excess heat

resulting in

pumping

a fluid can

be

avoided

by

determining t}re minimum

flow

required to

allow

proper

heat dissipation. At low flow

rates

or

shut-

off conditions,

heat is transferred to the

liquid contained

in the

pump

casing

at

a rate representing

the

power

losses

of the

pump.

The

power

loss is the difference be-

tween

the

brake

horsepower consumed and

the water

horsepower developed.

The remnant energy in the

pump

bearinss

and that

lost

to convection

to the outside atmo-





36

Mechanical

Design

of Process

Systems

Pump Hydraulic

Design

Calculation Sheet

Liquid

Viscosity

at P.I

(Pumping

Temp.)

Vapor

pressure

at PT

Sp.

gr.

(7)

at

PT.

psra

gpm

gpm

gpm

Flow at ambient temD.

Operating flow at

PT.

Design

flow at

PT.

_

Suction

Source'pressure

Static

(+

headx- lifi)

=

-

APr line loss

Suction

pressure

-

Vapor

pressure

NPSH

avail

NPSH

avail

NPSH req'd

[,lin

NPSH

avail

>

NPSH

req'd

+

2

fl

psra

psi

psi

psra

psra

psra

ft

ft

Terminal

pressure

Static

(head)(lift)

-

APr

discharge

Piping

system

Other

Discharge

press.

-

Suction

press,

TDH

TDH

Discharge

psia

psl

psi

psi

psia

psia

psra

leet

'lnilial

press.,

e.9.,

ATM

or

O

unp at Duty condition

ono"

_

(gpmXTDHXr)

_

(3,e60Xr)

66o.,"

=

(gpm)CrDHXr)

(3,960Xri)

TDH

=

discharge

press.

-

suction

press.

4 =

pump

efficiency,

o/o

PT.

=

pumping

temperature

@

Onp at Maximum Capacity

Condition TDH

=

total

dynamic

head

sphere

is negligible.

The

temperature rise

per

minute is

computed by the following

relation:

42.2(bhp,")

(6-1)

W*Cp

where

At

:

temperature rise

per

minute,

oF/min

bhp,"

:

6.u1" horsepower

at shut-off

W*

:

weight

of liquid in pump,

lb

Co

:

specific

heat of liquid in

pump

The break

horsepower

of the

pump

is

given

by

..

OH"y

bhp

=

-,::--r

(6-2)

J,vou4

Figure

6-5,

Pump

hydraulic

design calculation

sheet.

which is the

power

required

if

the desired head

at the re-

quired

capacity could be

produced

with

zero losses.

For

water

flowing

through the

pump,

conditions be-

come stabilized

and

the temperature

rise is determined

by the following:

".

_

(bhp

-

whp)

2,545

m

(64)

where 2,545

:

Btu equivalent

of

I hp-hr

ir

:

mass

flow rate- lb/hr-

Another variant

of

Equation

6-4 that

relates

the tem-

Derature

rise

to

the

total

head

is

(6-5)

In Equations 64 and 6-5 the compressibility

of

water is

neglected.

To

prevent

overheating of the

pumped

liquid, a bypass

piping

arrangement is used

to

have

the

pump

operating at

full

capacity. Such an arrangement

is

shown

in

Figure

6-6.

It is

always desirable

to

pass

the

bypass

liquid

^

=

ô(;-,)

here

Q

H

"v

q

:

flow

rate,

gpm

=

total

head,

ft

=

specific

gravity

=

pump

efficiency

(fraction)

The water horsepower is

given

by

who

:

QHI

'

3,960

(6-3)



through

an

intercooler to cool the

fluid

before it enters

rhe

pump

suction

port.

Under no circumstances

should

the

bypass

line connect directly from the

pump

discharge

to the

pump

suction.

So

faq we

have

not considered

the

pumping of viscous

liquids.

For

a

liquid

that has

viscosity

greater than about

10

cp,

a

viscosity correction

must be made,

because the

pump

motor

must

work

harder to

pump

the

fluid.

All

pump

manufacturers'

pump performance

curves

are based on

pumping

water.

To

correct

for the

pumped

liquid's viscosity, Figures 6-7 and 6-8 are

used to ap-

proximate

the equivalent water

performance.

The fig-

ures, developed by the Hydrauiic

Institute, are used

by

entering

the bottom

with the viscous flow

rate

(gpm),

moving vertically upward

to

the desired

viscous head

(head

per

stage for multistage

pumps),

then

moving

hori-

zontally

to

the

left

or

right to the viscosity line, and

pro-

ceeding

vertically

upward to the correction-factor curves

for the head and capacity. The equivalent

water-perfor-

mance values

are then obtained by dividing

the viscous-

performance

values

by

the correction

values. Thus, the

pump

selection

can

be

made on

those

ratings established

for water. The efficiency of the

viscous

liquid

pumping

conditions

can

be calculated

using

the

efficiency correc-

tion factor multiplied by the

pump

efficiency

for water.

In this manner the viscous

performance

of the

pump

can

be

determined using the

manufacturers'

performance

curves, which

are always

based on

pumping

water. This

procedure

is illustrated in the examples later

in

this chap-

ter.

Positive

Displacement

(PDl

Pumps

Positive displacement

(PD) pumps

are usually selected

after it

has

been determined that a

centrifugal

design can-

Rotating Equipment 37

not

meet the requirements.

Thus,

PD

pumps

are used

where centrifugals

cannot

operate-under low

NPSH re-

quirements

or

handling

a

highly viscous liquid.

There

are

several types

of

PD

pumps,

as

previously

mentioned,

and

their

positive

attributes are that they

1. Operate at

relatively

high efficiencies when

pumping

viscous liquids.

2.

Operate under

low

NPSH conditions

and

produce

high suction

lifts.

3

. Operate with

high

heads at

a

wide

range

of

capacities

.

4.

Have

a

wide speed range,

which

is limited

by the

liq-

uid's

viscosity.

5.

Are inherently self-priming.

Selecting

the

fype

of rotary

pump

is

primarily

a

func-

tion of

cost

and the

particular

requirements that

are

to

be

met.

1.

Vane

ptmps-normally

have a capacity

up to

about

380

gpm

and operate

by trapping liquid

within vane

differential

pressures,

usually at around 50

psig.

The

practical

limit

on viscosity is approximately 100,000

SSU.

Vane

pumps

are subject to

wear

and

should

not

be used

with

a

liquid

that

has

poor

lubricating

quali-

ties.

2-

Gear

pumps-normally

are

used

up to about

1,000

gpm

and can handle liquids

with viscosities up to 5

x

106

SSU.

These

pumps

operate at approximately

1,200

rpm with

liquids

of

10

to

500

SSU viscosity

(see

Figure 6-9).

It is

desirable to

have

internal tim-

ing

gears

and bearings since only one shaft

sealing

area

is required.

A variant of

a

gear pump

is shown

in

Fieure

6-10.

INT€RCOOLEA

Figure 6-6. Excessive

heat

build-up

is

often

caused

by operat-

ing

pumps

at reduced

flow

rates. To

prevent

overheating the

pumped

liquid,

it

is advisable to

pass

the liquid through an

in-

tercooler before

it

enters the

pump

suction

port.



Mechanical

Design

of

Process

Systems

l

00

.90

.ao

.70

.60

.50

_40

.30

.20

."n,

B

S9

vrscoslTY-ssu

'.

s

s

u";*t*

g;*1"

I

15

20

25 30 40

50

60

CAPACITY.GALLONS

PER I\4INUTE

(At

bEP)

Figure

6-7. Performance

correction

chart for viscous liquids.

(Courtesy

of

the

Hydraulic

Institute,

Cleveland,

Ohio.)

o

z

.icF

CP

.\$

?p

r_':

\9,

'6

rd

^

3cP

g

1s"

Hp

Zro

o o

-co

g



Rotating Equipment 39

ol

fil

-l

v,

l(

o

F

()

[>l

z2l

ogl

trol

HEI

t

ol

8el

l

FI

gl

o-l

5l

gt

<l

FI

:l

uI

o-l

rr

lrl

I

:l

<l

lrl

I

-l

4

6 810

15

CAPACITY

IN lOO

GPM

Flgure

6-8,

Ferformance

correction chart for viscous

liquids.

(Courtesy

of

the Hydraulic Institute,

Cleveland,

Ohio.)




Figure

6-9. This

drawing of

a

rotary

gear pump illustrates the


principle. The

fluid

is

captured

in

the

gear teeth and displaced

to the

suction

port.

The crescent

acts

as

a seal

between

the suction

and discharge

ports.

An

applica-

tion of

this type

of

pump

is illustrated

in

Example

6-2.

3.

Screw

pumps-these

pumps,

depicted

in

Figure 6- 11,

are

used where

large flow

capacities,

4,000

gpm

and

3,000

psi,

are

required. Screw

pumps

can

handle

vis-

cosities

up to 10

x

107 SSU and

have bearing

and

timing

gear requirements

sirnilar

to

gear pumps.

Screw

pumps

come

in

various designs,

and one

type,

shown

in

Figure

6-12,

can

handle

highly

viscous,

non-Newtonian

fluids such as

glues,

molasses,

tar,

asphalt,

and wastewater

with

ease.

Positive

displacement

(

PD)

pumps

come

in a

vast

vari-

ety and

you

should

refer to the

manufacturers'

literature

to

best

determine

the selection

of

the

particular

pump

to

be

used.

However,

PD

pumps

are

sized

very much

like

centrifugal

pumps,

and the

calculation

sheet

in Figure

6-5 can

safely

be used

for

sizing

either

type.

Pump sizing

is focused

upon

here to illustrate

the

various

ways in

which

a

pump

may be specified.

Figure

Gl3 shows

vari-

ous installations

for a

pump.

Some

properties

and

char-

acteristics

illustrated

in

Figure

6-

13

are

Static

suction

lfi-the

vertical distance

in feet

(ex-

pressed in

psi)

between

the

liquid

level ofthe

liquid

to

be

pumped

and

the centerline

of

the

pump

suction

port

when the

pump

is

located

above the

liquid

level

of the

'

liquid

to be

pumped.

Static

suction

head-the

vertical distance

in

feet

(ex-

pressed in

psi)

between

the

liquid

level

ofthe

liquid

to be

pumped and the

centerline

of the

pump

suction

port

when

the

pump is located below

the

liquid

level of the

liquid

to be

pumped.

Figure

6-10.

The internal

bearing

gear pump

is

a variant

of

the rotary

gear pump in Figure

6-9.

(Courtesy

of

Worthington

Pumps,

Mccraw

Edison

ComPanY.)

Friction

head-the

pressure

(psi)

required

to

over-

come

frictional

resistance

of

a

piping

system.

Velocity

head-expressed

in

psi,

see

Chapter

1.

Tbtal

suction

/r/-the

total

pressure below

atmo-

spheric

(in

Hg

or

psi)

at

the

pump

suction

port

during

pump

operation

and equals

the

following:

1. Static

suction

lift

plus the

frictional

head,

or

2.

Frictional

head

minus

the

static

suction

head

(only

if

the frictional

head is

greater

than the

static

suction

head).

Total suction

head-the

total

pressure

(psi)

above

at-

mospheric

at

the

pump

suction

port when

the

pump is op-

erating

and

is equal to

the

static

suction

head minus

the

frictional

head

.

Static

discharge

head-expressed

in

psi,

is the

vertical

distance

in

feet between the

centerline

of

the

pump and

the

point

of liquid

discharge.

Total

discharge

head

(TOH)-the

sum of

the

frictional

head in the

discharge

line

(discharge frictional

head)

and

the

static

discharge

head.

Tbtal

static

head-the

difference

between

the

static

discharge

head and

the static

suction

head or the

differ-

ence

between

the static

suction

lift and

the static

dis-

charge

head.

Toial

dynamic

head-the

sum

of

the

total discharge

head and the

total

suction

lift or the

difference

between

the total

discharge

head and the

total suction

head'



Et=

P:;

'6'y

-o

E.i

aa

E

E

E

'T

I

35

,^.c

q

.=

.:..

-o

0)

9q)

.E

AE

F.q

.?3H

;6o

ao

c

s

.2

o

(L

:

r

".

33

r_d

?E

&:

49

E0-

6:

r

*

xE ;i

P:

EP

E=

=;

F

i

o

r].1

o

;

oo

€

.=o

9?

E=

o

-

PU(J

thJ

rDt

5.s,b

F>\

DDq O

x

rE F

=';

I

ai dE 6 crt

gl'"

dd

E

'i-

-oi

d)

6--

E9g

$Eg

'EE

q

P H:1

:..6

=. eb

.

o

CDY)

(.)c,

3t*

*(5

.g

E

i:

-oo

dz

;-F

B

o-

t

.F

9E

b5

",

o.;

-o

: \

d

9

o.:

i -P I

E.EE

E';e

qIb

E9

EE3

s

g

=

=E

-

3

il

bX-

9

b;d

9=Y"t

o

I

cg

E:0i

il

(sYE:,

E

Xe.d"

=;=

6-d



Mechanical Design

of

Process

Systems

t|'r$|lhF..Dl$hra

Figure 6-13. The

principal

parameters

of

pump

selection.

(Courtesy

of

Viking

Pump

Division, Houdaille

Industries,

Inc.

)

Figure

6-12.

A

cavity

screw

pump

is

ideal

for

handling

higbly viscous

non-Newtonian

liquids.

(Courtesy

of

Moyno@

Industrial

Products,

Fluids

Handling

Division,

Robbins

and

Meyers,

Inc.)

When

using PD

pumps

where

a

suction

lift

is required,

remember

that

the theoretical

height

to

which

a

liquid

can be

lifted

at any temperature

is the difference

between

atmospheric

pressure

and the vapor

pressure

of the liquid

at that temperature, when

both values of

pressure

are ex-

pressed

in

feet

of liquid.

However,

the suction lift

practi-

cal

for

actual

pumping

installations is

somewhat less

than

the

theoretical

value.

Figure

6-14 shows the theoret-

ical and

practical

suction

lifts for

water.

Also, remember

that

the

higher

the installation is

above sea level, the

lower the vapor

pressure,

and the lower

the

maximum

suction lift.

Application

of

PD

pumps

to

practical

installations

is

given

in

the examples. The unit conversions

included in

Appendix

D

are

helpful in

pump calculations.

Pressure

Protection

For PD

Pumps

By

definition, a


pump

transfers

fluid at a

rate

proportional

to the speed of displacing

ac-

tion and this rate of transfer is independent

of the

pres-

sure

differential

across

the

pump.

Thus, means must

be

provided

to

limit

the

pressure

and the

pump

discharge

side should

the

discharge

piping

become restricted

or

blocked.

There

are

various methods

used to

prevent

overpres-

sure:

1. Install

a relief

valve

at the discharge of

the

pump

with

the

relief

valve

discharge being

piped

back to the

pump

inlet in

which

an intercooler

is

placed

in

the

line.

Such a

configuration

is shown in Figure

6-15.

In

such

an arrangement a temperature

sensor

device

is

placed

at the

pump

discharge to detect excessive tem-

peratures.

The intercooler,

or heat exchanger, is

used

to cool the

pumping

fluid.

Normally, temperature be-

comes a

problem

when

the

instantaneous discharge

and

inlet flows

are equal.

Gear and

multiplex



(plunger,

diaphragm,

and

piston)

pumps

are examples

of

such

pumps

in which

this situation occasionally de-

velops.

2.

Place

a

pressure

switch

in

the discharge side of the

pump piping,

interlocked

to shut

off the

pump

driver.

Since pressure

switch

set

points

are

not

as

reliable

as

relief valves,

a

relief

valve

must

be added to the dis-

charge

piping

and set at a

pressure

slightly

greater

than the

pressure

switch

to

ensure adequate

protec-

tion. The relief valve would

be

piped-up

similarly to

that shown in

Figure 6-15.

3. Install a torque{imiting

device

in

the

pump

driver

when

a

relief

is not

practical,

such as slurry

service.

A torque{imiting

device can come in

the

forms

of

a

shear-pin

or

torque

limiting

coupling. These devices

have

advantages other than protecting

the

system

against overpressure;

they

protect

the

pump

against

foreign material

or

whenever

the

pumped

fluid

might

tend to solidify.

Overpressure

protection

is

essential

in

positive-dis-

placement

pumps.

Relief valves

applied should be

added

to

the

discharge

piping

itself, because built-in

relief

valves

on the

pump

that are not removable

for testing

are

undependable.

COMPRESSORS

The

three types of compressors

used in the

process

in-

dustries

are centrifugal, reciprocating, and

axial

flow

compressors. Like

pumps,

depending

on the application,

the

type of compressor is roughly

a function of the

gas

capacity, action,

and discharge

pressure.

Figure

6-16

shows

the operating ranges

of

the

three basic types

of

compressors. As

clearly shown, one

type of compressor,

despite

its disadvantages

or advantages compared to

other types,

is

usually the obvious choice.

Reciprocating

compressors

are

normally

used

when

a

relatively low

flow rate is required,

but high discharge

pressures

are expected.

This

situation is common

in

the

gas processing

industry

where high

discharge

pressures

are

needed

for

process

conditions. The

need

and use

of

reciprocating

compressors

is unavoidable

in

many

pro-

cess system

applications.

Centrifugal compressors

are

the most

common typ€

in

hydrocarbon processing

plants

and

are

to some extent the

workhorse

of chemical

process

compression needs.

There

are four basic

advantages a

centrifugal compressor

has over a reciprocating

compressor:

1.

Lower

initial capital investment.

The cost advantage

is

increased

as the

power

demand is increased.

Rotating

Equipment

Figure

6-14.

The

theoretical and maximum

recommended

suction lift for water

at

various

temperatures,

'F.

(Courtesy

of

Viking Pump Division,

Houdaille

Industries, Inc.)

(B)

Figure

6-15, A temperature

switch can

be used in lieu of an

intercooler

(heat

exchanger)

in which

the switch can shut off

the

pump

driver when

liquid temperatures

become

excessive

as

in

(A)

or

can be

used with an intercooler

in

(B)

to

divert

flow

through

the exchanger.

In

either case,

a

pressure

safety

valve

should be used on

discharge.

(B)

assumes

the

suction

temperature is constant.

To prevent

overheating

on low flow

rate

conditions,

a flow switch is

often

used.




of

Process

Systems

2. Lower

operating

and maintenance

cost. The operat-

ing and maintenance

cost

of

a centrifugal is approxi-

mately one-third that

of

a

reciprocating

compressor.

3.

Compactness of

size. Centrifugals

occupy less space

and make

much

less noise.

4.

Simplicity of

piping.

Reciprocating

compressors

can

cause

severe

pulsation

shock response

in

piping

sys-

tems. The

cost

in

preventing

the effects of

pulsation

in

piping

systems

can

entail many hours

of

engineer-

ing and a healthy

capital investment for

either

analog

or

digital

simulation tests.

Centrifugals do not have

this

problem.

Axial-flow

compressorc

operate at

greater

capacities

and are

often

used

in

series

with

centrifugal

units. Axial-

flow compressors are

governed

by the

same

formulas

that apply to centrifugals.

The axial units

are more

effi-

cient

than

the centrifugals,

but

the latter

have

a

much

wider

operating range.

Axials are

used

primarily

for

clean

gases

such as

air,

because

they are much more

sus-

ceptibie

to corrosion, erosion,

and deposits than

centrif-

usals.

INLET

FLOW,ACF

souRcE:DriroPLot{

t2l

Princlples

of Compresslon

The

general

gas

law

that applies

to

all

gases

can be

written

in several forms:

PV

=

zmRt

(6-6)

zmRt

mw

:

zM.Rt

PV:

zRt

.c"c"

K=---j=

c"

cP

-

1.986

(6-7)

(6-8)

(6-e)

where

P

:

absolute

pressure,

psra

V

:

volume

of

gas,

ft3

z

:

compressibility

factor

for

real

gases

(z

:

1 for

a

perfect

gas)

R

:

R/mw

:

gas constant

of

the

particular

gas

R

:

universal

gas

constant

:

1,545 ft-lbr/lb.

mole

-

'R

t

=

absolute temperature,

'R

:

'F

+

459.7

m

=

mass

of

gas,

lb-

mw

:

molecular weight

of

gas

Mo

:

number of moles of

gas

:

m/mw

v

:

specific

volume

of

gas,

ft3llb.

A very

important

gas

property

is

the specific

heat

ra-

tio, k.

This

property

is determined from the following:

(6-10)

where

C,

:

specific heat at constant volume, Btu/lb.-mole-

=

4.97

Btu/lb,-mole-"F

for ideal monatomic

gases

Cp

=

specific

heat

at

constant

pressure,

Btu/lb,-mole-

:

7.00

Btu/lb.-mole-"F for

most

diatomic

gases

Reverslble Adiabatlc

(lsentropic)

Compression

The reversible

adiabatic

(isentropic)

compression of

an

ideal

gas

is

obtained when no heat

is

added

to, or

re-

moved from,

the

gas

during

compression.

The

process

is

reversible

when no

friction exists.

The

formulations dif-

fer for

a

perfect gas

versus a real

gas.

Perfect Gas

(z :

1)

PrV,K

:

PtYtx

(6-11)

(6-12)

lgure 6-16, Approximate ranges of application for

recipro-

:

l&F

cating, centrifugal,

and

axial-flow

compressors

[2].

tr

\Pr/



t'

lP:l

"

t'

\Pr/

Real

Gas

(z

*

1)

P1V17

:

P2V2'y

(6-13)

(6-14)

where.

for any

system

of

units

P

:

absolute

pressure

V

:

volume

or specific

volume, v

k

:

specific

heat

ratio

-y

:

isentropic exponent

for

real gases,

Co/Cu

t

:

absolute temperature

subscripts

I

and

2

denote inlet and discharge

conditions,

re-

spectively

To determine the exponent,

T,

real

gas properties

must

be used. These

properties

can

be

obtained

from

gas

property

charts and

used

in

the

following

formulation:

r / \'l

I-I

=*l'*,lSll

(6-15)

-y

JCp

[

\atloj

where

J

:

mechanical equivalent

of

heat

:

'778

ft.-lbrl

Btu

/,el

rate

of

change

of compressibility

facror.

z.

with

re-

ll-l

:

spect

to the required temperature.

t. along a constant

\

d[,l

p

pressure,

P.

path

To determine

a mean value of the isentropic exponent

for

a

real gas,

?,

over

a

compression

range, Equation 6-

15

must be

solved by iteration.

In Equation

6-15 if we

have

a

perfect gas

in

which

l=l

=0andz:1.0

then Equation

6-15 becomes

Rotating Equipment

where

Q

:

gas

flow rate

in standard

cubic feet

per

minute

of

gas

(60"F,

14.7

psia)

P,

:

absolute

pressure

at suction,

psia

Pd:

absolute

pressure

at discharge,

psia

t.

:


at suction,

oR

?

JCP

'v=k

For

a compression

ratio PzlPr <

2.0,

t

is

approxi

mately equal to

k

for most real

gases.

For

isentropic compression

of an ideal

gas

the

theoreti-

cal horsepower

requirement is

as

follows:

_2"*24

:

mean comoressibilitv

factor

z.

:

compressibility factor at

suction

za

:

compressibility factor

at discharge

For a

gas

capacity of

Q

:

100

scfm, Equation 6-16

becomes

[,,,-, ],

h._

=6.42llPdl

k

_rl{r,l_

k-r

l\P,/

-l\520/-

kt

I

In applying

these

formulations

that deal with the isen-

tropic

compression

of

an ideal

gas,

efficiency factors

must be defined in order

to

apply the

equations to

real

world compressors. These efficiencies

are

as

follows:

4"

:

adiabatic

efficiency

:

the isentropic horsepower,

hp1, delivered

by the

actual

horsepower delivered to the gas,

or

hpr

gnp

qn

:

mechanical efficiency

:

the ratio

of

the

actual horsepower

delivered

to the

gas

to

the brake

horseDower.

or

shp

''''

bhp

4,o

:

overall adiabatic

efficiency

:

the

ratio

of

the isentropic horsepower,

hpr, for a

stage

of compression to the brak€ horsepower,

or

hD"

bhp

.

In

defining the

horsepower

input for

a single stage of

compression,

utilize the overall

efficiencv as follows:

R k-1

(6-17)

(6-18)

(6-19)

[,

'*-'

],

hp':ffi|('iJ

,

-'l['',J'

\k/t

t

6'6,

bhP=*=ffi[(,t-']F;'{*)

\

k

/r

(6-20)



Mechanical

Design of Process

Systems

For bhp

at

100

scfm,

Equation

6-20 becomes

bhp=ffiH=-'l'H

\-o

/1

J

The

isentropic

energy

transmitted

to the

compressed

gas

in

ftJb/lb-

of

gas

represents

the

adiabatic head, or

t \t ,[,,*-,

I

",:

llsl{IlllSlT

-

'la

mw/

\K-

r/

[\Ps/

I

(6-2r)

The compressor

driver

horsepower

(bhp

or

ghp)

is re-

lated

to the

adiabatic

head by

the following:

PV'

:

constant

(6-26)

When

Equation

6-26 is

expressed

between

the initial

and final

conditions

we have

PlVtn

=

PrYro

(6-27)

where

n:

the

polytropic

exponent,

n

+

I orn

+k

Expressing

Equation

6-27

in

terms of

temperature

and

pressure

we have

(6-28)

'

-

/p,\?

t'

-

\P,/

33.0001""

where rir

:

mass flow rate

of

the

gas,

lb./min

The

adiabatic efficiency

can

be defined in terms

of the

polytropic

efficiency

by

the following:

Equation

6-24 is discussed

in more

detail

below.

For

a

single stage

of compression,

neglecting any

changes

in

potential

and kinetic

energy, the temperature

change

from

the inlet

and discharge is

given

by

Af : r.

-

r :

6.33(2,547bhp

-

q)

(6-2s)

where

q

:

total

heat

energy lost

to the surroundings or

to

any

available

cooling water

or

cooling jackets.

This

value

does

not include thermal

enersv for inter-

coolers or aftercoolers.

For a multistage compressor,

Equations

6-20 through

6-25 must be applied

separately

for

each stage.

Polytropic Compression

This type of compression

occurs when a

gas

is revers-

ibly

compressed along a

path

that is defined by the fol-

lowins relation:

The

value

ofn depends

on whether

the

gas

is a

perfect

gas (z:

l)

or a real

gas (z

*

1)

as

previously

dis-

cussed.

For

a

perfect gas

the

relationship

between

adiabatic

and

polytropic

efficiencies

is

given

by

Equation

6-24.

Similarly,

the

polytropic

exponent,

n, for

a

perfect gas

is

related

to the

polytropic

efficiency

and adiabatic

expo-

nent.

k. as

follows:

n-1

k-l

lll

\4el

k-1

(6-24)

sincek:

ColC"

;H"

ghp:

bhp

:

33,000a"

frfl,

(6-22)

(6-23)

QCo

(6-29)

(6-30)

The relationship between

the

polytropic

efficiency

and

adiabatic

(isentropic)

efficiency of

a

perfect

gas

is shown

in Figure

6-17. The

polytropic

efficiency,

4p.

is usually

determined by the compressor

manufacturer using

either

an

old

design

or testing

a

new

design.

The

polyropic

exponent, n,

for

a real

gas

is

deter-

mined from real

gas properties

or with

using real

gas

data and using the following

expression:

_R

JCo

n

[z

/a'\]

_t_

+

t l_tl

JCo

lqo \at/l

n-

I

(6-31)

Equation 6-31 is identical

to

Equation 6-15 except that

the isentropic exponent for a real

gas,

7,

is replaced

by

the

polytropic

exponent, n,

and the compressibility fac-

tor for real

gases,

z,

is

divided by the

polytropic

effi-

crency,

?p.

Similarly to Equation 6-15, Equation

6-31

must

be

solved by

iteration for

a mean value of the

polytropic

ex-

ponent,

n,

over

a

compression range.



6A

70

72

74

POLYTROPIC

767880

EFFTCTENCY

lp

Rotating Equipment

47

Figure 6-17. The relationship

between

the

polytropic

efficiency and the adiabatic

effi-

ciency for a

perfect gas (Z

=

1).

In

Equation

6-31, if

we have

l3l

:

ouno z

=

I fora

perfect

sas

\r/p

then,

n-l:j-:=

@32)

n

JCp4p

,

/r

\

K

l-l

\4pl

For most real

gases

below

a compression

ratio

of

ap-

proximately

2, then

n

-

I

_k

-

1

n

'

ll\

K

l-l

\ql

The

basic horsepower

and head

expressions

for

poly-

fopic

compression

are similar to

those for isothermal

compression,

Equation

6-20. Thus,

we have

*'ffilett'le

''

For

ghp

at

100

scfm,

(,C,tffi

(6.33)

,no

:

H -1J_j1?L

[tfl

(il

t=l

_

]

(,$,,,

H,6.34,

\;/\

k

/t

The

equations for

polltropic

head

are

similar

to those

for adiabatic head.

Equation

6-21. Thus.

.

:

(.*_)t^J

IH(l

FJ

,]

"

(6-35)



48

Mechanical

Design

of

Process

Systems

If the

polytropic

head

is known,

the compressor

horse-

power

(ghp

or bhp)

can be obtained

from

the following:

mil

Equation

6-43 assumes

that the

heat

of

compression is

fully removed

by

cooling.

In

practice

this

is not

achieved,

because the heat

of compression

causes the

gas

to exceed the inlet

temperature.

The actual

performance

of a real

compressor

can be

evaluated

by

the

following:

bhp

:

ehp

:

33,000a*

riH

(6-36)

(6-37)

(6-3e)

(6-40)

(6-41)

(6-42)

33,00040

P1V1

:

P2V2,

OI

PV

:

constant

bhp

_

hp,

hpr

tlfl- tl:.

N

:

N(Q)o

5

'

H0.75

where N,

:

specific

speed, dimensionless

N

=

speed, rpm

Q

:

capacity

of flow

rate, ft3lsec

H

:

head, ft-lbrilb.

^

D(H)o

25

"":

e*

where D.

:

specific diameter, dimensionless

D

:

diameter

of

impeller

ot

rotor,

ft

H

:

head,

ft-lbr/lb.

(644)

(64s)

(6-46)

(6-47)

where

4oo

:

overall polytropic

efficiency

:

IpI.

The outlet

and inlet

temperatures

for

polytropic

com-

pression

are related

by the following

expression:

=

i&)H

F)

t,

\P,/

(6-38)

Equations

6-35 through

6-38 are

used separately

for

each

stage of a multistage

compressor.

Equations

6-38

and 6-39 can be used

to

calculate

the

polytropic

effi-

ciency

directly

(provided

t,

ta, P,,

P6

and k

are known

values):

7h

where

4,

Ia

:

isothermal

efficiency

:

overall efficiency

:

Itlln

After

applying Equation

644

and determining the

brake horsepower

(bhp)

for

a single stage

of

compres-

sion, the discharge temperature

can be determined by

Equation

6-25.

Dimensionless

Reference

Numbels

In sizing

and selecting

the type of

pump

or compressor

to

be

used, a logical correlation

is often

desirable. The

following dimensionless

parameters

apply to

pumps

and

compressors and are

the specific

speed and

specific

di

ameter, as defined as

follows:

/\

.

I

lk-ll

v

./

-

-t----t

--

4p\

K

/

\p

.

k-l

wnere y

:

k

Normally,

the

value

of

?e

is estimated from

data sup-

plied

by the manufacturer.

For initial

or

preliminary

val-

ues

of the

polytropic

efficiency,

10,

Figure

6-17 may

be

used.

lsothelmal

Gompression

This compression

occurs when

the temperature of

the

gas

being

compressed remains

constant

during compres-

sion. For

a

perfect gas

in

which z

:

1.0 and

(AzlAip

:

0 we have

The

theoretical horsepower

developed

during

a

revers-

ible isothermal compression

process

is

Figure

6-18

shows

the

dimensionless

parameters

as

originally

presented

by

Balje

[3].

This figure

is

the

graphical

combination

of

Equations

6-46 and 647. Past

experience often dictates what

type

of

pump

or

compres-

sor

is

to be used and in cases of uncertainty

or

new

appli-

cations, this figure

will be

useful in

equipment selection.

Figure

6-18

must be

applied to each stage separately,

as each

impeller

or stage must be chosen

with

each sepa-

rate

inlet

capacity or head for that stage.

ho,:atz hl&)

"

8.1l0

\P,/

(6-43)



Rotating Equipment

0.3

0.6

1

30

60 r00 3m 6m

1,000

3,0()()

10,000

Specific speed,

4

Figure

6-18.

The

initial

selection

ofa single-stage compressor

is made using

the

specific

speed

and

specific

diameter

parameters

t3l.

^.

10

E

G

I

Gentffugal

Gompressors

The centrifugal compressor

powered

the

first turbojet-

powered

aircraft and

is still used today

injet

engines as

a

supercharger.

The main

advantage

of

the centrifugal

compressor is that

it

produces

a large

pressure

ratio for a

single stage of compression, and is easily

manufactured.

Its

advantages over the

reciprocating

design

were cited

previously.

Most

centrifugal

compressors are designed so that the

gas

enters

the

impeller

axially-parallel

to

the rotating

shaft-as shown in Figure 6-19. The

gas

flow

is

then

changed

to

the radial

direction

and is accelerated in a

pe-

ripheral

direction as it moves along the impeller.

As

the

gas

exits the impeller, it enters

a

stationary diffuser

where

the

gas

velocity

is reduced.

This

process

is

re-

peated

at each stage on multistage compressors.

Most of

the

pressure increase

in

the

gas

occurs

in

the impeller

and

the

greatest pressure

drop

occurs

in the diffuser. In

multistage

compressors, cooling the

gas

between

stages

is

quite

common and many such compressors

have wa-

ter-cooled

separators

or diaphragms.

The

polytropic

relations,

Equations

6-26

through

6-

40, are usually

preferred

for centrifugal

compressor

cal-

culations. Figure

6-20 shows

why with

a schematic

plot

of

the

centrifugal compression

process

on

a

temperature-

entropy

graph.

Using the adiabatic

(isentropic) process,

the

actual discharge temperature is underestimated

Figure 6-t9A.

Centrifugal compressor-single-stage.

(Cour-

tesy of Dresser Industries, Inc., Roots Blower Operation.)

4

=

N

'/q/Ha1

D,=

DHltalJT'

/V

=

Speed,

rpm

O

=

Flow, fr3/s

D

=

lmpeller

diameter, ft



50

Mechanical

Design

of Process

Sysrems

1 Discharge

Nozzte

9.

Shaft

17

Intetsection

2. Casing

Cover 10.

Oi

Fterainer

j8.

impe|er

3 Sub

Cov€rSeclion

11

BeartngSrand

j9.

clideVane

Housing

4

Bearing

Stand Cap l2

Coupting

End

Beanng

ZO. In

er

Nozzte

5 SteelShim

13

tmpelerEnd

Bearino

21 cuideVane

6

r'rus

Bed

nq

4.

Or'Ferar-e.

-

22

curoevaneLrtdop

7

- run

Ho-s

nq

5eal t5

Sa.t

23

moe

er End

ptdl;

8. Spaci.g

Fing

16

Votute

24.

Intet

Wearing Fing

Figure

6-198.

Cross-section

of

a

single-stage

centrifugal

compressor.

(Courtesy

of Dresser

Industries,

Inc., Roots

Blower

ODeration.)

(ideal).

Since

the

polytropic

compression

process,

by

definition, is

the

path

connecting

the

inlet

and

actual dis-

charge conditions,

the

polytropic

formulations

are

pre-

ferred

by compressor

manufacturers.

This factor

be-

comes

extremely

important

in

sizing intercoolers,

since

using

the adiabatic discharge

temperature

would

result in

undersizing the

cooler. The

larger

the compression ratio

of the machine,

the more

severe

the mistake

ofundersiz-

ing

the cooler becomes.

Gas inlet

conditions

can change

and when

they do they

affect

a

centrifugal

compressor

differently

frorn

a

posi-

tive-displacement

compressor,

such

as a reciprocating

machine. Table 6-1 lists

the effects

of

changing

inlet

pa-

rameters on a centilugal

compressor

operating at

a con-

stant

volumetric

flow

rate

and a constant

sDeed.

Changing

the speed

of a

centrifugal

compressor

in-

volves

the

"affinity

laws,"

which

apply

to

single-stage

compressors,

multistage

compressors

when

each stage

is

considered

separately,

and to multistage

machines

over

a

narrow

speed

range representing

no more

thm

a

15%

change

in speed. These

laws

are

stated

as

follows:

1.

The

developed head

(feet)

varies

to the

square of

the

speeo.

2. The required

power

varies to

the

cube

of

the speed.

3.

The

capacity

(cfm)

varies

to the speed.

Figure

6-21 shows

the effect

of varying

centrifugal

compressor

speed.

In centrifugal

compressors

a

phenomenon

known

as

surge occurs when

the compressor

capacity

is lower

than

a specific

flow

rate.

This

specific flow

rate is shown

in

Figne

6-22

as

the

"surge

limit."

The

phenomenon

of

surging

is

manifested

by

cyclic

vibration

of

gas

flow,

which

can even result

in

reversal

of

flow

direction,

power

requirement,

and

discharge

pressure.

The

phe-

nomenon

normally

is

associated

with excess

noise

and

ENTROPY

s

Figure

6-20. Centrifugal compression process.

Atidd:

t2t-tl

At*,-r=t2-tl



Rotating

Equipment

5l

E

R

I

o-

ga.

:9

>, ;-n

<E

.ti

ry'E

oa.

<

=E

r^

ltY

=

c

oU;

?;

3;t

3i-

_11

-t

;L

,l

\

|t\

3\

rl

al

L

a

l

I

,/

t\

I

I

\

\

:-

.J

.J

\l

\

lstu Stnsslud

1N3?8ld

83q33P339

3Sll

]Unss rd

1i ltld

E33P

cY3ll

I

rillu

ld

3/VlodtsuoH lNlltld

3NrOd3S 08

1tt3U3C

3

9833P33

t3fl0dlst0lt

1r llttd

9

.46

,

i:

lPj

k'

:.1

il

.J

t{l

lsr

t\\

3A\

I

N

a\

1\

.1

ll

/.

t

/.

2

N;

\\N

\E.I

NN

"y

t;

w:,

tt

n

E

\

N

I

(V

N

\'l

A

I

s

,l

/s

t[t

u

\'l

$

)

4

'=\\

'I

>-Kl

J;-

Al

d

.t

AI

{

\l

il

*f

.lI\

aE:93Bs3B9BEig9S3P3e9

s I

8

L

\

\

61

I

N

\

5T--r-t

ii",-l I

\

\\



52

Mechanical

Design

of

Process

Systems

Table

6-1

E tects

ot

Varying Various

Inlet Parameters

on a

Centrilugal

Compressor

Increasing

Inlet

lncreasing

lnlet

Pressure

Increasing

Molecular

Weight

of

Gas

Increaslng

value

ot

Polytropic

n

or

Adiabatic k

Pressure Differential

Compression Ratio

Inlet

Density

Discharge Pressure

Discharge Temperature

Power Required

Head Developed

Mass Flow

Rate

Deateases

Decreases

Decreases

Decreases

Decreases

Decreases

Constant

Decreases

Increases

Constant

Increases

Increases

Constant

Increases

Constant

Increases

Increases

Increases

Increases

Increases

Increases

Increases

Constant

Increases

Decreases

Decreases

Constant

Decreases

Increases

Constant

Constant

Constant

CONSTIiIT

SPEED

COMPRESSOR

CHARACTERISTI(

B

I

I

I

,

7

I

...

I

c0MPn€ss0F

SURGE LIM I]

I

I

t

D

0 r0 20

30

40

50 60 70

80 90

too tn

PERCENT

CAPACITY

Figure

6-22,

Pressure vs.

capacity

for

a constant-speed

cen-

trifugal compressor

[4].

vibration of the compressor

and sometimes

the compres-

sor

piping.

Normal surge

limits are 40% to

90%

of

rhe

design

point,

with the higher

range

(close

to 90Vo) being

associated

with

multistage

mach

ines.

Controlling

surge

in

centrifugal compressors is more

difficult than

in centrifugal pumps,

but the

following

fac-

tors

ease

the

problem

considerably:

1. Throttling at the discharge

flange.

2. Throttling at the

inlet flange, which is usually

more

efficient than throttling at the

discharge flange.

Using

a

variable

speed

driver, usually

accomplished

by

the turbine driver.

Bypassing

or blowing

off excess

gas

to avoid

surge.

These steps

will

help

in

alleviating

surge

problems,

but

if a

variable rate

operation

is required,

the compres-

sor manufacturer should

be consulted.

Antisurge

devices can be incorporated

into

compres-

sor systems. For nontoxic

or inexpensive

gases

the com-

pressor

discharge

can be vented

to

the atmosphere

as

shown in

Figure 6-23. For

expensive

or

toxic

gases

an

automatic anti-surge system

can

be

installed

as

shown

in

Figure

6-24.

In

this

type of

arrangement

a

heat

ex-

changer

is

placed

in

the

system to remove

the heat

of

compression from

the vented

discharge

gas

to

prevent

a

loss

of

compressor

performance

caused

by

the tempera-

ture rise above the

design value

at the inlet.

Compressor

manufacturers

use standard

cubic

(scfm)

feet

to speciry compressor performance,

just

as

pump

manufacturers

use water

to determine

pump perfor-

mance. The manner in

which scfm

and altitude correc-

tion is handled is

discussed later.

Impellers

are critical

in

the selection

of centrifugal

compressors. The three

basic types

of impellers for cen-

trifugal

compressors are shown in

Figure 6-25.

The con-

ventional

closed impeller

shown in

Figure

6-25 is used

for

adiabatic

heads

up to approximately

12,000 ft-lbri

lb-.

The open, radial-bladed

impeller

shown in

Figure

6-25 develops more head

with the

same

impeller diame-

ter

and

shaft speed. The open

inducer impeller

can

produce

heads

up

to 20,000

ft-lbrnb*.

Whenever

the

head requirement becomes

too

great

for

a single impel-

ler,

then one must think in

terms of multistage compres-

sors. Each

stage of

compression of

a

multistage

com-

pressor

is

treated as a single

stage compressor and

the

same formulations hold.

r20

0

t0

0

J.

4.

s80

460

o.

40



Reciprocatlng

Compressofs

These

compressors

normally are sized according to the

adiabatic

expressions

of

Equations

6-11

through 6-25.

Normal

practice

in calculations for reciprocating com-

pressors

is

to use

the adiabatic exponent,

k

=

Cp/C,,

then adjust the results according to the specific

compres-

sor design and

configuration.

The

parameters

that

affect

the compressor horsepower,

cylinder

capacity, and dis-

charge temperature are length

of

stroke, shaft

rotation

speed, cooling efficiency, and fixed clearance of cylin-

ders.

All

of

these

parameters

vary for

each

given

appli-

cation, but have the same basic cylinder design and cy-

Figure

6-23.

comPressor.

Rotating Equipment

cle.

Figure 6-26 shows

the reciprocating

compressor

cycle.

This

cycle

involves

this

displacement

of

gas,

hence the classification of a

reciprocating compressor as

a

positive

displacement

type of unit. The compressor

is

unable

to exhaust

all

gas

from the cylinders and the

re-

sidual

gas

remaining in the compressor at

discharge con-

ditions

expands

to inlet

conditions. This

phenomenon is

shown

in

Figve 6-27 .

The clearance

voiume is usually set by the compressor

manufacturer

and

is

specified

to

match

the specified

ca-

pacity

with

the standard size compressor

unit.

Power

consumption is

not

affected

by

the

clearance

volume

or

the

volumetric efficiency.

The use of

"clearance

pockets"

is

used

in

some com-

pressors

to

vary

the

volumetric efficiency.

These

clear-

ance

pockets

can be sized to affect the capacity of the

compressor,

as in Figure 6-28. Power consumption at

re-

duced

flow rates is minimized by

use

of capacity control.

The

use

of

a

clearance

pocket

(additional

clearance

vol-

ume) reduces the

volumetric efficiency

of the

compres-

sor, because

the re-expanding

gas

fills most of the cylin-

der,

and

the

suction

valve opens further

in

the

stroke.

This mechanism is economical, because the energy ex-

pended

in

gas

compression is

retrieved

in

expansion.

The

clearance

pocket is

separated

from

the

cylinder by

a

stop

valve.

Figure 6-28

shows

how

varying

the

cylinder

clearance affects

the numeric

value

of the volumetric ef-

ficiency at constant

compression

ratio. The

volumetric

efficiency

for

a

reciprocating

compressor

is

given

by:

(6-48)

inlet

actual capacity

Manual

surge

control

system

for

centdfugal piston

displacement

flow monitor

centrifugal

compressor

Figure

6-24.

Automatic surge control with recirculating

by-

pass.

The

parameters

that affect the volumetric efficiency

are as

follows:

l. The

ratio ofa

relative clearance volume,

e,

which is

the

ratio of

clearance

to theoretical

displacement

ex-

pressed

as

percent.

2.

The compression

ratio, C., of discharge

to inlet

pres-

sure.

3.

The

various exponents

of the

polytropic

curve of

re-

expansion.

Such

a curve

is

shown

in

Figure

6-29.

Here the

cylinder is

normally cooled by

a

water

jacket

or

surrounding

air.

The

small volurne

of

gas

that

remains

in the

clearance volume expands

and

contracts with a cooling

surface. Consequently,

the

re-expansion

curve

(curve

3-4) is

initially steeper

than the adiabatic

curve

(curve

1-2).

With continuing

expansion

ofthe

gas,

the

gas

temperature

falls

below

that

of the

piston

and walls, and heat is transferred

from these surfaces to the

gas.

Thus, the exponent

of

the re-expansion

curve

(curve

3-4)

is

variable. For

re-

expansion

oflower compression ratios, Chlumsky

[5]

drscharge



Mechanical

Design

of Process Systems

OPEN BACKWARD.BLADED

IMPELLER

OPEN RADIAL-BLADED

IMPELLER

CLOSED BACKWARD.BLADED

IMPELLER

BACKWARD

LEANING

B LADED

IMPELLER

(PARAMETER-

%

SPEED)

e

63

s

si

o"o,

EH

o

-4,

?E

100

ao

RADIAL

BLADED

IMPELLER

(PARAMETER.

%

SPEED)

opi-l

RADIALACKWARD

LEANING

IMPELLER

AOJUSTABLE

IN

LET

GU

IDE

VANES

3

1?O

E

100

c'

ao

s

q

E

E

s

Vcc

d>

BLADED

IMPELLER

100

g'g

ro

.-B

so

s9

40

ADJUSTABLE

IN

LET

G UIDE

VAN

ES

20

40 60 80 100

120

oToFATED

INLET

VOLUME

Figure 6-25.

Basic

types of impellers for centrifugal

compressors.

uon.)

20 40 60 BO 100 120

obFATLO

l\-ET

VOIUMF

(Courtesy

of Dresser Industries, Inc., Roots Blower Opera-

'120

100

80

60

40

149

120

40

60

B0

100

120

qoRATEO

INLET

VOLUME

ol

l

ll

GUIDE

V

WIDE

T



Rotating

Equipment

55

P2

=

receiver

pressure

P1

=

inlet

pressure

Compression

Stages:

O =

start

@ =

comPression

@

=

discharge

@

=

expansion

O

=

intake

-tl

@@

Figure

6-26.

Reciprocating

compressor

cycle.

Clearance

volume

o/o

Clearance

=

Volume

Figute 6-27.

The effect

of clearance

capacity.

(100)



Mechanical Design

of Process

Systems

Clearance

volume

o/o

Piston

DisDlacement

Figure

6-28. A

clearance

pocket

(additional

clearance

volume) reduces

the volumetric efficiency

of the

compressor

because

the

re-expanding

gas

fills

most of the cylinder, and the suction valve opens further in

the stroke.

voLuME -.---------+

sourcE

:

cH

urMsl(Y

l5l

tts

l{-

ts

F

6.

It

rs-

l<\ |

rls

lrlo

I

115

100

CLEARAT{CE

:C

|

0O5L

+

O.Smn,

WHENE

L=STHOKE

L-ETGTH

Figure 6-29.

A

pressure-volume

diagram of a compresor

with clearance

(zero

flow

resistances)

[51.



recommends

fof

compression

ratios of appfoximately

2; the re-expansion

may be approximated

as an adia-

batic

process.

For the

volume,

Va-the

volume to

which

the

gas

expands during the

pressure

drop

from

P2

to

Pr-we have the expression

(64e)

Substituting

Equation 6-49

into the

expression

for

volumetric efficiency,

we have

Rotating Equipment

For

compression

ratios of 4 and higher, the

re-expan-

sion

cannot

be considered as an adiabatic

process.

For

these

compression

ratios

the

polytropic

exponent

m

(where

m denotes

the difference

between the re-expan-

sion

PV'

(constant)

and

the compression

PVn

(constant).

For diatomic

gases,

m

:

1.25.

The

value of the

polltropic

curve exponent,

m, varies

with

pressure.

Chlumsky

[5]

recommends for

a com-

pression

ratio of

3:4

the following values of

m be used:

., ..

/pl,

".

:

""

\p,/

-.[(,*i

-

-

'u'(*o]'

First stage

Second

stage

Third

stage

Fourth stage

Fifth

and further stages

m:l 20

m

:

1.25

m:

1.30

m

=

1.35

m:k

"+v"-v4

,lt

-

-----------=;--

-

or

-

-t

(6-50)

where

e

=

*

:

.utio of

the clearance

volume.

Vo.

to

vp

the

volume swept by the

piston stroke.

v"

V^

?"

=

#

:

expression

for volumelric efficiency.

vp

Equation 6-48, the

ratio of

gas volume

pumped

to

the

volume swept by the

pis-

ton

(compressor

displacement)

Figure 6-30 shows the

graphical

solutions

of Equation

G50 for

various compression ratios

and

exponents

of

the

polytropic

curve

of

re-expansion and clearance values.

34

_L-

c

These values are

given

at different

pressure

levels, as ex-

ist in multistage compressors

with

the suction of the

first

stage

at

atmospheric

pressure.

The

volumetric

efficiency

for

a

perfect

gas

(z

=

1),

not realistic,

is

given

by

4,r:100-c(cRr/k-1)

(6-s

l)

where

4,,

:

theoretical volumetric

efficiency

The

volumeuic

efficiency

for

a

perfect gas

(z

:

1)

with realistic effects.

4":100-cR

-

c(cR'/k

-

l)

Cs

:

compression

ratio

:

PzlPr

(6-s2)

80

I9n

1@Z

Figure

6-30.

Curves

for

determining

volumetric

efficiency

[5].



58

Mechanical Design

of

Process Systems

The difference between Equation 6-52 and Equation

6-51

is that the theoretical volumetric efficiency should

be

reduced

by

a value

equal to the compression

ratio

to

obtain

an

actual value for

a

perfect gas.

This is

a

factor

that

has

been

determined

from

field

experience.

For a

real

gas (z

*

1)

with

realistic effects,

lv:

100

_

cr

_

c1(cR)i"

_

I

where

zt, 22:

rnlet and discharge compressibility factors,

respec-

tively

As

stated

previously,

reciprocating compressors

fol-

low the expressions

for an

adiabatic

process.

The work

required

for

the adiabatic

compression

of

a

perfect

gas

(z

:

1) is

found by

the

following expression:

w:

PV

(-o_JhtJ=

-']

(6-54)

The theoretical

horsepower may be found

by

Equation

6-16 or bv the

followine:

6o.

:

(P

Vrrr44 u

[ll,I-

.

rl

[,,

*

,l

33.ooo

k-l

[\Pr/

'l

\

2.,

/

(6-5s)

For an

ideal

ga's,

21

:

22

where P1,

Pz

:

inlet and discharge

pressures, respectively,

psia

Vl,

V2

=

ir et and discharge

gas

flow

rates,

respec-

tively, acfm

In

Equation 6-55, the theoretical

horsepower

may be

varied by the

following

parameters:

l

lncreasing

the compression

ratio,

Cp

2.

Increasing the specific

heat ratio, k

3.

Increasing

the inlet

pressure

at a constant

compres-

slon rate,

4.

Increasing

the actual inlet volume

(nat

standard

vol-

ume).

Multiple

Staging of

Reciprocatang

Compressors

Multiple

staging

is

the compression

of

a

gas

from one

pressure

to another involving

more

than

one step.

Each

step acts

in

series

with

the others and entails

a

basic

ma-

chine element.

In multiple staging of

reciprocating

com-

pressors, increasing the cylinder size

is less expensive

than

increasing the number

of

cylinders,

thus

the

ten-

(6-53)

dency

has been to increase

the cylinder

size using a

smaller

number

of

cylinders.

Multistage reciprocating

compressors

have the

following

advantages:

1.

Operating

at high

speeds,

they

can

be

coupled

di

rectly

at

high shaft

speeds thus

utilizing

cheap

electric

motors.

2.

Better balance

of inertia

forces.

3.

The mass

of

the flywheel,

which rotates

at

high

speeds, can

be

made

smaller, resulting

in

a smaller

fluctuation

of

torque. The more cylinders, the

less

the fluctuation

of torque.

4.

Starting multistage

compressors

is easier

because

they

have small moving masses and

thus can

be

driven by electric

motors with

less

inertia torque

and

lighter

construction.

5.

Variations

of

pressure

and flow

velocity in

the

inter-

cooler

or

oil

separator

are

less, thus

making

these

parts

smaller.

6.

Machines

of

various

capacities

can

be

manufactured

using

identical

parts,

making

interchangeability effi-

crent,

7. Multistage

compressors are better suited

to automatic

operation.

Gas

Temperature for

Reciprocating

Compressoas

The discharge

temperature of a

positive

displacement

compressor,

a

class

of

which the

reciprocating

is

in-

cluded, can

be

predicted

by

the

following expression:

(6-56)

where

t

:

absolute

temperature for any system

P

=

absolute

pressure

for any

system

k

:

Cp/C',

adiabatic

exponent

1,

2

:

inlet and discharge

conditions, respectively

Axial

Flow

Gompressors

In axial flow

compressors, the flow enters

the unit

oarallel to the

axis

ofthe

shaft and the

flow

direction

es-

ientially remains

unchanged

from

the inlet to the

exit

of

the

unit.

Airfoil

blades are located

on the rotor shaft,

varying

in

pitch

and size

according to the

flow condi-

tions. The

gas

passes

through

the

airfoil

blades

in an ax-

ial direction.

Axial flow compressors

are used

for

applications

of

about

25,000 cfm upward.

The formulas

for

centrifugal

compressors

apply to axial

flow machines.

Axial flow

compressors

can

handle

greater

capacities,

which is the

primary

reason

why they

have replaced centrifugal

com-

,r

-

/P'\?

t-\Pj



pressors

in aircraft

gas

turbine units. The characteristic

curve

(head

versus

flow) for

an axial

flow compressor

is

much

steeper

than for

a

centrifugal

compressor

and the

surge

limit

is a function ofdesign capacity.

Contrary to

a

centrifugal

compressor, the

required horsepower

for an

axial

flow

compressor

at constant speed

and

pressure

de-

creases with increased

flow Axial flow

compressors

are

not

as common

in

the

process

industries as centrifugal

or

reciprocating types of

machines.

Fans

and

Blowers

Fans

and blowers

are basically compressors.

They

fall

under two types

of

compressors-centrifugal

and

axial

flow. If one understands the basics

of

centrifugal

or

axial

tlow

compressors,

fans

and

blowers

come easy,

for

they

are

less complicated than compressors.

Specifying Gompressor

Flow

Gondltlons

Specifying compressor

flow

conditions

is

a major

source of confusion in applying compressors

to

process

sl stems. There are three basic

ways

to specify

compres-

:or

flow conditions:

l. Mass

flow-define

the

mass

flow

rate of the

gas,

Ib./

in

the English system and

kg/hr-m in the Sl/metric.

3.

Actual, or inlet, volume flow-volumetric

flow rate

of

the

gas

at the

inlet conditions, expressed as acfm

or

icfm in the English system and m3/hr

in

the SI

and

MKGFS

systems.

-1.

Standard volumetric

flow-the

volumetric flow rate

of the

gas

at the inlet conditions expressed in terms

of

standard

cubic

feet

of

gas per

minute

(scfm)

or mil-

lions of

standard

cubic

feet

of

gas per

day

(MMscfd)

in

the English

system and m3/hr in

the

SI

and

MKGFS

systems.

Iass

Flow

The

method

of defining

the

mass

flow

rate of the

gas

h

terms of the inlet conditions of the comoressor is fa-

r

ored

by

many and

is

mandatory

in

calculating

gas

prop-

enies

between stages. Mass flow rate

,?2uJt

be

specified

as

either dry

gas

or wet

gas.

Ifthe

gas,

for example, con-

rains

water vapor, this could drastically change the com-

pressor

design. One of the

problems

of using mass

flow

is not

speciffing the

flow

conditions

as a

dry

gas,

which

ir reality

is a two-phase or multiphase flow.

Another

disadvantage to using mass

flow

is that

it

does

not

allow one

to appreciate the

physical

size of the sys-

rcm.

An

intuitive

feel

for

any system is essential

to

its

successful

desisn.

Rotating Equipment

Actual

or Inlet Volumetric

Flow

Actual

flow rate conditions

at the inlet to the

compres-

sor

is denoted as

acfm or icfm-acfm

meaning actual

cu-

bic

feet

per

minute

and

icfm meaning inlet cubic

feet

per

minute.

The

disadvantage

to specifying acfm is

in the internal

components

ofthe compressor, e.g., a sideJoad

refriger-

ation

compressor,

or in

a multistage compressor.

In

a

multistage

compressor the

previous

stage's discharge

temperature

is a

function

of

the

previous

stage's

com-

pression efficiency, and

mass flow rates are better

for

such conditions.

Acfm

is

best

for

plotting

compressor

performance

curves,

because the

impeller is sensitive only to

the

ac-

tual volumetric

flow and

is

insensitive

to

the

gas

state

conditions.

Mass

flow and

acfm

volumetric

flow

should be used

because

mass

flow is invaluable in communicating

with

tle

compressor

manufacturer and

in

dealing

with

inter-

nal

machine flow

conditions, and acftn

is

essential

in

getting

a

feel for the

physical

size

ofthe

system.

The

use

of

mass flow and

acftn should counter the disadvantages

of both

approaches.

In computing

pressure

drop through connecting

piping

systems

to

compressors,

it is imperative that acfm

be

used to avoid

any confusion

in

designing the

piping

sys-

tems.

Standard Volumetric

FIow

Specifying

gas

conditions in terms of standard

volu-

metric

flow

is done extensively throughout

industry. The

gas

flow

conditions

are based on standard inlet condi-

tions-pressure,

molecular

weight,

temperature,

and

compressibility-all

based

on

"standard"

conditions.

Thus, the

standard specific

volume

is constant

being

that

u.,.

:

"'+J'':

constanr

(6-57)

where z.,a

:

compressibility

factor at standard conditions

R:

universal

gas

constant,

which

is

a function

of

the

molecular weight of the

gas

tsld

:

temperature at standard conditions

P$d

:

pressure

at standard conditions

Volume

flow is expressed as

Q,ta

:

mV,ro

(6-s8)

where the standard

volumetric

flow

is directly

propor-

tional to the

mass flow rate.



60

Mechanical Design

of

Process

Systems

As with using mass

flow,

when

using standard

flow

conditions one cannot appreciate the

physical

size

of the

system.

And worse still, using

scfm does

not

provide

any

of the advantages of using either mass flow or acfm. To

specify

something

as

"standard"

one

thing is

essential,

that

all

parties

agree

on what is

"standard."

Unfortu-

nately,

this

is

not

the case

with

using

scfm,

as

the

follow-

ing

"standards" cited

by Lapina

[6]

indicate:

English

system

Metric system

1. P",a

:

14.7

psia

t'ta

:

60'F

2.

P,u

:

14.7

psia

t"a:70"F

3.

Pd

:

14.7

psia

t.to

:

32'F

1

P",a

=

101.3 k?a

t,ra

:

0'C

P"a

:

101.3 kPa

tsa:15'C

Thus,

what

is considered "standard," as Lapina

[6]

writes, varies from industry

to industry and engineer to

engineer. In the

net

result what is often

gained

is confu-

sion.

Properly Specifying Gompressor Flow

Gonditions

To

properly

size or select a compressor, the capacity-

no matter

how

it

is

given-must

be

converted to the

inlet

conditions. To

do this

the

following

expressions are

used:

PrVr

_

P2V2

tflt

tzzz

where

V: volurne

P

:

absolute

pressure

t

:

absolute

Iemperalure

z

:

compressibility

factor

In Equation 6-59, if

z

:

1.0 for

a

perfect gas,

and P

and

t are at standard

conditions, then

acfm

:

e_

=

rirV

: "'

(6-60.)

p

where ri

:

mass

flow

rate,

lb./min

V

:

specific

volume,

ft3llb,,,

p

:

density,

lb./fC

The specific

volume,

V, may be

determined by

/r sas\ /

'

\

v

=

z

l::_:l I::-::l

(6-61)

\

mw

/ \1,14Pl

where,

as before, mw

:

molecular weight

2.

scfm

:

(379.46)mh

(6-62)

60

where mh

=

moles/hour

and

rir

=

(rfi)(mw)

and

finally,

(6-63)

_

[(MMscrdx106)1 1,0

nu)1/f*l*)|/t)

"_*,

clm:

qs

=

t--aOoz,

t

\-pJ\460

+

rJ\il

."

-'

where

lie subscript, s, denotes

properties

at the

inlet

(or

suc-

tion)

conditions.

Equation

6-64

may be

expressed as follows:

e.=acrm=*-tltjHP*.,-J

(6-6s)

where the scfm is

based

on a dry

gas.

To

convert

the standard

volumetric

flow

to

mass

flow

the

following

relations are used:

English system:

(6-66)

Sl/metric system:

rir

:

scfm fP"o

'

ro'\

\zd

R.td t.ld/

(6-61)

PIPING SYSTEilS

FOR

ROTATING

EQUIPMENT

For rotary equipment

to be

functional

and

contribute to

the

process

system,

it

must be connected to the system

with

piping.

The science of connecting

piping

systems

to

rotary equipment is a relatively new field and has drawn



the stalwarts

of

academe

to

join

with industry

in

solving

problems

of

piping

and equipment.

The

two

problems focused

upon here are

nozzle

load-

ings and

pulsation

response

spectra distributed

to the at-

tached

piping

system

by reciprocating

machines.

Nozzle

Loadings

In earlier

years

various rotating

equipment

manufac-

turers would define

allowable

nozzle loadings

as "zero

force

and

zero moments."

Such statements

were

not only

ludicrous,

but

showed how

little confidence

some

rotary

equipment manufacturers

had

in

their

products.

Ulti-

mately, the

pipe

stress

engineer was

left

to

use

his

(or

her)

sole

judgment

to determine

if the

piping

loads

were

substantial enough to

damage

the

attached equipment.

There are several

standards

for handling

nozzle load-

ings

on

rotating

equipment,

and

probably

the best

known

are

those

of

NEMA

(National

Electrical

Manufacturers

-{ssociation).

NEMA

provides guidelines for

nozzle

Ioadings for steam

turbines

for mechanical

drive

service.

Unfortunately,

its

guidelines are appiied

to every

prece

of rotating equipment

by

eager

customers

and engineer-

ing

contractors.

For example,

what

is

valid

for steam

turbines is not valid for

inline

pumps.

Because

steam tur-

bines are

more

fragile than

most types

o[

rotary

equip-

ment, using the

NEMA

standard

produces over-conser-

vative

designs

for

most types

of

rotary equipment.

The American

Petroleum

Institute

(API)

also

has stan-

dards for rotating

equipment:

API 611-General-Pur-

pose Steam

Turbines

For Refinery

Service;

API

612-

Special-Purpose

Steam

Turbines For

Refinery

Service;

,\PI

617-Centrifugal

Compressors

For

General

Refin-

ery Services; and

API 618-Reciprocating

Compressors

tor

General

Refinery

Service.

Applying API standards

to

nozzle loadings

on rotating

equipment

leads

to

the

argument

in which

rotating equip-

rnent specialists claim

that the

API

standards

are only

in-

tended

for

procurement purposes,

and

the

pipe

stress

en-

gineers,

having no other

guidelines

to

follow,

assert

that

the

API

standards

are

what

is to be used

in

practice.

The

best

criterion

for

judging

nozzle

loadings is expe-

rience with a

given

piece

of

equipment.

For example,

my

several years

of practical

experience

with turbo

expand-

ers dictate they can

withstand

three times

the nozzle

loadings

allowed by

NEMA

(remember-only for

steam

turbines )

.{lowables for inline

pumps,

as above,

did

not

exist

a

tew

years

ago.

Such

pumps

were regarded

as

piping

components,

e.g.,

valves, and allowables

were consid-

ered unnecessary.

But "thinning-up"

casings

to

reduce

naterial and

costs makes such

allowables

possible,

al-

rhoush

controversiai

at times.

Rotating Equipment

Table

6-2

Typical

Manufacturer

Allowables

lor

Nozzle

Loadings

tor

Inline

PumPs

Mi=

Li

Mo=

Fo

Lo

PUUP

SIZB

(

in)

Fa lb

t-1

-tb

2x3x6

3x4x6

4000

6000

50

00

60 00

4000

5000

2x3xo

3x4xB

4x6xg

4000

5000

6000

5000

6000

7000

4000

5000

6000

4x6xl0

6x8x

0

5000

8000

7000

9000

5000

8000

6x6x20

| 0x1 0x20

12x12x20

500 0

800

0

r 2000

6000

9000

13000

5000

6 000

10000

F

*Miao * {oact 1 2.g

F"

Mi.o

to,n",

-

Hhere,

F

=

resultant

of actual force applied,lb

Mh.

u.tuut

bending

monent

on suction nozzle,ft-1b

Mou;,

actual

b€nding

nonent on

discharge

noz2Ie,ft-1b



62

Mechanical

Design

of Process

Systems

There are three

basic options

to

solving nozzle load-

ings on rotating

equipment.

1. A detailed

finite

element

study of

the equipment.

2.

Destructive testing

of

the

equipment.

3.

Close interface between

the rotating

equipment man-

ufacturer

and

the

piping

stress

engineer.

The

problem

with finite

element analyses

is

who

is

going

to

pay

for

it-the

client, the

engineering contractor,

or

the

rotating

equipment

manufacturer? Next,

can the

ro-

tating equipment manufacturer

disclose

proprietary

in-

formation often required

in finite

element analyses?

De-

structive testing

poses

the same

question,

who will

pay

for it? The third

option-the

pipe

stress engineer confer-

ring

with

the

equipment

manufacturer-is perhaps

the

most viable of the

three,

because if

the

NEMA

and

API

criteria

cannot

be

met,

then the

rotating

equipment man-

ufacturer can at least

expect extra loadings

and can de-

sign

for

it, if

time

permits.

Thus,

the rotary

equipment

vendor working

as

a team with

the

piping

stress engi-

neer(s) can help to

alleviate most nozzle loading

prob-

lems.

NEMA and API standards

are

very

safe and a

piece

of

equipment that meets

their requirements

should

not have

any

nozzle

loading

problems,

such as leaks. The

prob-

lem comes in modular

skid construction, where the val-

ues

provided

by the

standards are very

conservative.

Manufacturers often give

allowable

values for

their

equipment,

and Table 6-2

presents

some

typical

ones. A

generalized

standard taken

from several

pump

manufac-

turers'

allowable

standards is

shown

in Fieure

6-31.

Reasonable nozzle loadings

for

turbo

expandJrs

worked

out by the author

and several

turbo expander manufac-

turers

are listed

in

Table

6-3.

Neither

Thble 6-2 nor

Table 6-3 should

be substituted

for

the manufacturer's

allowables,

if

the

vendor

has his

own. However,

the information

can be a valuable

tool.

Rules

of thumb often are not

only invalid but

are often

based on special situations

that may not be

true for every

case.

One must be extra careful in

piping

steam turbines,

be-

cause these units are

usually

fragile.

Example

2-2

in

Chapter 2 illustrates

a

piping

arrangement

connected to

a

steam

turbine. If

expansion

joints

are allowed, the

con-

figuration

shown

in

Figure

6-32 is ideal.

PULSATION BESPONSE SPECTRA

INDUCED BY RECIPROCATING

EOUIPI'ENT

Reciprocating machinery

often

induces

pulsation

re-

sponse spectra in attached

piping

systems.

This

subject

alone is comprehensive to

fill

several volumes,

so we

will

just

outline the

problem

here.

Mno

=\fif,,T

Mfi Mfl MF"

=..ffi*r

N/-t+Tlg

MFN

=

greater

of Mpo &

Mp",

where

Mso

& MRs

are

resultant moments applied at nozzles

MRO

=

resultant bending moment

about

DM,

=

F"-(0") + FD,(dD) + M"y+ MDy

DM,{

=

F"y(d") + FDy(dD) + l\4"y + l\iDy

tr\arr 12

t\-a

i,

110.5.

.-.-L

LtAr'.-"t_

| lL,/..r,r- )l

-

FFs

=

[Fs2"

+ F , + F .]o5

;

Fno

=

[F2o*

+

FBy + FzD.]o

5

FB

=

greater

of FRs or FFD

*&*^ '*ffi.

z.o

Figure

6-31.

Generalization

of

forces, moments,

and allowable nozzle loadings.



Rotating Equipment

Table

6-3

Reasonable

Turbo

Expander

Nozzle

Loadings

Nozzle

Size

(in

g9

974

1

too

|,623

1,948

2,272

t

so5

l too

1,948

,

<o7

3,246

3,895

aa \

5,189

1 ?OO

I,948

)

5q1

3,246

3,895

La \

5,189

1,948

)

q))

3,896

4,869

5 R4?

6,817

7

,784

1,624

2,436

7 )47

d n{q

4,871

5,683

6,486

2,436

3,654

4,870

6,088

7

,306

8,524

9,730

M,

3,383

5,074

6,7&

8,455

10,146

11,838

l3,513

4,474

6,710

8,947

I

1,184

13,421

15,658

t7

,870

M,

4

6

8

10

'|.,

t4

l6

Nozzle

Size

(in

F,

4

6

8

10

12

t4

l6

&9

9',14

1

too

1,623

|,948

1 11)

t

so{

r too

1,948

t <o7

3,246

3,895

4,545

5,189

1,299

1,948

,)

<o?

3,246

3,895

A \A\

5,189

1,948

I O))

3,896

4,869

5,843

6,817

7

,784

1,624

2,436

a )L1

4,059

4,87r

5,683

6,486

3,383

5,074

6,7&

8,455

10,146

11,838

13,513

2,436

3,654

4,870

6,088

7

,306

8,524

9,730

4,474

6,710

8,947

11,184

13,42r

15,658

r7

,810

Nozzle

Size

(an

6

8

10

12

t4

l6

l8

20

24

648

972

|,296

|,620

L,944

2,268

, 50?

')

cll5

3,240

3,892

l,080

1,620

2,160

2,699

? )10

3,779

4,3t9

4,859

< 100

6,486

F,

1,080

r,620

2,160

2,699

1r10

3,779

4,319

4,859

< ?oo

6,486

|,659

2,488

3,318

4,147

4,976

5,806

6,63s

7

,464

8,294

9,964

1,620

2,429

l tlo

4,049

4,859

5,669

6,479

7

,289

8,099

9,730

2,699

4,U9

s ?oo

6,748

8,098

9,448

10,798

r2,147

13,497

16,216

2,699

4,O49

s lqq

6,748

8,098

9,448

10,798

12,t47

13,497

16,216

4,147

6,220

8,294

10,367

12,M\

14,514

16,588

18,661

20,735

24,912



64

Mechanical

Desisn of

Process Svstems

Table 6-3

(continued)

Compressor

Discharqe

Nozzle

Size

(in.) F, F, Fz Fs M, My

M'

Mp

4

6

8

l0

1''

14

16

18

650

974

1,300

1,624

I

q4q

) )74

, sqq

? ol

o

|,444

2,165

2,888

3,610

4 111

{ n{l

6,486

i

rqq

1,949

,

soo

3,249

3,899

4,548

5,198

5,838

2,048

3,072

4,097

5,121

6,145

7

,169

8,193

9,202

1,624

2,436

7 )49.

4,060

4,872

5,684

6,496

'7

1a-l

,

165

?

t4q

\

A1L

6,496

'7

a'7q

8,662

9,730

3,W7

4

at)

6,016

7,5r9

9,023

t0,527

12,030

13,514

4,046

6,070

8,093

10,116

12,139

14,162

16,185

l8,

181

PG:

Planar Guide

lA:

IntermediateAnchot

G: Guide

HEJ:

Hinge

Expansion Joint

GEJ:

Gimbal

Expansion Joint

Figure 6-32.

An expansion

joint

arrangement ideal

for steam

turbines

where nozzle loadings must be kept

low

(almost

al-

ways the case

with

steam

turbines) and

the use of expansion

joints

is

practical.

(Courtesy

of

Pathway

Bellows, Inc.)

Currently,

two

methods are used

to

predict pulsation

problems:

(a)

modeling the system on

an analog com-

puter

and

(b)

simulating

it

on a

digital

computer.

Basi-

cally, the

piping

system is modeled

with support and

soil

stiffness

vaiues input at every

pipe

support

as discussed

in Chapter

2.

Then the system

is

excited

with various

forcing

functions that represent the

reciprocating

ma-

chine

or machines.

The

piping

supports

are

moved

around, deleted,

or added to decrease the amplitudes

generated

by the

forcing functions. This analysis can

be

done

on

either

an analog or digital computer.

There are two

methods available on existing computer

software

that can

help head off

pulsation

problems.

These

methods arc modal ertaction

analysis

and time

spectra

(time

history) analysis. Modal extraction

is com-

puting

the natural

frequency

of

the

piping

system, after

modeling

the

pipe

support and

soil

stiffness

values,

and

comparing

this frequency to that of the shaft

speed of the

equipment.

Time

spectra

analysis

is

a

transient

analysis

that

basically does exactly

what modal extraction does

except on

a transient basis

for

every time interval over

a

specified

period

of time. In other

words, we compute the

system's

natural

frequency

for every

second

over

a

pe-

riod

of

one

hour. Over

the

period

of one hour

we

excite

the

system

with a forcing function that accurately

defines

the

rotating equipment.

Figure 6-33

shows a

piping

system

excited

by

pulsa-

tions

from

a reciprocating

machine. A complete

investi-

gation

of

the

pulsation

frequencies and surge

capacity

is

normally required,

which involves the

compressor

bot-

tles

(surge

drums), compressor

suction header,

and suc-

tion compressor

bottle,

the discharge

header, and dis-

charge

compressor

bottle. Two

companies

are

engaged

separately

in

investigating these

problems-Southern

Gas Association's

compressor analog

computer at South-

west

Research

Institute and the Structural

Dynamics

Re-

search

Corporation

(SDRC).

The compressor

bottle

(or

surge

drum)

acts

as a

pulsation

dampener. A typical

bot-

tle is shown

in Figure

6-34.

The compressor bottle

acts

as

an acoustic

filter designed for all frequencies

induced

as

the reciprocating

engine

speed

varies. The compres-

sor bottle

cannot damp out

all

frequencies, but should

store

energy

generated

from the

various

frequencies and

reduce

them to

produce

a

relatively smooth

and continu-



Rotating

EquiPment

Figure 6-33.

Piping

system

excited

by

pulsations

from

a

reciprocating

machrne'

ous operation.

Sizing

the

compressor

bottles should

be

done

by

a

specialist

who

has

worked

in this

field

for sev-

eral

years.

In the

days

before

analog

and

digital

simulations,

pul-

sation Droblems

were

solved

(and

still

are)

with orifice

plates.

These

plates were

placed in the

piping system

and

the

orifice diimeter

was approximately

0.53

times

the

internal diameter

of

the

pipe. These

plates' distributed

throughout

the

piping

system,

acted as

pulsation

damp-

eners.

Although

orifice

plates

produce

huge

pressure

drops, they

are

effective

in

many installations.

EXAMPLE

6-1:

HORIZONTAL'

CENTRIFUGAL

PUIIP

SYSTEM

DESIGN

A

food

processing

plant

is

having

a

cooking

kettle

in-

stalled

to

process

molasses

into refined

syrup

for break-

fast foods.

A

horizontal

centrifugal

pump is to

be in-

stalled next

to

a

fuel

tank to

supply

fuel

oil to a

burner

in

rhe cooking

kettle.

The

fuel oil

tank is

to have

a 50

psig

nitrogen

pad because the

tank cannot

be

raised

for higher

head at the

pump.

The

cooking

kettle

is

200

ft

down-

stream and

15

ft above the

discharge

flange

of

the

pump.

It

is

desired

to select

and

size the

burner

feed

pump

shown in Figure

6-35.

The

discharge

pressure at

the

burner end

is

to

be

40

psig.

Suction

Llne

Pressure

DloP

Fluid

:

tuel oil

TemDerature

:

90'F

Figure

6-34.

Typical

pulsation

bottle

(or

drum)

configura-

tions

that

act as

pulsation

dampeners.

Pressure

=

50

psig

p

:

54.725

lb^lft3

p:

139.53

cp

:

(139.53)(6.72

x

10-a)

:

0.094

lb./ft-sec

e

:

0.0018

L:1.0ft

Suction

line

=

3

"dSch

40,

Di

:

3.068 in

Q

:

150

gpm



Mechanical Design

of

Process

Systems

Figure 6-35. Hot-oil pump piping

scheme

for Example

6-1.

fuel

tank

cooking kettle

3" x 1tlz"

burner feed

pump

(r5o)sar

lrj,

ll]]ry\

min

\7.479

gal/

\60

s€c/

Entrance

and exit:

K:1.78

l-3-in.d 90" std ell

:

K

:

0.30

1-3-in.d

gate

valve

:

K

:

0.14

6.51I

ft/sec

(7.393)

in.:

I

t n'

)

\1,14

in.r/

l3

068li.,o.srr,

rt

(s4.72s)l9r

N.-=DVP-\

12l

sec

ft'

-

nur.,

r lh

r0.094;-1\

n-sec

With

NR"

:

969.1, the

flow

is laminar.

From Equation

1-6b

we

compute the

friction

factor as

follows:

6L

6A

f=j_:

-

:0.066

N*"

969.1

K.Values

(Velocity

Heads)

Referring to Figures

1-7 and 1-11 we have

the

follow-

ing:

\-.. *

From

Equation

1-4 we

compute the frictional pressure

drop as follows:

ao,

:

ILL

*

'l-

r.leY

'

\D -

I2e,

oo,

-

fro.ooorrts.oorrtzr,,.rrl

t

(3.068)

I

rsa.72sr

llr(6.511)?

tt2

I

'o',,l

tr

sec2

\144

in.2/

zr:z.zr

n-111

sec'-ln

Ap1

:

L524

psr



Rotating Equipment

Discharge

Line

Pressure

DroP

The

conditions

are the

same

as

the

suction

line except

for the

following:

Line size

=

2-in.

Schedule 40

for

which

Di

:

2.067

o,,:[ry.0"']

ts+.zzs1$

@.64D'?#Hh)

^."^

^.

fr-lb.

-'--

-'

sec2lbr

or 1l/z-in.

d

pump

discharge,

.

(150)#[+r-J(#,J

";;;m[

For 2-in.

S/40

discharge

line,

fr

=

23.642

:

sec

For

2-in.

d

S/40

pipe,

=r/.lR?

(o.os+)

-.1 .

n-sec

"

64

_

64

_^^^^

Nr"

1,438.3

K-Values

for 11/2-in.

Portion

Entrance

:

K

:

0.78

From

Thble

1-7,

for a 2-\t.

x

lll2-in.

diftuser,

K

:

0.055

E*: o.srt

L

:

3.0 in.,

d

=

1.610

in.

Apr

:

2.982

Psi

K-Values

lor

2-in.

Portion

2-2-in.-std90'

elbows

=

K

=

0.40

exit:K:1.0

EK:

r4o

L-200ft

^

^.

-

[{0.044)r200.0X

t

t,

.,

,.oOl

^r,r

-

[-o

06zr

-.

-l

th.

.

ft2

I

tfr:

I

rS+.225r

'il

r14.343).

_

.z

\raa

in.,]

fr-lh

S€C'-lD1

A*

=

63.72

psi

:

too

high-choose

a

1

r/z-in.

x 3-in'

diftuser

With 3-in.d

Sch

40

PiPe,

(lso)sa,

L+fu)(,**)

(ry)

-

(14.343)A(54.72rk

ft

sec

(7.3e3)

in.2

(r-

*--L)

K-Values

for 3-in.

d

PiPe

2-2-in.-std

90"

elbow

=

K

=

0.54

exit:K:1.00

[('-ryt

-

e3.642t

L,ro.rrr,

hl

r".

-

l\

tz

I

'-

tt'l

=

r.zzo.s

| 10.094;.'"'

I

I

tt-sec

I

@

:0.037

Nt"

Nn"

:

:

969.125

Dr:

t.so

l1 4lr

(6.511)

a,so.tts,

l :

\l2 l

sec

(0.094);lb'

n-sec

6A

f:

-

=

0.066

Nn"



68

Mechanical Design oI

Process

Systems

^_t

pf_[

(0.066x200.0)(12)

(3.068)

+

r.54]

(s4.7zs)t#(6.51rF

g

(,*

*-)

fr-lh

SeC'-lDr

Apr

:

13.309

psi

=

use

3-in.

{

S/40

pipe

New K-Values

for

1r/2-in. Pipe

Entrance

:

K

:

0.78

From

Table l-7

,

for a 3-in.

x

lllz-in.

diffuser,

K

:

0.337

E*:

r.ttt

L

:

3.0

in.: d

:

1.6i0 in.

Nn.:1,720.5;f=0.037

..^.: lQ

ji1Ii0.,

"rl

(1.610)

'-'

l

lh fr2 / rfr?

I

\s4.125)'+

(23.642f

::-

|

'"

,I

Irr sec? \

144

in.2/

V--

fr-1h

)/1t tr "

'"m

sec'-lDl

Apr

:

3.912

Psi

Total

pressure loss in

discharge

linc

-

13.309

-

3.912

=

17.221

psi

Using

the

pump

manufacturer's curve

in

Figure 6-36,

we

can

enter

data

on

the

Hydraulic Design Calculation

Sheet

in

Figure 6-37

to

size the

pump.

The

Effects of Laquad

Viscosity

on

Gentrifugal

Pumps

From the

previous

analysis

and Figure 6-36

we know

the

hydraulic performance required of

the

pump. Before

the actual

horsepower requirement for

the

motor

and

the

impeller

size can be determined, the

viscosity effects

of

the

liquid

being handled must be considered.

One re-

quirement

of a

centrifugal

pump

is that the

handled

liq-

uid

be relatively clean of suspended

particles.

Obviously,

for the same size

pump

and motor

a

highly

viscous liquid

will tax the unit

more

than

would

a

low

viscous liquid.

Thus,

the

viscosity is

an

important

property

that affects

the

horsepower of the

pump

motor. To account

for

this,

the

Hydraulic

Institute

has

prepared

charts shown

in Fig-

ures 6-38

and 6-39

for

determining viscosity

effects.

To

use the charts, the fluid being handled

should be

Newto-

nian. Gels, slurries, asphalt,

and other non-Newtonian

fluids should not be considered

with

these

charts. In han-

dling

such

fluids

a

positive-displacement pump

is

usually

required.

(Example

6-2 is an illustration

of

how

to

han-

dle

such a

liquid.)

To use Figure 6-39 we

must

convert the

absolute

vis-

cosity

io

kinematic viscosity.

This is done as follows:

p

:

139.53 cp at

90'F

w

:

54.725

lb/ft3

io.oooozog\tu-r..

rr-lh

(139.53)cpl

.

--"1

;;--(32.17)

-ij-i: -.

\

rcp

/

r(' rDr-sec'

th

154.'725)=

rt"

f12

z

:

0.0017-:-

sec

or

0.0017

ll

sec

centistokes

0.0000107639

i:

sec

v

:

159

.261 centistokes

Using Table

l-8

we

make

the

viscosity

conversion

from centistoke to SSU as

follows:

rq5

0.226r-::::=v

t

t2

-

704.695t

-

862.832

=

0

t

:

706 SSU

Now, looking at Figure 6-39

we

see

that for 150

gpm,

TDH

:

82

feet,

and

706

SSU

we obtain the

following

coefficients:

Cr:056

Ce:090

Cu

=

0.90 for 1.0

x

Q^*,

where

QNw

is the water

capacity

at which maximum efficiency

is obtained

The corrected

flow

rate becomes

^

sDm

150

Qc

=

"i...

-

:-:

=

166.61

=

167

spm

LO

U.YU



Rotating

EquiPment

69

O

o

@

(o

<o

{)

5lL

a



70

Mechanical Design

of Process

Systems

Pump Hydraulic

Design Calculation

Sheet

Liquid

fuel

oil

Viscosity at PT.

(Pumping

Temp.)

139.53

Vapor

pressure

at

PT

0.010

cp

psra

Sp.

gr.

(.y)

at

PT.

o.477

Flow

at ambient

temp. 150

gpm

Operating flow at

PI

150

gpm

Design flow at

PT.

150

_

gpm

Suction

Source

pressure

Static head

-

APr,

line

loss

Suction

pressure

-

Vapor

pressure

NPSH avail

NPSH

avail

NPSH req'd

1.52

65.08

=

-

0.01

=

65.07

_

171

t

=

82.017

64.7

'1.9

psra

psi

psi

psia

psia

psia

ft

ft

1.9

Discharge

Terminal

pressure

=

71-38

psia

Static head

-

APr

discharge

Piping system

Other

17.221

psl

psi

psl

psia

psra

psra

feet

Discharge

press.

=

96.201

-

Suction

press.

=

TDH

3'1.12

bhp at Duty Condition

DnpD

=

The total dynamic head

becomes

TNH

R'

Hc

=

'i-"

=;:91.

=

9l

fr

LH U.YU

Now, referring to the manufacturer's

curve

in

Figure

6-40,

for

Qc

:

167

gpm

and

TDH

:

91

ft,

we

deter-

mine the

pump

efficiency

as

n:63%

The

NPSH required

=

8

ft

To

correct the

efficiency

for

viscosity

we

have

r"

:

C,t

=

(63%)(0.56)

=

35.28% efficiency

The

brake horsepower

for

pumping

the

liquid

is

bho,,,"

=

QHl-

-

(167)19l)10.877)

-

9.53

ho

3,960

4.

(3,960X0.153)

Referring

to

Thble

6-4,

we see that the

next

larger

mo-

tor size

is a

10

hp rnotor,

thus we select a

3

x

lllz-in.

=ffi =

515hP=5v+hP

bhp at Back-Pressure

Condition

or'c*

=

Sffi

=

*AlrffiB

=

3.7o6hp

-

4hpwithwater

Figure 6-37.

Pump

hydraulic

design calculation sheet for Example 6-1.

centrifugal

pump

with a l0-hp motor

and a

5-in.

impel-

ler. In selecting a centrifugal

pump

it is desirable for the

required flow rate

to

fall

in

the middle

of

the

pump

curve.

Avoid extreme

sides

of

the

manufacturer's

perfor-

mance curves. Select an impeller that is at least two sizes

below the

largest

size available

for

the

pump,

because

if

greater

head is later required, e.g.

,

if additional

piping

is

added to the system, changing impellers is much cheaper

and expedient than

purchasing

a

new

pump.

In the final analysis the design engineer must not for-

get

the

potential problem

of back

pressure

that

the

pump

could

be

exposed

to under varying conditions. For exam-

ple,

if

the discharge line contained a bypass valve that

diverted

flow

to either the cooking kettle or to

a reser-

voir that collected

water,

the

reservoir would be used

if

and when the

pump

and

piping

system are cleaned

with

water

or a

cleaning agent. In this situation the

pump

would

have

to

be

sized

for

handling water

or

whatever

cleaning

is

to

be

used. When

the bypass

valve is

shut

off,

closing the

discharge

piping

connecting the

pump

to the

cooking

kettle,

the

flow

conditions are changed,

result-

ing in a lower TDH. With the same size impeller,

as

the

TDH lowers- the

flow rate increases as

the curve shifts



Rotating

Equipment

300

26

150

1(n

80

60

40

30

20

15

10

8

10,000

8,000

6,000

tO 15

20

25 30

40 50

60 70

80

90

100

CAPACITY-GALLONS

PER

MINUTE

Figure 6-3g.

Viscosity

corrections

for

capacities

of 100

gpm or less

(Courtesy

of

the Hydraulic Institute, Cleveland

Ohio.)

'4,000

3,000



72 Mechanical Design of

Process Systems

Figure

6-39.

Performance

correction

chaft for

viscous liquids.

(Courtesy

of the

Hydraulic Institute,

Cleveland, Ohio.)

i

F>

*.2

;t

?E

P,Z

E<

o6

;



Rotating Equipment

Table

6-4

NEMA

Frame

Dimensions

___o

Ir

r--i

F-

E

=q-

E

-->l

H-SIZE

HOLE

Source:

Goulds Pumps,

Inc.




of

Process

Systems

to the

right

in Figure 6-40.

Since

the impeller

does not

change, more horsepower

is required for

the lower

TDH. This condition

is known

as

the break horseoower

(bhp)

required at the

end

of

the pump

curve.

or

maxi-

mum

flow

capacity

condition.

In

our

case

we

have a

minimum

TDH

of approximately 45 feet in which

the

bhp becomes

bhp

=

{llE(s){l

0)

:

3.706 or 4 hp with water

'

3.960(0.46)

Thus,

we

see

that our 10-hp motor is

sufficient

against

back

pressure.

Often,

the

water

condition requires more

horsepower, and

thus a larger

motot

than

the

process

liquid

condition. The

design engineer must be always

cognizant of any other fluid that

the

specified

pump

may

have

to

handle.

EXAUPLE 6.2: POSITIVE

DISPLACEIIENT

PUMP

DESIGN

A

positive-displacement pump

is

required

to

transfer

a

adhesive coating mix from

a

storage tank to

a

bin

in

which

the

mix

is dropped

onto

a

nylon

sheet (see

Exam-

ple

3-6). The adhesive coating mix adheres the

particles

together to form roofing

shingles.

First,

we

must

perform

a fluid analysis of

the system

shown

in Figure

6-41.

Suctaon

Line

Pressure

Drop

N*"

:

DVP

:

{lP}n

(3.78r) a

tes.soer

k

\tzl

sec n"

Fluid

=

coating mix p

Temperature

:400'F

L

Pressure

=

20

psig

a

Suction

line

=

4

in.

Schedule

40

e

:

0.0018

p

:

938.08

cp

=

(938.08)(6.72

:0.6:0

lb'

ft-sec

4

lb.

(0.630)

_'

-

ft-sec

:

193. t 16

From Equation

1-6b we

compute the

friction

factor as

f:

-:-

=

0.332

Nn"

K.Values

(Velocity

Headsl

for

Suction

Line

Referring to Figures

1-7

and 1-11 we have the follow-

tns:

En'irance andexit

:

K

:

1.0 + 0.78

:

1.78

2-4-in.

plug

valves

:

K

:

2(18X0.017)

:

0.612

1-4-in.-90" standard elbow

:

K

:30(0.017):0.510

\-r

LtK

:

2.9O2 velocirv

heads

From Equation 1-4 we compute the

frictional

pressure

droo

as

follows:

oo, [<o.zs:xgo.ox

r2)

*

a.M8l

L

(3.068)

I

(e5.eoe)

k

(6.5ilr

g

F- -,-

fr-Ih

SeC'-lD1

Apr

:

40.822

psi

Referring

to

the

pump

hydraulic

calculation

sheet,

Figure 6-42,

we

summarize our

results. From this we

compute a total dynamic

head

(TDH)

of

93.76

feet. Past

experience indicates that a

rotary

gear pump

of the type

shown in

Figure 6-43 is excellent for handling

high

vis-

cosity

liquids.

The

pump

manufacturer

has

the

perfor-

mance

curves

rated

in

terms

of

kinematic

viscosity in

SSU. Now converting our

viscosity to SSU's we have

(rso)

sar

(_ri'

)tr_ry

min

\7.479

gat/

\60

sec

:

95.909

lb*/fc

:

11.0 ft

:

150

gpm

+

Dr

=

4.026

in.

x

10-4)

(t2.73)h.2H*l

:

3.781

ftlsec

Ssu

:

ll(.1,]1

635)

(938.08X4.635)

=

1.459.78 SSU

w/g

195.9091

l-l

\

32.2

l



L)

o

q

o

r)

(o

o

lt

a

z

E

o

o

ro

o

to

N

o

o

GI

o

lo

o

o

o

to

ir

o

o

o o

o

izu(o @sl-





Rotating

Equipment

Pump Hydraulic

Design

Calculation

Sheet

Liquid

adhesive

mtx

VG;o;itrt

PJ.

(Pumping

Temp.)

938

08

cp

\/.^^r

^;aee,,ra

at

PT

-

PSla

qn

^r

/_ I .r PT

.1.537

rioriat

ihbient

temo. -

not

lEn

Operating

flow at

PT.

j:X

YI:

{^n,.r PT

150

Suction

Source*

pressure

Static

+

(headlift)

-

APr

line

loss

Suction

pressure

-

Vapor

pressure

NPSH avail

NPSH avail

NPSH req'd

2+2=4ltrcqulred

14.7

psia

psi

psi

psia

psra

psia

ft

ft

Terminal

pressure

=

16.70

Static

tift

=

2g.g1-

-

aPr

discharge

Piping system

Other

13.74

Discharge

press.

=

53.75

-

Suction

press.

=

-8.70

psia

psl

psi

psi

psia

psia

psia

feet

-

4.0

-

2.O

=

8.70

=

8.70

ri

na

=

6-90

TDH

TDH

=

67.58

lrin

NPSH avail

>

NPSH

req'd + 2

lt

bhp at

Duty Condition

nr"^

_

(gpm)CrDHXr)

*

(150X67.58X1.537)

=

n

=

3g.g4o/o

(3,960Xrr)

(3,s60X10)

bhp at

Maximum Capacity

Condition

We

now refer

to

the

manufacturer's

performance

curves

which, in this

case,

are rated to the

viscosity

of

the service

fluid. The closest

curve is that

shown

in Fig-

ure

6-41.

As a starting

point,

it

is always desirable

to

start at the

middle

of

the curve.

Extreme ends

of any

pump

performance

curve

should be

avoided,

as the

pump's performance varies

significantly

at either end

of

the curve. Thus,

we select a very common

speed

for this

type of

pump-155

rpm.

Now for 150

gpm

and 62.45

psi

TDH, we find that

we need approximately

an

1l-hp

motor. Solving

for the

pump

efficiency

we have

bhp

=

Q(rDH)"y

(6-2)

(3,960)rt

Thus, we

have

,,

_

(150X93.76X

1.537)

:

0.496 or 49.6%

'

(3,960)(10)

This efficiency rating

is

quite

common

with a rotary

gear pump

handling

a

highly

viscous

liquid.

Now, refer-

..

(oom)ffDHX'v)

bnp"c

=

:(38;bX4-

TDH

=

total

dynamic

head

TDH

=

discharge

press.

-

suction

press

4

=

pump

efficiency,

o/o

ring to

Table

6-4

one can observe the

classifications

of

electric motors.

From Figure

6-44 we

see

that

the

viscos-

ity

of

our

fluid, 1,460 SSU,

is about mid-way

between

the

two curves shown.

Thus. the

required horsepower

is

between

8 hp and

l0

hp. Looking

at Thble

6-4

we see that

electric

motors

are lUz hp and

10

hp.

To meet

our re-

quirements,

we select

a

lO-hp motor, because

7llz hp

is

too

small.

Notice that

the

pump

has built-in

jacketed

en-

closures

to match the

piping,

which is hot-oii

traced, to

keep

the

fluid

in the

piping

and

pump

liquid.

These

jack-

eted systems

are discussed

in Chapter 3.

In

this

problem

we

have a suction

lift

on the suction

side

of the

pump.

It

is important to remember

that the

theoretical

height

to

which a

liquid

can

be lifted at

any

specified

temperature

is

the

atmospheric

pressure

at the

installation site

minus the vapor

pressure

of

the liquid

at

the specified

temperature

minus the friction loss

in

the

piping.

The theoretical

and maximum suction

lift for

wa-

ter

is

shown

for

various

temperatures

in Figure 6-14.

For

non-volatile

liquids, the maximum allowable

suction

lift

should

never

exceed

15

in.

Hg

(7.4

psia)

under

ideal

conditions.

For

volatile liquids, the maximum

allowable

Figure 6-42.

Pump

hydraulic design

calculation

sheet

for Example

6-2.




Complete

jacketing

ol

casing,

head

and rotor

bearing

sleeve

for

heating

or

cooling

liquids.

Hich ten

Dronze

for

long,

rugged

service.

on

head

for

handling

hot

liquids.

Figure

6-43.

The

type of

gear

rotary

pump

selected

in Example

6-2.

(Courtesy

of

Viking Pump Division,

Houdaille Industries,

Inc.)



Rotating Equipment

Figure

6-44.

Rotary

gear

pump

performance curve.

(Courtesy

of

Viking

Pump Division,

Houdaille

Industries, Inc

)

suction

lift

should never exceed

10

in.

Hg.

If

these

val-

ues

are

exceeded,

then

the suction

source

should

be

pres-

surized

with

a

neutral

gas

(inert

nitrogen)

to

offset

any

pressure

that

may

fall below

the

vapor

pressure

of

the

liquid.

At

the

liquid

vapor

pressure,

vaporization

occurs,

resulting

in

possible

cavitation

and

pump

damage.

A

Word

About

Prlming

A


pump,

like the

rotary

gear

pump

in this

example.

must

be

primed

when

pumping

low

viscosity

liquids.

This is done

by a

vacuum

device

or

by

using

a

foot valve.

Also, with

a

low

viscous

liquid,

the fluid drains

back

to

the

suction

when the

pump

is

idle. For

a viscous

liquid, like

the

one

in

this

example,

the

liquid

is

retained

in the

rotary

gear

clearances

and

thus acts

as

a seal

when the

pump

is restarted.

However,

before restarting

the

pump,

the

liquid

being

pumped

should

be introduced through

the

discharge

side

of

the

pump

to

lubricate the rotating

components.

Since the coating

mix is not

a

clean

service,

a centrifu-

gal pump

is impractical

because it cannot

handle a non-

Newtonian

fluid

containing

suspended

particles.

EXAilPLE

6-3:

CENTRIFUGAL

COiIPRESSOR

SELECTION

A

centrifugal

compressor

is to

be

specified

for

a

gas

plant,

which

is at sea

level.

The

unit

is to compress

3,000

lb./min

of

gas

mixture

at

50

psia

at

60'F

to

150

psia.

The

gas

mixture

is

composed

of

40%

ptopane,3O%

ethane,

and

30%

methane.

The

reduced

pressure,

P", the

reduced

temperature,

L,

the

molecular

weight,

and

the specific

heat

of

the

mix-

ture

is

determined

as

shown

in Table

6-5.

Using

the

data

in

the table

we calculate

the

ratio of specific

heats

for the

mixture

as

follows:

c-.

(6-10)

cp.

-

1.986

13.08

=

1.18

13.08

-

1.986

The compressibility

factor

for

the mixture

is deter-

mined

from

the

reduced

pressure

and reduced

tempera-

ture. Thus.



80

Mechanical Design

of Process Systems

Gas Mol

o/o

Table

&5

Tabulation of Gas

Mixture Properties

P"

(psia)

t"

("R)

Gas

Mixture

Pc

Propane

Ethane

Methane

44.t0

30.07

16.07

17

.64

9.O2

4.81

246.q

212.40

20Q.40

659.20

40

30

30

666

550

343

616

708

668

266.40

6.86

165.00

3.68

102.90

2.54

534.30

13.081.47

Table 6-6

16l

Typical Centrifugal Compressor

Frame Data*

Nominal lnlet Volume Flow

ffi

Frame (icfm)

(m3/h)

Nominal

Nominal

(lt-lbl/lbm)

(k.Nm/kg)

Nominal

Polytropic Rotaiional

Efficiency

Speed

(%)

(rpm)

Nominal

(in,

(mm)

lmpeller Oiameter

English

Metric

B

c

D

E

F

l,000-7,000

6,000- 18,000

13,000-31,000

23,000-44,000

3

3

,000-

65

,000

48,000-100,000

1,700-12,000

10,000-31,000

22,000-53,000

39,000-75,000

56,000-110,000

82,000- 170,000

10,000

10,000

r0,000

10,000

10,000

10,000

l

l,000

7

,700

5,900

4,900

4,000

3,300

406

584

914

1,120

|,370

30

30

30

30

30

30

l6

30

36

44

54

76

76

77

77

78

78

*Wite

this table is based on a survey of currently

available equipment, the instance

of an, machinery duplicating

this table

woud

be

purely

coincidenml.

P

.D

:0.076

0

Computing

the

compression

ratio

we have

^

P,

150

''

Pr 50

Assuming

that we have a

perfect gas

(z

:

l),

we can use

Equation

6-14 to

find

the average

discharge

temperature.

Thus, we

have

(6-14)

Now from

Eouation 6-32

we

have

659.20

t 60

+ 460

534.30

:

4.97 3

Now from

Figure

6-45,

we have

zr

:

0.972:

inlet compressibility factor

Using Equation

6-6

the

inlet

volumetric flow is

-

-

,mRt,

V:

----"

(6-61

(mw)Pi

,,

(0.972x3,000)(

l.545x60

-

460.)

(144)(31.47Xs0)

t1

n-r

/r-

r\

T=\-o

1"

From above,

Y

:

1O,339.276 icfm

(or

acfm at the inlet)

kr

:

l'18

Using Table

6-6

from

Lapina

[6],

we find our

unit

to

ltp'

=

0'76

he a Frame

B

with nominal

values

to be

as

follows:

Thus,

Hp"

:

10,000

ft-lbfnb.

N"

:

7,700

rpm

rp^

:

76%

l0

18\

r0.i6l

-

0.116

u.18/



from

which

n-I

0116

t?

=

tr(C

"

:

(60

+ 460X3.0)

or

t,

:

590.68'R

:

130.68'F

Now,

the

average

compressibility

for

the

gas

mixture

must

be

obtained.

From

above

the inlet compressibility,

zr

:

0.972

Compression

ratio,

p-

150

rc^r,=:j:

'-"

=

O.228

P,

659.20

Temperature

ratio,

,_,

_tz_560.68

_

r

^<

rR,2-L-534-30-'"-

Using

the

compression

ratio and

pressure

ratio

we de-

termine

the

outlet

compressibility

factor

from

the com-

pressibility

charts

in Appendix

E.

Thus,

zz

:

0'93

v

=zt

zz

_0.972

+

0.93

_

0.95

-22

In determining

the

polytropic

head

we

use

Equation

6-

33,

where

Pz=Pa

and

the average

ratio

of

specific

heat,

k, is

k

=

1.18

=

inlet

conditions,

which

is

an

approximation. Thus,

'

=

(-*-)

(*,{,)

[[&J-"*-"

-'],u

Rotating

Equipment

81

(6-33)

f-

1.00

compressibility

tactor,

Z

=

PV/RT

0.92

0.94

0.91

0.02

0.03 0.04 0.05

0.06

reduced

pressure,

Pr

Figure

6-45.

Compressibility

curves

for

very

low

values of reduced

pressure.

(Reprinted

by

permission of Chemicql Engineering,

Mc-

Graw-Hill

Company,

July

1954.)

N

---1

------J

-t

401

=

2.00

1.60

--

S

=

\

N

\s

iK

(\

S

-->

=

-'----

=

'1

0

\

\-"%_

ii(

><

\

x

-tl

N

ilxl

riP{

/-

x

\

-0.85

>i

'r;{

\*r-1

\l

\t

I

BO

-x

I

0.60

Y

'oS

*al

\"'r

\

\

\

0.01

0.07

0.08

0.09

0.10



a2

Mechanical Design

of

Process

Systems

from which

r

:

[(t't31srq'01

(8.62r)

r(3.0f

,,6

-

r]

H

=

29,913.143

ft-lbr/lb.

The

required

number

of compressor

stages

is determined

by

where

Ho.

:

maximum polytropic

head per

stage, ft-lb/lb.

(see

Figures

6-46 and

6-47)

Using

Table

6-5, we

have

Thus,

*-

='ni?l

lo'

-

2.ite

=

3

The required

rpm

is

I u

\05

N:N"l

..P

I

'

\Ho.

N.J

(6-10)

I rqqrr

lo'

N

=

r7 TOOr l,:-"'' |

-

7

l3l

rpm

Lr

r t.ooox

r)l

The required shaft

power

is

^

rir

H.

(3,000)(29.913)

r.t:

'

"

33,000

4o

(33,000X0.76)

P,r

:

3,578.11 hp

Using

Table

6-7

to

determine the mechanical

losses, L.,

we find

that

L.

:

(0.02sx3s78.11)

:

99.453 ta

(P.rL"*r

:

P.r

+

L.

:

3,578.11

+

89.453

:3,667.563hp

II

N.,

:

-q

.H-

(6-68)

(6-69)

[t26.

tttm'ntl

I

tzo.

r

lt:

r

.+zr

I

L

krzrrr

I

L(

| . t8

)(0.972)(520t

0

=

1.377

From Figure 6-46,

He.

:

11,000

ft-lbfnb.

12,000

11,000

10,000

9,000

8.000

7,000

6.000

5,000

4,000

3,000

E

1.0 1.1

1.2 1.3 1.4 1.5 1.6 1.7

1.8 1.9

2.0 2.1

0

6:

lNTuw

ltl

I

limit

for miled

yield

slress

I

mpeIers

I

Figure 6-46. Maximum

polytropic

head

per

stage-English

system

[6].



Rotating

EquiPment

Eru

=32

ot

928

Ezc

e20

o

.-

16

5rz

'i^

't.0

1.l

1.2

1.4

1.5

1.6

1.7

0

Figure

6-47.

Maximum

polytropic

head

per

stage-metric

system

[6]'

Table

6-7

[61

1.9

.8

Approximate

Mechanical

Losses

as

a

Percery

trl

u=@

-

I'n,J,,'*Lon"N

v

krzlTt

ttl

slress

impellers

English

(hp)

Metric

(kw)

Mechanical

Losses,

L,'n

(ohl

0-3,000

3,000-6,000

6,000-10,000

10,000+

0-2,500

2,500-5,000

5,000-7,500

7,500+

3

2.5

2

1.5

nents.Thistablewitt'howewr'ensurethatmechanicollossesareconsideredandtiea

uselul

valuas

for

estittutinS

purposes.

The

discharge

temperature

becomes

tz

=

rr(C

("

')/"

=

(520X3.0)0.r'6

:

590.68'R

tz:130.68'F

This

example

demonstrates

how centrifugal

compres-

sors are

estimated.

The

reader

should

be cautioned

as

when to

use

inlet

values for

the

values

of

k and

z.

The

value of

k will decrease

during

the

compressron

process

and calculations

for

the

polytropic

head

and

discharge

temperature

should

be

made

with average

values

of

k,

including single

stage

compressors.

Compressor

manu-

facturers

use

the

inlet

values

at

each

stage

of

compres-

sion,

but

the

inlet

values

for each stage

wi1l

be different.

In

calculating

the

polytropic

head,

the inlet

value

of

k

can

be

used

to

achieve

an

approximate value

of

the

head

with

some

error,

because

the

polytropic

head

is

insensi-

tive to

the

value

of

k

and

thus

n/(n

-

l).

The discharge

temperature

is

much more

dependent

on

the

value

of k.

Using

the

inlet

value

of k

will

yield

a con-

servative

value of the

discharge

temperature,

generally

25-50'F

in

extreme

cases.

For

a

more

detailed

discussion

of

the

specification

and

design

of

centrifugal

compressors,

the

interested

reader

is referred

to

Lapina

[6].



84

Mechanical

Design of Process

Systems

EXAMPLE

6-4:

INSTALLING

A

COMPRESSOR AT ELEVATION

A reciprocating

air

compressor is

to

be installed

in a

food

processing

plant, which

is

at an

elevation

of

6,562

feet.

The

desired

capacity is 33.3 m3/min. The machine

to be used is to

be refitted

and is

of

Polish

make. From an

elevation-barometric conversion

chart, such as Figure

6-48,

we determine that the atmospheric

pressure

at the

site

location

is

11.53

psia.

The compressor is to com-

press

the air to

7

atmospheres, or

102.87

psia.

Now,

r'

/^-

^.

.

rP\

v

:

33.3

r

l3s.314

",1

:

I.175.96 cfm

mtn \

m"i

Compression

ratio:

Pr

=

11.53

psia

Pi

:

102.87

psia

C-

:

t02

g

=

8.92

>

6.

thus requiring two-stage

I l.)J

compresslon

With an

intercooler,

you

must consider the

gas pres-

sure

drop across

it.

The minimum

horsepower is devel-

oped

when the ratios of compression

are

equal

in all cyl-

inders. The

ideal

case

is

with

no intercoolins

in which

Ludwig

[7]

suggests

Pr=Pr=&:...:

P"

Pr

P2

P3

Pn-r

and

with intercooling,

Po1

_

Pa2_ Po3_ Pr.

n

-P__,'-4-

4-'

(6-71)

(6-'72)

where

subscripts 1,2,3, ..., n

:

gas

conditions

across a cyl-

inder in which

I

represents

the

first

stage,

2

represents

the

second

stage, etc.

subscript

d

:

interstage discharge

pres-

sure condition, directly

at

the cylinder

prime

(')

:

represents

the actual

pres-

sure

to

the suction

of the

succeeding

cylinder,

which

rs the interstage

discharge

condition that

is reduced

bY

pressure

drop

over the

in-

tercooler system

subscript

f

:

final

discharge

pressure

from a multistage

machine

t4 t3 t? tl

Alfr

o3ph.ric

Pn33ur., lb./sq.

in.

Figure

6-48.

Atmospheric and barometric

pressures

at

vari-

ous altitudes

[7].

For

a

multiple

stage

unit,

the compression

ratio

is

p.

wnere

LD

:

i

'',

P.

D.

-z

p:

rol

p.

cD-

:

--:l

'.J

Dl

P,

^n

D1

'o.-l

Thus, for two

stages,

/P.

\0.5

t_,21

LRI

:

LR2

:

l;l

\r

l,r

Cnr

:

Crz

-

CR3

:.-":[bJ

(6-73)

(6-74)

Thus, the compression ratio

per

stage

is

approximately

CR:(8.92)05=2.99

and

for

the

first

stage,

Pr

:

11.53

psia

5

Pdr

:

(2.99x11.53)

+

i

=

36.94

psia

For

second stage,

Por

=

(2.99r(11.53)

-

i:

31.97

psia

8p00

2,000



Pr

:

102

87

Psia

The discharge

temperature

the

first stage

is

by

Equation

6-5s

ta,

:

ttFJ?

for

k

=

1.406,

tu,

:

(85

+ 460)(2.99)0'?8e

='147.94"R

or

tnt

:

287 94"F

The

discharge

temperature

for

the

second

stage

is

based

on the

discharge

temperature

from

the

intercooler.

The

intercooler

cools

the

air to

90'R

which

is the suction

temperature

to the second

stage.

Thus

kz

:

tiiR"G

tvr

=

(90

+ 460)(2.99)0

287

:

754.80'R

tr2

:

294.80"F

Selecting

the

Reciprocating

Gompressor

A reliable

and

quick

method

to

approximate

the

com-

pressor

size

is to

use

the

"horsepower

per

million"

iurves

depicted

in

Figure

6-49.

The

"horsepower

per

million"

ii the

bhp/MMcfd

and

is used to

determine

the

horsepower per

stage

by

the

following

relation:

rr:#:b(MMcrd)F,,

(*)

(6-75)

where F"n

is determined

in

Figure 6-50,

converting

the

acfm

to MMcfd

we have

MMcrd

=

(r.r75.e6){60x24)

('-lr;(14_:#.

J

:

I,421,068.508

For the

first

stage,

F.,

=

69.6

bhp

:

(6e.6)

Hi+Hfl

:

e8.eo6

hp

For the second

stage,

/,,

..\

1f9_ Jl)

Mcfd

: (

t.421.068.s08\j-r:)

touo

_

uu

/

Rotating

EquiPment

-.

11.203,486.3721

bho

=

(69.6)

l

','-

'=.

l

=

83

763

hp

'

\

l.u

x

lu"

/

Total

horsepower

:

98.906

+ 83.763

=

182.669

or 183

hp

Mechan

Gas

vek

3,000

f

Gas

ref(

intake

I

ical efiiciency,

95j

'city

through

valv€

(

|

(APl

equat(

to

'14.4

psia

.

rfll

$

z

i.?_

2

t

7/.,

/,,/'

v

lllllll

Ratios

below

1-4 are

subiect lo

signiticanl

etror, consult the

manufacturer

foa best

dala.

ttttttl

1.5 1.6

1.7

1.9

1.9

2.0

2.1 2.2

2.3 2.4 2.5

Ratio

of

comPression

Figure

6-49.

Power

requirements

for

reciprocatmg

compres-

sors.

(Courtesy

of Ingersoll-Rand

Company.)

1.5

2.0

2.5

3.0

Ratio

of

compr€ssion,

Figure

6-50.

Horsepower

correction

factors

for specific

grav-

ity

[8].

Equation

6-75

is

based

on a

given

compression

ratio,

Cp,

6rake

horsepower/

106

ft3ld

at 14.4

psia

and

suction

ternperature.

F,s

is

a

constant

which

is a

factor

for the

specific

gravity of the

gas.

9

I

q:

'

I

60

l:

58fi

561-

:

54f

1

521

50|.-

48r

46r

Ml

o2l

40l-

*l

36

l-

3o

l-

"'l

30

l-

28f

26

l-

24Y

22u

0.60

:1,203,486.372



86

Mechanical Design

of

Process

Systems

Next, the cylinders must be sized. This

can

only be

done after the interstage

temperatures and

pressure

are

defined

.

Because

of the clearance required to allow oper-

ation

and

permit

the

provision

of

passages,

the

piston

does not sweep the

entire

volume

of

the cylinder.

Thus, the actual

cylinder capacity is lower than the

displacement

of

the

cylinder.

Relating

this

in terms of

volumetric efficiency we have

where

4"

:

volumetric

efficiencY

Q

:

capacity at

inlet

conditions, acfm

Cp

:

cylinder displacement, ft3/min, where

""

=

I4*l ",)E'|"

\

144

I

\121

(6-77)

where

L

=

piston

stroke, in.

,46"

:

ar€r of

head

end

of

piston,

in.2

A""

:

area ofcrank end

piston

(,46"

minus the area

of

the

piston

rod), in.2

N: Ipm

A convenient

formula

recommended

by

Neerken

[8]

is

n.

=

o.si

-

..

[eU:l

I

zdtzs

I

(6-78)

where

C"

:

cylinder

clearance

Cp

:

compression ratio

k

=

ratio

of

specific

heats

2., za

=

colllpr€ssibility factors at

the

suction and dis-

charge conditions, respectively.

For our

machine

we

have the

following

design:

L

=

220 mm

:

9.661 in.

:

piston

stroke

N

-

500

rpm

Dr

:

500

mm

:

19.685 in.

=

diameter

of first stage

cylinder

Dz

:

300

mm

=

11

.81

1

in.

:

diameter

of

second

stage

cylinder

For the

first

stage,

piston rod

diameter

=

65 mm

:

2.559 in.

/r o

<rs\t

A,-

=

r l'-

"""1

=

304.341 in.2

-

\21

..

=

lrogL:

,2.]2t

)lr

uu'),roo,

"\

t44

l\t2 I

:

1,512.514 ft3lmin

For the second stage,

piston

rod diameter

:

60

mm

:

2.362 in.

o

LD

10e.563

in.2

_

*(.9)'

,n.,

:

105.181

in.'z

10e.563 + ro5.r8r

'l

{gjutl ,roo,

r44

l\

t2

I

(6-i6t

o,.

=

"

(";t")'

:

roe.563

in.?

c":l

:

538.165 ft3/min

The

volumetric

efficiency is approximated

by Equation

6-76

as

n,

=

o.si

-

(0.lr)[(2

ee)'i

-

r]

=

0.81i

:8t.iEa

This analysis is

only a

preliminary

estimate

of

what

the compressor

design is to be,

although in this example,

data is drawn from an

existing unit. The actual

selection

of

a

compressor can

only

be

accomplished

using the

manufacturer's

data

on

such

items

as

piston

displacement

and

the

volumetric efficiencies

of

the

cylinders. The

manufacturer's data

should always be used before

at-

tempting

a

final

design. The actual unit

in this example is

similar to

the one shown

in Figure

6-51

.

A more

detailed discussion on how

to specifr

and de-

sign reciprocating

compressors is

given

by

Chlumsky

t5l.

EXAMPLE

6.5: NAPHTHA PUMP

SYSTEiI

DESIGN

A

cosmetic

manufacturer

of

women's

lipstick

con-

tracted

a

chemical

company to formulate a chemical that

satisfies certain specifications.

The

chemical

process

en-

gineers

determined that a

light

cut

of

naphtha would

make an excellent

base

for

the

lipstick.

The

pump

in this

application can also be used to supply the

naphtha

to

a

small chemical

company nearby

for manufacturing

paint

thinner. This

second

application

is

called

the "maximum

capacity condition" and will be discussed after the

pump

is sized

for

the

first application. The

pump

must be sized

for both cases.

/r

sso\'

&.

:

304.34r

-

"

\;)

=

2ee.

re8 in.'



Figure

6-51.

Two-stage

reciprocating

compressor

with

a

shell

and

tube intercooler.

The

first stage

is achieved

with

the

vertical

.yiinder and

the

seconl

stagi

with tiie

horizontal

cylinder.

Pistons

of the

first

stage

are

aluminum

and

the

second stage

are cast

iion.

(Courtesy

of

Zaklady

Budowy

Maszyn, Aparatury

im Szadkowskiego,

Poland

)

In

the

first

case,

a

rail

switcher

transports

the

naphtha

to the chemical

plant from

a

nearby

refinery

The

plant

only

needs

to

send

one

50,000-gallon

railroad

tank

-car

once every

four

months

to

meet

the

cosmetic

manufac-

turer's

needs. The

light

naphtha

cut is

68"API.

The task

is

to

design

a

pump

and

hydraulic

system

that

will store

and

transport

the

naphtha

according

to

the

configuration

shown

in

Figure

6-52.

The

reservoir

is

large

enough

to

consider

the

fluid as

having

a

constant

head.

The

plant manager

estimates

that

the

naphtha head

required

is

12 feet,

but

wants

to

have

it

evaluated.

The basic

process involves

the

naphtha

passing

throush a scrubber

that

contains

caustic

soda

(NaOH).

The ciustic

soda

removes

the straw

color

in

the

naphtha,

resulting

in

a

colorless

liquid.

Next, the

naphtha

is

pro-

cessed

through

an activated

charcoal

filter

to

remove

the

fuel

odor.

Finally,

the

finished

process

liquid

is

loaded

into

the 50,000-gallon

tank

car.

In

the petrochemical

industry, the

specific

gravity

of

petroleum

is

given

in

terms

of

hydrometer

termed

'API.

The

relation

for API

is as follows:

"4p1

=

141.5:131.5

^tp

7w

where

.yo

:

the specific

gravity

of

the

petroleum

product at

60"F

l*

:

the specific

gravity

of

water at

60"F

(6-79)



88


Process Systems

9"

g',

g',

$+-Llj

',g',2,-O,g',

9"

5',-O"r

_L

NLL

=

normal liquid level

Figure

6-52. Pump-piping

scheme

of light

naphlha

cut used 10

manufacture women's

lipstick.

(Example

6-5).

The

relationship

between

the

'API,

temperature

is

given

in

Figure 6-53.

For our

case

of

68oAPI,

using Equation

6-79,

we

have

ro:141

5:

o.zo9

7*

199.5

in

which

7o

:

(0.709)(62.4)lb/ft3

:

44.26lbifC

at

60'F

The

maximum

pumping

temperature

is controlled

at

90'F.

The coldest

pumping

temperature

is at

34'F Since

the

density

is higher at the

lower temperature,

that is the

one

used

for

frictional

pressure

drop calculations.

Thus,

referring

to Figure

6-53

"Yp

:

0'13

and

p

:

45.55

lb/ff

The

Flow

from

the Reservoir

to

Naphtha

Storage

llank

The reservoir

is of such large

magnitude that

the

head

of

liquid

is considered

constant,

because

the

railroad

switch

engine

delivers the

naphtha regularly

to the

plant.

The

flow

rate

from the reservoir

to

the storage

tank

in

gallons

per

minute

is determined

from the following

ex-

Dresslon:

o

:

rr.os o'(\)"

(6-79)

The velocity

heads

on the

line from

point

@

to

point

@

are as

follows:

Values

of f1

are determined

from

Figure

1-7.



Entrance:K:0.78:0.78

2-4-in.

plug valve: K

:

18

fr

:

18(0.017X2)

:

0.612

Exit:K:1.00:1.00

sr-

LtK

:

2.392

Rotating Equipment

APr

:

1.223

Ot'

12

ft

of

naphtha

head

=

(12X0.325)

psi

:

3.900

psi

3.90

psi

<

10.5 psi

nitrogen

pad

This

pressure

differential

will

cause the

naphtha to be

forced back

into

the reservoir. The number of feet

re-

quired

to

deliver

the

liquid to

the tank

will now

be deter-

mined.

Since

we already

have

12

ft

in the tank,

then

x

+

3.90

)

10.5

psi

x

=

6.60

psi

=

20.308

ft

Adding

an

additional 26

ft

of

head

we

have

38

ft

:

12.35

psi

-

1.223

psi

:

I l. 127

psi

>

10.5

psi

The new

flow

rate

is

o

ft

:

4.46'1

:::-

sec

-.

DVp

l\Re

:

-

=

p

:

93,088

fi

=

'7

.948

-:-

sec

f

l rnin

I

\oo

r""/

2.640

lb.

I

I

hr

I

ft-hr

\3,600

sec/

r-05:

-2r"r.[+.

rt=*)

(1-6a)

Nr":

0.0884 ft,

(rgl*

(7.e48rx

a

r+s.ss1$

\ul

sec

r"

:

165,633

t-- \

,,,

aP,

=

ILL +

l- x.l4I

t.t

H

|

16

\s I

'Etc

o,,:

[rr1*opu"'

*

r.,nl

t(Hft I

(4s.5s)k

$.46if

##j

fr-lh

SeC"-lDf

"*n

tb

lt* \

-

-

-

ft-hr

\:.ooo

sec/

Applying

Equation

l-6a, f

:

0.0319

r.-"1

aP.

=

l(o.o3l9x1o5.83)

ft

+

2.3921

|

4.026

-

|

|

-n

I

t12J

rrelr'

ft-lb.

SeC'-lOf

AP1

:

3.364

Or'

38

ft

-

APr

=

12.35

psi

-

3.864

psi

:

8.486

psi

8.486

psi

< 10.5

psi

pad

Select a 6-in.

{

Sch

40

pipe

Q

:

1e.65(4.026)

[rrrt*J"

:

177.2

wm

a

=

r. r co [2

4

lb./ft-hrl

:

z.u

lb^

' '\

lcp

/

ft-hr

0.0884

ft,

(1 4|-(a.a67;x

A(45.ss)k

Q

:

1e.6s(4.026)

(#r)0'

:

315.317

gpm

Using Np" to check

the

friction factor,

f:

0.03198

Now,

2.51

(93,088)(0.17875)

(1-4)



90

Mechanical

Design

of

Process Systems

Repeating

the

hydraulic

analysis

we have

Entrance

and

exit:

K

:

1.78

2-6-in.

{

plug valve: K

:

18

fr

:

18(0.015)

:

0.27

[]-::Yl(a.8a)(a5.55)

N*"=''-';

i

:100.863

2.6401

r

I

\3,600/

f+

:

0.032

(from

Equation

1-6a)

r l.---.

./

r\

or,

-

l,o.orrt,zr.rtr,

* ,.rrnlI

tt'd'*l@

-ll+.oze\

|

2(32'2\

t I

't

I

I

APlo

=

9.5110t'

For a

3-in.

Iine,

L

:

1.0

ft

,rnr,l-l\[)

_

\7.47e1

\601

-

(0.0s130)

-

|

3',0168),s.:0,,+s.:s,

tl".: I}- ..

=

t32.449

2.6401

1

I

\3.600/

8.34 ft/sec

e

:

1e.6s(6.065)

(,

or--)"

=

513.107

gpm

19'ut\,r.rouo,

,r,

\

12,

rr ne

-

--

----7------ -lll

t*tr-oj

f:0.01803

t

rll

l(0.0r803x105.83)

,

^ ^.^lt+s

ssxs

zot'?\r++/

APr:l

-

-r

-.U)Ur-

-

ffi-

I l6.o6sl I

z\rz'z

I

r

\-rzl

I

APr

:

9.939

Ot'

38

ft

-

APr

:

12.35

psi

-

0.930

psi

:

11.42

psi

>

10.5

pst

So

there

is 0.92

psi (11.42

-

10.50)

net

positive

pres-

sure

head of naphtha entering

the storage

tank.

Naphtha

Pump

Hydraulics

Suction

Line

For

4-in.

Sch

40

portion

of line, L

:

23.313 ft

K-Values

0.2006

Dr :

z.oso

fr

5.700--

sec

a4

-

1'79

0L')

To

determine

the flow rate we must

consider

what

the

system

is to

service. Plant operations dictate that the

loading of the tank car

must not

take

longer

than

four

and

one-half hours.

The

rail

tank

car

capacity is 50,000 gal-

lons. We select 4.35 hours, which

yields

a flow rate of

s9.990

eat

{+ )

:

re

|

.57

r

=

re2

spm

4.35

hrs

\60 min/

We

will

size a centrifugal

pump

with

192 gpm

capacity. For

192

gpm,

0D(#(*J

:

4.84 ftlsec

(0.0884)

Entrance and

exit:

K

:

1.75

1-4-in.

plug valve: K

:

0.306

4-in.

x

3-in.

reducer: K=0.163

sr-

LtK

=

2.219

For

3-inch

Sch 40

portion

of line:

K-Values

3-in.d

plug valve: K

:

(18)(0.018)

:

0.324

4-in.

x

3-in. diffuser: K:Kr:0.055

D*-*"

f:0.0344

t

la5.5sl,t.r,'(,;)

re,.

=

l(o

ol'l4t(t

or

-

o 3791-

''

l /r.oos\ |

2\32

2)

r

\,2/

r

APi.

=

9.175

ntt



Rotating

Equipment

The total

pressure

drop

for

the

suction

line

:

AP,

AP.

=

APlo

+ APr3

:

0.686

Psi

Discharge

Line

L

=

60.708

ft,

4-in.

Sch

40

K-Values

(0.0884)

1-4-in.

swing check:

K:

(100)(0.017)

:

170

1-4-in.

gate

valve:

K

:

(8X0.0i7)

:

0.136

4-4-in.

plug valves:

K

:

(4X18X0.017)

=

1.224

5-4-in.

std

90'

elbows:

K

=

(5)(30)(0.017)

:

2.550

Entrance:K=1.0:i.00

D"

:

oôto

(45

5r(8

34r(*)

2(32.2)

t

ar,,

=

l(o

ol2'(6oi7o8)

+

o.oro

|

14.0261

t

I

r? |

bhe

-

(19?)(6172)(9

73)

=

3.i

or

a4

hp

motor

(3,960x0.61)

The

Maximum

Capacity

Condition

The small

chemical

company

nearby

that

manufac-

tures

paint

thinner

needs

the

naphtha

only

about

once

a

year.

However,

when the

naphtha

is

needed,

it

must

be

delivere.d

quickly.

Consequently,

delivery

time

is

crucial

to

the client.

i1,uti,,

,0,,0,

,r,

I 1) |

N^".'"-';

i

=

132'449

-ttl

z

e+o[3

roo/

fi

=

0.0344

rl

aP,,

:

l(o 934'(i

o)

* o.e43l

I lr.068l

I

t

I

17

I

r

AP1,

:

0.460

Psi

The

total

pressure drop

for the

discharge

line

-

APo

APp

=

AP1,

+

APi.

:

|

42'7

+

0

460

:

1

887

psi

From

Figure

6-54,

the

pump

hydraulic

design

calcula-

tion

data

sheet,

it is obvious

that the available

NPSH

is

much

higher

than

the required

NPSH.

This

means

that

the

10.5

psi

pressure for

the

nitrogen

pad

is

excessrve.

The

minimum

pad

pressure

required is

ATM.

pressure

(psia)

+ x

+

static

head

(psi)

/

friction oressure

\

i/

tiquio

uupot

'i

=

\arop

on

suction

line

tAP.rf

=

\preisure

rpsiarJ

where

x

:

minimum

pad pressure required,

pslg

14.7+x+3.557:0.511

+

20.85

+ 21.361

psi

select

x

=

Tpsig

Referring

to

Figure

6-55

and

6-56,

we

re-evaluate

the

pump

performance.

Since the

light

naphtha cut

has a

low

viscosity

bhp

:

QHr

(6-2)

3,96Ou

14

026'1,+.

s+x+s.ssr

t*.

=

E*'

too.8b3

2.6401

I

I

\3,600/

f

:

0.032 (from

Equation

1-6a)

APlo

:

1.427

Ot'

For the

3-in.

portion

of

the discharge

line,

L

:

3

ft.

For

3-in.

Sch

40

pipe,

d1

:

3.068

rn.

K-Values

Entrance:

K

:

0.780

4-in.

x

3-in.

reducer:

K:0.163

DK

=

o,sa3



Mechanical

Design of

Process

Systems

Equivolents

of

Degrees

APl,

Specific

Grovity,

Weight

Pounds Per

Gollon ot

Degrees Boum6,

Densily,

ond

6OF/5OF

Degrees

API

Baum€

Scale

Values

for API

Scale

oil

Values for Baum6

Scale

Liquids Lighter Than

water

Liquids

Heavier

Than

Water

peci6c

s

W€ight

D€nsity,

LblFt3

Pounds

per

Gallon

Specific

Gravity

s

Weight

Density,

Lb

/Ft3

Pounds

pef

Gallon

Specific

Gravity

.s

Weight

Density,

Lb/Ftx

Pounds

per

Gatlon

8.337

8.454

8.574

4.697

4.8L4

8.955

9.0E9

9.XX8

9.371

9.518

L67r

9.828

9.990

10.159

10.332

10.512

10.698

10.891

1r.091

11.297

11.513

11.737

11.969

12.2rO

12,462

12.998

13.244

r3.583

13.895

14.22?

14.924

t5.302

15.699

16.118

..'

'..

l0

t7

I4

l8

0

2

6

8

20

22

'),4

'].6

28

30

32

34

36

38

40

42

46

48

50

54

5E

60

64

66

68

70

7X

78

80

a2

84

E6

88

90

92

94

96

98

100

'.:

,.oooo

0.9861

0.9725

0.9593

0.9465

0.9340

0.9218

0.9100

0.E984

0.8871

0.8762

0.8654

0.8550

0.8448

0.E348

0.8251

0.81 55

0.8063

o

.797 t

0.7883

0.7796

o.77tl

o.7624

0.7547

0.7467

0.7389

0.7313

0.7238

0.7165

0.7093

0.7022

0.6953

0.6E86

0.68r9

0.6754

0.66S0

0.6628

0.6566

0.6506

0.6446

0.6388

0.fl3r.

0.6275

o.6120

0.6166

0.6112

,,.

61.50

60.65

59.83

59.03

57

,87

56.03

54.64

52.69

52.06

5l .46

50.86

50.28

49.7?.

49.

l6

48.62

48.09

47.57

47 .07

46.57

46.08

45.61

45.14

44.64

44.23

43.79

43.36

42.94

47..53

42.12

4t .72

41.33

40.95

40.57

40.20

39.84

39.4E

39-

13

38.79

38.45

38.12

8.337

a.xll

8.108

7

.998

7 .891

7

.787

7.587

7

.490

7

.396

7 .305

7.124

7

.043

6.960

6.879

6.799

6.646

6.499

6.429

6.359

6.292

6.160

6.097

6.034

5.973

5.913

5.854

5.797

5.474

5.424

5.274

5.186

5.r41

5.096

','

1.0000

0.9859

4.9722

0.9589

0.9459

0.9333

0.9211

0.909r

0.4974

0.8861

0.8750

0.8642

0.8537

0.8434

0.8333

0.8235

0.8140

0.8046

0.7955

0.7865

0.7778

0.769X

0.7609

0 .7 527

0.7447

0.7368

0.7292

o.7216

0.7143

o.7071

0.7000

0.6931

0.6E63

0.6796

0.6731

0.6667

0.6604

o.6542

0.6482

0.6422

0.6364

0.6306

0.6250

0.6195

0.6140

0.6087

oa.s

6t.49

60.63

59.80

58.99

58.20

56.70

54.57

53.90

53.L4

5L.60

5t.97

50.76

50.18

49.61

49.05

48.51

47 .97

47.45

46.94

46.44

45.95

45.4E

45.00

44.10

43,66

43.22

42.40

42.34

41.98

41.58

4t.19

40.80

40,42

40.05

39.69

39.33

38.98

3E.63

34.29

37 .96

''.

8.337

E.Zt9

8.105

7 .994

7 .886

7

.781

7 .679

7.579

7.48X

7 .387

7

.295

7 .205

7.117

7 .03r

6.947

6.786

6.708

6.484

6.413

6.344

6.209

6.143

6.079

6.016

5.955

5.836

5.774

5.722

5.666

5.506

5.454

5.404

5.306

s.ttl

5. 1r9

r.0000

1.0140

1.0284

1.0432

1.0584

l.o74l

1 0902

1.1069

1.1240

l.t4t7

l.1600

1.1789

1. 1983

1.2185

1.2393

1.2609

1.2832

1.3063

1.3303

1.3810

|

.4078

|.4356

1.4646

1.4948

1.5591

1.5934

1.619L

1.6667

1.7059

1.7470

1.7901

I . E354

1.EE31

1.9333

,'.

62.36

63.24

64.t4

65.06

66.01

66.99

67

.99

69.03

70.10

7r.20

73.52

75.99

77 .29

7E.64

80.03

8t

.47

42.96

86.13

67.80

89.53

91.34

95. l9

97 .2]

99.37

101.60

103.94

106.39

108.95

I I1.64

r

14.46

t17.44

120.57

Figure

6-53.

Relationship between 'API

and temperature.

(Courtesy

of

Crane Company.)



Rotating Equipment

Pump Hydraulic

Design Calculation Sheet

Liquid

Viscosity at

PT.

(Pumping

Temp.)

Vapor

pressure

at

PT.

Sp.

gr

(-y)

at

PT.

Flow

at ambient temp.

Operating flow at

PT.

Design flow at

PT.

Light

Naphtha Cut-68'

API

'1.1

_cp

0.73

20.85

psia

gpm

gpm

gpm

't92

't92

Source

pressure

Static head

-

APr,

line loss

Suction

pressure

-

Vapor

pressure

NPSH

avail

NPSH

avail

NPSH req'd

=

28.248

=

-

20.85

Terminal

pressure

=

Static

head

(litt)

=

-

APi discharge

Piping system

Other

Discharge

press.

=

-

Suction

press.

=

Discharge

6.313

1.887

20.o

44.90

28.244

Suction

25.20

psra

psi

psl

psl

psra

psia

psia

feet

psra

psi

psi

psia

psra

psia

ft

ft

-

0.51

'l

7.398

1.3

bhp at Duty

Condition

bhp at Maximum Capacity Condition

onpo

=

Q{l(IPr)1

onp""

=

QSTrylI1

(3,960Xn)

(3.960Xr)

Figure 6-54.

Pump

hydraulic design calculation

sheet

for

Example 6-5.

Referring

to the

pump

manufacturer's

pump perfor-

mance

curve, Figure 6-55,

we

see that

approximately

400

gpm

is the maximum

limit.

Using this

flow

rate we

re-evaluate

the

pump

for the maximum capacity

case.

Suction

Line

Referring to

previous

calculations on

the suction side

we

have the

following:

u.

:

{gl

r4.84r

=

io.o8jl

\tvtl

sec

I 1n ORI

N^.

:

l',"i,'l

r100.863)

-

210.062

\

+.d4

/

From

Equation 1-6a we obtain

f

:0.0315

(4s.ss)(ro.osf(1-L;

APr,,

:

2.200

ps

For the

3-in.

portion

of

the suction line,

u.

:

lq)

r8.34r

=

r7.37sa

\t92l

sec

**.

:

(,rr4)

032,44s)

:275,e35

From Equation 1-6a,

f:

0.03395

APr

:

9.759

Ot'

AP,:APso*APi.

AP,:2.29*a.trt

AP,

=

2.959

Ot'

on,.

=

['o

o:'n''lt

o'

*

o

rrnl

rffi l

It

I

2(32.2)

or,.

:

[ro

o1le1'"'r

* r,nl

rffi I

2(32.2)





Rotating

Equipment

Pump Hydraulic Design Calculation Sheet

Liquid

Viscosity al PT.

(Pumping

Temp-)

Vapor

pressure

at PT

Sp-

gr.

(1)

at

PI

Flow at ambient temp.

Operating flow at PT-

Dosign

flow at P.T.

Light

Naphtha

Cut-68o

API

1.1

cp

psra

gpm

gpm

gpm

20.85

0.73

't92

192

'192

Suction

Source

pressure

=

21.70

Discharge

Static

head

-

APr,

line loss

Suction

pressure

=

-

Vapor

pressure

=

NPSH

avail

NPSH

avail

NPSH req'd

=

3.559

Terminal

pressure

=

Static head

-

aPr

discharge

Piping system

Other

Discharge

press.

=

-

Suction

press.

=

TDH

psia

psl

psi

psra

psia

psra

ft

tt

'16.7

psra

psi

psi

psi

psra

psia

psia

leet

6,313

-

0.51

1

24.744

1.847

-

20.85

12.3

24.744

1.3

20.'152

=

63.77

bhp at

Duty

Condition

bho"

=

(SPmXTDH)(?)

(3,960X4)

bhp at

Back-Pressure

Condilion

bho""

=

(gPm)[rDH)(?)

(3,s60)(a)

Re-evaluation of

pump hydraulic

design

calculation

sheet

of Example

6-5.

igure

6-56,

f:0.03395

v.

:

{g}

(4.84)

=

ro.o8l

\t921

sec

N*

:

(lryrt

(100,863)

:

zto,o6z

II

l(0.03395X3.0)

^ ^.^l

l-h

=

l------l---------

+

U,y4Jl

I l4ql

I

'

\

12

/

J

APlr:1.939

,or.rr,,rr.rrrr(r{)

Discharge

Line

Referring

to

previous

calculations

on

the

discharge

side

we have the

followins:

f:0.0315

t.^

^".-

..^

-^^

l,or.rrxro.oo,{*}

ap,.

_

l(0.031sx60.708) +

6.610l

--"-'--'\r44l

-l

[+.ozo\ |

2(322\

I

IrrI

)

AP6n

:

6.143

O.'

For

3-in.

portion,

*""

:

(,1q;

032,44e)

:275,e35

2(32.2)

APp

:

AP6o

* AP1,

=

6.143

psi

+ 1.989

psi

APo

:

3.132

ntt

Referring

to Figure 6-57, we reevaluate

the

pump

for the

maximum

capacity condition.

Normally,

we would

use a 9.5-in.

impeller,

as

indi-

cated

on

the

pump

manufacturer's

curve, Figure

6-55. In

this

case, being that

the application

is infrequent, we

keep the

8.Gin. impeller.

As the

flow rate

increases

with

the

same size

impeller,

the TDH

decreases

and the

re-

quired

NPSH

increases.

As we see

on Figure 6-55, the

available NPSH

of 4.589

ft

is

slightly

exceeded at

400

u,

=

lgl(8.34):

r7.37sa

\t>Ll

sec



96

Mechanical

Design

of

Process

Systems

Pump

Hydraulic Design

Calculation

Sheet

Maximum

Capacity

Condilion

Reevaluaiion

Light

Naphtha

Liquid

Viscosity at

PT.

(Pumping

Temp.)

Vapor

pressure at

Pl

Sp.

gr- (r)

at

PT.

Flow

at ambient temp.

Operating

flow at

PT.

0.73

cp

psia

gpm

gpm

4no

Desion flow at

pT.

4uu

gpm

Discharge

Terminal

pressure

=

16.70

Psia

Suction

Source

pressure

=

21.70

Psia

3.559

tatic head

-

APi,

line

loss

Suction

pressure

=

-

Vapor

pressure

=

NPSH avail

NPSH

avail

NPSH req'd

-

2.959

psl

psi

psia

psia

psia

ft

ft

Static

(lifi)

-

APr discharge

Piping system

Other

Discharge

press.

-

Suction

press.

6.313

8.132

=

20.00

=

51.145

psl

psi

psl

psra

psia

psia

feet

TDH

TDH

-

20.85

1.45

4.589

=

22.300

=

24.845

=

91.282

20.85

bhp

at

Duty

Condition

ono"=9##

gpm.

It

is suggested

that

a

flow rate

of375

gpm

be used

to

avoid

cavitation.

From Figure 6-55 the actual

TDH

is

TDH

:

34

ft

The

required

brake horsepower

rs

..

(375

x34.0X0.73)

-

J'v'

ttv

'

(3,960X0.65)

A

4-hp

motor is sufficient

for normal and

maximum

ca-

pacity

operations.

Re.evaluation

of Reservoit

Line

Since the

nitrogen

pad

on the

naphtha storage

tank

was

decreased

from

10.5

psi

to 7.0

psi,

we must

reconsider

the line

size.

With 38

feet of head in the

reservoir,

we incurred

a

pressure

drop

of

3.9

psi, yielding

an entry

pressure

of

8.5

psi.

In the

back-pressure condition,

we

need

a

flow

rate of

375

gpm.

The new

presure

drop

in the line

con-

Figure

6-57.

Maximum capacity

re-evaluation

of

pump hydraulic design calculation sheet

of Example

6-5.

bhp at

Maximum

Capacity Condition

.

.

(oom)fiDHXr)

bnp""

=

=.(3GbX4.

necting the

reservoir

to the storage

tank, considering

the

pipe

to be

4-in.

schedule

40, is

as

follows:

(3?r

(ciL)(=*-]

tl+tl

,,

_

"

-'

\min/ \7.a79

eau

\60

sec/

=

9.45j

a

0.0884 ft?

sec

:

196,992

2.640

lb'

fchr

From Equation

l-6a,

f-05

:

-lr"c.

[+

.

*]-tt)

lnr

f:0.031s



From

Equation

1-4

we

have

aP.:lLL*rrl

Pv'

\d

-

l2e"

or,

=

[,o

olualgrl,o

*

r.rnrl

I

(r flr

I

1+s.ss1

p

1r.+s:;,#

(*q-J

fr-lh

SeC'-lDf

APr

:

5.41f

Or'

With

38 feet

of

head

in

the

reservoir

we have

an entry

pressure

to

the storage

tank of

38ft:12.008psi

Entry pressure

:

12.008 psi

-

5.411

psi

:

6.597

psi

Because

6.597

psi

<

7.00

psi

pad,

we

keep the 6-in.

schedule 40 pipe.

The

6-in. line was

evaluated for 513

gpm,

so it is adequate

for the 375

gpm

in

the

4-in.

line.

The system is now

completely

designed for hydraulics,

using

a

4-in.

x

3-in. horizontal

centrifugal

pump.

NOTATIOl{

acfm

=

actual

cubic

feet

per

minute,

ft3lmin

bhp

:

6.u1"

horsepower, hp

e

:

clearance

volume,

in.3

Co

:

specific

heat

at

constant pressure,

Btu/lb--

mole-"F

C.

:

compression ratio

C"

:

specific heat

at constant volume,

Btu/lb.-

mole-'F

D

:

diametef

of

impeller

or rotor, in.

D"

=

specific diameter,

dimensionless

ghp

=

gas

horsepower

:

horsepower

delivered

to

gas,

hp

H

:

head

:

energy

per pound

of mass,

ft-lb/Ib.,

or better known

as feet

of head, ft

icfm

:

actual

cubic feet

per

minute

at compressor

in-

let, ft?/min

J

:

mechanical

equivalent

of

heat: 778 ft-lbrl

Btu

Rotating Equipment

97

k

:

ratio of specific heats

:

CplC,, dimensionless

m

:

mass,

lb.

and re-expansion

polytropic

expo-

nent

dl

=

mass flow rate, lb-/hr

fiIo

:

moles

of

gas

:

m/mw

mw

:

molecular weight

n

:

polytropic

exponent

N

:

speed, rpm

N,

:

specific speed, dimensionless

NPSH

:

net

positive

suction head, feet or

psia

P:

pressure,

psi

Q

:

flow rate,

gpm

or ft3/sec

R

:

R/mw

:

gas

constant

of

a

particular

gas

R: universal

gas

constant

:

1545

ft-lbr/lb. mole-

scfm

:

standard cubic feet

per

minute, ft3lmin-see

discussion under standard volumetric

flow

t:

temperature,

"F

At

:

temperature differential,

oF

V

:

volume

of

gas

or cylinder,

ft3

v

=

specific volume of

gas,

ft3/1b*

w*

:

weight

of

fluid

whp

=

*ur".

horsepower, hp

y

:

constant

:

(k_

lyk

z

:

compressibility factor,

dimensionless

Greek

Symbols

?

:

specific

gravity,

dimensionless

4

:

efficiency, expressed

as

percent

€

:

ratio of

clearance volume

to the volume sweot

by the

piston

stroke

p

:

density, 1b./ft3

REFEREilCES

1.

Buchter, H. Hugo,

Industrial

Sealing Technology,

John

Wiley

&

Sons, New York,

N.Y., 1979.

2.

Dimoplon,

William, "What

Process

Engineers Need

to Know About

Compressors,"

Compressor Hand-

book

for

the

Hyd,rocarbon Processing

Industries,

Gulf

Publishing

Co.,

Houston, Tx., 1979.

3.

Balje,

O.8.,

'A

Study on Design

Criteria and

Matching of Tirrbo-machines-Part

B,"

Trans. ASME,

J. Eng. Power,

Jan. 1962.







100 Mechanical

Design

of Process

Systems

WARM WATER OUT

KEBOSENE

IN

KEROSENE

OUT

(cooLED)

COOL

WATER IN

Figure

7-1. An

example

of a fixed

tubesheet

heat exchanger.

(Courtesy

of Howell

Training

Company.)

ISOBUTANE

VAPOF

LEAVING AT

2OOOF

orL

ENTEBTNG

AT

6650F

LIOUID ISOBUTANE

LEAVING AT

2OOOF

LIOUID

ISOBUTANE

ENTERING AT I95OF

Figure 7-2. This

U-tube exchanger represents a kettle type reboiler.

(Courtesy

of Howell Training

Company.)



FRACTIONING TOWER

(DE

ETHENIZERI

PAOPANE

&

PAOPYLENE

50%

VAPOR

-

50%

L|OUTD

PROPAN€ AND PROPYL€NE

50%

vaPoR

50%

LroulD

PROPANE ANO PFOPYLENE

100% Ltouto

Design

Classifications

of Heat

Exchangers

Typical

shell and

tube

heat

exchangers and their func-

tions

are

as

follows:

Reboiler-transfers

heat to

a liquid

to

produce

a two-

phase, gasJiquid

mixture

used

in

a

distillation

col-

umn.

Thermosiphon Reboiler-provides natural circulation

of

the boiling fluid by a static

liquid

head shown

in Fig-

ure 7-3.

Forced

Circulation

Reboiler-a

reboiler

in which

a

pump

is

used

to force

the

liquid

through the

heat

ex-

changer

(reboiler)

into

the

distillation

column.

Condenser-a heat exchanger to

condense

vapors by re-

moving heat from a

gas.

Partial Condenser-only

partially

condenses a

gas

to

provide

heat

to

another medium

to

satisfy a

process

The Mechanical Desien

of

Shell-and-Tube Heat Exchangers

101

Figure 7-3. Iilustration of a

thermo-

s]phon reboiler.

(Courtesy

of

Howell

Training Company.)

condition.

The residual

gas

is

recirculated through a

heater and recycled.

A

common

application

is using

excess

steam

to

heat

up a

process

fluid.

A typical

ap-

plication

of

a

partial

condenser on a distillation col-

umn

is to condense only enough

liquid for the reflux

when

the overhead

product

is

vapor.

Final Condenser-an exchanger where

all

the

gas

is con-

densed

and

all

the heat

is transferred to the other me-

dium.

Steam

Generator-a

device

that

generates

steam, such as

a

boiler.

to

provide energy

for

process requirements.

The most classic example is

the old stearn locomotive,

which is

a shell and tube exchanger

"mounted

on

wheels" with the

steam used

to Dower

the

locomotion.

(This

unit

is

a

fired

vessel

and is not

covered by ASME

Section VIII

Division.)

Vaporizer-an exchanger

that

fully

or

partially

vaporizes

a liquid.

Chiller-an

exchanger

in which

a

process

medium is

cooled by evaporating

a

refrigerant,

or

by

cooling and

heating

with

little

or

no

phase

change.

CONDENSATlON



102

Mechanical

Design

of Process

Systems

'AIIONARY

HEAD

IYPIS

A

ANO iEA{OVA8TI

COVEP

B

BONNST

(INIEGRAI

COVER)

c

CHANNET INTECFAL

WITH

IU8E.

SHETT

AND

RE/nOVASIE

COVTR

N

CHANNEI

INIEGRAL

WIIH

''UBT-

5HEET

ANO

REITOVABLE

COVER

D

SPEC|AL

hICH

PREsSURE

CTOSUI€

F

WTh LONGIIUOINAT

3AFFIE

G

H

J

K

x

These classifications

are the

major

types of

services

that

shell and tube

exchangers provide

in

the

process

in-

dustries.

Process requirements

dictate

the type of

design

to

be

used. Figure

7-4 shows

some

of the

major types

of

con-

struction.

The

standard

TEMA

classification

of ex,

changers is

to use

the shell

identification

and number

with

the exchanger

designation

type. For

example,

an

18- 150 BEM

is

an

exchanger

having

an 18-in.

shell

with

150

tubes, a bonnet

(integral)

cover with

a fixed

tube-.

sheet at

one end

(B

in Figure

7-4),

a

fixed

tubesheet

and

a stationary head

at the

other

end

(M),

and a

one-pass

shell

between

both ends

(E).

Figure

7-4. Nomenclature

of shell and

tube heat

exchangers.

(@1978

by Tlrbu-

lar Exchanger

Manufacturers

Associa-

uon.)

Fixed

Tubesheet

Shell

and Tube

Heat

Exchangers

Fixed

tubesheet

shell

and

tube

heat

exchansers

are the

simplest

of

the

shell and

tube designs.

They

ionsisr of

a

tube bundle

attached

to a

tubesheet

on each

side of the

tube

bundle. The

tubesheets

are welded

to

the

shell

pro,

viding

an absolute

seal

to

prevent

the shell-side

fluid

from

leakage.

Often the

tubesheets

extend

beyond

the

shell diameter

and have

flange

bolt holes

that allow

the

tube heads

to be bolted

to the tubesheets.

In fixed

tubesheet

exchangers,

tubes

can

fill

the

entire

shell

to

achieve

maximum

heat

exchange

(of

course, this

I

Uff

'

" SIAIIONARY HEAO

tn

Ul(E

"4"

STATIONARY HEAO

N

LIKE

'1T

STAIIONARY

HEAD

P

OUI5IOE PACKED

FTOA'ING

'IFAO

s

T

PUIT TIiROUGH FIOATIIIG

HEÔ

U



also increases

shell-side fluid

pressure

drop) such that

tolerances

between tubes

are minimum. However, this

factor limits the

shell-side fluid to

a relatively clean ser-

vice,

because the exterior

of

the

closely-packed tubes

cannot be mechanically

cleaned

or

inspected. Another

limitation

to

the design

is

that there

is

no allowance

for

thermal

growth

of the tubes

,

except if an external expan-

sion

joint

is

used, which is

quite

common

for this

type of

exchanger. Normally,

single convoluted bellows

are

used since the maximum

temperature differential is

200"F

and the cyclic loading is

insignificant.

Tube-side headers,

channel covers,

and internals

of

tubes can be cleaned

quite

easily and the

shell side can be

cleaned only by

circulating

a cleaning

fluid or backwash-

ing.

U.Tube Shell

and Tube

Heat Exchangers

U+ube shell and

tube heat exchansers

consist of one

tubesheet with

tubes bent

in

a

U-shipe atrached to rhe

single tubesheet. This

type

of

exchanger

is used for large

temperature

differentials

where there

is

a

lot

of

tube

growth.

This type

of

design

allows

for easy access to the

The Mechanical Desisn

of

Shell-and-Tube Heat Exchansers

103

shell

side of the tubes

and

removal

of the tube bundle.

The inside

of

tubes must be cleaned

with

soecial

tools

and then only when the bending

radius is fairly large.

This

tne

of design is also very suitable for

chemical

cleaning.

The

maximum number

of

tubes per

tubesheet

is

less

than the fixed

tubesheet

design

beciuse

of

the minimum

bending radius required

to

form the

U-shape.

The U-

tube design is also

very

applicable

to high-pressure ser-

vlces.

Floating

Head

Shell and Tube

Heat

Exchangers

This

type of shell and tube heat

exchanger

has

a float-

ing

head

that

is

designed

to accommodate thermal expan-

sion of the tubes and to

provide

access

to the

tube-side

and shell-side exchangei components.

This

type

of

de-

sign

is

expensive and its use should

be

considered against

other

possible

designs.

Packed

Lantern

Ring

Exchanger

(Figure

7-5a).

This

construction

is

normally limited

to design tempera-

floating-head

cover

shell

(A)

Packed

lanternring

exchanger

backing ring

flange

(B)

Outside-packed floating

head

exchanger

f

tlange

gasket

floating

tubesheet

floating-head cover

floating-head

cover

shell

cover

floating

tubesheei

gasket

(C)

Internal floating head

exchanger

(D)

Pull-through

lloating head

exchanger

Figure

7-5. Several

configurations

of floating

head exchangers.

gland

tollower



'lO4


Process

Systems

tures

<

370"F

and

design

pressures

< 300

psig.

This

type

of

design

is used

only

for mild

services,

such

as

steam, air,

low

viscous

oils. In

this design the shell-side

and

tube-side

fluids

are

sealed

by

separate packings

which, in turn, are separated by a lantern

ring.

The lan-

tern

ring fits between the

packings

that separate the shell

and

tube-side

fluids

and normally contains

weep

holes

that

accommodate any leakage through the

packing.

Such

leakage, which is

passed

to the outside and drops to

the

foundation below, will not cause shell and tube-side

fluids to

mix.

The tubesheet must be designed such that it

is

large

enough

in

diameter to encompass

the

packingJantern-

ring ensemble and differential thermal expansion of the

tubes.

Occasionally,

a

skirt

is attached

to a

thin tubesheet

to act

as a bearing surface for the

packingJantern-ring

ensemble.

Outside-Packed

Floating Head Exchanger

(Figure

7-56).

Rings of

packing

contain

the

shell-side

fluid,

which

is

compressed

by

a

gland

follower

that

is

guided

by a tube

sheet

skirt. The skirt is integral to the

floating

tubesheet.

This

removable-bundle construction allows

for differential

expansion between

the shell

and

tubes.

This

design

is normally

limited

to 600"F and 600

psig,

which

is one reason why

it

is the most commonly

used

removable-bundle

type exchanger in the

petroleum-

chemical

industry, even though usage

has

decreased

over

recent

years.

Internal

Floating-Head Exchanger

(Figure

7-5c).

This

design consists

of

an

internal floating tubesheet

held

by an internal backing ring,

which is bolted to an

internal

floating head cover.

The internal backing ring

and internal

shell cover are beyond the end

of the shell

containing

the tubes. To remove

the tube bundle, the

shell cover,

split backing ring, and internal

floating

head

cover

must be

removed.

The

internal floating head cover

acts as a return

cover for the tube fluid

with

an

even

number of

tube-side

passes.

with an odd

number of

tube-side

passes,

a nozzle must

be

extended

from the

in-

ternal

floating-head

cover through the

outside shell

cover.

Clearances

between the shell and the

outermost

tubes are

1rla in. for

pipe

shells

and 17re

in. for medium-

sized rolled

plate

shells. This design

is more suitable

for

higher

shell-side temperatures

and

pressures

than

for

pull-through bundle types

of

construction.

This design

has been

used extensively in the

petroleum-chemical in-

dustry,

but there

has

been a decline

of

use

over the

past

few

years.

Pull-Through

Bundle

Floaiing-Head

Exchanger

(Figure

7-5d).

This

design

consists of a floating

head

di-

rectly

bolted to

an

internal floating head

cover. The tube

bundle can

be removed without removing

either

internal

floating

head

cover or shell cover

when bundle is

pulled

out

an opposite end of shell cover facing internal floating

head.

This

feature

reduces

down

and maintenance

time

during

inspection and

repair.

The clearance

between

the

outside of

the

tubes and

shell

inside must be

sufficient to allow

space

for both the

gasket

and

bolting at the internal floating head cover.

This clearance

is

usually twice that

required for

the

split

ring

design used

in the internal floating head in the

pre-

vious section.

This

type

of

design

is normally limited to

services

where leakage of the internal

gasket

is tolerable.

With an odd number of tube-side

passes,

a nozzle must

extend

from the internal floating-head cover through

the

shell cover. The

number

of

tube-side

passes

is simply

limited

by

the number

of

tubes. This design

is

generally

suited

for

lower

temperatures

and

pressures than

that

of

the

internal floatine

head

exchanger described

earlier.

General

TEIIA

Exchanger Glasses-Rr

Ct

and

B

There

are three basic categories

of

shell and tube

heat

exchangers

in

TEMA-Class R, Class

C,

and Class

B.

The difference

in class is the degree

of severity

of ser-

vice the exchanger

will

encounter.

Descriptions

of the

three classes

are

as

follows:

Class

R includes

heat exchangers specified

for the most

severe

service in the

petroleum-chemical

pro-

cessing

industry. Safety and durability

are

re-

quired

for exchangers

designed

for

such

rigor-

ous conditions.

C/css C includes

heat exchangers

designed for the

gen-

erally

moderate services

and requirements.

Economy and overall

compactness

are the

two

essential

features

of

this

class.

Class

B

are exchangers

specified

for

general

process

service.

Maximum economy

and

optimum

compactness

are

the main criteria

of design.

Rubin

[3]

described the

TEMA classes of

exchangers

in

terms

of the

various components

and

how they

vary

from one class

to another.

This

data is

given

in Table 7-1.

Ludwig

[4]

described

various types

of

heat exchangers,

their applications

and

limitations,

which include shell

and

tube exchangers

as

well

as

other types.

This data is

oresented

in Thble 7-2.

-

tbles 7-1 and

7-2

provide

a comprehensive

view

of

the

various types

of heat exchangers

and their

applica-

tions,

so we

can now focus

on the components

of the

shell

and tube design.







Baslc

Gomponents of Shell

and Tube

Heat

Exchangels

There are various components to

a shell and tube heat

exchanger, but the following are the essential ones:

1.

Tubes

2.

Baffles

3. Tie

rods

4. Tubesheets

Tubes

There

are basically two types-finned

tubes and bare

tubes. Finned tubes have

external fins mounted by vari-

ous mechanical means.

The necessity

of

having

external

fins mounted on tubes is

to

provide

more heat transfer

area and thus more heat influx

to

the

tube

fluid.

Finned

tubes

are most common where there

is

a

gasJiquid

or

gas-gas

transfer of heat with

the

gas

always being exter-

nal to the

tubes. Typical applications

of finned

tubes are

waste heat recovery

exchangers, waste heat boilers,

gas

turbine regenerators,

and air-cooled

exchangers. Exam-

ples

of

some finned tube

designs are shown later.

Plain or bare tubes are the most

common

in

shell and

tube design. These tubes come in

two basic

types-solid

wall

construction

and

duplex construction. The duplex

design consists

ofa

tube

within

a tube in which the outer

tube is mechanically

drawn over

the inner tube.

The

solid wall tube is what the name implies,

a simple tube

of

solid wall construction.

Tubing is available in almost

as

many

materials as

piping

and

is

available

in

standard

gauge

sizes

listed in

Table

7-3,

along

with

diamerers

and

section

properties.

In

applying the U-tube

exchanger

design, tubes must

be bent 180'. Thble

7-4 lists

the

recommended

minimum

bend

radii.

Baffles

Baffles

serve several functions

and consequently

the

design

of each is dependent

on its

purpose.

Baffles

can

act

as:

l

Structural

supports

for

the

tubes.

2.

Dampers

against

vibration.

3.

Devices

1o control

and direct

flow

Datterns

of

the

shell-side liquid.

Baffles

as Tube

Structural

Supports. Like

piping,

tubes

behave as

structural

beams

and consequently will

develop excessive

deflection, or

sag, if

left

unsupported.

Baffles

act

as

the

structural supports in

the shell

and

tube

exchanger.

Another structural

function

of

baffles is to

add

stiffness to the tubes so that each

tube. in effect. is

The Mechanical

Design

of

Shell-and-Tube Heat Exchangers 1o7

constrained

at

each

baffle. Thus, the hole in

the

baffle,

being larger by

varying amounts than the outside tube di-

ameter,

acts as a limit stop

for

the

tube. In

piping

me-

chanics

(see

Chapter 2)

a

limit stop is

a

restraint

that

lim-

its the

amount

of

pipe

(in

this case, tube) movement

to

the

distance between

the hole diameter and the

outside

diameter of the tube. In other

words,

the tube

can trans-

late

in the lateral direction

perpendicular

to

the

tube

axis

only

by

the amount of clearance between the tube

OD

and

the

hole

diameter.

Translation is mentioned instead

of rotation because even though the tube rotates,

it is

in-

significant.

Thus,

the

baffle

hole acts as a limit stop and

prevents

lateral buckling of the tubes when they are

in-

duced to thermal

expansion by temperature

differentials.

In this

sense the tubes are much

stiffer

and stronger than

they would be

without

the baffle supports. The conse-

quences

of strengthened tubes

affect

the integrity

of

tube

joint

connections

in

the

tubesheets and

this

will

be dis-

cussed shortly.

We

see

from

this discussion that the baf-

fle

plates

act as both structural supports

and as

buckiing

stabilizers.

Baftles as

Tube

Vibralion

Dampers.

Figure 7-6

shows

baffles of circular rings with

rods that run

verti-

cally

in the

first

two

rings

and

horizontally

in the second

two

rings,

thus damping

vibration

much in the same

way

as helical

vortex

strakes on stacks

(Chapter

5). The rods

break up forming vortices that induce

vibrations, a

phe-

nomenon discussed in Chapters

4

and

5

called vortex

shedding. The rods also reduce

turbulence to below

res-

onant

levels

of

the natural

frequency

of

the

tubes

and

they reduce fluid elastic

vibration.

Baffles

Conlrol and Direct the Flow

Pattern

of

the

Shell-Side Fluid.

There

are

various

types

of

baffles

that

direct and/or

control

the

flow ofthe

shell side fluid.

Fie-

ures

7-l

and 7-2 are

examples

of baffles

guiding

or d'i-

recting

the

flow in

the vertical

direction. Fig]ure

7-7

shows baffles

diverting

flow

in the horizontal

direction.

The

flow

direction is

a

function of

the orientation of the

baffles

and their respective

geometries

and

is

dependent

upon

process

requirements.

The

arrangement in Figure

7-7

is

said

to

be

vertically

cut

and

the

arrangements

in

Figures 7-l

and 7-2 arc said

to

be horizontally cut.

Often,

process

conditions require

the

shell-side fluid

to

flow

horizontally,

parallel

to the longitudinal axis of

the exchanger. This

arrangement,

called

a longitudinal

baffle,

is shown in Figure 7-8. Figure

7-8a shows a two-

pass

shell-side arrangement and Figure

7-8b

shows

a

four-pass

shell-side arrangement. The

baffles control the

flow

in the sense that both the

direction and

flow

rate are

dependent

on orientation and number

of

passes,

respec-

tively. With the same inlet

flow

rate, the fluid

velocity





The Mechanical

Design of

Shell-and-Tube Heat Exchangers 109

Table 7-4

Minimum

Tube Bend Radii

l4l

Tube

Outside

Dia.

(in.)

Bend Radius

(in.)

Center-to-Center Oistance

(in.)

Duplex, all

sizes

*Plain:5/s

I

*For

bends this sharp, the tube

wall

on the outer circumference

of the tube

ma\

thin down lt/z

to

2

gauge

rhicknesses. dependin| on condition and specific

tube materiaL Morc

genercus

ndii

\9ill reduce this thinning. TEMA

presents

a

formula for

calculating the minimum

wall

thickness.

Figure 7-7. Baffles

can divert flow horizontally.

(Courtesy

of

Howell Training Company.)

3 times

Tube O.D.

t3/te

1

131t6

6 times Tube OD

15/s

2

2z/s

Figure

7-6.

Although complex, this design

eliminates

tube

vi-

bration.

To use

this configuration, one

must be cognizant

of

pressure

data

[5].

(Courtesy

of Heat Transfer Engineering,

Hemisphere

Publishing

Corporation,

New

York,

Washington,

D.C.)

Figure 7-8. Longitudinal

baffles direct

flow

in the

axial di-

rection.

(Courtesy

of Howell Training

Company.)

VAPOR

IN LET

FLUID

IN LET

FLUIO

OUTLET

CONDENSATE

OUTLET



1 10 Mechanical

Design

of Process

Systems

increases as the flow

area decreases,

that is, the velocity

increases with

an increase in

the number

of

oasses.

The

control

of

flow in

exchangers is

accomplished as

well with orifice

baffles. Figure 7-9

shows an annular

orifice baffle.

To

utilize

this type of design

a

very

clean

shell-side

fluid

is required,

since

the fluid must flow in

the annular

space between

the tube outside

diameter and

the

hole in

the

baffle

forming the

orifice.

The flow at the

orifice is very turbulent

and the

pressure

drop

through

an

orifice-baffle

arrangement

is very high.

Consequently,

these baffles are

not used

often in industry. Also,

since

the orifice baffle requires

a very

clean

fluid,

non-New-

tonian fluids are

completely ruled

out.

We will

see

later

in the chapter that the

plate

fin type of

exchanger is supe-

rior to the shell

and

tube

design

for

many clean

services.

The

reason

for the shell

and

tube

desisn

to

be

dominant

is

because

of

the

wider

variery

of

fliids

it

can

handle

versus

any other design.

Other baffle

arrangements

are

possible

with

varying

baffle shapes and orientations.

Figure

7-10 shows baf-

fles

in disc and

doughnut shapes,

which disperse the

flow

in

a

radial direction. Baffles

can be cut to

allow for

horizontal or

vertical

flow in varying

amounts

as

shown

in Figure 7-11.

Tie

Rods

These are structural rods that run

oarallel to the ex-

changer tubes through

the

outer perimeter

of

the

baffles.

fastened to the tubesheets such

that they space and sup-

port

the

baffles. Tie rods,

being attached to the baffle

plates,

also

prevent

them from vibrating

and damaging

the tubes. Table 7-5 lists what

TEMA recommends

as

a

minimum number of

tie

rods and

rod

diameters

for

a set

of

shell diameters.

Tubesheets

These are the structured

plates

in which

the tubes are

connected at each end

ofthe exchanger.

Tubesheets

come

in

two basic types-single

and double. Double tube-

sheets consist

of

two tubesheets mounted together at each

end of the tubes with a

clearance between

the

two

sheets.

The

reason

for

using

two tubesheets at each end is to re-

duce the

possibility

of

a

leak of

the tube-side

fluid.

Dou-

ble

tubesheets

are

quite

common

with

highly

toxic

ser-

vices, where a leak cannot

be

tolerated.

Single tubesheets are much more common than double

tubesheets because

ofprocess applications

and economy.

Typical tube-tubesheet

connections are shown in Figure

1 1a

Of

great

immediate

concern in tubesheet design is the

loading induced

by

the tubes thermal movement, which

Figure 7-9.

Annular orifices

between tube outside surface

and

hole in

baffle

plate

[6].

Figure 7-10. Doughnut and disc type baffles

[6].

Table 7-5

TEMA Tie Rod Standards

(in.)

"c" & "8"

Exchanger

Tie Rod

Dlameter

8-15

r6-27

28-33

34-48

49-60

Nominal

"R"

Exchanger

"R"

Exchanger Tie Rod

ShellDiameter

Dlameter

irinlmum

Number

of

Tie

Rods

3/z

3/t

tlz

tlz

3/a

tlz

rlz

tlz

4

o

o

8

10





112

Mechanical

Design

of

Process

Systems

for

unsupported

tube lengths

between

two

tubesheets

for

unsupported

tube

lengths between

a

tubesheet and a

baffle

for unsupported

tube

lengths

between two

baffles

Et

:

modulus

of elasticity of

tube material at mean

tube metal

temperature,

psi

4

:

outside

diameter of

tubes,

in.

oc

:

allowable

tube compressive

stress,

psi,

for

the

tubes

at the outer

periphery

of the

tube bundle

Equation 7-1

is

based

on

Euler's

columl

equation and

Equation

7-2 is based

on the short

column formula

de-

veloped by Professor

J. B. Johnson

during the nineteenth

century.

Other

TEMA

formulations

are summarized

in

the

fol-

lowing sections. The

reader is urged

to be

familiar with

the TEMA

standard and

follow

its

guidelines

in

design-

ing a shell and tube heat

exchanger.

TEMA

Formulations

Baffles and

Support Plates

Natural Frequencies

ot

Straight Tubes

on

Multiple

Equal

Spans

3.36C

where

f"

:

tube natural

frequency,

Hz

C

:

mode

constant

from

Thble

7-6

I

:

span

length,

in.

E

=

modulus

of

elasricity. psi

I

=

moment

of

inertia, in.a

(Table

7-3)

W

:

Wr + Wn

+

MWr",

lbs/ft

Wt

:

weight

of empty

tube

(Table

7-3)

Wq

:

weight

of

fluid

inside

tube

0.00545

p1d1,

W6o

:

weight

of

fluid

displaced

by

tube 0.00545

p"d"'?

M

:

added

mass

coefficient

from Table

7-6

p

:

fluid

density,

lbs/ft3

d

:

diameter

of

tube, in

subscripts:

i

:

inside

o

:

outside

r

{o'

['o

Allowable

Tube Compressive

Stress-Periphery

of

Bundle. The allowable

tube compressive

stress,

psi,

for

the

tubes at the

periphery

of

the

bundle is

given

by:

-28

a,:ffi

when

C. s

kf/ror

-r

-.

I

s"=\l

r

-

(kur)l

whenc

>kur

21 2C"l

/:*

where

C"

= l/

^

Vsr

Table 7-o

Mode

Constant-C

[21

No.

of

Spans

Extreme

Ends Supported

Fr-l-'-l*,.1

|--___l

/T-7\--lzf-R

Extreme

Ends

ClamDed

,l-r+r

Extreme Ends

Clamped-Supported

r-fr-fr

lst

Mode

2nd Mode lst

Mode

2nd Mode lst Mode

znd Mode

I

2

3

4

5

6

7

a

9

to

31.73

31.73

3r.73

31.73

31,73

31.73

126.94

49.59

&.52

37.O2

34.99

34.32

33.67

33.O2

33.02

72.36

49.59

40,52

37.O2

34.99

34.32

33.67

33.02

33.02

33.02

198.34

72.36

59.56

49.59

44.r9

40.52

38.40

37.O2

34.99

49.59

37.O2

34.32

32.37

31.73

31.73

160.66

63.99

49.59

42.70

39.10

37.O2

35.66

34.99

34.32

33.67



KT:

yield

stress,

psi,

oftube material

at design metal

temperature used.

radius

of

gyration

of tube

0.25

.vu

+la"

-

2tJ1,

in.

(Table

7-3)

equivalent unsupported buckling length

of

the

tube,

inches.

Use

the largest value

considering

unsupported tube spans.

unsupported

tube span, in.

0.6

for unsupported

spans between two tube-

sheets.

0,8 for unsupported

spans

between

a

tubesheet

and a baffle.

1.0 for unsupported

spans between

two

baf-

fles.


of

Shell-and-T[be Heat Exchansers 113

quency,

assuming

simple supports

and

for the first mode

only, may

be calculated

as

follows:

2.74C"

=

U-tube natural frequency,

Hz

:

mode constant

for

U-bend

:

bend radius, in.

Note: For

other than simple

support conditions

the calculated

frequency may be

estimated

by multiplying the

above

value

for f,, by

the

appropriate ratio

of mode constants

from Thble 7-6

using single

span values.

ASME

Tube

Joint

Load

Grlteria

The

ASME

Secrion

VItr

Division

I

Dressure vessel

code

lists formularions in

evaluating

tube forces exerted

on

tubesheets. Referring

to

Figure

7-13

and Table 7-7

the

formulas

for the maximum

tube force

are as

follows:

For

joint

types

a, b,

c, d, e:

F,

:

A,o,11f,

For

joint

types

f,

g,

h,

i,

j,

k:

(7-3)

R2

where fnu

R

Note:

The value

of S"

shall not

exceed the Code allowable

tensile

stress

of

the tube material at desisn metal tem-

perature

used.

Effect

ot Longitudinal Tube

Stress

where fnp

:

tube

natural

frequency in

stressed

condition, Hz

P

=

axial force, lbs

(positive

for tensile,

negative

for

compressive)

Natural

Frequencies of

Straight Tubes on Unequal

Multiple

Spans

f"

:

10.83

t'z

For

a

tube on multiple

unequal

spans

with

the extreme

ends

fixed

and simply supported

at the

intermediate

sup-

ports,

ki can be obtained

by

solving the following

char-

acteristic determinant for

an n span

system.

Natural Frequencies

of

U-Tubes.

It

must be recog-

nized

that

each

tube is

a continuous beam

that has a

sin-

gle

fundamental

frequency.

This

frequency

may

be

largely

governed

by

the lowest "stand

alone" frequency

of either

the

longest

straight

span

or the

U-bend.

It

is

suggested

that both be calculated

and

that the lower value

be used,

keeping in

mind

the

approximate

and somewhat

conservative nature

of

the result.

The

straight

span

fre-

quency

may be

determined from Thble

7-6 using

the ap-

propriate

mode

constant.

The

U-bend

out-of-plane fre-

F,

:

A,o"11f,f"f,

where

Ft

:

oall

:

f=

f.

(no

tesg

=

f,

(teso

:

(7-4)

maximum tube

joint

force, lb1

cross-sectional

metal

area of tube, in.2

ASME maximum

allowable

stress.

psi

joint

reliability

factor

maximum value

without

test

given

in

Table

7-'7

maximum value

with

test as

specified in

the ASME

Section

VIII Division 1

code, per

section

UA-002

Figre

7-14 shows how

the tube

joint

load varies

for

various

tube

gauges

of various

process

conditions.

Natu-

rally, as

the tube

wall increases,

the

tube stiffens

and,

consequently,

the force exerted

by the

tube on the

tube-

sheet

joint

increases. The

engineer

should

evaluate

the

tube loads

with

the

various process

conditions

possible

and use the

worst

for

determining

the maximum

tube

joint

force, as shown

in Figure

7-14. The

TEMA stan-

dard

gives

the formulations

to determine

the

tube

ioint

lorces

and the user

is referred

to

this

standard for

these

expressrons.

The buckling

of

exchanger tubes

can be

a

problem

if

thermal

expansion

is not

properly

accounted

for in

de-

Dt2

'Er.,j



114

Mechanical Design

of

Process

Systems

Table 7-7

Reliability Factors, f,

[71

Type

Joint

Descriptions

Notes l.

(tesr)

f,

(no

test)

(1)(7X8)

(1X2)

(1X3)

(1X6)

(1X7X8)

(1X4)(s)

(7)

(

l

)(4)(s)

(7)

(l)(4)(5)

(7)

(l)(4xs)

(l)(4x5)

(l)(4)(5)

a

b

c

d

f

c

h

I

j

k

Welded

only,

a>

1.4r

Welded

only, tsa<L.4t

Brazed, examined

Brazed, not

fully

examinable

Rolled,

welded,

a> l.4t

Rolled, two or more

grooves,

and

welded,

a< l.4r

Rolled,

single-groove,

and

welded, a

<

1.4r

Rolled, no

grooves,

and

and

welded, a

<

1.4r

Rolled,

two or

more

grooves

Rolled,

single

groove

Rolled, no

grooves

1.00

0.70

1.00

0.50

1.00

0.95

0.85

0.70

0.90

0.80

0.60

0.80

0.55

0.80

0.40

0.80

o.75

0.65

0.50

0.70

0.65

0.50

Notes:

(l)

The

use

of

f.

Ceso

factor

requires

qualification in

accordance

with

UA-003 and UA-004.

(2)

For

welds where a is less than

t,

fi

(no

test)

-

0.

Tubes with Type

(b)

joints

where a<t may be considered as acting as stays

and

contributing

to

the

strength of the tubesheet only when the

joint

is tested in accordance

with

UA

003 and UA-o(X.

(3)

A value of 1 00 for f,

(test)

or .80 for f,

(no

test) can be applied only to

joints

in

which

visual examination assures that the brazing filler metal has

penetrated

the entire

joint

[see

UB-14(a)] and the depth

of

penetration

is not

less than three times the

nominal

thickness

of the

tube

wall.

(4)

When the

ralio of

OD.

to

LD., using

nominal

tube dimensioos,

is less than 1.05 or

geater

than l-410,

qualification

in

accordance with UA403 and

UA-oO1 is required.

(5)

The nominal

pitch

used in the desigo of tubesheets for roller expanded

joints

shall

not be less than the

following:

P

=

d" + 0.165

(d"

+ 2r)

=

nominal

pitch (center-to-center

distance of adjacent

tube holes), in.

=

tube

o.D_,

in.

=

nominal

thickness of

average

wall tube, in.

except that:

(a)

nominal

pitch

shalt

not

be

less

than

4

+ 2t unless the

joint

is

qualified

in accordance

with

UA-003 and UA-004; and

(b)

96% of the

ligaments

between tube

holes

throughout

the

thickrcss of

the rnachined tubesheet

shall not be less than 0.85

(P-4).

Ligaments

which

do not

meet

this

requirement

shall

be evaluated and

€orrections made as may be

necessary.

(6)

A value

of

.50 for

f,

(test)

or

.40 for

f,

(no

t€so shall be used

for

joinls

in which

visual

examination

will

not

provide proof

that the brazing filler metal

has

penetrated

the entire

joint

Isee

US-14(b)1.

(7)

The

value of f.

(no

test) applies only to material combinations as

provided

for under Section

IX.

For

material combinations

not

provided

for

under

Section

IX, f. must be determined by test in accordance

with

UA-003

and LIA-0O4.

(8)

For

joint

types

involving

more than one fastening method, the sequence

used in the

joint

descriptions

does not necessarily indicate the order in

which

the

oDerations

are

Derformed.

I

sign. One

such formulation to

predict

the critical buck-

ling

load is as

follows:

P.,

- ,

"

q''

,,

t0.5216r

t7-51

I

L**

l'

\Ns

+

t/

where L,u6"

:

total length

of tubei

between tubesheets

NB

:

number of baffles

Equation

7-5 is based on the Euler column

formula.

In

situations

where there are several baffles, such

that the

effective

length,

L",

divided

by the radius of

gyration,

k,

is between

30 and 120,

exclusive,

then the

Johnson short

column equation

is more accurate. For

a

tube

to be con-

sidered

as

a series

of short columns constrained

by fixed

ends, one

must be certain that the

baffles constraining

the

tubes

allow practically

no translational

or

rotational

movement.

The stiffness

of

the baffle

plate

should be

analyzed,

as

small

translational

and rotational

tube

movement

allowed

by

the

baffle

plate

could considera-

bly

alter the

buckling characteristics

of the

tube.

The

evaluation of a

baffle

plate

containing several

tubes

can

be a somewhat

detailed

analysis,

and

it

may

be faster to

consider the

tube as a continuous

beam

in determining

buckling

characteristics.

For

further details on

the

mechanical design

of ex-

changers, the

reader is referred

to TEMA.

We will

dis-

cuss tube

vibrations shortly.



PBOCESS EVALUATION

OF

SHELL AND

TUBE EXCHAI{GERS

We are concerned here only

with

any

particular

heat

exchanger and determining whether

it can transfer heat

energy

as

required.

How

the

unit

affects process

condi-

tions

of

the entire system is not

our concern

here,

be-

cause we are interested only in the

proper performance

of

the unit. Evaluating

the exchanger in relation

to

the

process

system

is

the

primary

concern

of

the chemical

engineer. The thermal evaluation of

the exchanger

is

one

area where chemical

and

mechanical

engineering over-

lap;

just

as in Chapters 2 and 4

we

saw

how

civil and


of

Shell-and-Tube Heat

Exchansers

t15

mechanical

engineering coincide. Thus,

the

mechanical

engineer must be

cognizant

of

process

evaluation of heat

exchangers in order to

design these units.

A thermal

evaluation of shell and tube heat

exchansers

concerns

primarily

two modes of heat

transfer-conJuc-

tion

and

convection.

In Chapter 3 we considered

heat

transfer

through

pip-

ing

and vessel components

as well

as

jacketed

systems.

As described in Chapter 3,

the

basic expressions

used

in

conveetion are as follows:

q

:

rhcpat

q

:

UA(LMTD)

(3-24)

(3-26)

{1t

Some

ecceptable weld

geometriea

t2l

where t

is

not

less

lhan l.4t

(61

l7l

(81

Figure 7-13. Joint

types

[7].

(Courtesy

of ASME.)



116

Mechanical

Design

of Process

Systems

J

;

sooo

l

F

=

-

7t)00

U

ul

.o

*

6000

.o5

.o5

st .oa .o9 Jo

11 12 13 .1+

.15

16 t7

Equation 3-9 is a variant

of Fourier's heat law of con-

duction

in

which,

q:

KAAI

(7-6)

The treatment of

shell and tube exchangers requires

the same basic

theory

for

use

in

Chapter

3, but

a

differ-

ent application. In these

types

of

exchangers

we

are

pri-

marily

concerned

with

the heat

duty

or heat load re-

quired

in the same

general

sense as the

jacketed

vessels

TUBE WALL

THICKNESS Iin|

Figure 7-14.

Tube

joint

loads.

q

=

r;cp(ao

q

:

rimrg

in

Chapter

3.

Process requirements

are the

criteria

used

to determine the

heat

duty. The two basic components of

heat

transfer in

the shell and tube exchanger are sensible

heat

and

latent

heat. These

concepts are described

math-

ematically with the use of Equation 3-24. Using this rela-

tion

we

have:

(7-7)

(7-8)





118

Mechanical

Design

of

Process

Systems

r.0

P

.TEMPERATURE

EFFICIENCY

5

o.s

F

2

(l

o -.'

:

oa

o

O.9

F

z

9^"

o

0.7

=

o.6

/tL--.....-.-,

lr-t'

l.-+<_

-l/

LMTD

CORRECTION

FACTOR

I

SHEIL

PASS

EVEN

NUMBER

OF TUBE

PASSES

D

-.:l-J

'

T,-t,

Gl=

r

'2

03

0.5

0.6

P

.

T€MPERATURE

EFFICIENCY

LMTO

CORRE

2

SHETL PASSES

4

OR MUTTIPLE

OF

4 TUBE

PASSES

P'++

I:I

Q-tr

Figure

7'16.

LMTD

correction

factor.

(@1978

Ttrbular

Exchanger

Manufacturers

Association.)





'120

Mechanical

Design

of

Process

Systems

t.o

E

P

o.g

z

tr

o.8

tr

o.7

:

5

SHELL

PASSES

10

OR

MORE

EVEN

NUMBER

OF

TUBE

PASSES

.t

-+

r'#

"=

Tr-Tr

o.3 0.4 0.5 0.6

P

=

TEMPERATURE

EFFIoIENcY

LMTD

CORRECTION

FACTOR

6

SHELL PASSES

T2

OR

MORE

EVEN

NUMBER

OF

IUBE

PASSES

9:-]3-J

'

T,-t'

R

=

-l--3

Figure

7-16.

Continued.



I

o.g

z

P

o.t

o

o.7

F

=

o.6

The Mechanical Desien of Shell-and-Tube

Heat Exchangers

121

P.IEMPERA

LMTD

CORRECTION

FACTOR

SPLIT

FLOW SHELL

2

TUBE

PASSES

e'f{

''r-rE

P

=TEMPERATURE

EFFICIENCY

I

DIVIDED

FLOW

SHELL

PASS EVEN

NUMEER OF TUBE

PASSES

o.

-13--:

'

T,-t,

I-I,

Figure 7.16.

Continued.





The Mechanical

Desien of Shell-and-Tirbe Heat Exchansers

123

E

g

F

o\

o

o

t\

t-

o

tt

tlll

.9

.f

J

s'

ut

e

3

t-

4

ul

4

=

,l

F

e

3

|'|-

ul

()

e

ul

110

J0

rr|ivr9

'l 'd'v

;l;

lJ

:



124

Mechanical

Design

of Process

Systems

kn

:

thermal

conductivity

of

foreign

deposits

on

inside

of

tube,

Btu/hr_ftr_.F

T"

=

tube

wall

thickness,

ft

k*

:

thermal

conductivity

of

tube wall,

Btu/hr-ft2-"F

ho

:

outside

tube

film

coefficient,

Btu/hr-ftr-.F

Tro

:

thickness

of

outside

tube

deDosits.

ft

k,o

=

rhermal

conductivity

of

deposits

on outside

of

tube, Btu/hr-ft2_oF

The

terms

in Equation

7-17

,

llh, T/kf,

and

T*/k*,

are

known

as film

resistance,

fouling

resistance (we

will

re-

fer

to

this as

fouling

factors),

and tube wall

resistance,

respectively.

These parameters

represent

the

resistance

to heat

flow

through

the

fluid

film,

foreign

deposits,

and

the

tube

wall.

This

is

shown

in

Fisure

7-18 where

the

temperature

is shown

varying

throGh

the

various

resis-

tance

zones.

This figure

is

a

conceptualization

of

the

temperature profile,

as

the degree

of

gradient

change

in

temperature

is a

function

of

the

flow

conditions

daminar

versus

turbulent)

and

on

the type

and amount

of foreign

deposits.

To understand

Equation

7-17

we will

discuis

each resistance

separately.

Fouling

of Inside

and

Outside

Tube

Surfaces

Fouling

occurs

when

deposits

are made

on

the

walls by

particles

contained

in

the

fluid

medium

or

bv the

fluid

itself forming

a layer

on the

tube walls.

This

can occur

two ways,

either

by

adhesive

characteristics

of

the

de-

posited matter or by

the

foreign material

being bonded

to

the tube

surface

by

thermal

gradients

between

the

tube

wall

and the

foreign

material,

so that

the

latter

chanses

phases

when it

contacts

tube

surface,

resultinq

in

a coat-

ing effect.

Thus,

the depositing

of foreign

miterial

adds

to the

resistance

of heat

flow

from the

tube

and

she

side

flows.

Fouling

can

occur

inside and outside

of

tube

sur-

faces.

The complexity

of fouling

and

how it

occurs

does

not

easily allow

this

phenomenon

to be treated

analyti-

cally.

There are

far too

many

variables

involved

for

one

to accurately

compute

fouling

factors.

Thus,

this

phe-

nomenon

is treated

in

a more subjective

light, using

ex-

perience

as

a

guideline.

Years

of

experience

with

various

services

have

resulted

in the

use

of accurate

foulins

fac-

tors.

Fouling

factors

are very

important

in

the design

of

shell

and tube heat

exchangers.

Bare

or

plain

tubes,

which

are almost

always

used, generate

low

U-values

when

compared

to

those

generated

by

tubes with

fin at-

tachments.

Finned tubes, especially

those

with

fairly

high

fins,

experience

very

little

fouling

unless

the depos-

Its cover

an appreciable portion

of the

fin

height.

With

the normally

accepted

long

periods

between

tube

clean-

ing in

plants,

fouling

certainly

must

be

considered

in the

calculation

of

the

U-value.

One must

be aware

of

the

shell- and

tube-side

fluids

and select

those

foulins

fac-

tors thar

best reflecr

the

op{imum

fouling

thar

williffect

thermal

duty.

The fouling

factor

in Equation

7-17

is

T/fu.

This

term

is the inverse

of the

thermal

conductance

of heat

throush

the foreign

matter.

denoted

by k,/T,.

Thus,

the

reciproial

of

the

thermal

conductiviry

of the foreign

material

is

known

as

the

fouling factor. Fouling

can exist on

both

or

one

side of the tube.

Typical

values

for

fouling

factors

for

common

services

are

siven

in

Table

7-8.

At1

=

Temperature

drop

through inside

turbulent

boundary

rayer

Atz

=

Te6p"tu,ur"

Orop through

laminar

boundary tayer inside

tube

Ats

=

Tsrnpsr.lrra

drop through

fouling

layer

inside tube

At4

=

Temperaiure

drop

through

tube wall

Ats

=

Tsrnpg,.1r,a

drop through

outside

touling

layer

At6

=

Temperature

drop through

outside

laminar

boundary

rayer

Atz

=

T66p"r"rrr"

drop

through

outside

turbulent

boundary

taver

Direction

+

-----T

Att

Att

Atr

At.

At"

At,

Figure

7-18.

Temperature

profile

through tube wall.





126

Mechanical

Design

of

process

Systems

where

n

=

0.4 for

heating

n

:

0.3 for

cooling

And

the temperature

differences

are

as

follows:

At

:

pipe

surface

temp-bulk

fluid

temp

At

<

lO'F for

liquids

At

<

100"F

for

gases

Outside Tube

Film

Coefticients,

Forced

convection

around

immersed

bodies

is

a complex

subject,

especially

when

a bundle

oftubes

is involved.

We

will

only

give

L

rather

brief

discussion

of

how

one

can

obtain

a

s;neral

order

of magnitude

of

film

coefficients.

The

-reader

should

be

aware

that

process

design

is

not

addressed.

Thus,

for

solving problems

dealing

with

condensation,

nucleate

boiling,

and

film

boiling-to

name

a few_the

reader should consult other

sources

that treat

Drocess de-

sign

in detai[

[4.81.

For

gases

flowing

normal

to

circular

cylinders

a

sim-

ple

relationship

is

contrived

by

M.

Jakob

[1]

using

an dy-

ercge Nusselt

number

for

the

gas.

An

empirical version

of

this expression

is

given

by

Nr,

-

hd'

:

C(PJ'/r(NRJn

Forced

convection normal

to

tube

bundles

is

mucl:

more

complex

than

that

of

a

single

tube.

The

size

of the

bundle

and how

the

tubes

are

oriented

(tube pitch

ar_

rangements)

in

the bundle

are

of

prime

importance.

First.

we

will

discuss

an

approach

io

determining

the

film

coefficients

for

bundles

and

then

discuss

the

mr-erits

of

arranging

tubes

in

various

geometries.

There

are four

basic

types

oT

tube

arrangements-tri-

angular

pitch,

inJine

triangular

pitch,

inJine

square

pitch-,

and diamond-square

pitch.

These

four geomelries

are

shown in

Figure

7-19.

Tubes

arranged

in bundles

are

more

complex

than a

single

tube

becaule

the

flow vorti-

ces formed

by

the

flow

around

the

first

tubes

affect

the

flow

around

the

tubes

farther

inside the

bundle.

Mose

researchers

agree

that

this transient

effect

is substantially

dampened

after

the flow passes

over

the first

ten

tubei.

Numerous

research

studies

have

been

made

that

ana-

lyzed

flow effects

on tube

bundles.

E.

D.

Grimson

[12]

concluded

from

several

studies

that

for

tube bundlei

ai

least

l0

tubes

in depth

the

following

expression

can

be

used

to

predict

the

film

coefficient:

Kf

hd,

=

C(NIJ"

('7

-2r)

hd,

:

B(pvd"irr.r)"

(7

-23)

where

h

:

average

film

coefficient

for

gas,

Btu/hr-ft2-.F

dt

=

tube diameter,

ft

ks

:

gas

coefficient

of

thermal conductivity,

Btu/hr-ft-.F

C and n

:

parameters

from

Thble

7-9

A

variant

of

Equation

7-21

is widely

used for

forced

convection

ofair

normal

to a

cylinder

is

given

by

the

fol-

rowrng:

Table

7-9

Parameters

for

Fluid

Flow

Normal

to

Circular

Cylinders

Range

ol

Reynolds

Numbers

The Reynolds

number

in Equation

7-23 is

evaluated

at

the

maximum

fluid velocity.

This velocity

is obtained

at

the

minimum

flow

passage

between

the

tubes.

This

min-

imum

distance

is shown

in

Figure

7-19. Tbe

minimum

distance

is expressed

in

terms

of the

tube bundle

geome-

try

for

each of the

four

configurations.

as

follois

isee

Figure

7-19):

pf

:

absolute

viscosity,

lb-/ft-hr

V

:

velocity

of

air,

ft/hr

Nx"

:

Reynolds

number

at maximum

fluid

_

velocity,

V.",

h

:

average

film

coefficient,

Btu/hr-ftr-.F

p

:

air density,

lb./ft3

ki

:

thermal

conductivity

of fluid,

Btu/br-ft-'F

B

and

n

:

constants given

in

Table

?-10

do

=

tube

outside

diameter

Triangular

pitch,

d-,"

:

*

-

.

2''

InJine

triangular pitch,

dni"

=

W

-

d,

InJine

square

pitch,

dmi.

=

W

Diamond

square

pitch,

d.;"

:

P

cos 45'

-

D

:

0.707p

-

D

k1

where

(4,

0.40<

NR"

<4.0

4

<

NR"

<40

40

<

NR"

<

4000

4000<NRe<40,000

40,000

<NR"

<400,000

0.989

0.91

I

0.683

0.193

0.027

0.330

0.385

0.466

0.618

0.805

(b)

(c.l

(d)





128

Mechanical

Design

of

Process

Systems

The

cross-flow

are

for

various

types

oftube

bundles

is

shown

in Figure

7-20.

.

From

the concept

of

continuity,

where

for

two

points

along

a

flow

path,

or streamline,

VrAl

:

V2A2

e-24)

where V1

:

velocity

of

fluid at point

I,

ftlsec

Al

:

cross-sectional

area, ftz

we

can deduce

With

all

tubes being placed

at

a

constant

pitch

and

Vr

:

Vr

:

fluid

velocity,

we have

v.,,

=

v'l+l

\o.,"i

'A

</

\@

or staggered and iniine

tube arrays,

o,

=

* [o"

-

p,"

*

9 :

o'10

-

o,yl

,rc

]44

[ " Pn

-

"l

For triangular layouts,

.

B

[^

-

o,^-dr.

..1

.^

n.

=

r++

[D"

-

D,"

+

+---i

(P

-

dJl

,rt'?

where,

DL

=

OD of tube

bundle

D"

=

lD

ol

shell

dr

=

OD of

tube

B

=

baffte spacing

Ar

=

flow

area-cross-llow

area

for

one s€ction

tween

two baffles

Figwe

7-20. Tube

bundle

cross-flow

area.

Equation

7-25

represents

the fluid

velocity

that would

be

used in Equation

7-23.

For

tube

bundles

containing

less than

l0

tubes,

values

of the

film

coefficient

in Equation

7-23 must

be

multi-

plied

by the

correction

factors

in

Table

7-1

1.

Each tube

pitch

arrangement

has

its

own advantages

and

disadvantages.

A

listing of

these facts

is

given

in

la-

ble 7-12.

Whatever

the tube

arrangement

selected,

the

tube

arrangement

in

the

tubesheet

should

be made

verv

carefully.

Clearances,

which

could

be

such items

as

im-

pingement

baffles,

channel

and

head

baffle

lanes, must

be

considered.

Table

7

-13

is a compilation

of

various

in-

dustrial standards

for

tube sheet

layouts.

Fipure

7-21

shows

a typical

tube

sheet layout.

One

of the

easiest

and most

common

methods

used

to

calculate

shell-side

film

coefficients

is

that

proposed

by

Kern

[9].

The

Kern

correlation,

which

is used for

all

flu-

ids. is

as follows:

h"&

-

o

ro lq"o)"'l,9url'

'lu

l'

'

k

\p/

\t/ \pJ

or

h"rD":

o.:orN""f

t,*rr"t

(")o''

Equation

7-26

is divided

into

two

components,

jH

and

Np" in which

(7-2s)

(7-26)

Figure

7-21.

Typical

tubesheet

layouts.



:iw

The Mechanical

Design of

Shell-and-Tube

Heat Exchangers

j":+H'(,+)

129

('7

-27)

Table

7-1 1

Kays and London

Constants

for

Tube

Bundles

Containing 9 Tubes or Fewer

Number

of Tubes

123456789

In-line

0.64 0.80 0.87 0.90 0.92 0.94 0.96 0.98 0.99

Staggered

0.68 0.75

0.83

0.89

O.92

0.95

0.97 0.98

0.99

where

h.

:

outside tube bundle

film

coefficient,

Btu/hr-ft

"F

G,

:

mass

flow

rate

of

fluid, Iby/hr

k

:

thermal

conductivity of shell-side

fluid,

Btu/hr-

ft-"F

D.

=

shell-side

equivalent tube diameter,

in.

C,

:

sPecific

heat of

fluid,

Btunb-"F

Table

7-'12

Pros and Cons

of

Various Tube

Arangements

Tube Pitch

Arrangement

Advantage

Disadvantage

For a square

pitch

tube arrangement,

l(p:

-

nd;

)

l

i-

?iorn

For

a

60'equilateral

triangular arrangement,

.1(0.-13o:

-

0.5rdi

ilt

D.

:

-_:

in

(7-28)

Yields higher

film Medium

to

h

igh

coefficients than

pressure

drop.

in-line square Cannot

be used in

pitch.

More tubes foulrng

serrice..

can be contained in

Can

only

have

shell

becau.e of

chemical

cleantng.

compact arrange-

ment.

Film

coefficients

are

not as high as

triangular

pitch,

but

greater

than

in-

line square

pitch.

Suitable for fouling

conditions.

Good

for

condi-

tions

requiring low

pressure

drop.

Ar-

rangement allows

for easy

access

of

tubes for mechani-

cal cleaning. Good

for fouling

service.

Better

film

coeffi-

cients

than inline

square

pitch,

but

not

as good

as

tri-

angular or in-line

pitch.

Easy access

for

mechanical

cleaning. Good

for

fouling service.

(.7

29)

uhere

p

:

tube

pitch.

ir.

d,.

=

ID

of shell.

in

a,

:

flow area of

tube bundle, ft:

g

rcB'

ft:

17-30

r

p(144)

D,

=

ID of

shell,

in

c:

clearance

bgtween tubes

nleasured along

tube

pitch, in.

B

:

baffle

spacing,

in.

G,

:

mass flow

rate

of fluid, lb,/hr

G.:th

as

p

:

viscosity

of

the shell-side fluid

at the

ca-

loric temperature,

lb/ft hr

p*

=

viscosity of the shell-side fluid at the

tube

wall

temperature, lb/ft-hr

The

parameter

js

is

plotted

against Nx"

in

Figure

l-22a.

The value ofjH

is determined from

the

figure af-

ter the

Reynolds

number is calculated. Then from

Equa-

tion

7

-27 the

film

coefficient

is

determined.

The

use

of baffles

is extremely important

in

directing

the

shell-side

flow,

tube support, and controlling

the

shell-side

flow rate.

As

the

number

of

baffles

is

in-

creased,

the

flow rate increases. Likewise

with

an

in-

creased

flow

rate, the

pressure

drop increases

substan-

tially

with an increasing

number

ofbaffles,

with the

film

coefficient

increasing

as

well.

Ludwig

[4]

reports that

for

a

constant flow

rate,

the

velocity across the bundle

is

doubled

with an

increase

in

the film coefficient

of

ap-

proxirnately 44% .

(text

conttuued

on

page

139)

a

(a)

Triangular

Pitch

(b)

In-line

Triangular

Pitch

(c)

InJine

Square

Pitch

Medium

to

h igh

pressure

drop. Can

only

have

chemical

cleaning.

Relative

low film

coefficients.

Relative

low

film

coefficients.

Does

not have as

low-

pressure

drop

as

the

inline square

prtcn

arrangement.

(d)

Diamond

Square Pitch



130

Mechanical

Design

of Process

Systems

Table

7-13

Tube

Count for

g/a

in.

OD

Tr.rbes

on

13^6-in.

A

pitch

TEMA

TEMA

TEMA

TEMA

LorM

P

Type

s

Inside

ixed

Tubesheet

Outside

Packed

Head

U

U-Tube

ead

No.

ol

Passes

No. of Passes

No, ot

Passes

No.

of Passes

hell

lD in.

5.047

6.065

7

.981

10.02

12.N

13.25

15.25

17

.25

19.25

21.25

23.25

25.00

27

.00

29.00

31.00

33.00

35.00

22

68

0

170

212

283

3&

454

562

668

922

t9

JI

6l

104

151

178

24r

316

396

490

588

812

70

t6

30

28

66

60

106

96

164

148

196

r88

270

252

348

332

440

420

554

524

646

612

18

t2

26

24

52

48

98

84

142

t28

168

156

232

220

798

292

388

3s2

484

456

570

s48

t9

14

t2

31

26

16

56

52

44

96

90

76

lsl

138

t28

187

184

160

258 242

224

336 326

304

421

412

392

s26

502

480

608 s98

556

868

836

804

t152

lt24

t088

1496

1468

1424

902

868

808 764

1230

l2t2

lt72

1590

1560

1516

l106

1092

1040

1438

1430

1336

Tube

Count for

s/s

in.

OD Tubes

on

Z8

in,

A

pitch

TEMA

TEMA

TEMA

TEMA

LorM

Fixed

Tubesheet

P

Outside Packed

Tvoe

S

Inside

U

Head

Head

U-Tube

No.

of Passes

No.

of Passes

No,

ot

Passes

No.

of Passes

hell

lD in.

5.047

6.065

7

.981

l0.02

12.N

13.25

15.25

t7

.25

19.25

21.25

23.25

25.W

27.W

29.W

31.00

33.00

3s.00

22

31

61

96

151

187

241

396

482

568

19

26

55

88

130

151

206

270

JJO

418

506

704

92

780

752

1062

1030

1008

13s6

t346

13c4

700 660

18

t6

30

24

s2

48

94

80

138

132

176

168

232

224

302

292

384

352

472

456

554

536

14

t2

26

16

48

44

82

76

124

t12

148

t32

196 184

266

252

334

312

4t6

396

492

472

14

t4

t2

22

20

16

51

48

40

85

76

72

130

120

112

163

152

144

216

2r4

196

288

282

264

358

350

340

450 436

416

526 506

484

724

720

696

994 978

948

1288

1252

1220

946 930

896

1234

1220

n80





132 Mechanical

Design

of

Process

Systems

Table

7-13

Continued

Tube

Count for

3/a-in.

OD Tubes

on 1-in. A

pitch

TEMA

TEMA

TEMA

TEMA

LorM

P

s

U

Fixed

Tubesheet

Outside

Packed

Inside

Head

No.

ot

Passes

Head

U-Tube

No.

ol

Passes

No.

ol

Passes

No.

of

Passes

hell

lD in.

5.047

6.065

7

.981

10.02

12.00

13.25

15.25

17

.25

t9.25

21.25

23.25

25.00

27

.00

29.N

31.00

33.00

35.00

37

.00

39.00

42.00

45.00

48.00

51.00

54.00

60.00

14

14

t2

22

20

16

42

40

36

7372&

109

86

80

139

134

124

187 180

168

241 232

220

296 290

280

372 354

344

434

420

404

507 489

476

604 s94

s68

689 679

660

808 804

772

906 891

860

1030

1026

1000

1152 1134

1090

1273

1259

1222

1485

1461

\434

r72r

1693

1650

1968

l94l

1902

2221

2187

2134

2502

2465

2414

3099 3069

3010

10

108

19

18 16

1652 1620 1586

1894

1861 1820

2142

2101.

2060

2417

2379

2326

29W

29s7

2906

10 104

19

14

l2

2

40

64

98

122

t&

212

270

330

404

482

582

672

64

108

38 36

32 28

37

32 28

26

28

24

&62

6058

61 60

48 46

56

44

95 94 84

78

96 94

80 78

86

72

12L

ll0 100

98 r21

ll8 104

98 106

96

151 146

140 138

163 1&

144

140

148

136

208 196 188 160 216 214 196

158

200

184

258 242 232 230 276 270 260 235 254

240

320 316

296 298

338 338

324

3m

3r4

300

380 372

364 33s

396 396

376

339 388

368

475

466

452

430

460 440

420 4r4

452

432

530 526

508 49s

558 554

536 494

538

524

653 &2

620 610

624 605 s89

581

632

612

724

696 688

669 7s6

744

'116

669

732

708

859

848 818

805 818

797 783

771

838

808

946 922 9M

880 980

978 944

880

950

916

1106 1081

1054

996 tU1

1039

1001

996 1074 rO40

1218 1208

tr74

1r2s

rt72

1164

1130

1125

1200

1164

1426

1399

1376 1306 1367

1350 1322

13M

1406 1364

1635

1608 1536

1s04

1632 1s84

1887

tUz

1768

1740 1870

1832

2143

2lA4 2019

1992 2122

2076

2399

2366 2270

2244

2396

2340

2981

2940 2932

2800 2992

2936

Tube

Count for

g/q-in.

OD Tubes

on

f-in.

Pitch

TEMA

TEMA

TEMA

TEMA

LorM

P

Outside Packed

s

U

Fixed

Tubesheet

lnside

No.

ol

Passes

Head

No.

ol

Passes

Head

No.

ol

Passes

U-Tube

No.

of

Passes

hell

lD

in.

5.U7

6.065

7

.98r

10.02

12.00

13.25

).5.25

t7

.25

19.25

2r.25

23.25

25.00

27.N

29.N

31.00

2l

38

61

97

117

158

zlo

262

J10

370

442

524

602

698

l2

t2

16

16

38

32

60

52

90

88

I 16

112

158

148

208

188

256

244

316

308

372

368

432

428

524

500

596

580

692

688

12 t2

16

16

37 32

)/

)t)

89 82

9'7

94

t37

128

177

176

224

216

274

270

333

332

414

406

464

456

570

562

628

620

4

12

32

12

52

24

76

56

88

80

120

lt4

164 160

208

198

268

260

316

308

392

344

448

424

548

496

612

576

98

16

16

56

52

89 82

104

104

r45

140

188

184

238

236

304

292

344 332

398

386

484

472

554

532

650

@8

4

t2

32

t2

<t

'tA

80

56

96

80

140

tl4

180

160

232

198

284

260

332

308

366

344

468

424

510

496

640

576

64

88

24

20

44

40

68

68

90 88

128

120

176

168

112

108

138

134

340

332

400

388

472

460

554

544

640

624



The Mechanical

Design

of Shell-and-Tube Heat Exchangers

133

Table

7-13

Continued

Tube

Count

lor

s/+-in.

OD Tubes

on

1-in'

n

Pitch

TEMA

TEMA

TEMA

TEMA

LorM

P

Outside

Packed

s

lnside

U

Fixed

Tubesheei

No. of Passes

No. of

Passes

Head

No.

ol

Passes

U-Tube

No. of

Passes

Head

Shell

lD

in.

33.00

782 768

768

35.00

894 892

880

37.00

1004 978

964

39.00

I102

1096

1076

42.00

1283 1285

1270

45.00 1.484

1472

1456

48.00 l70l

1691 1610

51.00

1928 1904

1888

s4.00

2154

2138

2106

60.00

2683

2650

2636

742 732

732

668

'130

112

682

816 8r2

804

760

848 828

824

952 931

928

8'72

931

918

882

1062

1045 1026

972

1048

1028

996

1232 1222

1218

1140 1224

1200 1170

1424 1415

1386

1336 1421

1394

1350

1636

t634

1602

1536 1628

1598

1548

1845 1832

1818 1764

1862

1823 l7'/9

2080 2066

2044

1992 2096

20.+8

2010

2582 2566

2556 2476

2585 2552

2512

668

724

't20

160 836

8L2

8'72 940

924

9'72

1048

10/10

1140 1222 1204

1336 1420 1400

rs36 1624 1604

t'164 1852 1820

1992 2084 2064

2416

2596

2564

Tube Count

tor

3/4-in.

Oo

Tubes on

f

in.

'

Pitch

TEMA

TEMA

TEMA

Type

L

or M

Type

P

TYPe

S

TYPe U

Fixed

Outside

Packed

Inside

Tubesheet

Head

Head U-Tube

No.

of

Passes

No.

ol

Passes

No.

of

Passes

No.

ol

Passes

Shell

lD in.

5.047

6.065

7

.981

10.02

12.00

t3.25

t5.25

17 .25

19.25

21.25

23.25

25.00

27

.00

29.00

31.00

33.00

35.00

37.00

39.00

42.OO

45.00

48.00

5l .00

54.00

60.00

8

16

28

48

84

104

136

184

236

294

352

416

486

568

654

756

850

958

1066

1250

1440

1650

i868

2098

261.2

\2 l0

2r

18

37 32

61

54

97

90

113 108

156

146

208 196

256

244

314

299

379

363

448

432

522

504

603

583

688

667

788

7'70

897

873

1009

983

1118

1092

1298 1269

1500

1470

1714

1681

1939

1903

2173 2135

2692

2651

t2

10 8

16

t2 8

32

28

24

52

46

40

81 74

68

9't

92

84

140 134

128

188

178

168

241

228

216

300

286

272

359 343

328

42t 404

392

489 472

456

575 556

540

660 639

624

749

728

708

849

826

804

952

928

908

1068

1041 1016

1238

t2t6

tt96

1432 1407 1378

1644

t6Il

1580

1864 1837

1804

2098

2062

2026

26W

2560 2520

108

24

20

42 36

66

64

86

80

124 116

174

164

2t8

202

272

260

334

320

390

380

468

452

550

532

626

608

'720

700

818

796

928

904

1036 1016

1220

rr92

t4t2

1384

804

788

1834 1804

20'72

2036

2584 2544



134 Mechanical

Design of Process

Systems

Table 7-13

Continued

Tube

Count

for

l-in.

OD

Tubes

on 11/a-in.

A

Pitch

TEMA

TEMA

TEMA

TEMA

Type L

or

M

Type

P

Type

S

Type

U

Fixed

outside

Packed

Tubesheet

Floating

Head

Inside

Floating Head

U-Tube

No.

ol Passes

No.

of

Passes

No.

ot

Passes

No.

of

Passes

hell

lD

in.

5.M7

6.065

7 .98r

10.02

12.00

t3.25

15.25

17 .25

19.25

21.25

23.25

25.00

27 .U)

29.OO

31.00

33.00

35.00

37.00

39.00

42.00

45.00

48.00

51.00

54.00

60.00

864

14 148

26 26

16

42 40

36

64

61

56

85

'76

72

ll0 106 100

147 138

128

184 175

168

227

220

212

280

265

252

316 313

294

371 370

358

434

424

408

503 489

468

576 558

534

643

634

604

738 709

6U

8M 787

772

946

928

898

1087 1069

1042

1240

1230

rl98

t397 1389 1354

1592

1561

1530

1969 1945

t90/.

74400

l0 104

44

22

18 16 14

l8 t4 812

14

8

38 36 28 24

33 28 16 18

26

24

56 52 48 46

51 48 42 44

44

36

13 72 60 44

73 68 52 44

56

52

100

98 88 80

93 90 78 76 86

76

130 126 116 104 126 122 112 192 114

104

170 162 148 140 159 152 132 136 152

136

2r2 20r

188 176 202 r92 182

172 19?

176

258 2s0 232

220 249 238 21.6

2t2 232

220

296 294 276

250 29r 278 250

240 270

256

3ss 346

328 300 345 330 298 288

322

3U

416 408 392

360

400

388 356 348

378

3U

475 466 446 420 459

450 414 400 444

424

544 529 510 498 s26

514 484 4&

508

492

619 604 582 s66

596 584 548 536 s78

560

696 679 660 646 672 68 626

608 660

632

768

753 730 723 756 736

'7M

692 740

'1r2

908 891 860 840 890 878 834

808 872

836

1041 1017

990

968 1035

lm8 966

948 1010

980

1189 1182 1152

1132

1181

l162

lll8

tO92 1156

tt24

1348

133'1

1300 1280 1350

1327 1277 1254 1322 1284

l53i

1503

1462

r4r'iO

1520

r49Z

1436

1416

1496

1452

1906 1879 1842

1802 1884 1858 1800 1764 1866

1828

Tube

Count

tor

1-in.

OD

Tubes

on

11/4-in. v

Pitch

TEMA

TEMA

TEMA TEMA

Type

L or

M

Type

P Type S Type U

Fixed

Tubesheet

Outside

Packed

Inside

Floating

Head

Floating Head U-Tube

No.

ot

Passes No.

ot

Passes No. of Passes No. of Passes

Shell

lD in.

5.O47

6.065

7 .981

10.02

12.00

13.25

15.25

17 .25

19.25

21.25

23.25

25.00

27.W

29.00

31.00

86

12

l0

24

20

37

32

57

53

70

70

97

90

r29

r20

t62 r52

205 193

238

228

275

264

330

315

379

363

436

422

54

12

l0

2t

18

l) lt

52

46

61

58

89 82

113

1r2

148 138

180

r74

22r

210

261

248

308

296

359

345

418

40r

4

8

l6

28

48

&

84

1t2

142

184

220

256

300

360

410

4

8

16

28

40

56

76

104

128

168

2N

236

286

336

388

00

44108

24

i0

36

32

50

44

70

64

96 88

124

t20

156

152

200 188

232

220

282

268

330

320

382

368



{l

The

Mechanical Design

of


Table

7-13

Continued

Tube

Count

for

1-in.

OD

Tubes on

11/4-in. 0

Pitch

135

TEMA

TEMA

TEMA

Type

L

or

M

Type

P

Type

S

Type

U

Fixed

Tubesheet

Outside

Packed

Floating Head

lnside

Floating Head U-Tube

No, of Passes

No.

ol

Passes

No.

ol

Passes No. ol Passes

Shell

lD in.

33.00

35.00

37.00

39.00

42.00

45.00

48.00

51

.00

54.00

60.00

495

478

472

556

552

538

632

613

598

705

685

672

822

'799

786

946 922

912

1079 1061 1052

1220 t159

1176

1389 1359

1330

1714

l69t

t6&

477 460

448

540 526

508

608 588

568

674 654

&O

788 765

7s6

910

885

866

1037

1018 1000

1181 1160 1142

1337 1307

1292

1658 1626

1594

440

424

498

484

562

548

630

620

144

728

872

852

1002

980

1138 1116

1292

12@.

r@4

1576

Tube

Count

lor

1-in.

OD

Tubes

on

1tA-in. Pitch

TEMA TEMA

TEMA

Type L or

M

Type

P

Type S

TEMA

Type U

Fixed

Tubesheet

Outside

Packed

Floating Head

lnside

Floating Head

U-Tube

No.

of

Passes

No.

ol

Passes No.

ot

Passes

No. ot

Passes

Shell

lD in.

5.O4'7

6.065

7 .981

10.02

12.00

13.25

15.25

17 .25

19.25

21.25

23.25

25.00

27.N

29.N

31.00

33.00

35.00

37.00

39.00

42.00

45.00

48.00

51 .00

54.00

60.00

964

L2 l2

12

22

20 16

38

38

32

)o )b Jz

69 66

66

97 90 88

t29 \24

120

t64 158 148

202

l9l

184

234

234

222

272

267

264

328

317

310

378 370

370

434 428

428

496 484

484

554 553 s32

628 621

608

708 682

682

811 811

804

940 931

918

1076

106l

l0,l0

1218 1202

tt92

1370 1354

1350

1701

1699

1684

544-544-00

1264-126444

21 16 16 12 t7 12 812 12

8

32 32 32 18 30 30 16 18 24

20

52 52 44 24 52 48 42 24 38

36

61 60 52 50 61 56 52 50 52

48

8984806485786264.7268

l

13 112 rt2

96

108

108

104

96 98

96

148 144 140 114 144 136 130

tt4 t28

124

t'18 178 t'72 156 1',73 166 154

156 166 156

216 216 208 192 217 208 194 192

200

196

258 256 256 212 252 240 230 212

240

232

302 300 296 260 296 280

2'70

260

284

276

356 353 338 314 345 336 310 314 332

332

4r4 406 392 368 402 390 366 368 290

384

476 460 460 420 461 452 432 420 442

436

542 530 518 484 520 sr4 494 484 254

248

602 596 580 550 588 572 562 548

574

560

676 649 648 625 66r &0 624

620 W

628

782 780 768 730 776 7s6 738 724 758

748

9M 894 874 850 900 882 862 844 872

868

1034 rO27

101,2 980 1029

i0l6 984

9'72 1002

988

1178 1155 1150 1125 1170 1156

tt26

1114 1146 ll40

1322

1307

1284 1262

1310

1296

t268

1256 1300

1288

1654 1640

1632

1585

t64t

1624 1598

15'76

1620

1604



136 Mechanical

Design of Process Systems

Table

7-13

Continued

Tube Count

tor 11/4-in, OD Tubes

on 1sfi6-in.

A

Pitch

TEMA

Type

L

or

M

Type

P

TyPe

S

TYPe

U

Fixed

Tubesheel

Outside

Packed

Inside

Floatlng

Head

Floating

Head

U-Tube

No. of

Passes

No. of

Passes

No. ol Passes

No.

ol

Passes

shell

lD in.

5.047

6.065

7 .981

10.02

12.00

13.25

15.25

1',7

.25

19.25

21.25

23.25

25.00

27.W

29.W

31.00

33.00

35.00

37.00

39.00

42.00

45.00

48.00

51.00

54.00

60.00

744

864

t9

14

12

29

26

20

423834

52

48

44

69

68

60

92

84

78

121 1l0

104

147 138

128

r74 165 156

196 196

184

237 226

224

280 269 256

3t3

313

294

357

346

332

4t6

401

386

461

453

432

511

493

478

596 579

s70

687

673

662

790

782

758

896

871

860

1008

994

968

1243

1243

l2l0

000-

764-

14

148-

22

20 16

37

36 28

22

44

44 36

28

64

62

48

45

85 78 72

69

109 w2 96

86

130 130 116

1r2

163 152

r44 130

184 184

1"12

164

22r

216 208

196

262

252 242

228

302

302 280

270

345

332 318

305

392

383 3&

3s7

442

429

4r2 407

493

479

460

449

576

557 544

5r2

657

640 628

596

756

745

728

696

859 839

832

820

964

959 940

892

1199

1195

1170

1116

00

00

64

14

12

22

20

32

28

48

44

64

60

86

80

tt4

104

138

132

t62 152

196

184

232

220

268

256

310

296

356

3M

4M

388

452

440

534

522

626

6t2

720

700

822

800

930

908

1160

1140

Tube Count

for

11/4-in,

OD

Tubes on

11fi6-in'

Pitch

TEMA

fvoe U

TEMA

Tvpe

S

TEMA

l'vDe

P

TEMA

LorM

Fixed

Tubesheet

Outside

lnside

rtinq

Head

Packed

I

Head

No.

of

Passes

No. ol

Passes

No.

ol

Passes

No. of Passes

Shell

lD in.

000

000

064

12

128

18

20

20

24

28

28

48 42

36

50 56

56

80 74

68

96 98

96

114

124 120

136

140

136

0000000

6640664

t2 12 lZ 0r2 12

4

21

16

16 12

21

12

8

32

32 32

18

29

28

16

38

38

32

24 38

34

34

52

52

52

48 52

48

44

70

7o

68

s0

70 66

56

89

88

88

80

85 84

70

rr2 112

ll2 96

108

108

100

138 138

130

114

136

128

128

164 l@

156

136

154

rs4

142

444

664

12 12

12

24 22

16

37 34

32

45

42

42

61 60

52

80

76

76

97

95 88

t24

124

t20

t45

145

144

172 168

164

5.04',1

6.065

7 .981

10.02

12.00

t3.25

15.25

17 .25

19.25

21.25

23.25

25.00



The Mechanical

Design

of

Shell-and-Tube Heat

Exchangers

'137

Table 7-13

Continued

Tube

Count

lor

'l'tlq-in.

OD

Tubes on l/rs-in. Pitch

TEMA

TEMA

Type L

or M

Type

P

TEMA

TEMA

Type

S

Type

U

Fixed

Tubesheet

Outside

Packed

Floating Head

lnside

Floating

Head

U-Tube

No. of

Passes

No. of

Passes

No, of

Passes

No. of

Passes

Shell

lD

in.

27 .OO

29.00

31.00

33.00

35.00

37.00

39.00

42.00

45.00

48.00

51.00

54.00

60.00

210 202

202

24r 234

230

272

268

268

310 306

302

356

353

338

396 387

384

442 438

434

518 518

502

602 602

588

682 681

676

7'.70

760

756

862 860

8s6

1084 1070

1054

193 184

184 172 184 180

158 r12

176

176

224 224

216 198 2t7

212 204 198 200

196

258 256 256 236

252 248 234 236

232

232

296 296 282 264

289

2',76

270 264

272 268

336 332 332 304 329

316 310 304

312 296

378

3'70

370 358

312 368 354 340

348 348

428 426 414

408 420

.102

402 392

396 392

492 492 4U

464 485 116 468 464 472

456

570 566 556

544 565 55J 5+6 544

552

536

658 648 648

620 653 616 628 620

628 620

742

'729

722 7t2 738 126 ?20

',705

'7t2

708

838 823 810 804 837 820

811

80.+

808 804

1042 t034

1026 1008 1036

l0lE i0r2

1008

l0t2

992

Tube

Count

tor

1tlc-in.

OD

Tubes on

1el16-in.

,

Pitch

TEMA

Type L or

M

TEMA

Type

P

TEMA

Type

S

TEMA

Type U

Fixed

Tubesheet

Outside

Packed

Floating Head

lnside

Floating Head

U-Tube

No, of

Passes

No.

ot

Passes

No.

of

Passes

No.

ol

Passes

Shell

lD

in.

5.047

6.065

7

.981

10.02

12.00

13.25

t5.25

t7

.25

r9.25

2t .25

23.25

25.O0

27 .O0

29.00

31.00

33.00

35.00

37.00

39.00

42.O0

45.00

48.00

51.00

54.00

60.00

544664

13 108

24 20 16

37 32

28

45 40

40

60 56

56

'79

76

'16

97 94

94

t24 tt6

ll2

148 t42 t36

174

166

160

209 202

t92

238 232

232

275 264

264

314 307

300

359 345

334

401 387

380

442 427

424

522 506

500

603 583

572

682

669

660

1"t"t

'762

756

875

857

850

1088 1080

1058

< ,l ,tl

12 108

2t 18

16

32 28

28

3',7 34 32

52 52

48

70 70

64

90 90

84

tt2 108

104

140 138

128

162 162

156

191 188 184

442

,130

416

26t 249

244

300 286

280

34t 330

320

384 372

360

428 412

404

497 484

4'72

5',75

562

552

660 648

640

743

'728

716

843 822

812

1049 1029 t0t6

0000

44

12

t2

20

20

26

24

40

36

56

52

74

68

96 88

120

ttz

142

136

170

164

200

192

228 220

268

256

306

296

346 336

390

380

456

448

542

528

618 604

708

692

802

'784

l0l0

984



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'



Sallle Pilch

or Stoci.q

The Mechanical Design

of

Shell-and-Tube

Heat

Exchangers 139

Kern

[9],

where

the expression

for the shell-side

pres-

sure drop is

given

as follows:

iGlD.(NB

+ 1)

Polh

ot

Fluid

A.

Shell side fluid baffling

showing

segmental cut baffles.

Fluid Flo13

Poroll.lhTubrs

os,t

Po3r.3

/From

one

Bolll.d A..o

io N.rr.

;9

]

eorrrr

"wintoi'

or

"co'l

O

O

J

Eac$ed

03

%cu ,{hkh

is

(5.22X10)roD"7

r

where

(p/p,,).

G,,

D.,

D.

NB

^l

t.

(7

31)

:

are

previously

defined

:

number of bafiles

=

specific

gravity

of

shell-sidc

fluid

:

combined friction

factor deter

mined

from Figure

7

23

TUBE

VIBRATIONS

Chapters I and

-1

described

how

fluids moving

around

objects can produce

r

ibrations. The

same

thing

happens

in shell

and tube heat

exchangers,

but it creates

a

differ-

ent

problenr.

Chapters

I and

,1

were primarily

concerned

with Yorte\

sheddrng.

This

chapter

covers vortex

shed-

ding and sereral

olher t\pes of vibration

phenomena.

Also.

the

problen

is difterent from

rhat

in Chapter 4 be-

cause

the boundarr

conditions

of

the

system have

chansed.

Chapter

I

used

a cantile\er

beam

to

show

how

a til\\ er or srack is

restrained

several

different

ways

at

the ends."

There

have been nrany

research

studies made

in the

field

of

tube vibrations.

Probably

the most

numerous

stem

lrom

the

nuclear

industry.

The

problem

is

complex

and

no

one

method proposed is

a

full

and

complete anal-

ysis

of tube vibrations.

Consequently,

research

is still

being

done

to better

understand

the causes

and

preven-

tion of

tube vibrations. Here

we

will

outline the causes

of

the

phenomena

and

present

some quanlitative

ap-

proaches

to the

problems.

Presented

first is a simple

and

quick

approach to

pre,

dict

tube vibrations

caused

by

shell-side

flow.

This

ap-

proach

was originally

developed

by

John

T. Thorngren

[14]

in

1970

and is

called the

"maximum

velocity

method."

We will

present

a modified

version

of

the

method

proposed

by Thorngren

to

encompass

a wider

range

of

applications

and to specifically

define

all the

variables

in

the

equations.

This

method

addresses the

tube

vibration

caused by vortex

shedding

when

the

shell-

side fluid

alters direction

at the

baffle

plate

and strikes

the

tubes. The arrows

in Figure

7-l

show

how flow

di,

rection of a fluid

turns

at

the baffle

plate

and strikes the

tubes

midwal

between

rhe bafile

plates.

Thi5

causes rhe

tubes

to deflect and

the hole

in

the baffle

plale

acts

as

a

fulcrum

for

the tubes to

deflect against.

Two

types

of

problems

can

result

a

fatiguing

of

the tubes at the baf:

fle

hole

and eventual

tube rupture,

or

the tubes colliding

{%XSherl

10.).

Ner

Fror

Ateo

ol Wiido* is

Full

Windor A/.0 Diius

Ar.o

B. Segmental

baffles

showing window

are

for

fluid flow.

Iilol.:Ar0o Avo,l0bla

tor Cror Flor L3ed

Cort,r, rr r

.'

peler*ce

I

. oth.t

i:l:1if.'

4,i

*,

orhe'

a,,olq -e.rt

ro

ooh

.

E3se"',- ||e

so-.

C. Cross

flow

area

for iube

layouts.

Figurc

7-228.

Various

baffle

*indou

schemes

[,1].

Baffles

neveq

except for unusual

designs such as ori-

fice baffles,

extend a

full

360" around

the

shell.

The

baf-

fle

plate

is cut such

that the

shell-side fluid

can

flow

around

its edge.

The open

area between

the baffle

edge

and the

shell

wall

is

known

as

a

baffle "window." Baffle

windows

are commonly

referred

to in

terms of

percent-

ages of

the

entire

circular shell

area. Figure

7-22b illus-

trates

various

baffle window

schemes.

Shell-Side

Pressure

Drop.

There

are several methods

to calculate

the

frictional

pressure

drop across tube

bun-

dles,

and the reader

is referred

to

Ludwig

[4]

or Kern

[9]

who give

comprehensive

discussions

of

the

various

tech-

niques.

The method

we will

use is

the one developed

bv

Boftr.s

@0,'d@

sotlrs

Ooid

O

oootoo

ooolooo

ooooo

ooooo

ooooo

ooooo

)OOOooO

oo



^z,Ez.E

-

. .

l-€-e;;

=l

-

--

g

3: ;

rtx :-

-al

--

?

.3

:E

3

3l;

F

"-

-i

xla

E I

tree

li:E

i;

fr

q

xl- o :i

-.j

-.

X

'rli

3;

--:'i

5

=

lu

5c;

=F=

;

e

"el

,r, <

.

---

--=:

-=

=

6

'r

-c6;

"'59

-tr

E

î'

=-Ft*

{SS

s5-

}

tl+

c*

E

E

:lr=-E

ES.=

E

ld

P;

=

E

i;t--E

E\.E

E

zl

x

Sooo 9=

eElDF;

-o-

xld.FsE:;.E-E€

g::

-E+

;l:

@

;9-e

A;o.9

g

ol

-

-:;-

-

:1:;5# =sji5Fi;

^6+

-

oo

-oo

c(r(-)oE 6r9

0'--jzz, o,<

j:

P

r:

=

-

ri=\=

=t

i

d

>

;E

"l

,:

99

o

::

e: E=.E

:5 EEP

;€

:d*+

rj6tr

o._

2.f

--es

ii;

;==*

3=

i;

--;;;

{E-

iE

s

r

'EE

Ess_ii3;S

;

t

-:

a*:L:L+:

6

<tQ

oooF

lo

|.l rt

n)

<t

For low-Finned

Tubes

f"

'

(sq.

ft)/(sq.

in.)

oooF

iol) a ||'

-ooFtoI'

t ro

N

-';

;

@-1.

E5

x:

.. A

.9:

fri

{.)

N

":

tl.E

oti

ii,

F

o.

-8

I

go





'142

Mechanical Design

of Process

Systems

F.

:

aFrL

^

bFI L4

EI

As the

shell

fluid

exerts

pressure

on the

tube,

the tube

deflects at mid-span forcing

the tube at the baffle

against

the baffle hole. The

stresses

induced in

the tube are a re-

suit of localized forces

at the

tube-baffle contact

points.

At these

points

the tube

behaves

similarly to

a

horizontal

vessel

such

that

only

a

portion of

the tube

wall

offers ef-

fective resistance

against collapse.

Thus, Equation 4-2

predicts

the amount

of

tube

wall

that effectively resists

the baffle wall reaction,

and

is written

as

(4-2)

(7

-32)

(7

-33)

The

values

for a and b

are dependent upon the bound-

ary

conditions of

the

continuous beam. Typical values

are

presented

in Figure

7-25

and

are

fairly

comprehen-

sive for most shell and

tube exchangers. For

cases

not

covered in Figure

7-25, the

specific values must be

solved for using the analysis

for

a continuous beam.

a:11{

+

:ol

180

\12 I

Continuous

Beams

dmar br"r

I

2

3

4

5

6

l .200

0.550

1.100

1.223

0.572

1.143

0.0059

0.0099

0.0069

0.0094

0.0097

0.0065

4

[.r.

(0,130

r fton A]

=

t.005r ,rrlsl

il.r.

(0.tt

, tioE

^

.'

o)

a

0,00n &,4/El

a

r.&

o.raa

I

rroi

A o.

D)

-

0,0rl r.

r

A rr.r.

(0,415

r koh E,

5

o.m a

r./al

€

comtruous

BEAM-FoUR

Eeual spaNs-LoAD

FtRsr aND THrRo spANs

a

tl.r.

(t

az

rlr.n A)

E

0.6tt {,r/al

6.

coNTlNuous

BEAM-FoUR EeuAL

spANs--{LL spaNs

LoAoEo

^

.L

(Gaa

I lr.h A

.na

a)

-

O.Ol5 d./s

Figure 7-25.

Boundary conditions

of

continuous

beams

u5l.



where

d

=

angle of contact

A

=

radians

Thorngren

[14]

proposed

that

about 40%

of

the tube

metal is effective

in resisting

wall membrane

stresses.

In

Equation

4-2 this would make

the value

of 0

:

144'

,

greatef

than

most

saddle-shell connections

for

horizontal

vessels

.

To take the

problem

furtter

we consider the

tube

wall

as

a

ring

shown in Figwe

716.

The assumption is

that neither

the tube

nor

baffle hole

will

deform

to

re-

duce

stfesses, which

is the worst

condition. For

deter-

mining

contact

stresses between

the

two

bodies,

Ti-

moshenko

[16]

has

shown

that for the case in

Figure 7-26

the

diameter

of

the circle

of contact

is

d

:

1.76E(qi_e")g4"

l"'

[

2EEB(d, +

dJl

Q-34)

From Equation

7-34 one

can

deduce

that

the tub€-baf-

fle interface

should

be

alLalyzed

as

point

loadings.

For

such loadings

as

shown in Figure

7-26

the contact force

representing

the

shear

of the tube

against

the

baffle

plate

ts

*.

(#,.

J"L,,*',J'

Q-3s)

Evaluating

the

relationship

for shear

in Equation

7-32

we

have

F.

=

aaFrl-

The

Mechanical Design

of

Shell-and-Ttbe Heat

Exchangers

149

where

q

:

constant that

represents

the

amount of effective re-

sisting tube wall

area

Now combining Equations

7

-32

and

7-35 we

have the

followins:

"

:

I--Lil--qe-)||--t')'

\aF,L/

\4

+ dB

/ \0.798/

r-v

l-u$

where

c

=

(7-36)

(7-37)

z,

=

hisson

ratio

for

the tube

material,

dimension-

less

/B

=

hisson ratio

for the

baffle

plate

material,

di-

mensionless

E =

modulus of

elasticity for

tube

material,

psi

Ea

:

nodrlus of

elasticity

for

baffle

plate

material,

pst

c

=

constant, in./lbr

cr

:

constatrt,

dimensionless

To

arrive at the modified

damage

numtrr

for

baffle

damage we solve

for

F1 in Equation

7-32:

Now dividing

this relation

into

Equation

4-80 we obtain

Cpd,p\Palc

_

2g"F.

Letting the baffle

damage

number

be

represented by

Nss,

a dimensionless

parameter,

we have:

EB

F,:

F"

-

o'al-

1.0

*,

-

Cpd.pV2alcv

^""

- --fdE-

Q-38)

where

Nss

(

1.0

If

NBE

>

1.0,

then

tube

damage

at

the

baffle

is

very

probable

and a

tlicker

tube

should be

selected

and the

analysis

repeated.

The analysis

of determining

the dimensionless

param-

eter,

NsD,

which

governs

tube

damage

induced

by exces-

sive displacements

in

tube movements,

is

similar to that

for

the

baffle

damage

parameter.

Solving

for

F1

in Equa-

tion

7-33 we

determine F1

as

follows:

Figule

7-26. Fluid

foroe

causing

tube

to impinge

on baffle

F"

:

E

plate.

"

bL4



144


Dividing this expression

into Equation 4-80

we have

Ded,pV'?bLa

_

1.0

2g.6E,I

We define N6p as

rrcD

-

CDd,pv2bL4

(7

-39)

where

NcD

<

1.0

Once again,

if

NcD

>

1.0,

then thicker

tubes should

be selected

and the analysis repeated.

Equation

7-39 is similar to that obtained by Thorngren

[14]

and

Coit

[17].

The

dimensionless

parameters,

Nss

and

N6o,

in

Equations 7-38

and

7-39 should be regarded

as mere

rules of thumb. Even though they are dimen-

sionless,

they do

not

have the same firm

basis as do di-

mensionless

parameters

used in

fluid

mechanics and

transport

phenomena.

One can

approximate the tube behavior

by using the

principles

in

Chapter

2,

Example

2-6.

Using the

baffles

as

supports

and spacing them

(either

equally

or

un-

equally),

one can simulate the

tube displacements.

How-

ever, since

we are

not dealing with

a single

tube, vortex

shedding

around tube bundles

can

presently

only

be ac-

counted

for in

design

by

being conservative.

Flow-induced

vibration of exchanger tubes

is another

mode

different

and

distinct

from

vortex

shedding.

In

vortex shedding

a component of the flow,

the

vortex,

is

the contributing

cause to the tube

vibration. In

flow-in-

duced

vibration, forces are exerted on

the

tubes that

are

caused

by flow

field

interactions

around the tubes.

Fluid

that

flows normal to the tubes

is forced into a smaller

area

between

the

tubes

resulting in a

Venturi

effect

known as

"jetting" or

"jet

switching."

This

phenomenon

is shown

in Figure 7-27

where a control volume

of

fluid

is shown

being compressed between

two

tubes. The

re-

sult

of this

'letting"

effect

is the fluid exiting

the narrow

area

between

the tubes diverges

into

a

diffused

mass

that

whips

or

whirls around remaining tubes.

This

"whirl-

ing"

effect is

another mode

of

vibration.

Vibration

induced

by

turbulence

is the most

common

mode. This

phenomenon

is commonly

confused

with

the

other

modes because the term

turbulence

is viewed syn-

onymously

with

fluid flow

and

vibration

resulting

from

such

flow.

However,

vortex shedding,

jetting,

and

whirling

are different

from turbulence

because even

though they

exist

in

turbulent

flow,

they

can all be

final

causes

of

failure and

each must be controlled.

Turbu-

lence can

be best

viewed

as

a

pressure

field

around a

tube

shown

in Figure 7-28. Herc

we

see

a

pressure

dis-

FigUJe

7-27.

Jet switching in tube

arrays.

2g.6E I

F-+

6=

futr)

p

=

p_(t) where

t=

iime

Figure

7-28.

The magnitude

of

the

direction

of

the fluid

strik-

ing

the

cylinder

can

be

thought

of mathematically

as

a

forcing

function,

F-,

mapping

a

pressure

distribution

around the

cyl-

inder

over

region

R.

(r\






of

Process

Systems

V

:

fluid

velocity

of

fluid

external

to

tubes, ft/sec

m

:

mass

density of fluid external to tubes, slug/ft

dr

:

tube OD,

ft

L

:

tube

length

between

baffles, ft

Lr

:

total

length of

tube

between tubesheets,

ft

fN

:

fundamental natural frequency

of

tube

portion

between

baffles,

Hz

I

:

sum

of

structural damping

and

the

fluid

dynamic

damping

x

:

distance along tube,

ft

d"

:

4Rs

:

4(hydraulic

radius)

:

4

(flow

area between tubes)

wetted perimetel

port

end conditions, and tubes that have equal spans

and

unequal spans.

These expressions were

presented

earlier

in

this chapter

and in Thble

7-6.

Equation

7-44

is

sim-

plest

to use because

it

requires less

input.

However,

when the

information

is available

and

time

permits,

the

expressions

recommended

by TEMA

should be used.

The

phenomena

of

"jetting"

and

"whirling"

are

not

as

well founded as

vortex

shedding and turbulence.

This

does

not

say

that vortex

shedding

and

turbulence are sol-

idly based,

but relatively speaking, they are compared to

the

other

vibration

modes, such

as

jetting

and

whirling.

From Figure 7-28 one can

predict

that when the tubes

are inclined to the fluid flow, the

results

are force com-

ponents

about

the

x

and

y

axes. Equation 4-80 illustrates

how one can

determine the force induced

per

unit length

of

a

circular cylinder. In the

case

of

whirling

and

jetting

the term CD

is a

variable.

This

term is

called the force

coefficient

and

is

used

in Equation 4-80 to evolve

the

fol-

lowing expressions:

-

;]

-

".

tubes on an equilateral

triangular pitch

of

P

_

+0,

[/r\

-;

t\-dJ

-

f]

-

ro.,"0", on a square

pitch

ot

P

-

pv':d,

-.

16,l

ru

:

--

N"

l=l

'

zE"

\o,/

/\

."

_

pv'0,

6" 16,l

^

2e,

"

\d,/

Using

Figure 7-29

the

value

of

thejoint

acceptance

for

the

appropriate mode and the first mode are obtained.

The

ratio of

the

joint

acceptance

of

the mode being con-

sidered to that

of the first

mode

is

multiplied

by

the

value

of

6.*, obtained from Equation 7 41. The relationship

in

Equation

7-41 is

based

on

the

theory of tube turbulence

developed

by Wambsganss and Chen

[9],

which

yields

the

followins

maximum

stress

value:

where K,

:

Kr:

2T-

D

/nV r

l:l for

:

)

\T/

D'

tn

T

l

5

<

1.5

(7 -4s)

(7

-46)

(P\'",

r

\T/

D

t2

o*":

E-Cp1*-y

(7-42)

where Ce

:

drag coefficient of tube

surfaces

6.*

:

2.586.-,

(for

x

:

L/2)

(7-43)

Equation

7-42 represents the maximum tube deflec-

tion

to be

incurred.

The factor 2.58 represents the

ampl

tude

of the highest one

percent

of cycles.

The

value

for

the

natural frequency

at

the tube in

Equation

7-41 takes on several forms. The easiest

to

use

is the

formulation

developed

by Blevins

[18]:

K,

=

C'(D/T)

-(,n)'.,(,n)'

"7(

N

-

;;;

6L-

F

i,l2

-L

,lZ r0

5

where E,

:

modulus of elasticity of tube metal,

psi

mr

:

mass

density

of

tube metal.

slugs/ftl

4

=

tube OD, in

dti

=

tube

ID,

in.

TEMA

gives

a listing

of expressions for

the

natural

frequencies of the

tubes based

on several types

of

sup-

where D and

T

are

parameters

defined

in Figure 7-30 and Fig-

ure 7-31.

Values

for K,

have been

plotted

against the

parameter

T/D.

These

values

are

shown

in

Figures 7-30 and 7-31

to

represent the

whirling

parameter

2(2?r)0

5/(C"Kr)0

'?5.

Ex-

periments

indicate that the

lower the whirling

parameter

the

greater

the

probability

that

whirling

(and

jetting)

will

occur.

To determine

if

the

tube

deflections are

within

a safe

range one

must estimate

the

components

F,

and F*

at

their

maximum

values

using

Equation

4-80.

From

the

tube spacing

determine the

force coefficients

K,

and C*

from Equation 7-46.

Then solve for 6,

and 6" and deter-

mine if those deflections

are acceptable.

After determin-

(7

-44)



ing

that the deflections

are

in

a safe

range,

use

Figures

7 -30 and7

-31

to determine the whirling

parameter.

If

the

parameter

is on

the

low

side,

then

the tube

spacing

should be increased

to raise the whirling

parameter.

Un-

fortunately,

at the current

state

of technology,

there are

no critical values to

decide whether the whirling

parame-

ter is critical.

One manner in

which

to avoid nroblems

with whirling

is to use Table 7-14

in

derermining

the

maximum

shell-side fluid velocity flow. This

table and

the

previous

discussion

will

eliminate

any

problems

with

jetting

or whirling.

If

the

velocities

cannot

be controlled,

because

of someone else's

design

or

a

client's

requests,

then this

procedure

can

give

one an idea

of

whether

whirling

can

be anticipated. The

main focus is

to

keep

the tubes spaced such

that the maximum velocity

will

be

reasonable.

It

has

been confirmed

bv exneriment that the

critical

velocity

for whirling

increises'rapidly wirh

the

minimum

spacing

between the

tubes

and

that

inline

tube

arrangements have lower

critical

velocities than

stag-

gered

tube

arrangements

(refer

to

Figure

7-19

for

the

various

illustrations of

arrangements).

PLATE.FIN HEAT EXCHANGERS

These units use have

been on the increase the

past

sev-

eral

years

because

of

an increasing number

of liquified

gas

and cryogenic

plants.

The

plate

fin

heat exchanger is

The Mechanical

Design

of Shell-and-Tube Heat

Exchangers 147

more

efficient

than the

shell and tube

exchanser because

the comparable shell and

tube exchanger

req-uired

to

re-

place

a

plate

fin would be eight

times the volume

and

twenty-four times the weight

of the

plate

fin

if

con-

structed

of

aluminum. The reason for

this is that

if

the

plate-fin

is made

of

brazed

aluminum,

the

aluminum

conducts heat better

than most materials

and can be used

down

to

absolute zero

(-460'F).

Since the

ductility

of

carbon steel is lost

at

-20"F,

one must

revert to

expen-

sive

nickel

alloys or stainless

steels in

the shell

and

tube

design. Thus, for cold

services, the

plate-fin

offers some

advantages.

It

is here that the advantages

of the

brazed

plate-fin

ex-

changer end. For the

plare-fin

to

be applied,

a very clean

service is required. Even

in clean

services, these units

can accommodate

certain

thermal shock

and

fatisue. It

is

quite possible

after continued

and repeated

therrial

load-

ing in

excess

of

differential

temperatures

of

50'F

that

in-

ternal components

can

fail. In

addition, because

these

units are

aluminum.

external nozzle

loadings

induced by

the

piping

can

cause

pipe

stress

problems.

One

must

be

extremel

careful ho\\'

much loading

is

induced

to

the

nozzles.

because

even

if

failures

do not

occur, leaks are

common if overloading

exisrs. Thus,

if

the service is not

clean.

a shell

and

tube

design must

be used.

In

gas processing

and cryogenic

services,

the

plate-fin

exchanger suffices because

in

these

applications the

ser-

1

Figure 7-30. Whirling parameter

of

a

tube

row

expressed as a function

of transverse spacing.

(From

Flow-lnduced

Vibration

by R. Blevins

@1977 by Van Nostrand

Reinhold

Company,

Inc.

Reprinted

by

permission.)

--loF

\JT

rl-L

o

Oo

./

,-7

./

-rl-

-

i,

.

-2

.

5,onr-3ro'2

'

(0,1,3

A

---_

xY

-lDt'3



M

o

"F

A

o

,r1

1'h

.-

>;

-/l

--r.

h

--

u+

I

o

o

o

148 Mechanical

Design

of Process

Systems

Figure 7-31.

Whirling

parameter

for

tube ar-

rays.

(From

Flow-lnduced

Vibration

by

R.

Blevins

Oi977 by

Van

Nostrand

Reinhold

Company, Inc.

Reprinted by

permission.)

With newly developed techniques in vacuum

brazing,

stronger

bonds have

been achieved that reduce failures

of

internal

components subjected

to

thermal shock and

fatigue.

The

aluminum flanges

used

on

these

units

are

de-

signed

per

ASME Section VIII

Division

I

and,

quite

commonly, are identical to ANSI 816.5 flanges.

For

further

discussion on the thermal analysis and de-

sign of

plate-fin

units, the reader is referred

to Kays and

London

[20].

EXAMPLE

7.1:

REGENERATED GAS

EXCHANGER

DESIGN

A

gas-gas

shell and tube heat exchanger is to be

de-

signed. The exchanger

is

to

be used

to

exchange heat

be-

tween a hydrocarbon

process gas

and a

gas

used for re-

generation.

The unit is

to

be designed

per

specification

sheet in Figure 7

-34.

The exchanger is shown in Figure

7-35.

The

process gas

is

to be cooled from 965'F to 705'F.

The regeneration

gas

is to

be

heated from 200"F to

661'F

in

a

parallel

configuration. Thus,

Table

7-14

Maximum Recommended

Shell-Side Velocities

All liquids in

10

fusec

Gases and

Vapors-in fl/sec

Pressure Molecular

Weight

(psi)

18

30 50 100

150

200 400

2'7

-tn.(vac) 250

185 160 110

100 90

77

15-in.(vac)

130

100 85 65

60 52

45

0 100

80 70 50 45

40

35

50 65

55 45 35

30 25

20

100

200

500

1000

55 45

35 25 20

18 16

50 40

30 23 19

t7

40

30

20

20

15

vices

are

relatively

clean. However,

it

must be noted that

shell and tube exchangers

are more

popular

because of

their

flexibility

ofuse.

Certainly

with

moderate to heavy

viscous fluids, the shell

and

tube

exchanger

is the only

design

to

use.

Figtre 7 -32 shows

a

plate-fin

exchanger with rectan-

gular

boxes

containing

an assortment

of

plates

and fins

resembling honeycomb structures. Fluids

flow

in

tubu-

lar channels

formed

by

fin attachments between

plates

(Figure

7-33).

The

plates

that

separate

the

two

services

vary from

approximately

0.006 in. to 0.023 in. in thick-

ness, depending

on

the

pressure

of

the service.

This

de-

sign is commercially available

at a temperature and

pres-

sure

of approximately

-

452"F at

1,400

psig.

975'F

200'F

750'F

625"F

GTTD:775"F tiITD

:

125'F



.M

Figure 7-32. The

plate-fin

exchanger.

(Courtesy

of Albraze

International, Inc.)

The

Mechanical Design

of


'7'75

-

125

:356"F

h

(E,l

u25/

Turning

Distributor

Fin

q

:

riCo(LMTD)

The shell-side

mass

flow

rate

:

22,050

lb,/hr

for

the shell-

side

gas,

Co

:

1. 10 Btu/lb.-'F.

The

required

heat

duty of the

unit

is

q

=

122.050r

l

rr. ror

j'l=

1:so.r"r

'

hr lb",-'F

Rfr

I

q

:

8.634.780

--

nt

The available tube area in the exchanger is

determined

as

follows: From Table 7-3, we determine

that

for a l1/+-

in.

tube

the

square

feet

of

external

surface

per

foot

of

tube

is 0.3272

ft:.

Thus.

Available area

=

(0.3171)

T

(ZS:),u0.,

(tr)

,,

'ft

=

1.38E.95 it:

ng

Sh€el

Bar

149

LMTD:

now,

Figure

7-33. Tubular channels in

plate

surfaces result in excellent

heat

transfer

in

plate-fin

heat

exchangers.

(Courtesy

ofAlbraze

International. Inc.)



150

Mechadcal

Design of Process

Systems

HEAT EXCHANGER

SPECIFICATION

SHEET

I

2

5

5

7

a

9

lo

ll

t2

l3

l5

l6

t7

t8

t9

20

2l

22

23

?1

27

2E

?9

30

3l

33

34

35

36

38

39

40

41

42

43

1t6

47

4E

19

50

52

53

57

5a

59

6l

T"b"-T,rb".h".t

J.i.t

Bundle

Entranc€

Bundtc Erir

Figure 7'34.

Heat

exchanger specification

sheet.

(O1978

Tubular

Exchanger

Manufacturers

Association.)



The

Mechanical

Design

of Shell-and-T[be

Heat Exchangers

For the

tube-side

gas, 1%-in.-11

gauge

tubes

Np"

:

0.7,

obtained

ftom

Process

data

k

:

0.03

Btu/hr-ftL'F

P

=

0.01 Cp

:

0.024

lb/ft-hr

Tirbe-side

mass flow

rate

=

41,884 lb./hr

For each

tube,

.

-

41:qq4

9./hr

:

148 rb-ihr

--

283 tubes

P

:

O.1524lbJft3

'

4

=

l'25

in"

di

:

1'010

in';

Ar

:

0.8012

in''?

:

48.48 ff/sec

151

sa-tua-600

osME)

r.gu;riil{

sa-ra8-6lrt

(^snE)

Flgure

7-35.

Vertical

gas-gas exchanger.

Shell-side

nozzles

C and

D are

16

in.

in diametel

which

makes the

flow

area

o.os

r

n.396) ft,

ftr

From Table

7-14 we observe

that this

is a

reasonable

velocity.

ftrbe.Slde

Film

Coellicient

For turbulent

flow

inside tubes

we use Equation

7-19,

the Sieder-Thte

correlation,

From

Table 7-14

this

velocity is

reasonable

_-

(48.4D

a

(1.oro)

in.

ffi

,o

tou

*

l.

=

a'(16)'z=

2ol.o6

in.2

:

t.396

ftz

Shell-side

mass

density

:

p.

:

0.09 lb./ft3

rr.

/ rr,.

\

22,050

+ l=.:;r-l

nr Ijbtt, secl

v

:

-------

j::--l:i:-

:

48.75 ff:/sec

:

ro lt

Btu

--

--

hr-ft2-"F

Shell-Side

Fllm Goefficlent

Nu"

=

0.027(NrJ03(Np.)18

(rJrJ''4

N.,"

:

?

=

o.:o

(Ps,

)"'rN*,',,

(;)"

Nr"

:

93,278

>

10,000

and Equation 7-19

applies

Nr"

:

0.027(93,278)0.8(0.7)t/3(1.0)

:

226.78

h..1.

Nr"

::+:1

From

which,

Q-26)



152

Mechanical Design of Process

Systems

For

60"-4

arrangement,

p

:

1.75 rn.

^

_

8[0.43P'z

-

0.52'd"'z/4]

""

-

"dr,

^

_

8[0.43(1.75)7

-

0.5rr1.25r']l41

_,/1rr-

'

r(l .25)

or

D"

:

0.119

ft

c

=

tube clearance

:

L75

-

1.25

:

0.50

in.

B

:

distance between baffles

._

-.

-

-^.

ln.

n

B:

tnnll

1.r.

=

1100) l^',',"=l t0.8tr/\tt

\u.

r

lvl

Btu

-- -

hr-ft2-o

F

For

gases

used

in this

application the

fouling factors are

0.001

shell-side

and 0.001 tube-side. Solving

for the

overall

heat transfer coefficient,

(7-2e)

IT

_

:

22.50 rn.

I

_

+

0.001

80.83

Btu

I

:

l/'\l-

-

"

--

hr-ft2-'F

+

0.001

+

I

23.40

8,634,780

+

r

^rn

as

8 baffles

Computing the

flow

area of tube bundle

=

a,,

D.(cXB)

. ,

a\=-ll-

(7-301

"

p(t44)

(40)

in.

(0.50)

in.

(22.50)

- l. t9 rt'

in

2

(1.75xt44)

-

t'

For the shell-side

gas, p

:

0.09 lb-/fC

average

for tem-

peratures

specified,

and

p

:

0.05 lbm/ft-hr

lh

rt n <n

'"m

c"

-

.=^.+

=

12.348.00-15

"

1.79

11'

hr-ft2

^,

_

D.G,

_

{0.119)

fr

(12.J48.00)

lb./hr-ttz

_.

4

0.05 lb./ft-hr

NR":29,388.24

The

exchanger

has

baffles with

25

%

cut,

thus from

Fig-

ure

7-22,

jn

:

100

/

\o

t+

n"

:

ff

rr.re,r"t

[aJ

r

Area required

:

1tz.st,)

.

j\

1.lso;"n

hr-ftr-"F

:

1,384.91

ft':

From

previous calculation,

Available

area

:

1,388.95

ft'z

In most applications

the available

area should

not

be

10%

greater

than

the required area, such

material is

not

wasted.

Shell.Side

Pressu:e

Drop

Ap-

f

C.rD,(l_.,t8__t

t)

(7_Jt)

(5.22X10) oD"1d

Ns

=

8 baffles

D.

:

shell

ID

:

40 in.

:

3.333

ft

G,

:

12,348.00

lb./hr-ft2

For

Np"

=

29,388.24,

f

=

0.0022

from Figure 7-23.

For

plain

and bare

tubes,

f

f,=::=0.00t8

t.1.

D"

:

0.119

'y

:

specific

gravity

of shell-side

gas

=

0.9

Np.

=

0.8

from

process

data

/ \o

t+

d

:

r.0:

tl]



-

The Mechanical

Design

of


From Equation

7-35 we

compute the

shear

force induced

aP"

=

(0.0018)(12,348.00f

(3.333X8

+ 1)

(5.22X10)ro(0.

l 19)(0.9X1.0)

AP.

=

0.0015

psi

<

<

10

psi

allowed

on data sheet,

Figure

7-34

EXAIIPLE 7.2:

VIBRATION

CIIECK

FOF

REGENERATED

GAS

EXCHANGER

The exchanger in Example 7-l is to be checked

for

possible

vibration

problems.

To

accomplish this we com-

pute

the damage numbers

of

equations

7-38 and 7-39.

B

=

tube

span

between

baffles

:

22.50

in.

Shell-side

gas

density

=

0.09 lb-/ft3

4

:

1.25 in.

From ASME

Section VIII Div. I

(see

Chapter

4)

for the

tube

rnaterial at design temperature,

o"n,

:

18,052

psi

at

shell-side conditions

ds

=

1.25

in. *

1/o+

in.

=

1.156, where

t/e+

in.

is the baffle hole

clearance

(s€e

Figure 7-34)

From Figure

7-20 we compute the

shell-side

gas

velocity

bgtween

tubes.

D.

:

210.0

in.,

D,o

:

37.125, P

=

1.75

A,

:

+ [o,

-

o," +

D'=

d'te

-

al

nz

t44t

"

*

P

I

on the

tube at the baffle hole,

I

^"

\ /-

V

R:l#kltto-.tnr]

(7-35)

"=

t=''

+

--l

--f",&:

Es

=

27.0

x

1trpsi

qEB

c

_

2(1.00

-

0.?33)

=

4.941

x

10_E

27xlop

'.

=

[.'

?]il''"l,o

*

*,0-',

(oryJ

=

16.015

lbr/ft

Frorn Figure 7-25, a

:

1.10 and b

:

0.0069

Fr:

crPP4

<olito.ri

p

orr.rrur$

(9

o

z1tzz1

":

ft=

tDf

-

SeC"

Fr.

:

1.200 lbr/ft

F.

:

eaF,

L =

q:

F

aFrL

16.015

lbf/ft

^,

=#[oo

-

rr.,r,

.tf#6.?s

-

r.2s)]

Ar

:

2.051

ft2

l|1 lh

G":

18,522.0

-;l

:5.14s

.:"'

tt'-hr

ft'-sec

(5.145)

5

(2.0s1)

ft,

n'-sec

0.0e

5

:

117.236

A

betwe€n

tubes

sec

Nse

(7-38)

(7-39)

2(32.,

-:

J-(l6.0lt

+

lDf

-

SeC- n

(1.

1ox1.2oo)

+

(zig*")

-

^,

-

CpdlpV2ale

rtBE

-

------;-----

zE"r

'

=

6.471

,

rr.rr

(lJoJ

ft6.471)

NeB

:

1.00

",

_

Cedp\Pbla

,.."

_

_EJE;_



154

Mechanical

Design

of Process

Systems

^

bF,

L4

E,I/\

(o.oo6ex

t.2o)

ry l]i I

(22.50)a

fr

rt

\Lz ln.i

th-

(27

\

106)::]

(0.06881

in

4

f(

6

:

9.520

x

l0-5 in.

N.o

:

Nco

:

1.000

With

NBE

and NcD not

exceeding 1.0,

we do not expect

vibration

trouble.

To be certain

we compute

the

maxi-

mum tube deflections

as

follows:

For

the

tube-side,

M,:1.448f:o.o+sf

0,..

:

o.o36cV2d,

H

(;9"

[.J",r [.)

d,

=

1.25

in.

=

0.

104

ft

dt

:

1.084

in.

:

0.090

ft

'"

-

8(rrjo4:

(7

-4r)

(1-44)

Shell-side

(hot

oil)

Tube-side

(chlorine

gas)

GTTD

:

250'F in.

176'F out

77"F

in.

158'F out

173'F

LTTD:

l8"F

(27

x.

ro6)-,-]k

ttt.zsF

in.2 +

(1.084)2

in.2

LMTD

=

crrD

-

LTTD

_

173

-

t8

:68.496.F

. /crro\

irz:\

'"

\tttol

'n

\-tr

/

(with

a

parallel

exchanger no

correction

is

needed

for LMTD)

Tube-Side

Film

Goefficient

For

chlorine

gas,

Co:0.

l16

Btu/lb.-'Fi

p:

|.667

rb-/fc

q

:

rirCog-Uf O;

",

lb.

-

fr

I

9.520

x

l0-5

d-,

:

0.036

r4.94r

x

t0-)(t,r.rru,2

lit,ltl F

ol

\

12

/ \0.04s/

6..,

:

7.553

x

l0

7

ft

:

9.063

x

10-6 in.

With

this

magnitude

of

tube

displacement

and

Nss and

Nsp being

in

the

safe zone,

we

conclude

that

the

ex-

changer

will

not

have

vibration

problems.

EXAMPLE

7.3:

CHLORINE

SUPERHEATER

DESIGN

A

plant

wishes

to

use

hot oil

to heat

chlorine

sas.

The

exchanger

unit. a chlorine

superheater.

is to

be

i TEMA

18-150

AEL.

The

chlorine

gas

is

to

be heated

from

77oF

to

158'F

and the hot

oil is

cooled from

250.F

ro 176.F.

The

exchanger

is

to

be

rated and

analyzed

for

tube-

tubesheet

loading. The

exchanger

specification

is

shown

in Figure

7-36.

The

thermal

duty is

600,000

Btu/hr. The

exchanger is

a

parallel

flowing

unit.

rt2 /rr <n\n

tItT.236f

"

,

(0.00691

l':::l

fi.

sec'

\

12

|

.,

,0""

lb.

/

r ,rug

\

/rzza

in.r\

"

"*-

*l

1:z-z

ruJ

\--

ft-/

f"

=

1.710

Hz

For shell-side

fluid,

,'^ /,^,..^\

p

=

o.oe

r: ,llils

|

:

o.oo3:15

'

frr

\

32.2

lbJ

ft'



I

2

3

5

6

7

a

9

to

ll

72

14

t5

l6

t7

l8

l9

20

2l

22

23

21

25

26

27

2E

29

30

3l

32

33

31

35

36

37

3a

39

1to

4l

4Z

13

11

45

46

47

4a

49

50

5l

The Mechanical

Desisn

of

Shell-and-Tube

Heat

Exchansers

155

55

57

59

50

5l

H

EAT

EXCHANGER

SPECIFICATION SHEET

Add.€ss Prcposal No.

Plaht

Locarion Dale Rev.

Siz. TypG

(Horlvert)

Connected

In

Pa.allet

Series

Surf/Unii

(Gross/Eff.)

So Ft: Shells/ Unit

Surr/Sh.ll

(Gross/Eft.)

So rl

PERFORMANCE OF ONE UNI'I

ShcllSid€

Tube Side

Ur:T

otL

EEDfuflE

GA-

Ffuid

Ouantitv. Total

Lb/Hr

Liquid

T€mper.tur.

(lnlo l)

,

'F

7{6

soecific

cravitv lC

ÎEg

I

^fiO

Viscosity,

Liquid

Cp

Molecular W6isht,

Vapor

Molecular

Weighl Noncondensable

Specitic

Heat

Btu/Lb

"F

o.+zao

O.11to

Thermal

Conducalvity Btu Ftltlt Sq

Fr

'

F

Latent Heat

Btu,/Lb

@ "F

Inlet

Pressure

Psia

Ftls

Pressur€

Drop, Allow.

Calc. Psi

Foulins

Resisranc.

(Min.)

Heat

Exchansed

(D

O

O.OOO

Bru/Hr: MTD

(Correcr€d\

b ,t,5

Transler

Rate. Service

CONSTRUCTION

OF ONE

SHELL

sletch

(Bundle,Noz:le

Orientation)

Shell Side

D€siRn,/T€st

Pressur. Psir

15U

t /,79

DesiEn

TemD€rature

'F

No.

Passes

Der

Shell

Corrosion Allowance

ln,

Sizo

&

Ralins

ln

Out

rube No.,5O

op

I

In.;rhk

(Min/^vs)

In.r r€nsrh

r5Ff':

Ft; Pitch

If{

In. +30

a.50€-so

€>a5

Tsbe

Type

Material

Shell

Channel or Bonnet

Tubesh€ct-Stationary

Tubesheet.Floating

FloatinE Head Cover

lmpins€menr Protectio

Bafites.cross

b

TvDe

4h

%

cut

(Diam/area)

1

.j/4"spacine:

cuc tnlet

In

Supports-Tube U-Bend

Type

Bypass

Seal

Arransem€nt Tube-Tubesheet Joint

Bundle Exit

Gaskers-Shell Side

Code Reo0irem€nts

TEMA

Class

Weight/sherl

Lb

Figure

7-36. Chlorine

superheater heat exchanger

specification sheet.



156

Mechanical Design

of

Process

Systems

700,000

P

ft:

nr

:

9.116

_

Btl,

(68.496).F

rDm-

-r

88.0ee.783'u.(

t*

)

^

hr

\3600

sec/

lh

| .667

:

ftr

For each of

the

150-l-in.-14

BWG

tubes,

ft3

14,680

:

v

=

-........-

J9c

=

25.'796

ft/sec

-

Reasonable

(0.0037941ft

2(

I

50)tubes

Nq.

=

j:l;

I

=

0.0148

Cp

=

0.036

lb./[t-hr

r2s.7e6t

Llg{'o)

',

(1.667r

r

{:ooo

"''l

sec\12

/

ftr

\

lhr

/

88,099.733

5

nr

frl

14.680

i-

sec

(0.036)

lb'

'

'frhr

:298,860.527

u".:l.Cp,k=5.0x to-r-gu

'

k

hr-fc'F

ro.036r

lb'

,0.116r

Btu

--

ft-hr

'

lb^-'F

r'

=_ _____ =Aal(

(5.0

x

l0

3)

Btu

'

'hr-ft-"F

N*"

:

0.027(NrJo

tN.,l,'

(re)'

'',

0.7

<

Np,

<

17,000

^

_

8[0.43

p')

-

0.5rd 4]

Nn.

=

NN"

=

0.027(298,860.527)0E(0.835)r/3(1.01

:

619.a64

N",:\ i

=h,

(610.464r(5.0

t

l0-,)

Bt'

'

hr-ft-'F

r(1.0)

D"

:

0.711 or De

:

0.059 ft

c: |.25

-

1.00

:

0.25 in.

B

:

30 in. for 6 baffles

D,

:

18.00

in.

:

shell ID

D.cB

(

18.00)(0.25

X30)

I

=

_-

:

t-t.11

ll.

-"

p(144)

\1.2s)(t44)

^rn

as

lh .a^

(6.664)

iI

(3.600):::

^

sec hr

--

lh

G.

:

______=-_:=_______j::

:

31.987.20

.+

(J.

/) rt'

hr-tt'

lqi,t\n

\12 l

h,

=

41.866

-+-

r-n'-

-f

Shell.Slde

Film

Goefficient

q

:

fiCP(LMTD)

cp

:

0.426

Btu/lb,-"F;

lh

p

:

62

46

-;T:

k

=

0.077

Btu/hr-ft-'F

(.42o

d+

(68.4e6).F

:

6.664

r

sec

6.664

l9r

E9

:

0.10? ft3/sec

For 60" A

arrangement,

Tdo

8(0.43)(1.2sf

-

0.5 r(l.0)z

l

4l

th

62.46=

ftr

The smallest shell-side nozzle is the

3-in. outlet, where

Ar

:

7.393

in.'?:0.051

ft2

fr3

v

:

_--

g

:2.092:

-

Very reasonable

0.051

ftr sec

r'l'.:

&,

u

=

2.544ltt^tft-hr

12.5+4t

lb^

ro.426r

Btu

Np,

=

ft-hr

lb.-'F

=

t4.075

0.077

Btu

hr-ft-

"F



,0#

P

z.su

lb^

ft-hr

The

exchanger

has

baffles with

45

%

cut, so

from Figure

1-JJ

in=

12

I \0.14

U

=Jx 11-trr:l

Pl

"

D.'

"'

\p",

h

-

(l2xo.o77)

64.025)r/36)

Btu

-

o.o59

tt

t

tt

;$:F

Since both

gases

are

relatively

clean, the fouling

factors

for

both

sides are

0.0001.

1

*

o.oor

+

o.ooo8

*

,1 ..

37.779

43.866

U=19

700,0m#

The Mechanical

Design

of

Shell-and-T\rbe Heat Exchangers

'v

=

1.001

d=1.0

Aa

:

ratio of OD

to ID of tube

Ar

=

1.199

r,.-

:

h'

=

73'629

-

ot.+os

Blu

Ar l.l99

--"-'br-ftL'F

h

=

196.720

-

6r.M

(196.720

-

99.680)

61.,109

+

107.480

t*

=

161.436"F

Maximum allovable

tube

joint

load

=

L"*"

1..*

=

A,ouf,

For SB-l6l-2fi) at

162"F,

o"n:

10,000

psi

r

-,

:

(0.239)in.1lo,m)

g

(1.0)

=

2,3e0.00 lbr

tn.'

The tube wall tenperature

is used in a method devel-

oped

by

Miller

[21],

which is

a

more exact approach

than most and

consequerdy results in a more economical

design.

P,

:

shell-side

pressure

:

100

psi

At

=

tw

-

ta; ta

=

ambient

air

temperaturo

:

70oF

At

:

161.436"F

-

70'F

:

91.436'F

D.

:

shell

ID

=

Ds

-

CA

CA

=

corrosion

allowance

=

0 for

pure

helium

(inert)(ero-

sion

is negligitle)

D"

=

18.0

-

0.0

:

18.0

in.

,

.

..

=

PR

- (looxg'o)

=

0.0558

in.

'st'ctt

-

og

-

qfp

<te"zooltt.ol

-o.ettool

Use

ta'a

=

34rc-in.

=

0.1875 in.

For the shell,

F.+

=

27.546

x

lffpsi

at 161.436'F

aPs

=

(0.w2zs)(3r,987.20f(1.5x6

+ l)

(s.22X10)ro(0.059)(1.00lx1.0)

=

0.008

psi

which is acceptable

Nn"

:

=

741.842

G

_

(o'059)ft(31'987.

Area

required

=

:

521.875 ftz

(re.582h;h(68.4e6)

Available

area

=

(0.2618X150X15)

:

589.050

ff3

This

implies

a

12.87

lo excess,

which

is

acceptable.

Plessure

Drop

f"G.,D"(N" + l)

^''

=

iSzrl

o)'D.rd

\

:

6

baffles

D,

:

shell

ID

=

18

in.

=

1.5

ft

C"

:

ft,gSZ.Z0,l\.

nr-n'

f

f"=

i=

0.w225

D"

=

0.059

ft



158

Mechanical

Design

of

process

Systems

Tube

Metal

Temperature

For

parallel

flow,

Atn:259-77:173.F

At: tla

-

158

:

lS.F

At": 18

=o,no

Atn

113

For hot

end,

Ul:

600,000

where

E1,

:

modulus

of elasticity

of

tubesheet

metal

T

:

tubesheet

thickness

:

1.1875

in.

ar

:

cross-sectional

area

of tube

(see

Table

7-3)

:

0.239

in.2

n:

number

of

tubes

:

150

na,

:

(ls0)(0.239)

:

35,850

in.2

rr lR Or2

A

=

':A

=

254.469

in.2

:

shell

cross-secrional

area

4

rt I Ol2

c

-

/rtr'vr-

tljOt

:

86.394 in.z

-

total

cross-sectional

4

Area or

tube holes

B

=

{

(Di"

-o3r

=

I

Ul8.J75P

.{l8.r2s),J

:

7 .1668

in.2

A

-

C

:

254.469

in.2

-

86.394

in.2

:

168.075

in.2

The ratio

of

the

inside

shell bore

area

to

the net

tube-

sheet

area minus

the tubes

is

the net area

that resists

the

tube

and

shell reaction

forces

and moments.

This

ratio is

referred

to

as

the ligament

or

deflexion

efficiency

and is

expressed

as

(A-C)

tl

:--

^

Ler

I

=

4a,

-

(13.79q

^

106X35.850)

- ) sosr

'

E,B

t21.546

y

106X7.1668)

Let APn

=

equivalent pressure

difference.

psi

:6.794

(s

10.510x173)

For

cold

end,

g-

=

6oo'o00

=

6s )q^

'

(s

10.510)(18)

ltl

^

_

luh

-u"l

_

6.794

-

65.2e41

'-l

U[

,-i-5i%

l=u6eo

From

Figure

7-1"1,

F"

:

O.28

Lr,=tr,o*F.(tr-tm)

t"i

-

176

+

(0.28X250

-

176)

:

r96.jZO.F

L":

t"i

+

F"(t""

-

t"i)

t-=77

+

(0.28X158

-

711

=

99.6t0",

t*=t"i,-,

.

tt.n-r".t

n,o +

n.

ct

:

coefficient

of thermal

expansion,

in./in.-.F

For

the shell

material,

o,

:

6.090 x

10-o

in./in.-'F

at 161.436"F

Pri

:

tube-side

(channel-side)

pressure

:

100

psi

At

=

161.436'F

-

70"F

:

91.436.F

dt

=

0.834

:

100

-

100

-

(100)(35

850)

168.075

(7-11)

(7

-t2)

D

rri

n

ar

:

-21.3298

psi

Computing

the differential

thermal expansion

:

Ac

Aa=e,A,-o,A.

4*:

(7.010

x

10

6X91.436)

-

(6.090

x

t0

9(91.436)

:0.000084

PE

:

the effective

pressure

differential induced

by the

equiva-

lent pressure

difference,

APs,

and thermal

expansion,

Aq

P,:P+(ao)

qna'

A_C

(7

47)



fl

The

Mechanical Desien of Shell-and-Tube Heat Exchansers

159

Pe:

-2r.32e8

+

(0.000084)

(13

7e?l

lq6)(3s

85)

168.075

:

226.263

psr

Assume

ihe normal

tube

projection beyond the tub€sheet

to be

r/a

in.,

L

:

(13X12)

-

2(1.1875)

-

2(0.125): 153.375 in.

Defining

the

dimensionless

parameter,

tr, as

|

-

t025

\

:

1.08

l---

:rr-l

D.

[Lr

-DTdA

-

L,J

(748)

1.08

'10.25

I

(18.125)

4[(2.50s8X1.ss) +

3.12]

o.1."";

:

-1,418.659

compression <

16,?00

psi

allowable

for the tubesheet material

"I

(13.799 x

i09(3s.8s0)

(153.37sX1.

12sf(27.s46

x

109(168.075)

\:2.696

q,(.-r

:

-415.968

psi

for

|

:

f+

:

-0.046

4r-*r

:

-415.968

psi

is well below the maximum allowable

stress, which means that the tubesheet

is of

sufficient thick-

ness.

One could

repeat the

process

if

it

was desired to use a

thinner tubeshe€t.

Had o.1-o*1

exceeded the maximum

al-

1oo

lowable stress

for

the.tubesheet material,

then

a

greater q

tubesheet

thickness would

have

to

be selected and

the-

process

repeated.

From Figures

'7

-37

,

7

-38,

7

-39

,

and 7 40:

f

I

:

1.55; lz:3.12l,

l:

:

-0.046;

f+

=

1.970

The

maximum

radial stress

in

the tubesheet is expressed

as

.'-.,,ffi11,9'

(7-4e)

4(Vfr

+ fr)

l,o,o.,,

-{

I00){2s4.469X2.sOs8)

I

(rg.rz5\t

f-"

--

(168.07s)

I \

1.125

i

a

4

6 A 1ot2

14

16

\

Figure

7-37. Tube stress

factor Ir versus

\.

2

4

6

a lo12

14',r6

|a

X

Figure 7-38.

Tube stress factor 12

versus tr.



160

Mechanical Design

of Process Systems

The maximum

stress

in the tubes is the

sreater of the follow-

ins:

:u-[^,.-

na,

I

",

(n,

-

A_C

APt*

or

(7-50)

(7

-sr)

o.2

q

o.o

-o2

-ot

-0.6

-oa

-1.O

[

",{t"

-:t'e

]l

=A

_clAP,_

-\-

(A-crl

nu,[

-

(*

+

|.4)

I

;.lr"-l-,,,,n

,lrr^ ro^

-

t-"

-'

T

Figure

7-39.

TUbe

stress

factor

f3 versus

\.

2

4 6 8 rO12

14

16

18

I

Figure

7-40. Tube stress factor f4

versus

\.

(2.5058

+

1.970)

o,1^

1:

-92.62

psi

for

Equation 7-51

EXAMPLE 7-4:

ASPHALT

COATII{G

lllx

HEATER-A

NON.IIEWTONIAN

FLUID

APPLICATION

A roofing manufacturer

needs a shell

and tube

heat ex-

changer

to

heat an asphalt coating

mix

from

425'F to

500'F to

improve

flow

characteristics.

The fluid to

heat

the

asphalt

coating

mix

is a leading

manufacturet's

hot

oil heat transfer

fluid. The asphalt

coating

mix is to be

tube-side

and

the

hot oil is to be shell-side.

Determine

the size of

unit required

with the

design

to

be counter-

flow.

The

process

is described

in Example 3-6. The

ex-

changer

heat duty is to be

1,000,000 Btu/hr.

See

Figure

7-41

for

complete

exchanger

specifications.

First

we

compute the

LMTD for a

counterflow

exchanger,

Shell-side

(hot

oil)

TLrbe-side

(asphalt

coating

mix)

650"F

in

500'F

out

550'F

out

425'F

rt

GTTD:

150'F

LTTD:

125"F

LMrD:

crrP

-

qrp

_

l5o_-

l?5

:

r.7.t2"F

. lcrrDl ,

11501

'"\rt-/

'\*/

({

+

lr)

(100x254.469X2.50s8)

168.075



The Mechanical

Design of Shell-and-Tube

Heat Exchangers

161

HEAT

EXCHANGER

SPECIFICATION SHEET

Job No.

Addr€ss Proposal No.

Plant Location Date Rev.

Siz. Type

(Horlv€rt)

Connected

In Parall€l

Se.i€s

Surf/Unit

(Gross/E

f.) Sq Ft: Shells/Unit

Su.r/Shell

(G.oss/Efi

.)

So

Ft

PERFORMANCE

OF ONE

UNIT

ShellSide

HOT

otr-

ASHJNLTCd'NN6FNIf,

Ffuid

OuantitY,

Total

Lb/Hr

Liquld

Tsmpe.ature

(lnlout)

//.60

5=

4z€

sDecific

cravirv

@ 656cp

t.60

a

7

l-lz,

Viscosity,

Llquid

Cp

h.4) a.7>

q3*

q4?

Molccula. Weipht.

Vaoo.

Molccula. Weisht,

Noncondensable

Specitic

Heat

gtu/Lb

"F

0.52b

d.<2b

a.7b

t

D:47

Thermal

conductivity

Btu

Ft Hr Sq Ft

'

F

Latent

tleaa

Btu/Lb

@

'F

Inlet

Pressurc Psia

Ftls

Pressure

Drop,

Allow.,/Calc.

Psi

IO

I

tO

t0

b

Foulins

Resisranc€

(Min.)

Heat

Exchanaed

atu/Hri

MIO

(Correcied)

"F

r-".r". n"r",

S"-i."

so

rt

"

r

CONSTRUCTION

OF

ONE

SHELL

Sketch

(Bundle/No:?le

Orientation)

Shell

Side

DeEisn/TestPressurc

Psis

l<D r

221

EO /

226

O.sasn Temperature

No.

Passes

D€r Shcll

Corrosion

Allowanc€

ln.

Si2G

&

Ratins

Out

luge

o.

5gt

op

74

In.:rhk

(Min/^ve

lIl 8ly6

In.;

Len$h

20

Fr;

Pitch

I.A

In.

<- 30

fgl+119

9

Tubc

Type

Material

sh€l

274l

tp op

In.

lshen

cover

0nt€s.)

(Rernov.)

Channel

or

Eonnet I

Channel Cover

Tubesheel-StationarY

Tubesheet-FloatinE

Flo.lina

Head

Cover

lmoineement P.otection

o/o

Cur

(Dia6lA.ezt

Spacina:

c/c

Inlet

In

Baftles-Lorg

Seal Typ€

Supports-Tube

U'Bend

Type

Bypass Seal ArranAem€nt

Tub€-Tubesh.et Jolnt

pvt-lnlet

Nozzl€

Bundle

Entranc€

Bundle

Exit

Gaskets"shelr side

Tube side

-FloatinE

Head

codc

Requirements

TEMA Class

I

2

5

6

7

a

9

to

t1

t2

t3

t1

77

t8

l9

20

2l

22

23

24

25

26

27

2a

29

30

3l

32

31

35

36

3a

39

40

41

42

11

16

47

4a

49

50

5l

53

55

59

50

61

Figure

7-41. Asphalt

heater heat

exchanger specification

sheet.

l@1978

Tubular

Exchanger Manufacturers

Association.)



6s0

-

425

=

0.333;

R

:

500

-

425

From

Figure 7-16,

F

:0.93.

Thus,

the corrected

LMTD becomes

LMTD

:

(0.93X137.12)

:

127

.522"F

Tube.Side Film

Coefficient

For

asphalt coating

mix

at 450'F we have

the follow-

ing

properties:

Cp

:

0.368

Btu/lb.-'F;

p

:

89.2321b.ift3;

p

:

933

"O

2,251.20 tb^/ft-hr

q

:

fiCo

(LMTD)

=

1,000,000 Btu/hr

P=

162

Mechanical

Design

of Process

Systems

In

a

counterflow

exchanger we

must correct

the

LMTD.

Using Figure

7-16

we have

for a one-shell-pass,

two-tube-Dass.

500

-

425

650

-

550

To obtain

the tube-side

film

coefficient

we must

obtain

the Reynolds number.

The

asphalt

base

coating

mix

is

a

non-Newtonian

fluid

(see

Chapter 1),

so Equation

1-6 is

not

valid.

So, to compute

the Reynolds

numbet

we

must

use Eouation

1-7.

Nn"

:

DiV2

-

ip

(1-7

)

When

working with non-Newtonian

fluids,

rheological

data are necessary. The reader

is encouraged

to refer to

Govier

[22],

but will

often find

that

rheological

data

are

not

available

in literature. In

this situation a

samole

of

the

fluid

must be sent

to

a

testing lab. Do

not

attempt to

approximate

a

non-Newtonian

fluid

with

Newtonian

equations

and

assumptions-the

results

can be

a

catastro-

phe.

At

the current state-of-the-art

there are no

simple

answers for

such

complicated

subjects

such as non-New-

tonian

fluids.

Samples

of our

fluid

were

sent to a testing lab

to have

the

properties

evaluated.

Some of these

properties

have

already been

given.

The

fluid

is determined by the

lab to

be a

Bingham

fluid, in which the

shear stress and veloc-

ity

gradient

ofthe fluid

particles

are

linearly related. For

a

Bingham

plastic,

n in Equation 1-7

is

l-4x13*xal3

1-xa

where x

:

ratio of

the

fluid particle yield

stress

to the shear

stress

in

the

fluid

particles

at the tube wall

Lab tests reveal that x

:

0.5 and

1

:

3.9 for

which

.

4

.^

-.

(0.5r

I

-

-

(U.))

+

-:--------

n=

1

:=

=:;

"

=

O.378

l

-

(u.)f

Now,

€r

th

(0.584)0

r78(0.059)?-"

-1

(89.232)

+

N.-

=

sec n"

:

0.092

'*

8.0

The

film

coefficient is

determined

from

Figure 7-42,

which

is the Metzner-Reed-Reynolds number

(Equation

1-7) versus friction

factot

f. From

this

figure we

obtain

f

:

180

Now, we must compute the

pressure

drop through

each

tube to determine if a

3/+

in.

-

14 BWG tube is adequate.

1,ooo,ooo

9

hr

:

2l,309.196

lb^/hr

rO 16Rr

-i:L

r

l t? {t)\oF

---'''lh

-oF

--

'---'

'

/\

th I thf I

21.309.196

"'

l

' '"

I

^

hr

\3600

sec/

^

^,,

tt3

i

u.uob

_

lh

sec

co,'tt'"m

It,

We

will

try

594-3lq-in.

tubes-14

BWG.

tube wall thickness

for internal

pressure,

PD)

t-,":

o"11E

-

0.6P

where

o1

:

maximum

allowable

stress

for tube material, psi

E

:

tube weld

joint

efficiency

:

l.g

P

:

internal pressure,

psig

ID

:

tube

ID,

in.

(

150)(0.584)

t-,"

=

:0.005

in.

"-

(17.s00x1.0)

-

0.6(

150)

t""r4

:

0.083 in.

Flow

velocity

through each

tube is

fi3

0.066

i-

v

=

--

-

=j L

:

0.059 frlsec

(0.0019)ft'z(594)tubes

Checking

the

150 psig



The

Mechanical

Design

of


with

a

viscosity of almost

1,000

cp.

The Prandtl number

for our fluid

is

N".:

f

(2251.20\

lb'

,0.368'

Btu

ft-hr

lb--"F

(o.lo)

Btu

'

-

hr-ft-"F

Np,

=

8284.416

For laminar flow,

the Sieder-Tate correlation

is

N",

=

T:

,

eo

[6*.16.,;[n)]'' kl''

N",

:

r.86

[,o.or,,rrro.o,u,

ffi]"',t.,

o

lt-

c

.9

.9

u-

hrD

=

k

2.r85

Meizner Reed

Reynolds Number' Re"*

Figurc 7-42. Friction factors for

flow

of non-Newtonian

flu-

ids

[22].

For our velocity

heads we use the

entrance

and

exit

loses

and

get

f

:

O.ZS

+

1.00

:

1.78

(see

Figure l-1

l)

Using Equation

1-4

we compute

the

pressure

drop over

a

Shell-Side

Film

Coef

ticient

q

:

rirCp(LMTD)

For the

hot

oil

at 600"F the following

properties

exist:

Ce

:

0.526

Btu/lb--'F;

p:

(O.997)(O.a)

:

62.213

tb^/tt'

Rr,r

(-

j6)

){{l

ll

-

hr-ft-'F

"L_,|

10.5841

^

\

12

/

Btu

'-''

hr-ft2-'F

2O-ftJong

tube as

op,

:

ILL

*

r* )qr

\d

-

lze,

(t

-4)

ae,

:

p g(zo{l?I'*

*

r.zr

]

(8e.82)k(o.o5eFg(,-iI--J

2(32.2)

It-lD;T

aP,

=

2.47gpsi

:

Acceptable

sec'-rDl

Looking

at

this

pressure

drop one realizes that a flow

velocity

of 0.059 ft/sec is not

so slow for

a

bulky

fluid

m

:

0.076

Btu/hr-ft-'F;

p

:

0.30

cp

:

0.720

lb.ift-hr

_/-\

t.000.000

l'tu

I

t

nr

I

hr

\J.600

sec/

,

...

lb-

Rf

{0.526)

"'-

1127.522\'F

tb.-'F

th

4.141

:

sec

^ ^.-

ftt

:

U.UD/

-

h

Sec

62.213

+

tt'



,,.,

-

ACp

rrpr

_

k

164

Mechanical Design

of

Process

Systems

The smallest

nozzle

shell-side is

a

3-in. nozzle, making

the maximum shell-side velocit)

fr3

0.067

i:

sec

'

-

boslF

:

1.305 ftlsec

-

Very reasonable

h-

=

155.959

Btu

"

hr-fta'F

Fouling factors

are as

follows:

Asphalt coating

mix

:

0.01

Hot

oil

:

0.004

1

rl

_

+

0.004

+

0.01 +

155.959

4.695

Elr,r

"

-"

hr-ft2-'F

1,000,000

nr

Area required

:

Tlr,r

(4.284)

-'.-

-

i27.s22f

F

nr-rt'-

-|:]

:

1,830.308

ft'?

Available area

:

(0.1963)

i

(zo)

rt

694)

:

2,332.94^

tt

It

Twenty-seven

percent

of the

excess

area

can

be elimi-

nated

by

reducing the

number

of

tubes.

This

would in-

crease

the

flow

rate

in

each

tube

and

thus the

pressure

drop,

which

already

is

at

2.5

psi.

For

non-Newtonian

fluids,

properties

can vary from

sample

to

sample and

extra margin

is needed,

so25% to 30% excess

area is

not

unreasonable.

For more

heat

exchange it would

be

better

to

consider a surge tank

with

interior and exterior heat-

ing

elements, since we are at t}te

limits of

the shell and

tube design and, with a more viscous fluid, a surge tank

of

the

type in Examples 3-3

and

3-4 is more

practical.

(0.720)

lb'

ro.526r

Btu

ft-hr lh

-oF

(o.o76t

Btu

'

'hr-ft-'F

Fora60'Aarrangement,

n

810.43

p']-

0.5rdl/41

810.4311.00)

-

0.5rrl.0r/41

r4

-

"(0

?t

D"

=

0.127 or De

:

0.011

ft

c:

1.00

-

0.75

=

0.25 in.

B

:

15.0

in. for

16

equally

spaced

baffles

over

20 ft

D.

=

27.00

:

shell

ID

D.(c)B

(27.00X0.25X

15.0)

^

-^^

"

,

"

144p

(1.00)(144)

.

4.t41jl

t3,600r

l

^msechr

<=-:

as 0.703 fl

th

=

21.201

.920

:'*L

nr-n'

*,

_

D.G,

_

1\Rc---

p

(0.011)

fr

(2t,201.9201

5

nr-It"

lh

' '

ft-hr

From Figure'1-21

jH:

12

for

baffles

with

15%

cut

=

323.918

/ \o

rq

n"

=

lo4rNr,f

':[aJ

From

laboratory

tests it

2.0.

(

l2)(0.076)

Btu- ft

hr-ftr-'F

was

determined

thar

plp*

=

Shell-Side

Pressure

Drop

^.

_

tGiD,{NB

+ l)

(slt(t0t6.1d

Ns

=

16

baffles

D.

=

shell ID

:

27 in.

:

2.25 ft

G,

=

21,201.92

lb./hr-ft2

Nr"

=

324 and

from Figure 7-24, f

=

O.0O75

0.011)

ft

(4.983)'/r(2.0)o

r4



For

plain

and bare tubes,

f o

nn75

F_'-"'"--nnn<t<

'

t.2 1.2

D"

=

0.011

ft

"y

:

specific

gravity

:

0.997

The Mechanical Design

of


165

:

L(t)

:

0

when

t

:

a,

and

if

the ratio of dT(t)/dt

to

dl(t)/dt

exists, then

T(t):(200-0-(140-60)

dT(t)

-:

_l

dt

L{r,

=

ln l2oo

-

t\

\80/

dl(r)

/

ao

\/-r\

dr

\200

-

ri

\80i

200-t

1

I'Hospital's

rule states that

,.

T(t.)

..

dT(r)/dt

1.'t

L(t)

i-=

d

dl(r)/dr

or,

witha

:

120'R

Now

using

Equation 3-23 we have

to-Ro

o

LMTD:

. /so\

o

lnt I

\80/

This

problem

is somewhat similar to that of

Example

3-4

in its formulation.

We must

define

the LMTD as the

ratio of two functions

T(t)

and

L(t) for which

r1

|

-1

|

liml

l=

lim

(200

-

t)

80"F

r.al

_I

I

t-uu

troo -

tl

Therefore, LMTD

:

80'F

With this value

of

LMTD,

the exchanger can be de-

signed,

using the correction

factor

in the case

of

a coun-

terflow

unit.

NOTATION

A

:

tube surface

area, ft2

At

:

cross-sectional

area of tube, in.2

a

=

constant

for

a continuous beam shear,

dimen-

sionless

b

:

constant

for

a continuous

beam

deflection

c

=

tube

clearance,

in.

c

:

constant,

in.2/lb1

(Equation

7-37)

C

:

constant

:9.7

x

10

a(sec)05/(ft)'5

(Equa-

tion

741)

g"

=

12fE

t

or)o

5

(Equation

7-2)

Cp

:

drag

or force coefficient

for

a

body

immersed

in a fluid, dimensionless

Cp

:

specific

heat

at

constant

pressure,

Btu/lbn'-'F

D:4

x

hydraulic radius. in.

D

:

tube diameter, in.

D

:

parameter

(Equation

7-27)

ds

=

diameter

of baffle hole, in.

di

:

inside tube diameter.

in.

Tube-side

Shell-side

1141P

:

T(t)

:

Lt)

200'F

in

120'F

out

at:

80'F

(200-0-(140-60)

.

1200

-

rl

ln l-l

\80/

140"F out

60'F

in

At

=

80'F

As temperature t approaches a certain

value

such

that

T(t)

and

L(t)

become zero being

divided

by

zero.

The de-

rivatives of T(t) and L(t) exist when t approaches this

value

of

t, so

we

can apply

I'Hospital's rule that if T(t)

/

\o

t+

d:

1.0:

(E

(0.0062s)(2

r,20 l. 9D,

Q.2s)

(r7

)

(s.22)(10)'0(0.01

1)(0.997)(1.0)

:

0.188

psi,

which is acceptable

EXAMPLE

7.5:

ZERO LMTD EXCHANGER

A candy

manufacturer

wishes

to

cool

hot

molasses

to

140"F for the food

processing

of

various

confectionar-

ies. The molasses is coming

from

a

heating-blend kettle

at 200'F. Spring water is to be used and it

never varies

(

+

t/+'F)

from

60'F.

The water is to be heated

to

120'F,

and

held

at that temperature

to heat honey. Determine

the LMTD. The exchanser is a counterflow desien.



166

Mechanical

Design

of

Process

Systems

do

:

outside

tube

diameter,

in.

4

:

outside

tube diameter,

ft

Ea

:

modulus

of

elasticity of

baffle material,

psi

4

:

modulus of

elasticity

of tube material,

psi

F"

:

correction factor,

dimensionless

(Figure 7-16)

F".

:

critical

buckling

strength

for tubes, lb.

Fr

:

force induced

by

fluid

flowing

around im-

mersed body,

lbg

F,

:

shear force

against

tube

at baffle, lbr

; I

constants

used in

determining

tubejoint force,

i'

I

lbs

(Equations

7-3

and

7-4)

f"

1

fundamental

natural

frequency

of tube, Hz

gc

:

gravitational

constant

:

32.2

lb.-ftilbr-sec,

GTTD

=

greatest

temperature

difference between

the

shell and tube

side fluids,

'F

h

=

film

coefficient,

Btu/hr-ft -'F

hi

=

film

coefficient

inside

tube,

Btu/hr-fl:,-'F

h"

:

film

coefficient

outside tube, Btu/hr-ft -'F

hi,

:

outside film

coefficient

of tube, using outside

tube

surfaces

temperature,

Btu/hr-ftl'F

I

:

moment

of inertia,

in

a

Ir

:

moment

of inertia

of

tube cross

section, in.a

k

:

structural

constant,

dimensionless

(Equation

7-2)

k

:

equivalent

effective

unsupported length

of the

tube,

in.

k*

:

coefficient

of thermal

conductivity

of tube

wall, Btu/hr-ft-'F

kr

=

thermal conductivity

of

fluid, Btu

kn

:

thermal

conductivity

of

foreign deposits in-

side of tube,

Btu/hr-fi-'F

kso

:

thermal

conductivity

ofdeposits on

outside of

tube,

Btu/hr-ft-'F

L

=

tube length

or

span length

of tube,

ft

LMTD

:

logarithmic

mean

temperature

difference,

"F

LTTD

:

lesser

temperature

difference between

shell

and tube-side

fluids,

'F

/

:

typical dimension

of body

immersed

in fluid,

n

rir

=

mass flow

rate, lb-/sec

mt

:

mass

density

of

tube metal,

slugs/ft3

NB

=

number

of

baffles

Nna

:

baffle

damage number,

dimensionless

Nco

=

critical damage

number,

dimensionless

(Equa-

tion

7-39)

Np,

:

Nusselt number,

dimensionless

Np.

:

Prandd

number,

dimensionless

Nr"

:

Reynolds number,

dimensionless

P

:

axial force,

lbl

p

:

tube

pitch,

in.

q

:

rate

of heat transfer,

Btu/hr

r

=

radius of

gyration

of tube,

in.

(Equation

7-2)

T

:

parameter (Figures

7-30 and

7-31)

Tn

:

thickness of inside

tube

deposits, ft

Tro

:

thickness of

outside tube

deposits, ft

T*

:

tube

wall

thickness,

ft

t""

=

caloric

temperature

of

cold

fluid,

'F

t"1

:

caloric temperature

of hot

fluid,

"F

Li

=

inlet cold

fluid

temperature,

oF

t""

:

caloric

temperature

of cold

fluid,

'F

thi

:

inlet

hot fluid

temperature,

'F

th.

:

outlet hot

fluid

temperature,

oF

t

=

tube

wali

thickness,

in.

t*

:

outside

tube wall

temperature,

'F

ar

=

temperature

differential

(tr

-

tz),

.F

U

:

overall heat

transfer

coefficient

for ex-

changer,

Btu/hr-ft2-'F

U,

:

the value

of

the overall heat

transfer

coeffi-

cient

at the caloric

temperature. Btu/hr-ft2-.F

V

:

flow

velocity,

ft/sec

Greek

Terms

ct

:

factor of effective

tube resistant

area,

dimension-

less

6

:

deflection

or displacements,

in.

p

:

dynamic viscosity

of

the fluid inside

tube,

lb./ft-

hr

p*

=

dynamic viscosity

of

fluid

at tube

wall, lb-/ft-hr

uB

:

Poisson

ratio for

baffle

material

ut

:

Foisson

ratio for

tube material

or

:

frequency

of

a

given mode,

Hz

p

=

density, lb*/ft3

d"1

=

allowable stress

for tube,

psi

o"

:

allowable tube compressive

stress,

psi,

for

the

tubes at the outer

periphery

of tube bundle

(Equa-

tions 7-1

and 7-2)

o,

:

minimum

yield

stress

of

tube material

at

design

temperatue,

psi

f

:

sum of structural damping

and the

fluid

damping,

dimensionless

REFERENCES

l.

Heat

Exchangers,

Howeli Training

Company,

Houston. Texas.

1975.

2.

Snndnrds

of

the Tubular

Exchanger Manufacturers

Association

(TEMA),

6th Edition,

Thrrytown, New

York,

1978.

3.

Rubin.

F. L.

.

"What's

the

Difference

Between

TEMA

Exchanger

Classes," Hydrocarbon Process-

ing, 59,

June

p.

92,

1980.

4. Ludwig, E . E., Applied Process

Design

for

Chemi-

cal

and

Petrochemical

Plants, Volume

3.

Second



Edition,

Gulf

Publishing

Company,

Houston,

Texas. 1983.

5.

Small, W. M. and R. K.

Young,

"The

Rodbaffle

Heat Exchanger," Heat Trans. Eng.,

I,

ro. 2,

Oct.

-

Dec.

(1979),

p.

21.

6.

Skrotzki,

B. G. A.,

"Heat Exchangers," Power,

June,

1954.

7. ASME

Boiler

and

Pressure

ry'essel

Code.

Section

VItr

Division 1, American

Society of Mechanical

Engineers, New York.

8.

Colburn,

A. P., Ind. Eng. Ch.em.,35,

pp.873-877,

1933.

9.

Kern, Donald

Q.,

Process

Heat

Tlansfer,

McGraw-

Hill

Book

Company,

New York,

1950.

10. McAdams, W. H., Heat hansmission, Third

Edi-

tion, McGraw-Hill

Book

Company,

New

York,

1954.

ll.

Jakob,

M.

Heat

Transfer,

Yol.

l,

John

Wiley

&

Sons,

New York,

1959.

12.

Grimson,

E.

D.,

"Correlation

and Utilization

of

New

Data on

Flow

Resistance and Heat

Transfer

for

Crossflow Over

Tirbe Banks

i

'Tiansaaions

of

the

ASME,"

Yol.59,

pp.

583-584,

1937.

13. Engineering

Data

Book, Wolverine

Division

of

UOP,

Inc.,

A

Signal

Company, 1959.

14.

Thorngren, John T.,

"Predict Exchanger Tube

Damage,'

Hydrocarbon Processing,

I*l,l.

49,

rc.

4,

p.

129,

r97o.

The Mechanical Design of Shell-and-Thbe Heat Exchangers 167

15.

American

Institute of

Steel

Constrtclion, Mantal

of

Steel

Construaion,

Eighth Edition,

AISC, Chicago,

trlinois,

1980.

16. Timoshenko,

S., and J.

N. Goodier, Theory ofElas-

tr:ciry, Second

Edition,

Engineering

Societies

Mono-

graph,

McGraw-Hill

Book

Company, 1951.

17.

Coit, R. L.,

C.

C. Reak, and

A.

Iohmeier,

"De-

sign and Manufacturc

of

Large

Surface Condens-

ers-Problems

and Solutions,"

American

Fower

Conference, April

1965.

18. Blevins, R.

D., Flow-htduced

Wration,

Van Nos-

trand

Rheinhold Company,

New

York,

1977.

19. lbmbsganss,

M.

W.,

and S.

S.

Chen,

"Tbntative

Design Guide for

Calculating

the Vibration Re-

sponse

of

Flexible Cylindrical

Elements

in Axial

Floq"

Argonne National

Labomtory

Report

ANL-

ETD.7l-{r/, l9r.

20.

Kays,

William

M.

and

A.

L.

Lofron,

Compaa

Heat

Exchangers,

Third Edition,

McGraw-Hill

Book

Company,

New York,

1984.

21. Miller,

K. A.

G.,

'The

Design

of

Tirbe

Plates

in

Heal

Exchangers," Proceedings

of thz

Institwion

of

Mechanical Engineers,

\bl.

lB,

pp.215-231.

22. Ctovier,

G.

W.

and

K.

Azrz,

Thc

Flow

of Complex

Minures

in

Pipes,

Robert

E.

Krieger Publishing

Company, New

York,

1977.

23. Metzner,

A. B. and J.

C. Reed,

AICLE

Joumal, I,

p.434,

1955.





External

Loadings on

Shell

Structures

In a book about the mechanical design of

process

sys-

tems it is impossible to ignore the

phenomenon

of exter-

nal

loadings

on

shell structures. Such

loadings occur

when

piping

is flanged to

pressure

vessels and the

vessel

nozzle

is

exposed

to

loads induced by the

piping,

and

when vessels are erected

and

the force

of

gravity

induces

loads at the

lifting

lugs.

We

have

already discussed external loadings

in the de-

sign of

piping

supports in

Chapter 2.

Vessels

require a

simiJ.ar analysis, but the

phenomenon

is

different be-

cause

in

a vessel

the

loadings are more

localized.

partic-

ularly in

a

large vessel.

In

the case

of external

loadings

on vessel nozzles one must consider

primary

stresses in-

duced

by

internal

pressure

and

secondarv

stresses

in-

duced

by

the

external loadings.

In

the design

of

the lift-

ing lugs only secondary

stresses need

to be considered,

since vessels

being

lifted almost never have internal

pressure.

The two

"standards"

that are most

widely

accepted for

external loadings

on

pressure

vessel nozzles are

the

WRC

(Welding

Research

Council)

Bulletin 107

[1]

and

the

WRC

297

l2l.

The latrer

is

an expanded

version

with

more curves to cover more cases, but it is only for cylin-

drical

shells. Neither

WRC

107 nor

WRC

297 are

con-

sidered standards

per

se.

Therefore,

one

must

take

the

results

of the methods outlined here and add the

primary

stress,

which

is the

internal

pressure

stress.

The reader is

cautioned that the WRC 297

Bulletin

is

under evaluation

at the time of this

writing.

Shell theory

was

used to develop the

WRC

297

,

and the results are

being compared

to finite

element

studies

currently

being

made. The reader is

especially cautioned to

use

the Bul-

letin

when

the

ratio

of the dianeter

of

the branch to the

diameter of

the

header

is

between 0.5

and

1.0.

exoressed

mathematically

as

0.5 <

db/DH

<

1.0

=

diameter of the branch

:

diameter

of

the header

Also.

\\'RC

197 and

WRC

107

do not

consider the

case

of

erternal

ioading

combined

with

internal

pres-

sure. Current studies are being made

to accomplish this

task.

Stress induced by internal

pressure

at

the

nozzle-shell

intersection

are

extremely

complex,

so an

analytical

so-

lution is impractical. Discontinuity

stresses

at the

nozzle-

shell

juncture

are caused

by

the change in

geometry

from

the

nozzle

shell

into

the vessel shell.

Consequently,

a

stress concentration factor, ko, must

be applied when us-

ing the following expression for

internal

pressure

stress:

(8-1)

:

internal

pressure, psi

:

inside

diameter

of shell, in.

=

shell thickness, in.

:

internal

pressure

stress

concentration factor,

dimensionless

Values of

\

are far too

exhaustive to be listed here,

but

are available in a work by Forman

[3].

For

many years

reinforcing

pads

have

been used

for

external loadings

and it has been accepted

practice

to

as-

sume

that

such

pads

remove

discontinuity

stresses at the

nozzle-shell

juncture.

While

this is true, one must real-

ize that the

reinforcement

decreases the flexibility of the

nozzle-shell

attachment. As shown

in

Figure S-la,

the

nozzle

with

the reinforcement will have

maxirnum mem-

brane

stresses

occurring

at the nozzle-shell

juncture

(as-

suming the circumferential bending

stresses

are negligi-

ble compared

with

the membrane

stresses).

As

Figure

8-1b shows

that as the reinforcement thickness increases,

where

db

DH

P(ID)k"

"n

2t

where P

ID

I

kP

169



c

170

Mechanical

Design of

Process

Systems

M-'x'i€m5'mm'n'|ll

ll

I

--------|-

B

w> 1.6s(arf.

ll

r5

l

HI

_-____1---r

R-r

I

I

i

II

tN_

and

finite element

studies.

--'_;J

Figure 8-1. simple

schematic

of maximum

combined

stress

disribution,

as

supported

by field

tests

F----'1 |

ti

tl

r

ryrcJ

r,

I

the

maximum

stress

shifts

towards

the edge

of the

pad,

and

as

the ratio

of

the

reinforcement

pad to

the

shell

thickness

approaches

a

"critical

value,"

the maximum

stress

induced

by external

loading

occurs

at the

rein-

forcement

edge-shell

juncture

point,

shown

in Figure

8-

lc. Considering

this it

would intuitively

appear

that

a ta-

pered

pad

would

ideally

be

the

best in

application,

especially

for

thick

pads

(pad

thickness

relative

to shell

thickness),

as

shown in

Figure

8-1d.

The disadvantage

of

such

a

pad

would

be

the increased

difficulty

and expense

to

fabricate

such

a pad.

Analytical,

finite

element stud-

ies,

and

field

experience

bear the

previous facts

out.

The

width

of

a

pad,

from the

nozzle edge

to the

pad edge'

should

not exceed

1.65VRT.

Beyond this

range

a

pad has

been shown

to

be ineffective.

Pads

can

be even

dangerous

on thin-walled

shells.

In

many

instances,

adding

a

t/z-in.

pad

to

a

nozzle on

a thin-

walled

pipe,

such

as Schedule

55

(0.083 in. on a

4-in.

pipe), is

prohibitive.

Such

a

pad

could

very

easily

trans-

ier

the

maximum

loading to

the

pad

edge

as shown

in

Figure

8-1c,

resulting

in

crack

propagation or

even

rup-

tuie.

Caution

should

be taken

in working

with thin-

walled

shells, where the

flexibility

of

the shell

is

often

sufficient

to

decrease

induced

stresses

from

external

loadings.

LIFTING

LUG

DESIGN

The design

of lifting

lugs

can

become

an

arduous

task

if

one

is

not

familiar

with

the

erection

of

equipment.

Lifting

lugs

must

be designed

to

withstand

the

stresses

inducad

from

all the

loading

conditions;

allow

lifting

and

setting

the

equipment

in one operation

without

readjust-

ing oi

re-rigging

the

crane or

other

equipment'

and

pro-

teit

equipment

and

personnel. The lugs

must

not inter-

fere

with

vessel components,

such

as

platforms, ladders,

or

piping.

Thi

advantage

to lifting

lug design

is that

only

second-

ary stresses

must

be considered-primary

stress,

such

as

internal

pressure stress,

can

be ignored.

We

can assum€

that

the vessels

are

not

lifted

while they

are

pressurized.

Consequently,

the

AISC

Manual

of Steel

Constructi.on

[4]

can

be used

in

which

the

factor of

safety

is

2:

I

(unlike

ASME's

4:l).

The

vessel

is

to

be considered

as

a simply

supported

horizontal

beam.

All

non-shell

components,

head,

lad-

ders,

etc. are

considered

as

concentrated

loads.

The total

erection

weight

is

the sum

of

the concentrated

loads

and

the

distributed

loads

of the shell

weight

and

internals.

Various

types

of

lifting

lugs are

shown

in

Figure 8-2.

Lifting

and'election

procedures are shown

in

Figure

8-3

Techniques

for

designing

the

lugs

are

given

in

the fol-

lowing

examples.

EXAMPLE

8-1:

LIFTING

LUG

DESIGN

At{D

LOCATION

A 96-in.

ID

shell

and tube

heat exchanger

is

to be

lifted

from

a dock

onto an

offshore

structure'

The

ex-

changer

weighs

158,750

lbs,

which

is the total

erection

weight.

The

objective

is

to locate

and design

the lifting

lugs,

and

determine

the

minimum

chocker

length

and

maximum

chocker

angle.





172

Mechadcal

Design

of

Process Systems

1

T

A

norizontal

lili

"1" or

"W"

beam

c

spreader bar

rig avoids

€xcessive

bending

moments

on

lilling

lugs

First, we

construct a free body

diagram,

as

shown

in

Figure

8-4.

Each

lifting lug

is located such that

the

point

of lift

is

located

on a hypothetical

vertical line that

passes

close

to or

through the centroid

of

the

ellipsoidal

head,

shown

in

Figure

8-5.

Summing moments

to

zero

and

solving

for the

reactions we have

GDt.

:

0

:

-Rnt46.542)

+

(346x44.000)

+

(2,283)(40.7

s)

+

(346)(2.542)

-

(2,094)(46.7

5)

+

(1s1,587)(23.27

r)

Rr

:

75,888.874 lb

and

Rr.

:

75,698'

126

lb

For

lug supporting

the

fulI

vessel weighing

158,750

lb,

referring

to Table 8-1

we write

t

t

+

.l

J\

U

Figure 8-3. Lifting

lug and erecting

procedure

(moments

in-

duced

by

lift

load at choker angle d can

be avoided

with a

spreader bar

or with the

lug

design

in Figure

8-28.

A

=

16.50in.,B

:

6.50in.,

C

:

4.50in.,

D

:

4in.,

E

:

6.50 in.

Hole

diameter

:

a

=

(4.50

+ 0.125)

:

4.625 in.

:

mmlmum

Lug width

:

Wr

:

3a

:

3(4.50)

:

13.50 in.

:

minimum

Lug

Thickness,

t1

wL

_

13.50

88

=

1.688 in.

r/

use 1.75

in.

w

_

158,750

1.6ao,

(1.6X4.625X38,000)

=

0.565

in.

+

tL

:

Larger

of





174

Mechanical

Design

of

Process Systems

Table 8-1

Anchor Shackles

For lug

material

SA-516-GR

7O,

o,

=

38,000

psi

Lug

Height

(assume

2

in.

fireProofing)

"

:

|

*

(?.,)['

-

(".

;'

;f

-

")']"

lug

height,

in.

insulation

thickness

shell

outside

radius,

in.

(a)

screw

Pin

n

tl

ll

/\

f-et

r

D'

(b)

round

Pin

n

tl

|l

1"b

lvt

t+l

where,

H:

R":

t:?

*({*

\2

,\

-)

"["-(

50.0

+

50

JO

6.

n

1.75

50-00

-

4.00

I'

Pin Dia.

THWP,

(in.)

(in.)

(in.)

(in.)

(in.) Load-|9

rh

trys

rlz

shd

tllrc

.79-9-

q*

lttn

71rc

ls

r3/rc

1'100

,1"

Itl,

5/s

llrc

|

1,675

,l*

lrl^

3/c

rlz

lrlrc

2,200

,1, lrl, r'lrc 5/a Lslrc

2,900

tl"

2rltu

ltlrc

3lc

le/rc

4,500

t1o

ZIz

lh

1ls

|1ha

6,400

,t"

Yt^

171rc |

2tla

8,700

1

33lq

111/to

1Vs

23lg

1I

'4OO

-/"

4r1o

tt3lrc

ltlc

25lL

___J3_5W

1tl4

43h

Ztlta

13/e

3

16'500

rr-r"

stlo

2tls

lvz

351rc

2 ,59q

lrlz

53lq

2tlq

lslt

35la

u,w

1t/"

1

27lt

2

451rc

33,600

2

7rl"

3tlc

211+

5

44,800

zrh

-9V^

3?'1"

2tlz

5Vq

56,000

it, - tou qu" 23lq 6

67

'2w

i1o

tluz

4lz

3

6t/+

81,ooo

6tlz

100,800

33/q

63/q

125,000

3t/z

15Vz

t

50,000

16l lz

41lt

r79,200

lTtlz

6tlz

7)lz

200,000

4118

181/z

43/q

224,N0

313,600

r,

)

*.

where R

=

greater

of

RR or

R; for

horizonal

or

w'

vessel

R

:

reaction

at

lug

when

lifting

at

skirt

and

lug

end

5(19.690)

in.

(75,888.874) lb

ll.

(38,000)

+

t13

50)2

in'

ln.'

:

1.079

in.

<

1.75

in.

Lug

thickness

is sufficient

13

D

Safe

Lilting

H

:

19.690

in.

Check

lug thickness

Minimum

Weld

Size

R[0.47

+

0.45(h/w)]

re(wr)(0.707)

r*

=

0.426

in.

minimum

Actual

weld

size

:

t*u

where

ra

:

allowable

shear

stress

in

weld

:

0 3q*

o,*

:

weld

minimum

yield

stress/

in

tension

I rr2jry\l

(7s888.874)

[o

ot

.

o.ot

\ir.roo1

15

tl+

6tlq

(0.30)(70,000)(13.50)(0.707)

4tlz

5t/z

21

6tt/rc

5tlz

73lc

448,000



f

t,

-

t/ro

in.

r*a

:

Larger

ot

I

,,

_

,t,u

in.

and twr

>

h

where

t, =

vessel

thickness, in.

In this

case, tL

>

tv, so

that

t*"

=

1.75

in.

-

0.0625

=

1.688

For

each side of weld

t-,:l'688:0.844

-2

since

t*"

>

>

t*, A

a/+-in.

weld is sufficient

Choker Angle

(0)

o

:

arctan

[----tlt'

-,

I

l3w(H.A.;ll

A,B,C,D,E

:

Ds:

du:

H:

Kp:

L.:

Ml=

Mr:

External

Loadings on Shell

Structures 175

NOTATIOil

constants

(Figure

8-6)

header

diameter, in.

branch diameter,

in.

constant

(Thble

8-l)

internal

pressure

stress

concentration

factor,

dimensionless

minimum chocker

length,

ft

moment

resolved about

the left end

(Figure

84),

ft-lb

moment resolved

about

the

right

end

(Figure

8-4), ft-lb

U:

"r*rI

(38,000x13.50)(1.75F

R"

,

50.00

t"

:

12

rin

(4.90t

=

'+6'/rl

n

Because of height

restrictions,

the lug had

to be low-

ered

from

19.690 in.

to 11.00 in. Thus,

we now have

the

following:

" I

(38.ooox

l3.soxl.7sy

I

.:qrt.grt-l

lrrtst.zso.ooy

{rt.oo

* ro.so

*

4ll

t

\

zll

0

:

6.327'

and

LC:

A

=

16t/z in.,

B

=

61/z

in.,

C

=

4 z in., D

=

4

in..

E

=

6t/z in.

Figure

&6.

Detail of

choker and shackle.

3(1,58750.00)

(rn.uno

* r6.s0

+

4t0)

0:4.905"

I.:

:

minimum choker

lensth

12 sin

d

=

37.807

ft

12

sin

(6.327)



176

Mechanical

Design

of

Process

Systems

:

constant

(Thble

8-1)

:

reaction at

left

side

(Pigure

8-4),

ft-lb

:

shell

outside

radius, in.

:

reaction at right side

(Figure

8-4),

ft-lb

=

shell thickness,

in.

=

lug thickness,

in.

:

weld size,

in.

:

lug

width, in.

Greek

Symbols

o,*

:

minimum

weld

yield

stress

in

tension,

psl

7A

:

allowable

shear stress

in

weld,

psi

0

=

chocker angle,

degrees

REFERENCES

Welding

Research Council,

Welding Research

Coun-

cil Bulletin

WRC 107 bcal Stresses

in

Spherical

and

Cylindical

Shells Due to External

Inadings,

Match,

New

York,

1979.

Welding

Research Cotncil,

Welding

Research

Coun'

cil

Bulletin

WRC

297,

Incal Stresses

in Cylindical

Due to

External

Inadings on

Noales-Supplement

to

WC

Bulktin

No. 107, New

York, August,

1984.

Forman.

B. Fred.

Incal Stresses

in

Pressure

Vessels,

Second

Edition,

Pressure

Vessel

Handbook

Publish-

ing,

Inc.

Tirlsa,

OK., 1979.

American

Institute

of

Steel

Construction,

Manual

of

Steel

Constructior,

Eighth

Edition,

AISC,

Chicago,

Illinois,

1980.

P

RL

R"

RR

t

t1

t*

wL

z.

t.

J.

A





178 Mechanical

Design

of

Process

Systems

(a)

Figure

A-2. Partial volume

of

vertical

hemispherical

head.

(B)

Partial

volume

of

horizonral hemispherical

head.

PARTIAL

VOLUMES

OF SPHERICALLY

DISHED

HEADS

Horizontal

Head

The

partial

volume

of a horizontal

head

(Figure

A-3) is

Example-Spherically

Dished

Horizontal

Head

A

spherically dished

head

with

a I l4-in.

{

OD is

spun

from

1-in. plate. Determine

the

partial

volume

of

10

in.

of

liquid.

From

vessel

head manufacturer's

catalog

we

determine

the following:

IDD

:

16.786 in.

(Figure

A-5),

p

:

108 in.

l14

-

)/t o\

R:

"

-

.'"'=

56.0in.

'2

e:

159.43"

:

2.78

L

:

108

-

16.786

:

91.21 in.

-_T---T

-+l

i

ln'

tv

I

tl

tf

Figure

A-3. Partial volume

of

spherically

dished

horizontal

neaos.

Vertical Head

The

partial

volume

of

a

vertical

head

(Figure

.,

nv(3x2

+

vr)

v='

-

6

ot

..

nv2(3o

v)

y:

-

.

3

atl

P"l

x

v----i-

-v----T

\:-7lTv

llDD

-<--E--------i-:--r

I

Figure A-4.

Partial volume

of

spherically

dished vertical

heads.

(A-3)

A-4) is

(A-4)

--

J___

--.-{,>--

_

(A-5)



Appendix A: Pressure

ry'essel

Formulations

179

Yr =

6.786"

Flgure

4"5.

_

(9t.2r)(562

-

6.7862)

V

:

38,893.21 in.3

=

168.37

gal

Example-

Spherically

Dlshed

Vertical

Head

For the

same head

above, determine

the

partial

volume

of

a head

of liquid

of 9 in.

x

:

55.456 in.

u

-

zr(9)[3(55

416)'? +

9'z]

=

A.874

in.t

=

64.4

gal

6

"'

PARTIAL

VOLUIIES

OF

ELLIPTICAL

HEADS

The exact

partial

volume

of

a

horizontal

elliptical

head

lV(r08,

--i86at

-

v?ro8r

-

5-dF

.,-

IJ

(Figure

A-6)

is

as

follows:

..

(IDD)q

Venical

Elliptical

Heads

Volume

of top

portion

@

of

Figure

A-7 is

-a

v,.'

=

'Ri'

l"

-

Y'

I

'"

2

l'

3(rDDFl

Volume

of

bottom

portion

O

is

. ,

2r(IDD)R,2

rRl I

u3

I

-

"-----:

lw

2

(

3(rDDll

t\

Itr

t\

t\

l

ti

;;=*--:-__T,

_-

(A-6)

(A-7)

End

View of Horizontal

Head

Figure A-6. Partial volume

of horizontal elliptical

head.

Figure 47,

Partial volume

of

vertical

ellipticat

head.

(A-8)



180

Mechanical

Design

of Process

Systems

Horizontal

Head

Example

Find

the

partial

volume

of a 2:

I

(R;/IDD

=

2)

ellipti-

cal

head

that is 108-in.

OD. The

level

of

the liquid is 35

in.,

and

the

head

is

spun

from

l-in.

plate.

IOR

-

?rl

O\

IDD

--

'"-______:rr:',

=

26.50 in.

From

Equation

,4-6

and

Figure A-8

we

have

the

follow-

lng:

y

=

(IDDI

a

vm7

--tl'-

6R,

a

=

138.80"

=2.42

v

_

(

19.0)(2.42t

463r-

*

{Iqy-rr

6(53)

V

:

17,512.94 in.r:75.81

gal

Vertical

Head

Example

For some head

above,

determine the

partial

volume

for

a

vertical

head

with

19

in.

ofliquid.

Using Equation A-8

we

have the following:

.,

_

2a'(IDD)R1'?

o

A

vertical head

KR

IDD

-x

B

horizontal

head

c

vertical

knuckle

region

H=IDD-KR

D

horizontal

knuckle

region

Figure A-9. Partial volumes

of torispherical heads:

(A)

verti-

cal,

(B)

horizontal,

(C)

vertical

knuckle region,

(D)

horizontal

knuckle resion.

v

_

2?r(26.s0x53.01

_

1(5i.0)

[,o

n

_

trq.or,

]

6

2

t--"

3(26.s0),.j

V

=

77,951.81 in.3

-

13i0.75 in.3

Y

:76,641.06

in.3

:

331.78 gal

Figure

A-8.



PARTIAL

VOLUIIES

OF

TORISPHERICAL

HEADS

For Figures A-9

and

A-10,

Vk

:

knuckle

volume

y

:

height of liquid

Vo

:

dish

volume

IDD

:

inside depth

of

dish

KR

=

knuckle

radius

p

=

inside

dish

radius

For vertical

heads

(Figure

A-9c) the knuckle-cylinder

Dartial volume

is

Appendix A: Pressure Vessel Formulations

Figure

A-1o.

end view

of dish

volume

Flgure

A-11, Sketch for example

partial

volume calculation

of horizontal torisoherical

head.

Flgure

A-12.

The

partial

volume

ofthe dish region of

a

vertical head

is

v*:

?rtJ

+

4ry2

+

r,2;

uo

=

"[#

+

Ri-

KR)

+

(R,-

KRr]

., -

?ry(3x2

+

y2)

vD_-6-

The

total

partial

volume in

a

verticil

head is

,,

nH

,.

.

Ty(3x2

+

y2l

vu

:

-6-

(ro'

+ 4rM' + ri') +

-----6-------:-

whereY=IDD-KR

Horlzontal Torlspherical

Heads

Partial Volume

of

Dish

@

(Figure

A-11)

VO:

o

./(p,

-y-il.t

=

V(pt-7F_L(Ri,.

yi,)

|

,o_,r.,

JZ

Volume of Knuck-Cylinder Region

@

(Figure

A-12)

(A-e)

(A-

l0)

(A-ll)

(A-13)

The total

partial

volume

for

a horizontal

torispherical

head

is

as

follows:

V1

:

V6+

V6

.

"lry

+

Ri-

KR)

+

(&

-

KR),]

wherel:

p

_

IDD

-

.vG,

-

R-iT

L(Rr2

-

yi2)

(A-14)



182

Mechanical

Design

of

Process

Systems

Horlzontal

Head

Exampte

A 102-in.

S

OD

flanged

and dished

(torispherical)

head

made

to

ASME

specifications (KR

)

0.60p

and

KR

> 3th,

tr,

=

head

thickness)

is

spun

from

l-in.

plate.

The

head is

horizontal

and the

liquid

level

is

35-in.

deter-

mine the

partial

volume.

From

the vessel

head

manufacturer's

catalog

and Fig-

ure

A-12

we

determine

the

following:

p

:

96

in.,

KR

:

6.125

in.,

IDD

:

17.562

in.

ltut

R,

=

:

=

50in..

L

=

96.0

-

17.562

=

78.438 in.

z

From Equation

A-14

we

have

vr

:

Q.532)

,4%t_

1s+

_\@6r:50it

_

(78.438X50'

-

ls)

r

r-

/,''

<1r,

14(6.125)

|

[

J?r'

+

(5o.oo

-

6.12s) +

(s0.00

-

6.l25fl

'J

Vr

=

34.093.44 in.r

=

147.59

ga.

Vertical

Head

Example

A

138-in.

d

OD

F&D

(flanged

and

dished)

head

nor

made

to ASME

specifications

is

spun from

I

l/z-in.

plate.

The

head

is

vertical

and

the liquid

level

is 18-in.

Deter-

mine

the

partial

volume.

From

the vessel

head

manufacturer's

catalog we

deter-

mine

the following:

p

:

132

in.,

KR

:

3 in., IDD

:

20.283 in.

l?R

-

trl 5l

R,

=

'-"

=-"'-'

=

67.50

in.;

2

x

:

67 .50

-

(3f

-

H2lo

5

=

66.446

in.

For

knuckle-cylinder

region,

ro:

Ri

=

67.50; 11

=

Ri

-

KR:67.50

-

3.00:64.50

in.

67.50

+

@.50

r.=

,-

=

ob.u;

h

:

120.283

-

(3.0

+ 15.0)l

:

2.283

in.

+() 19,4\

Yv

=

"

-;-"-'l(67.501

+ 4(66.0),

+

(64.5011

b

*

r(I'1

.283)[3(64.500)'?

+

(17.283)2]

6

vv:31,247.726

in.3 +

115,645.832

in.3

Vv

:

146,893.558

in.3

:

635.903 gal



I

INTERNAL PRESSURE

ASME

FORMULATIONS

WITH OUTSIDE DIMENSIOI{S

Cylindrical

Shelt

Longitudinal

Joint

Appendix

A:

Pressure

Vessel

Formulations

183

-

oEt

R

-

0.4t

Circumterential

Joint

.PR

oE +

O.4P

r=

PRo

2oE

+

'1.4P

^

2oEt

Ro

-

1.4r

2:1

Ellipsoidal

Head

r=

PDo

2oE

+ 1.8P

^

2oEt

Do

-

1.8t

-

2oEt

R.

-

0.8r

-2rE+0,8P

s

=

0'885P1

oE

+ 0.8P

Sphere and Hemispherical Head

ASME Flanged

and Dished

Head

when UR

=

164s

r

=-

0.885L

-

0.8t

When UB

<

16ry3

t=

PLM

2oE+P(M-0.2)

r=

PDo

-

2 cos o(oE

+ 0.4P)

Section

^

2SEt

cos d

^

2oEt

ML-(M-0.2)

Do

-

0.8t

cos

o



184

Mechanical

Design

of

Process

Systems

FOR VALUES

OF

M

SEE

SUPPLEMENT

INTERNAL

PRESSURE

ASME

FORMULATIONS

WITH

INSIDE

DIilENSIONS

1-\

i-

-----T;-'-

,-il

/l\

{,;ft

<=]li

}<T-t"._

PRi

'-rE-O.6P

t=

PRi

2oE

+

O.4P

oc-v.tr

pt

tu

Cylindrical Shell

Longitudinal Joint

Circumferential Joinl

2i'l Ellipsoidal

Head

Ri

+

0.6t

^

2oEt

Ri

-

0.4t

^

2oEl

Oi

+ 0.2t

-

2oEt

R +

0.21

P=

oEt

0.885L

+ 0.1t

<

164s

2^tr1

LM + 0.2t

Sphere and Hemispherical

Head

ASME Flanged

and Dished

Head

when

UR

=

1 6?3

I

Ft

-./L-

\

#+\

\-__=-2,

F---

q--l

F.-t

When

UR

Conical Section

PDr

p

=

2oEt

cos

o

Di

+ l.2l

cos

d

cos

d(oE

-

0.6P)



Appendix

A:

Pressure

Vessel Formulations

Supptement

for

ASME

Formulations

185

't.

For a

cvlindrical

shell,

when

the wall

thickness

exceeds

one

half

the

inside

radius

or

P

>

0.385dE,

the

tormulas

in

ASME

Code

AoDendix

l-2

shall

be used.

For hemisoherical

heads

without

a straight

llange, the

effi-

ciencv

ot

the

head-to-shell

ioint

is to

be ussd

it

il

is

less than

lhe efficioncy

ot

the seams

in the

head.

For

elliDsoidal

heads,

whsre

ths mtio

ot the

maior

axis is

other than 2:1.

retsr

to

ASME Code

Appendix

1'4{c).

To use

the

fomulalions

lor

a

conical

seclion

in

the table,

the

halt

apex

anqle,

d, shall

not

exceed

30o.

ll d > 30o'

then a

soeci;l

analysis

is

required

per

ASME Code

Appendix

4.

1-5(e).

5. Foian

ASME

flangsd

and dished

haad

(torispherical

head)

when

Ur<

164r the

tollowing

values

ot

M

shall

be used:

'

The maximum allowed

ratio:

M=

1

L-r

=

D

When

L/r

>

162/3

(non-ASME Code

construction), the

values

ot

M may be

calculated by

'('.

Values

ot

Factor

M

Ul

M

Ur

M

1.00

1.00

7.00

1.41

1.25

1.03

7.50

1.44

1.50

'1.06

8.00

1.46

1.75

1.08

8.50

1.48

2.00

'1.10

9.@

1.50

2.25

1.13

9.s0

1.52

2.50

1

.15

10.0

1.54

2.75

1.17

10.5

1.56

3.00

1.18

'|

1.0

1.58

3.25

1.20

11.5

1.60

3.50

1.22

12.O

1.62

4.00

1.25

r3.0

't.65

4.50

1.28

14.0

1.69

5.00

1.31

15.0

1.72

5.50

1.34

16.0

1.75

6.00

1.36

16?s

1.77

6.50

1.39





A

standard is a

collection

of current

practices, past

ex-

periences,

and research knowledge.

Standards

that

are

developed by consensus

groups (e.g.,

ASTM,

ANSD,

trade associations

(e.9.,

AISC, ACI),

or

government

groups (e.g.,

HUD, CPSC) carry

more authority than

other

standards because

they reflect wider ranges

of ma-

terials.

The ANSI A58.1-1982

is a

collection

of information

that is

considered to be the

state-of-the-art in

the desien

of buildings and other

structures. Local

and

region-al

building codes

adopt

portions

of the

ANSI

srandard

for

their

own use. These local

and

regional

codes are

devel-

oped

to

represent the

needs

and interests

of their

respec-

tive areas and are written

in legal language

to

be

incor-

porated

into

state and

local laws.

Because these building

codes

are regional or local

in

scope,

they often do

not

include everything

in the

ANSI

standard, which is

na-

tional in

perspective.

For

this

reason,

one must be cer-

tain

that a local code

written

for one

area

is applicable to

the site being

considered.

The

ANSI

standard

does

not

have

as much authoritv

as

the

ASME vessel

codes.

and,

unfortunarely.

does not

have a referral

committee or

group

to

officially interpret

the

document. Therefore,

one must

rnake decisions

based

on

past

experience

and

accepted

methods of

de-

sign. The

ANSI

standard

(Paragraph

6.6,

p.

16) states

that in

determining the value

for

the

gust

response factor

a

rational

analysis

can be used.

A note

below

the

para-

graph

states

that

one such

procedure

for

determining the

gust

response

factor

is in the

standard's

appendix.

The

note

at the top

ofthe

appendix

(p.

52)

states

clearly

that it

is

not

a

part

of

the ANSI 458.1

miminum

design

stan-

dard. What

all

this

implies

is that

one may follow

the

guide

of

the ANSI

standard's appendix

or use

another ra-

tional

analysis,

which includes

another

wind

standard.

Thus, one

care

use another

standard

for design

purposes.

Appendix B

National Wnd Design

Standards

One

of the most

widely

accepted

international

standards

is the Australian

Standard 1170,

Part

2-1983,

SAA

Loading

Code

Part 2-Wind

Forces.

The Australian Standard

I 170 is

more applicable

to the

process

industries

because

in it

are

shape factors for

geometries

that are more

common in

that industry,

e.g.,

circular

shapes.

However,

before applying

the

shape

fac-

tors of the Australian

standard to

the

ANSI

or any other

national

standard, one must

be

very

careful to correctly

convert

the factors. This

is because

the codes have dif-

ferent

basis

upon which these factors

are determined,

and

a

direct application

of other

parameters

is not

possi

ble. This is

discussed

later after we

discuss

the basis

for

the various

standards.

CRITERIA FOR

DETERMINING

WIND

SPEED

Wind

is caused by

differential heating

of

air

masses by

the sun. These

masses

of air

at approximately

one mile

above

the

ground

circulate

air

around

their

centers of

pressure.

At

this altitude,

the

velocity

and direction

of

the wind is

almost entirely

determined

by

macro-scale

forces

caused by large

scale weather

systems.

Below

this

gradient

height,

the wind

is

modified

by

surface rough-

ness,

which

reduces its

velocity

and changes

its direction

and

turbulence. A

secondary

criterion,

except

for

ex-

treme

wind

conditions,

is

the

temperature

gradient,

which

affects the vertical

mobility

of

turbulent

eddies

and therefore influences

the surface velocitv

and

the era-

dient height.

Therefore.

the exact

nutur"

of

the suriace

wind

at any

point

depends,

first,

on

the

general

weather

situation, which

determines

the

gradient

wind

and the

temperature gradient,

and,

second, on

the surrounding

topography

and

ground

roughness

which,

together with

147



188 Mechanical

Design

of

process

Systems

the

temperature

gradient,

modify

the

gradient

wind

to

the

surface

wind.

_

Wind

motion

is

lurrher

complicated

by

rhe rorarion

o[

the earth. which

induces

additional

forces

that cause

the

alr movrng

across

the

earth's

surface

to be

subiected

to

a

force

at righr

angles

ro

the wind

velocity

vecior.

These

additional

forces

are

known

as

Coriolis

iorces.

Each

country

has

adopted

its

own

standard

for

measur_

ing

wind

velocity.

The

U.S.

National

Weather

Service

and

U.S. codes

use

the fastest-mile

wind

speed,

which

is

defined

as

the

arrerage

speed

ofone

mile

ofair

passing

an

anemometer.

Thus,

a

fastest-mite

wind

speed

of

120

mph

means

that

a

"mile"

of wind passed

the

anemometer

dur_

ing

a 30-second

period.

Other

nations,

namely

Australia

and

Great Britain.

use

the

two-second

gust

speed.

This

is

based

on

the worst

2-second

mean

as measured

bv

a cuo

anemometer.

The

mean gust

speeds

are recorded

over

a

period

of time

such

that

a

mean

recurrence

interval

is

de_

termined.

The

mean

recurrence

interval

is the

reciprocal

of

the

probability

of

exceeding

a

wind

speed

of a'given

magnltude

at a

particular

location

in

one

year.

The risk.

or

probability.

R.

thar

the

design

wind

speed

will

be

equaled

or

surpassed

at least

once

in

the

life

ofthe

tower

is

given

by the

expression

R:l-(l-P,)"

where

P"

:

annual probability

of

exceedance

(reciprocal

of

the

mean

recurrence

interval)

n

:

life

of

the tower

or

stack

The

risk

that

a

given wind

speed

of

specified

magni_

tude

will

be equaled

or

exceeded

increaies

with

the De-

riod

of time

that

the

tower

is exposed

to the wind.

Values

of risk

of

exceeding

design

wind

speed

for

a designated

annual probability

and

a

given

design life

ofthe

structure

are shown

in Table

B-1.

_

For

example.

if

rhe

design

wind

speed

for

a tower

is

based on

an annual probability

of

0.02

(mean

recurrence

interval

of

50

years)

and

the

projected

tower

life

is 25

years,

there is

a

0.40

probability

that

the design

wind

speed

will

be

exceeded

during

the

life

of the

structure.

The

United

States and

Australian wind

codes

use

rhe

50_

year

recurrence

interval.

The

instrument

for

measuring

the

wind

in

the

United

States,

Great

Britain,

and

Australia

is

the

cup-generator

anemometer

shown

in Figure

B-1.

This

device

is

oper_

ated

by

rhe wind

striking

rhe

cups,

which

drive

a

small

permanent

alternator.

The indicator,

which

incorporates

a rectifier,

is simply

a volrmeter

calibrated

in

miles

oer

hour.

[n

most

recent

cup-generator

models

the

generator

output

is

used

to activate

a

pen-chart

recorder

w-hich

oro_

vides

a record

of continuous

wind

speed.

WIND

SPEED

RELATIOIISHIPS

As

stated

previously,

another

method

can

be

substi_

tuted

for

the

appendix

in ANSI

A59.1.

What

this

means

is

that

another

code

could

be used

instead

of

the

appen_

dix.

To

do this

one

must be

careful

to

utilize

the

correct

conversion

factors

between

standards.

To

accomplish

this we

refer

to Figure

B-2.

For

a 100-mph

fastest

mile

wind

speed

in ANSI

A58.

I we wish

ro

determine

the

equivalent

fastest

mile

wind

speed

for

a

2-second gust

using

either

the

Australian

or British

code.

From

Fis-ure

B-2 we

read

from

the

ordinate

1.54

for

2

sec. Knoiins

that

one

mile

of

wind

moving

ar

100

mph

will

pass

thi

anemometer

in

36

sec,

we read

36

sec

on the

curve

and

arrive

at V,/V366

:

1.30.

Thus,

the

equivalent

fastest

mile

wind

speed

is

I I

54t

"

:

tffil

(100y

rnp6

=

118.4

mph

for

a 2-sec

gust.

For

I l0 mph,

the

values

becomes

V:

(l.l8x1l0)

mph

=

129.8mph

Table

B-1

Probability

of

Exceeding

Wind

Design

Speed

Pr

=

1-(1

-

PJ"

PA

0.

l0

0.05

0.01

0.00s

r

5

l0

0.100

0.410

0.651

0.0s0

0.226

0.401

0.010

0.049

0.096

0.005

0.025

0.049

15

25

50

0.794

0.928

0.995

0.537

0.723

0.923

0.140

0.222

0.395

0.072

0.rr8

0.222

100

0.999

0.994

o.634

o.394

Figure

B-1.

Cup generator

anemometer





190

Mechanical

Design

of Process

Systems

Table

B-2

Maior

U.S.

and Foreign Building

Codes

and Standards

Used in

Wind Design

Code

or Standard

Edition

Australian

Standard

I170,

Part

2-Wind

Forces

British

Code

of

Basic

Data

for Design

of

Buildinss

(cP3)

Wind

Loading

Handbook

(commentary

on CP3)

National

Building

Code

of

Canada

(NRCC

No.

17303)

The

Supplement

to

the

National

Buildins

Code of

Canada

(NRCC

17724)

ANSI

A58.1-

1982

Uniform Building

Code

Standard

Building

Code

Basic

Building

Code

Standards

Association

of

Australia

British

Standards

Institution

Building

Research

Establishment

National

Research

Council

of

Canada

National

Research

Council

of

Canada

American

National

Standards

Institute

International

Conference

of Building

Officials

Southern Building

Code

Congress

International

Building

Officials

and

Code

Administrators

International,

Inc.

Address

Standards

House

80 Arthur

Street/North

Sydnev.

N.S.W.

Australia

British

Standards

Institution

2 Park

Street

London,

WlA

285,

England

Building

Research

Station

Carston,

Watford,

WD2

7JR,

England

National

Research

Council

of

Canada

Ottawa,

Ontario

KIA

OR6

Canada

1430

Broadway

New

York,

New York

10018

5360

South Workman

Mill Road

Whittier,

California

9060

I

900 Montclair

Road

Birmingham,

Alabama

35213

17926

South

Halsted

Street

Homewood,

Illinois

60430

1983

t972

1974

1980

1980

t982

1982

1982

with

1983

rev.

1984

Table

B-3

Reference

Wind

Speed

Australian

British

Canadian

United

States

Beletence

Averaging

time

1

1

2-3

second

gust

speed

I18.4

2-second

gust

speed

1

18.4

Mean

hourly

76.9

1

Fastest

mile

100

I

Equivalent

reference

wind

speed

to fastest

mile

100 mph



Appendix B: National Wind

Design

Standards

Table

B-4

Parameters

Used

in the Maior National Standards

191

Parametel

Australian

Brltlsh

(sAA,

1983)

(BSr,

re72)

Canadian

(NRCC,

1980)

Unlted States

(ANS|,

1982)

Wind

Speed

Terrain

roughness

Local terrain

Height variation

Ref.

speed

Wind Pressure

Pressue coefficients

Gusts

Magnitude

Spatial correlation

Gust frequency

Analysis

procedure

2-sec

gusts

tbles in

appendix includes

figures

Gust speed

Reduction

for

large

area

Dynamic consideration

for

h/b

>

5

This

standard is consid-

ered

by many

the

best

for

us€

in

the

process

industries. Figures

and

tables

are easy to

read.

The

standard actually

provides

the user with

equatrons to curves.

The

analysis

procedure

is straight-forward.

z-sec

gusts

Tables, includes

figures

Gust

speed

None

Dynamic

consideration

not

included

Overall

a very

good

code,

its weakest

part

is the

lack of

dynamic

consideration.

3

None

Yes

Mean

hourly

Figures and

tables

in

commentaries

Gust effect factor

Gust effect factor

Dynamic

consideration

for h/b

>

4

in. or for

h>400ft

An excellent wind

standard. The

analysis

procedure

is

straight-forward

and the

docu-

ments-code

and

supplement con-

tain tables and

fig-

ures easy

to

read,

None

Yes

Fastest

mile

Tables,

figures

and

notes

Gust

response

factor

Area

averaging

Dynamic

consideration

for

h/b

>

5

Although

the appendix

is technically

not con-

sidered a

part

of

the

standard, it

contains

figures difhcult to read,

namely

Figure

6.

For

many structures the

data extend beyond the

limis of

the curves

in

Figures

6 and 7. In the

method

in the appendix,

one

must

assume

an

ini-

tial natural frequency,

resulting

in

an

iterative

process.

This

method is

extremely

difficult

in

designing

petrochemical

towers

without the

use

of

a computer.

4

Yes

Yes

Yes

Yes



192

Mechanical

Design

of

Process

Systems

Table

B-5

Limitations

of Codes

and

Standards

Code

or

Standard

statement

ot Limitation

Location

Australian

Standard

I170,

Part

2

1983

National

Buildinq

Code

of Canada

-

(NRCC,

r980)

British

CP3

"Minimum

Design

Loads

on Structures"

"...EssentiallyaSer

of Minimum

Regulations

.

.

."

".

. .

Does Nor

Apply

to

Buildings.

.

. Thdt'Are

of

Unusual

Shape

or Location

For Which

Special

Invesrisations

May

Be

Necessary

. .

."

-

"Minimum

Design

Loads

. . ."

"Specific

Guidelines

Are

Giyen

For.

.

. Wind

Tunnel Investisations

...

ForBuildinss..

.

Havin--s

Irregular

Shapei.

. ."

"The

purpose

.

. . is

to

provide

minimumstandards.._"

"The

Basic

Minimum

Wind

Speeds

Are Shown

in Figure

912.1 .

.

."

"The

Purpose

of This

Code

is to

Provide

Minimum

Requirements

.

.

.',

"The

Building

Official

May

Require

Evidence

to Support

the

Desisn

-

Pressures

Used-in

rhe Design-

of

Structures

Not Includedln

This

Section."

United

States

ANSI

A58.I

Uniform

Building

Code

Basic

Building

Code

(BOCA,

1984)

Standard

Building

Code,

1982

(SBCCI,

t982)

Title

Guide

to the

Use

of the

Code

Section

I

(Scope)

TitIE

Paragraph

6.1

Section

102

Section

912.1

Preface

Article

1205.2(a)





194

Mechanical

Design

of

process

Systcms

i

weight

ol

pipe

per

toor

(pounds)

weighl

ol

wcter

'€r

toor

(pour&)

squdr€

leet

outside

iurloce

per toot

:

Bqucre

leet ilside

surloce p€r

toot

=

inside

qrea

(squqre

inch*)

olea

of Inetdl

(squcte

hches)

momert

ol

inertid

(inch6s.)

PROPERTIES

OF

PIPE

*

The

tollowinq

lormulds

C're

used

in ihe

computotior

ol

the volues

lhown in

the

toble:

i tbo fsrridc

steels

rlay

b€

qbout

S%

les.,

@d tbo

dultesitic

stoh.

l6ss

ste€ls

dbout

2/o qred'ler

th@

the

values

lhown

in

this

tqbl€

which

dre

bdsed

o

weights

lor

carbon

steol.

r

schedul€

Du.Ebers

Stotdord

weigbt pipe

ond

schedule

40

dle

the

sqme

in dll

sires

througb

lo-inch;

Irom

l2,iach

through

24-iach,

stondqrd

weight

pipe

hcB

a

wdll thicble$

oI

%-inch.

Ertro

Btlong

eeight

pipe (r|td

sch€dule

gO

q 6

the

sdme

in

sll

siz6

lhrough

8-inchr

trom

8-irch

thlough

Z4-irch,

ert

ci

sttoag

weight

pipe

hds c

wdll

rhjcLdess

ot

%-irch.

Double

enrd

stloEg

weight pip€

bas

no

cor*ponding

scbedule

nu.Eb6r.

o:

ANSI 836.10

steel pipe

schedule

Dumb€rg

b:

ANSI

836.10

steel

pipe

rtoEinql

wdl

ihicla€ss

designctioD

e

ANSI835.19

stainless

sloel

piF,e

scbedule

du.Dclors

10.6802(D-r)

0.3{05d

0.2618D

0.2618d

0.78sd

0.78s{Dr-d)

0.049r(Dr-d.)

A^n;

sectio

boduluB

(inchest)

rodius

oI

glrotion

(illches)

=

0.0982(D.-d.)

D

=

o.zs

l

ozlp-

A,

=

dreo

of Estcrl

(Equa.e

nocles)

d

=

inside

dida€ter

(iach€6)

D

=

outsids

didnete

(bchos)

R,

=

lodiu

ol

gFotior

(irches)

t

: pip€

wdU

thicloess

(inchss)

DoEinol

piF

rize

oulside

diclmeter,

ll|"

Bchedul€

wcll

thick-

in.

inside

didm-

in-

io"ia.

|

-.tot

|

"q.tt..

l

I

ouardo

cleq,

ldred, | ,

. I _ |

auddc.€

3q.In.

|3cr'l|r" I

pertr

Bq It

il|3id€

sur{dc6,

per

lt

weight

Fr

It,

lbf

lr'6ight

ol

wcler

p€r

lt,

lb

|'roEeDt

ol

inertio.

a6clioE

Erodu.

lus,

lodiu

gYrd-

UorL

in

%

0.405

40

80

std

xs

l0s

40s

0.049

0.068

0.095

0.307

0.269

0.215

0.0740

0.0568

0.0364

0.0548

o.0720

0.0925

0.r06

0.106

0.r06

0.0s04

0.070s

0.0563

0.186

0.245

0.3ts

0.0321

0,0246

0.0157

0.00088

0,00106

0.00I22

0.00437

0,00525

0-00600

0.1271

0.1215

0.1146

%

0.540

;;

80

std

l0s

40s

80s

0.065

0.088

0.119

0.410

0.364

0.302

0.1320

0.1041

0.0716

0.0970

0.12s0

0.1s74

0.141

0.I4t

0.141

0.1073

0.0955

0.0794

0.330

0.425

0.535

0.0572

0.04s1

0.0310

0.002?9

0.00331

0.00378

0.01032

0.01230

0.0r395

0.1594

0.1628

0.1547

%

o.675 40

80

t;

xs

ss

l0s

{0s

80s

0.065

0.(E5

0.091

0.t26

0.710

0.54S

0.493

0.423

0.396

0.2933

0.19t0

0.1405

0.1582

0.1246

0.1670

0.2173

0.220

0.t77

o.t77

0.t77

0.1859

o.t427

0.1295

0.1r06

0.538

0,423

0.568

0.739

0.1716

0.t011

0.0827

0.0609

0.01197

0.00586

0.00730

0.00862

0.0285

0.01737

0.02160

0.02554

0.2150

0.2159

0.2090

0.l9sl

%

0.840

40

80

160

;;

XS

xxs

l0s

40s

80s

0.065

0.083

0.109

0.147

0.187

0r9{

0.710

o.6't4

0.622

0.546

0.466

o.2s2

0.3ss9

0,357

0.304

0.2340

0.1706

0.u99

0.1583

0.1974

0.2503

0.320

0.383

0.504

0.220

0.220

o.220

0.220

0.220

0.220

0.I859

0.1765

0.1628

0.1433

0.t220

0.0660

0.538

0.671

0.851

1.0€8

r,301

't.7t4

0.17t

0.rs47

0.1316

0.10I3

0.0710

0.0216

0.0120

0,01431

0.0ttl0

0.02010

0.022\3

0-t2125

0.0285

0.0341

0.0407

0.0478

o.0527

0.0577

o.2750

0.2692

0.2613

0.2505

o.2102

0.2t92

1.050

40

80

160

;;

xs

xxs

l0s

10s

80s

0.065

0.083

0.1l3

0.I54

0.2t8

0.308

0.920

0.884

0-s21

o.?42

0.614

0.434

0.655

0.6t4

0.533

0.432

0.2961

0.1479

0.2011

o.2521

0.333

0.435

0.570

0.718

o,275

0,273

o.275

0.275

o.275

o.275

0.2409

0.2314

0.2157

0.1943

0.1607

0.

37

0,684

0.857

l.l3t

t.414

1.937

2.441

o.2aa2

0.2661

0.2301

0.1875

0.1284

0.0541

0.02451

0.02970

0.0370

0,0448

0.0527

0.0579

0.0467

0.0566

0.0706

0,0853

0.1004

0.1t04

0.349

0.343

0.334

0.32r

0.30{

0.2810

I

.1.3r5

40

80

t60

;;;

xs

xxs

l0s

40s

80s

0.065

0.109

0.133

0.1?9

0.250

0.358

1.185

1.097

1.049

0.957

0.815

0.599

1.103

0.945

0.864

0.719

0.s22

0.2818

0.2553

0.113

0.4s4

0.639

0.836

1,076

0.344

0.344

0.344

0.3{{

0.3{{

0.341

0.310

0.2872

o.2716

0.2s20

0.2134

0.1570

0.858

1.404

1.679

2,t72

2.811

3.659

0.478

0.{09

0.374

0.311

0.2281

0.t221

0.0500

0,0757

0.0874

0.I056

o.1252

0.1405

0.0760

0.ll5l

0.1329

0.1605

0.1900

0.2137

0.443

0.42A

0.42t

0.407

0.387

0.361

r%

J.660

{0

80

160

;;

xs

xxs

55

r0s

40s

,::

0.065

0.109

0.t{0

0.250

0.382

1.530

t.442

r.380

1.27a

1.160

0.896

r.839

r.496

r.283

1.057

0.63r

0.326

U.53I

0.669

0.8b

I

1.107

1.534

0,434

0.434

0.434

0.434

0.134

0.434

0.401

0.378

0.361

0.335

0.304

0.2346

1.r07

1.805

2.273

2.997

5.2t4

o.797

0-7al

0.648

0.458

0.2732

0.1038

0.1605

0.1948

0.2418

0.2839

0.341

0.1250

0.1934

o.2346

0.2913

0.312

0.41I

0.564

0.550

0.540

0.524

0.506

0.472

r%

1.900

l0s

0.06s

0.t09

t.?70

1.682

2.461

0.37s

0.613

0.197

0.497

0.469

0,{40

t.274

2.085

1.067

0,962

0.I580

0.2469

0.1663

0.2599

0.649

0.63{

tCt,kne\) ,'f ITT

Ctinkll.



Appendix

C: Properties

of PiPe

195

PROPERTIES

OF

PIPE

(Continued)

noEitrol

prpe

rir.

outside

diomelet

ia.

rchedule

qumber'

wcll

thick-

646.

in.

inrid€

diqa-

iriide

3q.

in.

metol

rq.

rD.

eq lt

outsido

suatcce,

po.

ft

rq

It

itrlide

EurIqce,

pe.Il

rrreight

per

lt.

w€isht

ol wsler

p€r

It,

lb

oI

in€diq,

in..

a6ctioE

rnodu-

lus,

inJ

rodiue

gYrc-

tio|1.

q

b

1%

L90{)

40

80

160

srd

xi

xxs

.:

40s

8os

0.I45

0.200

0.281

0.400

0.52S

0.650

1.6r0

1.500

1.338

1.I00

0.850

0.600

2,036

1.767

r.406

0.950

0.567

0.283

0.799

r.058

I.{29

1.885

2.247

2.551

0.497

0.497

0.497

0.497

0.497

0.{97

o,42r

0.3s3

0.350

0.288

o.223

0.I57

2.7t8

3.631

1.859

6.40€

7.7tO

8.6?8

0.882

0.765

0.608

o.4tz

0.246

0.t23

0.310

0.39r

0.483

0.568

0.6140

0.6340

0.326

0.{12

0.508

0.598

0.6470

0.6670

0.623

0.50s

0.58I

0.549

0.5200

0.4980

2

2.375

;;

80

160

;;

xs

xxs

5'S

los

{0s

80s

0.06s

0.109

0.154

0.218

0.343

0,436

0.s62

0.687

2-245

2.157

2.081

I.939

1.689

l.5m

r.251

1.001

3.96

3.65

3.36

2.953

2.210

1.774

t,229

0.787

0.472

o.176

r.075

1.411

2.1s0

2.556

3.199

3.641

0-622

0,622

D-622

0.822

0.622

0.822

0.622

0,622

0.588

0.565

0.541

0.508

0.442

0.393

0.328

0.262

1.60d

2.638

3.653

5.O22

7.444

9.029

t0.882

12.385

r.715

1.582

1.455

1,280

0.971

0.76S

0.533

0.311

0.3rs

0.499

0.868

1.163

I.312

L.442

1.5130

o,z6s2

o,420

0,561

0.731

0.979

1.I01

1.2140

t2110

0.817

0.802

o.741

0.755

o.129

0.703

0.6710

0.6410

2tl

2.875

;;

80

160

.''.

";;

s

)o(s

l0s

40s

80s

0.083

0.120

0.203

0.274

0.3?5

0.552

0.675

0.800

2.r09

2.635

2.469

2.323

2.L25

t.771

1.525

t.275

4.19

4.24

3.55

2.184

1.826

1.276

o.724

1.039

r.704

2.945

4.03

4.663

s.2t2

0.753

0.?s3

0.753

0.753

0.753

0.753

0.753

0.?s3

0.709

0.6s0

0.646

0.608

0.556

0.464

0.3s9

0.334

2.175

3.531

5.793

7.661

I0.01

13.70

15.860

t1-729

2.499

2.361

2,016

1.837

1.535

1.067

0.792

0.554

0.710

0.988

1.530

1.925

2.872

3.0890

3.2250

0.4s4

0.68t

1.064

1.339

l.ss8

2.I4S0

2.2430

0.988

0.975

0,947

0.924

0.894

0.844

0.8140

0.7860

3.500

1;

80

160

;;;

xi;

10s

40s

80s

0.083

0.120

0.2I6

0.300

0.437

0.600

0.725

0.850

3.334

3.260

3.068

2.900

2.626

2.300

2.050

r.s00

8.73

8.35

7.39

6.6r

5,12

4.15

3.299

2,5,13

0.89r

1.2?4

2.224

3,02

1.2L

5.4t

6.317

7.O73

0.916

0.916

0.916

0.916

0.916

0.916

0.916

0.916

0.873

0.s53

0.803

0.75S

0.687

0.602

0.537

o.171

3.03

4.33

7.58

10.25

tl-32

18.58

zt-447

24.0s'l

3.78

3.6r

3.20

2.864

2314

1.801

1.431

1.103

1.301

LazZ

3.02

3.90

5,03

5.39

6.50r0

6.8530

o.111

r.041

t.724

2-226

2.476

3.43

3.7t50

3.9160

1.208

t.t96

1.154

1.136

1.094

1.0,17

1.0140

0.9810

3h

1,qn

40

80

i;

xs

xrs

10s

40s

80s

0.083

0.I20

0.226

0.318

0.636

3.834

3,760

3.548

3.364

2.72A

11.10

9.89

8.89

5,845

1.021

1.463

2.680

3.68

a,721

t.o47

1.047

1.047

t.o41

1.047

1.004

0.984

0.92S

0.881

0.7t6

3.41

4.91

9.r

12.51

22.850

5.01

4.81

4.28

3.8S

2.S30

1.960

2.756

4,19

6.28

s,8d80

0,980

1.378

2.394

3.14

4.9240

1.38s

L.312

1,337

1.307

1.2100

4'JU)

;;

80

120

r60

;;

xs

:o,s

IGS

40s

s0s

0.083

0.r20

0.188

0,237

0.337

o.437

0.500

0.531

0.674

0.800

0.925

4.334

4.260

4.L24

4.026

3,826

3.626

3.S00

3.138

3.152

2.900

2.650

14.7S

11.25

13.357

t2.73

It.50

r0.33

s.521

s.28

7.80

6.602

5.513

2.547

3.17

4,41

6.283

6.62

8.10

9.294

10.384

1.178

t.178

1.178

1.178

1.178

1.178

r.178

1,178

1.178

1.r78

t.178

l.ll5

1.082

r.054

1.002

0.94S

0.916

0.900

0.825

0.759

0.694

3.92

8.560

10.79

14.98

r8.96

21.360

21.54

31,613

35.318

6.40

5.800

5.51

4.98

4.48

4.160

4-O2

3.38

2.864

2.391

2.811

3.96

5.8500

123

11.65

t2.17tO

13.27

15.29

16.66t0

t7.?130

t.249

L.762

2.6000

3.21

4.27

5.6760

5.90

6,79

7.4050

7.8720

1.562

1.549

1.5250

t,510

t.177

1.445

1.1250

t.116

1.37{

1.3380

r.3060

5.563

;;

80

t20

160

;;;

xs

)o(s

r0s

4os

80s

0.109

0.134

0.258

0.375

0.500

0.62S

0.7s0

0.875

1.000

5.34S

5.29S

5.(X7

1.813

4.563

4.313

4.063

3.813

3.553

22.44

22.02

20.01

18.19

I6.35

14.6r

t2.97

rt.4l3

1.868

2,285

4.30

6.ll

7.95

9.70

I1.34

12.880

1t.328

1.456

t.456

1.4s6

1.456

1.456

1.456

1.455

l.{s6

1.156

1.399

1.386

r.321

1.260

1.t95

1.129

r.064

0.998

0.933

6.35

7.77

l{.62

20-74

27.O4

32.96

38.5S

43.8t0

17.7s1

9.73

t.89

7.(x)

6.33

s.62

{.951

4.232

6.95

8.43

15.17

20.68

25.74

30.0

36.6450

39.lll0

2.494

3.03

5.15

7.13

9.25

10.80

I2.10

13.1750

11.0610

1.920

1.878

1.839

1.799

1.760

1.6860

1.5s20



196

Mechanical

Design

of

Process

Systems

PROPERTIES

OF PIPE

(Continued)

pipe

Biz€

in.

schedule

wall

thick-

inside

diom-

rn.

inside

sq.

in.

tItetol

3q.

rL

aq lt

outside

pe

It

sq

ft

inBide

surrcc

per

lt

weighl

per It,

lbf

w€ighl

per

It,

tb

oI

inertia,

luB,

rddius

gyra-

tion,

in.

6

40

80

t20

160

sia

xs

xxs

l0s

40s

80s

0.109

0.134

0.219

0.280

0.432

0.562

0.7I8

0.864

L000

L

t25

6.407

6.357

6.187

5.761

5.50r

5.189

4.897

4.62S

4.37S

32.2

3t.7

30.r00

28.89

26.07

23-77

18.83

16.792

Is.02s

2.231

2.733

4.4I0

5.58

8.40

10.70

15.64

t7.662

19.429

1.734

1.734

t.734

t.734

I.734

1.734

1.734

1.734

t.734

t-734

r.677

1.664

1.620

1.588

1.508

L440

1.358

r.282

r.211

1.t45

5.37

9.29

15.020

18.97

28.57

36.39

45.30

s3.16

60.076

66.0S4

r3.98

t3.74

r3.100

12.51

It.29

I0.30

8.17

7.284

ll.8s

14.40

22.6600

28.\4

40.5

49.6

5S.0

66.3

72.r190

76.5970

3.58

4.35

6.8400

8.s0

t2.2s

14.98

r7.8I

20.03

21.7720

23.t240

2.304

2.295

2.2700

2.245

2.I95

2.153

2.r04

2.060

2.0200

1.s850

8

8.625

20

30

40

60

80

a;;

XS

I0s

4;;

80s

0.109

0.I48

0.219

0.250

0.27',|

0.322

0.406

0.s00

4.407

8.329

8.187

8.125

8.07r

7.991

7.813

7.625

s4.s

52.630

51.8

51.2

50.0

47.9

45.7

2.916

3.94

5.800

6.58

8.40

10.48

t2.78

2.258

2.258

2.258

2.258

2.258

2.25A

2.258

2.25A

2.2A1

2.180

2.150

2.t27

2.1t3

2.089

2.045

1.996

9,91

r3,40

19.640

22.36

24.70

28.55

35.64

43,39

24.07

23.59

22.500

22.48

22.t8

21.69

20,79

19.80

26.4S

35.4

sr.3200

63.4

88.8

r05.7

6.13

8.2I

ll.s000

I3.39

t4.6S

r6.81

20.58

24.52

3.0r

3.00

2.9700

2.962

2.953

2.938

2.S09

2.578

I

8.625

100

t20

l{0

160

0.593

0.718

0.8I2

0.906

1.000

7.439

7.18S

7.001

6.813

6.625

6.375

43.5

40.6

38.5

34.454

3L903

14.96

t7.44

19.93

2t.9?

23.942

26.494

2.25a

2.258

2.2s8

2-2s8

2.258

2.258

L948

1.882

L833

1.784

t.?34

r.669

50.87

60.63

74.69

81.437

90.1r4

18.84

17.60

r6.69

15.80

14.945

13.838

t21.4

140.6

1s3.8

177.t320

r90.62I0

28.t4

32.6

35.7

38.5

4r.0140

44.2020

2.847

2.807

2.117

2.7 4A

2.7I90

2.68I0

l0

)0.750

;;

30

40

60

80

100

120

140

I60

;,;

xs

l0s

40s

80s

0.134

0.t 65

0.219

0.250

0.307

0.365

0.500

0.593

0.718

0.843

0.87S

L000

t.125

1.2s0

1.500

t0.482

r0.420

r0-312

10.250

r0.r38

10.020

9.750

s.564

9.314

9.064

9.000

8.7S0

8.500

8.250

7.750

86.3

85.3

83.52

82.s

80.7

78.9

'14.7

7t.8

68.I

64.5

60.1

56.7

s3.45

47.r5

4.52

5.49

7.24

9.25

10.07

l l.9l

16. t0

t8.92

22.63

26.24

27.t4

30.6

34.0

37.3r

43.57

2.815

2.815

2.815

2.815

2.815

2.81s

2.815

2.815

2.815

2.815

2.815

2.815

2.815

2.815

2.8I5

2.744

2.728

2.70

2.683

2.654

2.623

2.5S3

2.504

2.438

2.373

2.36

2.:91

2.225

2.16

2.03

r8.70

24.63

28.04

34.24

40.48

54.74

64.33

76.93

89.20

92.28

104.13

t26.42

148.19

37.4

36.9

36,2

35.8

35.0

34.1

32.3

31.1

28.0

27.6

26.1

24.6

23.2

20.5

63.7

76.9

100.46

I13.7

137.S

160.8

2t2.0

244-9

248.2

324

333.46

368

399

428.t'I

478.59

I1.85

14.30

r8.69

21.I6

25.57

29.90

39.4

45.6

53.2

60.3

62.04

58.4

74.3

79.66

89.04

3.75

3.74

3.72

3.7r

3.69

3.53

3.60

3,56

3.52

3.50

3.47

3.43

3.39

3.31

t2

12.750

;i

30

40

;;

80

I00

t20

l{0

150

;;;

.-.

l0s

4;;

80s

0.156

0.180

0.2s0

0.330

0.375

0.406

0.500

0.562

0.687

0.7s0

0.843

0.87s

1.000

r,125

1.250

r.3t2

12.438

12.390

2.250

12.0S0

12.000

11.938

u.750

I1.626

11.376

1r.250

11.064

I1.000

10.750

10.500

10.250

10.126

rzt-4

r20.6

u7.9

ll4_8

I

t3.l

llt.9

108.{

106.2

r0t.6

99.40

96.t

95.00

90.8

86.6

82.50

80.5

7.11

9.84

r2.88

14.58

1s.74

19.24

2t-s2

26.04

.28.27

31,5

32.64

36.9

4l.l

45.16

41.1

3.34

3.34

3.34

3.34

3.34

3.34

3.34

3.34

3.34

3.34

3.34

3.34

3.34

3.34

3.34

3.34

3.24

3.21

3.17

3.14

3.08

3.04

2.978

2,94

2.897

2.88

2.8t4

2.749

2.68

2.651

20.99

24.20

33.38

43-77

49.S6

53.53

65.42

73.16

88.51

96.2

07.20

t0.9

25.49

39.68

53.6

4D.27

52.7

52.2

5r.l

19.1

4S.0

48.5

47.0

46.0

44.0

43.1

41.6

4I.l

39.3

3?.S

35.8

34.9

t22.2

I40.S

191.9

248.S

279-3

300

362

401

475

510.7

562

578,5

642

70r

755.5

781

19.20

22.t3

30.1

39.0

43.8

47.1

56.7

62.8

? 4.5

80.1

90.7

100.7

I09.9

118.5

122.8

4.45

4.44

4.42

4.39

4.38

4.37

4.33

4.31

4.27

4.25

4.22

4-21

4-t7

4.I3

4.09

4.01



'1

Appendix C: Properties

of

Pipe

197

PROPERTIES OF

PIPE

(Continued)

aoniaal

pipo

riz.

outtide

rchedule

woll

thicL-

11646,

iD.

iDsid€

diqn-

iD-

inside

3q.

h.

metal

sq.

it .

sq It

outside

rurlqce,

Frlt

sq lt

ingide

rurldce,

per

lL

weight

per

lL

lbt

woisht

trrr

fL

tb

trlo|ne|''t

ol

i|'ertiq,

iD..

aectiorr

modu-

luB,

in.t

rcdiu6

9Yra-

tioD.

i -

l{

t1.000

to

;;

;;

40

;;

80

100

120

140

180

l0s

0.1s6

0.188

0.2r0

0.219

0.2s0

0.281

0.312

0.344

0.375

0.437

0.469

0.500

0.ss3

0.625

0.750

0.937

1.093

1.2s0

1.406

13.688

13,624

13.580

r3.562

t3.s00

13.438

13.312

13.250

13.126

13.082

13,000

12.8I4

12.750

12.500

12,t28

It.8l4

I1.500

lI.l88

147.20

145.80

141.80

144.50

143.I

141.80

140.5

139.20

137.9

I35.3

1s4.00

t32-7

129.0

t27.7

t22.7

109,6

103.9

98.3

6,78

8.16

9.10

9.48

10.80

l2.tt

t3.42

14,76

16.05

18.62

19.94

24.94

26,26

31.2

44.3

50,1

55.6

3.67

3.67

3.67

3.67

3.67

3.67

3.57

3.67

3.67

3.67

3.67

3.6'r

3.58

3.57

3.55

3.53

3.52

3.50

3.48

3.4J

3.44

3.42

3.40

3.35

3.34

3.27

3.17

3.09

3.01

2.929

23.0

27.1

30.9

32.2

36,71

4t.2

45.68

s0.2

s1.57

63.37

67.8

72.09

84.91

89.28

106,13

130.73

150.67

t10,22

I89.12

63.1

62.8

62.1

6I.5

60.9

50.3

59.7

58.7

s8.0

57.5

55.9

53.2

s0.0

47.5

45.0

42.8

r62.6

194.6

2t8,2

225.1

285-2

3r4

34{.3

429

456.8

484

562

s8s

687

825

tr21

I0l7

27.8

30.9

t2.2

36.S

40.7

14.9

49.2

53.3

61.2

55.3

69.1

80.3

81.1

94.2

117.8

132.8

146.8

159.5

4.90

4.88

1.87

4.47

4.86

4.85

4.84

4.8s

4.82

1.80

1-79

4.18

4.14

4.73

4.69

4.63

1.58

4,53

4.18

16.0U)

i;

20

30

40

60

80

100

120

t40

t60

;;

xs

l0s

0.16s

0.188

0.250

0.312

0.37S

0.500

0.656

0.843

1.03r

1.218

1.437

1.593

IS.670

15.624

r5.500

1s.376

1s.250

15.000

14.688

14.314

13.938

13.564

13.126

12.814

I92.90

191.70

188.7

185.7

182.6

t76.7

169.4

160.9

1s2.6

144.5

t35.3

129.0

8.21

9.3{

t2.3?

15.38

18.4I

24.35

40.1

48.5

65,7

72.1

4.

ts

1.19

4.19

4.IS

4.I9

4.19

4.19

4.19

{.19

4.19

4.19

4.I9

4.10

4.09

4.06

4.03

3.99

3,93

3.85

3.75

3.65

3.55

3.44

28

32

42.05

52.36

62.58

42.71

10r.50

136.45

164.83

192.29

223.81

245.11

83.5

8S.0

81.8

80.s

79.1

73.4

89.7

66.1

58.5

25?

292

384

473

562

?32

933

ll57

1365

I?60

1894

36.5

48.0

59.2

?0.3

9t.s

114,6

170.6

194.5

220.0

236.1

s.60

5.59

5,{8

5.43

5.37

5.21

5.17

5.12

t8

18,000

;;

20

30

;;

80

r00

I20

140

r60

5S

l0s

0.r65

0.188

0.2s0

0.312

0.375

0.437

0.500

0.562

0,750

0.937

r.r56

1.375

1.562

1.781

17,670

t7.624

I7.500

u.376

17.250

17.126

17.00

16.876

16.500

16.126

r5.688

r5.250

r4.876

14.438

245.20

243.90

240.5

237,r

233.7

230.4

227.0

223.7

213.8

204.2

193.3

182.6

173.8

163.7

9.24

r0.52

13.9{

t7,34

20.76

24.11

21.49

30.8

40.6

s0,2

61.2

71.8

80.7

90.7

4.71

4.',1L

4.71

4.71

4.71

4.71

4.71

1.7r

4.71

4.7

r

4-71

4.7 |

4.7

4.63

4.61

4.58

4.55

4.52

4.48

{.45

4-42

4.32

4.22

4.ll

3.9S

3.89

3.78

36

41-39

59.03

70.59

82.06

93.15

r04.75

138.17

t70.75

207.96

244.14

274.23

308.5I

106.2

105.7

104.3

102.8

t01.2

99,9

98.{

97.0

92.7

88.S

83.7

79.2

75,3

7

r.0

368

4t7

5{9

678

807

93r

1053

rt72

1834

2180

2499

2',150

3020

40.8

46.4

61.0

75.S

89.6

103.4

117.0

130.2

168.3

203.8

242.2

z'17.6

306

335

6.31

5.30

6.28

6.25

8.23

6,21

6.19

6.10

6.01

s.97

5.90

5.84

5.77

20

20,000

l0

20

30

40

60

80

100

s;

l0s

0.188

0.218

0.250

0.375

0.500

0.s93

0.812

0.875

1.031

1.281

I9.634

19.564

r9.500

r9.250

t9.000

18.814

I8.376

18.2s0

17.938

17.438

302.40

300.60

298.6

291.0

283.5

278.0

265,2

261.6

252.7

238.8

I1.70

23.t2

30.6

36.2

48.9

52.6

61.4

s.21

5.24

s.24

s.24

5.24

5.24

5.24

5.24

5.24

5.24

5.14

5.12

5.ll

5.0{

4.97

4.93

4.8r

4.78

4.70

4,57

40

46

s2.19

78.60

104.I3

r22.91

I66.40

178.73

208.87

256.10

131.0

r30.2

129.5

126,0

t22.8

120.4

115.0

Ir3.4

109.4

103.{

574

663

7S?

1I l4

t457

1704

2257

2409

2772

3320

57.4

7S-7

lll.4

145.7

170.4

225.?

240.9

277.2

332

7.00

6.99

6.98

6.94

6.90

6.79



198

Mechanical

Design

of

Process

Svstems

PROPERTIES

OF PIPE

(Continued)

nominol

pip6

rire

schedule

wcll

lhick-

in.

inaide

dicm-

iD.

inside

sq

in.

metdl

aq rr'"

Bq

lt

oubide

gurlqce,

per

lt

sq lt

in8ide

surlcce,

perlt

lreight

per

Il,

tbt

n

eight

per

lt,

tb

lnoEent

oI

inerlid,

rection

Erodu-

lus,

rqdiur

9yra.

lion,

in.

20

20.ooo

r20

140

160

1.500

1.750

t.968

17.000

16.500

16.064

227.0

213.8

202.7

87.2

I00.3

lll.s

s.24

5.24

s.24

4.45

4.32

4.21

296.37

341.10

379.01

98.3

92.6

87.9

4220

4590

376

422

459

6.56

6.48

6.41

22

22.004

l0

20

30

;;

80

r00

120

140

160

;;;

xs

5S

I0s

0.188

0.218

0.250

0.37s

0.500

0.625

0.750

0.875

l.t2s

1.37s

r.875

2.t25

2r,624

21.564

2r.500

2t.250

21.000

20.750

20.s00

20.250

I9.750

19.2s0

18.7S0

r8.250

17.750

367.3

363.1

354.7

346.4

339.2

330.r

322.1

306.4

291.0

276.1

26t.6

247.4

12.88

14.92

17.18

25.48

33.77

41.97

50.07

58.07

13,7A

8S.09

104.02

118,55

132.68

5.76

5.76

5.?6

5.76

5.76

5.76

5.76

5.56

5.50

5.43

5.37

5,30

5,17

5.04

4.91

4.78

44

5l

87

lls

t43

1?0

I97

2Sl

303

351

403

45t

r59.t

1s8.2

157.4

153.7

150.2

146.6

143.t

132.8

t26.2

119.6

t07.2

756

885

l0l0

1490

1953

2400

2829

3245

40i29

4?58

5432

6054

69.7

80.4

91.8

135.4

t77.5

2t8-2

237

-2

295.0

366.3

432.6

493.8

550.3

602.4

't.71

?.70

7.69

7.65

7.61

7.52

7.47

7.39

7.31

7.23

7.t5

7.07

24.000

l0

20

30

io

;;

80

100

t20

140

150

XS

0.250

0.375

0.500

0.562

0.62s

0.687

0.750

0.218

0.8?5

0.968

L2l8

1.53t

1.812

2.062

2.343

23.500

23.250

23.000

22.816

22.750

22.626

22.500

2s.564

22.250

22.064

21.s64

20.938

20.376

1s.876

19.314

434

425

415

4

406

402

398

436.1

388.6

382

365

344

326

310

293

18.65

21.83

36.S

41.{

45.9

50.3

54.8

16.29

63.54

70.0

87.2

108.1

126.3

I42.1

159.4

6.28

6.25

6.28

6.28

6.28

6.28

6.2S

6.28

6.28

5.28

6.28

6.28

6-28

6.28

6.28

6.r5

6,09

6.O2

5.99

s.96

5.92

5.89

6.17

5.83

5.78

5.48

5.33

s.20

5.06

63.41

s4.62

125.49

140.80

156.03

t7t.I?

186.24

55

216

238.11

2S6.36

367.40

429,39

483.13

541.94

188.0

183.8

180.1

178.1

t76-2

174.3

172.4

r88.9

168.6

r55.8

158.3

149.3

141.4

t34.S

t27.0

1316

1943

2550

2840

3140

3420

37I0

I152

4256

4650

s670

6850

7830

8530

9460

109.6

16I.9

212.5

231-0

26t.4

285.2

309

96.0

354.7

388

t73

571

719

788

8.10

8.35

8.31

8.29

a.z7

8.25

8.22

8.41

8.18

8.15

8,07

7.95

7.47

7.79

7.10

28

26.000

t0

20

srd

0.2s0

0.3I2

0.37s

0.500

0.625

0.750

0.875

1.000

1.t25

2S.s00

25.376

25.250

2s.000

24.750

24.500

24.250

24.000

23.7s0

510.7

505.8

500.7

490.9

481.1

471-4

461.9

452.4

443.0

19.8S

25.18

30.19

40.06

49.82

59.49

69.07

78.54

87,91

6.8r

6.81

6.81

6.81

6.81

6.81

6.8I

6.81

6.68

6.64

6.54

6.48

6.41

6.35

6.28

6.22

88

I03

202

235

267

299

22t.4

2t9.2

217,1

2t2-8

20s.6

204-4

200.2

r96.1

ts2.t

1646

2076

2479

3259

4013

4744

5458

6149

6813

126.6

r59.7

190.6

250.7

308.7

364.9

419.S

473.0

524.1

s.l0

9.08

9.06

9.02

8.98

8.93

s.89

8,85

8.80

2A

28.000

l0

20

30

std

xs

0.250

0.3r2

0.375

0.500

0.625

0.750

0.875

r.000

1.r25

27.500

27.376

27.250

27.000

26.750

26.500

28.250

26.000

2s.750

594.0

588.6

583.2

572.6

562.0

s51.5

541.2

530.9

520.8

21.80

z',t.t4

32.54

43.20

53.75

64-21

74.s6

84.82

94.98

'1.33

7.33

7.G)

7.33

1.20

7.t7

?.13

7.07

7.00

6.34

6.87

6.74

71

92

lll

t17

183

2tg

253

288

323

2s1.3

2S5.0

252.6

248.0

243.4

238.9

234.4

230.0

225.6

2098

2601

3l0s

4085

5038

5964

6855

714D

8590

149.8

185.8

22t-A

23 1.8

359.8

426.0

490.3

6t3.6

9.81

9.79

9.77

9.68

9.61

9.60

s.55

9.51

30

30.000

l0

20

30

std

xs

t0s

0.250

0.3I2

0.375

0.500

0.62S

29.s00

29.376

29.250

29.000

28.750

683.4

477.8

672.O

660.5

649.2

23.37

29.19

34.90

46.34

57.68

7.85

7.85

7.85

7.8s

7.8s

7.72

7.69

7.66

7.59

7.53

79

99

119

158

96

296.3

293.7

291.2

286.2

281.3

258S

3201

3823

s033

6213

t72.3

2t3.4

254.8

335.5

4t4.2

10.52

10.50

10.18

I0.43

10.39



Appendix

C: Properties

of

Pipe

199

PROPERTIES

OF PIPE

(Continued)

nominol

pipe

size

oulside

didmeteL

schedule

wcll

thick-

inside

dicm-

irBide

sq.

in,

metal

Bq.

in,

sq It

outside

per

It

sq

It

inside

per

rl

weighl

pe

ft,

lbt

weight

per

It

lb

ol

ilrerlio.

lus,

in.3

(rdiug

gvrq-

iion,

30

30.000

40 0.750

0.875

I.000

l.l2s

28.500

28.250

28.000

27.',750

637.9

620.7

615.7

6D4.7

68.92

80.06

9t.Il

r02.05

7.85

7.85

7.85

7.85

7.46

7.39

7.33

'1.26

234

272

310

347

276.6

271.8

2E',t.O

242.2

137 |

84S4

9591

10653

491.4

566.2

639.4

7 t0.2

10.34

10.30

10.26

t0.22

32

32.000

l0

20

30

40

std

xs

0.250

0.312

0.375

0.500

0.625

0.688

0.750

0.87s

1.000

L l25

31.500

31.250

31.000

30.750

s0.624

30.500

30.250

30.000

29.?50

'179.2

773.2

766.9

7 54.7

7

42.5

736.6

730.5

718.3

706.8

694.7

24.93

3I.02

37.25

49.48

61.59

67.68

73.63

85.52

s7.38

109.0

8.38

8.38

8.38

8.38

8.38

8.38

8.38

8.38

8.38

8.38

8.2S

8.21

B.l8

8.l l

8.05

8.02

7.98

7.92

7.85

7.',19

85

106

t27

168

209

230

250

291

33t

371

337.8

335.2

332.5

321.2

321.9

319.0

316.7

306.4

301.3

3l4 t

38gl

4656

6140

7578

8298

8990

10372

I

I680

I3023

196.3

243.2

291.0

383.8

473.6

518.6

561.9

648.2

730.0

814.0

11.22

11.20

11.18

u.l4

I1.09

11.07

I1.05

lr.0l

l0.ss

10.92

34

34.A00

t0

20

30

40

st;

XS

0.250

0.312

0.375

0.500

0.62s

0.688

0.750

0.875

1.000

t.125

33.500

33.376

33.250

33.000

32.7s0

32.624

32.500

32.250

32.000

3t.750

881.2

874.9

867.8

s5s.3

841.9

835.S

829.3

816.4

804.2

791.3

26.50

32.99

39.61

52.82

65.53

72.00

78.34

91.01

I03.67

lI5.I3

8.90

8.90

8.90

8.S0

8.90

8.90

LS0

8.90

8.90

8.90

8.1',|

8.7 4

8.70

8.64

8.57

8.54

8.51

8.44

8.38

8.31

90

1r2

t79

223

245

266

310

353

395

382.0

379.3

370.8

365.0

3M.l

359.5

3S4.1

348.6

343.2

3173

4680

sssT

7385

9124

9992

1082s

12501

l4l t4

15719

22t.9

2',t5.3

329.2

434.4

535.7

587.8

637.0

735.4

830.2

924.7

IL33

I t.9I

11.89

r 1.s5

I1.80

I

I.78

11.76

tt.12

I1.67

I1.63

36

36.000

l0

20

30

40

XS

0.250

0.312

0.37s

0.500

0.625

0.750

0.875

I.000

1.125

35.500

35.376

3s.2s0

35.000

34.750

34.500

34.250

34.000

33.750

s89.7

982.S

s75.8

962.1

948.3

934.7

920.5

907.9

a94.2

28.11

34.95

42.D\

55.76

69.50

83.0I

96.s0

109.96

123.19

L42

9.42

L42

9.42

9.42

9.42

9.42

9.42

9.42

9.29

9.26

9.23

9.16

9.10

9.03

8.97

8.90

8.89

96

lIs

143

190

236

242

324

374

419

429.1

426.1

423.1

417.l

4lt.t

405.3

399.{

393.6

387.9

4491

5565

6654

8785

10872

12898

I4903

I6S5I

18763

24S.5

309.1

310.2

488.1

504.0

7I6.5

82',t.9

936.2

t042.4

t2.84

12.62

12.59

12.55

12.51

12.46

t2.42

I2.38

12.34

42

42.000

2i

30

40

std

XS

0.250

0.375

0.s00

0.62S

0.750

1.000

1.250

1.500

41.500

41.250

41.000

40.7s0

40.500

40.000

39.500

39.000

1352.6

1336.3

t320.2

1304.r

1288.2

1256.6

t225.3

1194.5

32.82

4S.08

65.18

81.28

97.23

128.81

160.03

190.85

10.99

10.99

t0.99

r0.99

I0.99

10.99

t0.99

10.99

10.86

10.80

10.73

10.67

10.60

10.47

10.34

10.21

tt2

I67

222

276

330

438

s44

649

586.4

s79.3

s't2.3

565,4

558.4

544.8

531.2

517.9

7 r28

I0627

I4037

17373

20689

210a0

33233

39181

339.3

506.r

668-4

427.3

985.2

r2s9.5

rs82.5

1865.7

14.?3

t4.7r

t4.67

14.62

14.59

14.50

14.41

14.33



200

Mechanical

Design

of

Process

Systems

INSWATION

WEIGHT FACTORS

To determine

the

rveight per

foot

of

any

piping

insulation, use the

pipe

size

and

nominal

insulation

thickness

to

find the insulation

l.eight

factor

F in

the

chart shorvn

belorv.

Then

multiply

fl by the

density

of the insulation

in

pounds

per

cubic

foot.

Example.

For

4"

pipe

rvith

4"

nominal

thickness

insulation,

f

:

.77.

Il the insulation

density

is

12 pounds

per cubic

foot, then

the

insulation

rveight

is .77

X

12

:

9.24lb/lr.

Nominal

Pipe Size

Nominal

Insulation

Thickness

1%"

2rt"

3%"

4%"

5t4"

I

1%

1%

2

.051

.066

.080

10

12

ll

l4

.21

.21

.23

.29

.29

.31

.30

.38

.40

.39

.48

.47 .59

214

3

3%

4

.09r

.10

.r9

.17

.24

.31

.30

.36

.34

,41

.39

.46

.44

.58

.56

.63

.70

.68

.78

.83

.81

.96

.s7

1.10

6

8

10

. 7

.24

.34

.43

.34

.38

.59

.45

.o.t

.66

.58

.64

.80

.93

.71

.83

t.12

.88

.97

1.17

1.04

1.13

1.36

1.54

1.20

1.34

1.99

t2

l4

16

18

.50

.6{

.68

.70

.78

.87

.88

.90

1.0r

l.t2

1.07

1.1I

1.24

1.37

1.34

1.49

1,64

1.52

1.57

1.92

1.74

1.81

2.01

r.s9

2.07

2.29

2.24

2.34

2.58

2.82

2.50

2.62

2.88

3.14

20

24

.70

.83

.96

1.13

1.44

1.50

1.77

t.7s

2.10

2.09

2.44

2.40

2.80

3.16

3.06

3.54

3.40

3.92

LOAD

CARRYING

CAPACITIES OF THREADED HOT

ROLLED

STEEL ROD

CONFORMING TO ASTM A-36

Nominal

Rod

Diameter, in.

%

lz

V+ %

1

.1ya,

ry4

1y4

2

2l+

2 2y4

3 3r/t

3

Root

Area of

Thread,

sq.

in.

.068 .126

.202 .302 .419

.693 .889 1.293 1.144 2.300 3.023

3.719 4.619 5.621 6.124

?.918

Max, Safe Load,

lbs.

at

Rod

Temp.

of 650'F

610 1130 1810

27t0

3770 4960 6230 8000 11630 15?00

20700 21200 33500 41580

50580 71280





202 Mechanical

Design

of Process

Systems

PIPE

r.660"

o.D.

Tempcrature

Renge

'F

Fiber-

Sodium

WEIGHTS

OF

PIPING

MATERIALS

Yn"

w'

4\

di

t_L_,

z

F

z

F

z

F

z

-J

tr

Nr$

ts-ts$

{l.-.-tis

Boldface

.ty"pe

is

\eight

in

pounos.

Lrghflace

t)pe

benexth

weight. is weight

factor

for

Insulation

thicknesses

and

weights arc

based

on averaqe

mnditiors

and

do Dot

constituie

a recommendation

for

specific

thicknesses of

materials-

Insula-

tion lveights are

based

on.85/p

magnesra

ano

nl drous

c3lclum

silicate

at

11

lbs/cubic

foot. The

listed thicknesses and

neights

of

combination

coverinq are

ihe

sums of

ihe

inner

layer of dia-

tomaceous

earth

&t 2l

lbs/cubic

foot and

the

outer

laycr

at

1l lbs/cubic foot.

Insulation

weiqhts include al-

lowances

for wiri,

cement,

can-

vas,

bands and

paint,

but

not

special surface fi nishes.

To find the

weieht

of coverine

on

flanges, vatvds

or fittings]

multiply

the weight factor by

the

\aeight

per

foot of

covering used

on

straight

pipe.

Valve

rveiqhts

are loproxi-

mate. When

possible,

-dbtain

lreights from

the

manuf&cturer.

Cast iron

valve weiqhts

arc

for

flanged.end

valves;

stiel

weights

lor weldrng

eno

valves.

All flanged

fitting, flanged

valve

and

flange weights include

the

proportionrl

weight

of

bolts

or

studs

to

makc up

all

joints,

,41

/A

#

,N

Jrtd

@

@

IrtJ

@

FsO

Nom.

Thick.,In.

*

16

lb

cu.

ft,

density.



.IVEIGHTS

OF

PIPING

X{ATERIALS

Appendix

C:

Properties

of Pipe

203

r.eoo'o.D.

l/2"

erce

Schedule No.

40

80 160

Wall

De,<igna.tion

std.

NS

xxs

lhickness-In.

.145

.200 .281

.400

Pipe-Lbs/Ft

2.72

3.63 4.86

6.41

lVatcr-Lbs/Ft

.88 .77

.61 .41

{,1

t2

nuj

>f\

i t />

LLP

e tij

i

-1/

4,

q---

1_

-0

dti

L.R.

90"

Elbow

.8 1.1

1.4

I.E

S.R. 90' Elbow

.6

.3

.7

.3

L.R.

45"

Elbow

_5

.2

.2

.E

.2

1

.2

Tee

.6

2.5

.6

3.L

.6

3.7

.6

Latera.l

1.3

5.4

Reducer

.6

,2

.7

.2

.9

.2

.2

c"p

.3

.5

.3

.7

.3

.7

Temper&ture

Range

'F

t00-199 200,29e

300,3c0

.100-.199

;00-it)9

000-0119 ;00-;,1,1

s00-sf)1r

1t00- r 9

11000-1099

1100-L:00

Nlaqnesia

Caliium

:

Siili.crp

\om. Thick., In. 1

I \)t 2 2

214

tl

i

3

Lbs,/Ft

.84

.84 1.35

2.52

3.47 4.52

4.s2 4.52

{

Combina-

z

\om. Thick.,In.

2tt 21

;

2)1

3

3 3

Lbs/Ft,

1.t0

1.20

1.20 5.62

5.62

5.62

Fiber-

Sodium

Nom.

Thick.,

In.

1

1%

1ta

2

2\l 2%

3 3

LbslFt

1.07

1.07

1.01 1.8s r.85 3.50

3.50 4.76

4-16

6.16

,MS

A rtr

za|

lg

tsrj_ri}

{rrTs

PressLrre

Ratiig

ps'

Casl

lron

blecl

Roldf.rcc tlpe is weight

in

pounds.

Lightfi.ce

tl

pc

bene&th

rveight is

rveight

iactor

lor

insulation.

Insul&tion thickncsses rnd

*eights

^te

based

on

:rverage

conditions

and

do

not constitutc

r rocommcnd&tion for spocilic

thicknesses

of

m"rtorial-q.

Insula-

tion

Neishts :rre

bstxl

on

85f6

mrgnesia ud

hrrlrous lrrlcium

l -.. , ,,,,1,i

^

f^^r Tl-

125

250 i

j;0

300

400

000

900 r500

2500

Screled

or

SIip-On 1.5

7

1.5

1.5 1.5

9

1.5

9

l9

1.:)

l9

1.5

\Yelrling

\eck

L5

I \2

1.5

l2

1.5

l9

1.5

l9

1.5

34

1.5

Lap Joini

1.5

8

9

1.5

9 t9

1.5

19

31

Rlind

3.5

1.5

7

I

5

3.5

t.5

9

1.5

10

1.5

l0

1.5

l9

1.5

19

3l

..4

a

/:)

Z

tt 4\

-

?41

| /A

3,\

S.R. 90"

nlbow

10

3.7

t2 23

3.8

26

3.9

46

listcd

lhiclinesses

orxl

\\'cights

of

combinltion

covering rte

the

sums

of the inner l.rver of dir-

tomaceous errth at

21

lbs .ubic

foot anrl the outcr

hl cr

at

11 ltls/cubic foot.

Insuhtion

\\'ci,ahts

inrluclc

cl-

louanr:rs

for \\'iro, ccmcnt. ern-

vlt'\, brnds

llnd

l)rint,

but not

st'ccirlsrrrf,,rc

ti

n

rs)'cs.

Tu lin,l

tlLe

\, iHl,t

.f,1,v,

ring

on

flugcs, vrlvos or

fittings,

multiplt

thc

rveight f.|rtor

l)y

thc

rvcight

lrcr

fooi

of

covcrir)g

uscd

or) strLright

pipe.

\'.rlvt} \ 0iJahts lrre appro\i-

mcte.

\\'hcn

lrossiblc,

obtrin

rveights

f|om

the munuf:rcturer.

(iust

iron

vrlvc \ eights:Lro for

lhnged cnrl vxlves:

stecl

$eighls

for

\eldins

end

vrlves.

,\ll

firLneed

fittins,

flrnjaed

vrlvc

ond

1|Lngc *cights

includc

iho

I)r'otxJrtional

\ 1'ighi,

of

bolts

or studs to make ur)

.lL

ioints.

L.R. 90' Elbow

4

45'Elbow

9

ll

23

39

Tee

t7

5.6

20

30

5.8

70

6

j=<l

s

k3J

lltn

++I

FrO

Ilanged

lJonnet

G:rt{

6.8

1.2

70

.l.il

125

Flanged

tsonnet,

GLrlrc

or

Angle

40

4.2

45

.t.2

t70

5

Irlanged

Bonnet

Clheck

30

4.1

35

.1.1

40

I l0

I)tessure

Seal

Borrrret-(-irte

42

1.9

t.2

Pressurc

Seal

Ilonnet

Giobe

*

16

h

cu.

ft.

density-

joints.



2O4

Mechanical

Design of Process

Systems

2"

ptpn

z',s,,

o.D-

wErcHTS

oF

pIprNG

MATERTALS

A

Schedule

No.

40

80

160

Wall

Designation

std.

XS

xxs

Thickness-In.

.154

.218

.343

.436

Pipc-Lbs/tr

1,

5.02

7.41 9.03

I4'ater-Lbs/Ft

1.46

1.2E

q

t -/

zf.

F t/>

w

'HJ

^

_l--__,

\i/

L.R. 90"

Elbow

1.5

.5

.5

2.9

.5

S.R. 90'

Elborv

1

1.3

.3

L.R.

45' Elbow

.E

.2

r.1

.2

1.6

1.8

Tee

.6

.6

.6 .6

Lateral

5

1.4

7.8

1.4

Reducer

1.2

.3

1.6

1.9

crp

.5

1.2

,+

t,2

.+

Temperaiure

Range

"F

100-199

200-299

300-399 400-499

500-5s9 600-699

700-7c9

800-899

900-9s9 1000-1099

1r00-1200

Megnesia

z

Calcium

I silicate

Nom.

Thick.,In.

I

I

L% 2%

2%

3

3

3

3%

Lbs/Ft

1.01

1.01

t.7l 2.53

2.53

3.48

3.48

4.42

4,42

4.42

*

uomDlnx-

;

tion

z

Nom,

Thick.,

In.

2%

2%

3

3

3%

Lbs/Fb

4.28 4-2E

5,93

5.93

7.80

Fiber-

Sodium

Silicate

Nom.

Thick.,In.

I

I I

1%

1% 2

2

214

3

3

Lbs/Fb

1.26

1.26 1.26

2.20

2.20

4.57

4.57

5.99

5.99

sffi

d-ir

Z

trLrlS

6N_l-M

ryi:-s

Pre-ssure

Rating

psl

Cast Iron

Steel

Boldface

type

is

weieht

in

pounds.

Ligh[flce

type

bineath

weigii.

is

yreight

factor

for

lnsul&llon.

fnsulotion

thicknesses

and

weights

a,re based

on average

COnOrtlons ancl

do

not

constitute

a

recommendation

for

specific

thicknesses of materials. I-nsula-

tion weights

are

based

on.85/,

magnes,a

anct

nydrous

c&lctum

silicate

st 11

lbs/cubic

foot.

The

listed

thicknesses and weiqhts

of

combination

coverins arl

the

sums of the

inner

Iajer

of dia-

tomaceous

eerth st 21 lbs/cubic

foot

and

the

outet layer

at

l1

los/cuorc

loo .

Insulation

weishts include al-

lowances

for

wird,

cement,

can-

vas, b&nds and

paint,

but

not

special

surface finishes.

To find

the

weisht

of coverins

on

flanqes.

valvds

or

fittincs]

muhipltth weisht factor

by tle

wergnt.per

too

ol coverrng usecl

on srrargn

prpe.

V&lve

weishts

are aooroxi-

mete.

When

possible,

-dbtain

weights

from

the

rnanuf&cturer.

C&st

ircn

valve weiqhts

are

lor

flanged,end

valves;

sGel weights

IOr

Welolng

eno

valves.

All flanged

fitting, flanged

valve

and

flange

weighls include

the

proportional

weight

of

bolts

or 6tuds to

make

uD

s.ll

ioints.

250

150

300 400

600 900 1500

2500

Scre* ed

or

SIip-On

9 6

9

ll ll

32 32

4E

'|1'elding

Neck

10

13

t3

3l

1.5

3l

{E

Lap Joint

9 12

1.5

4E

Blirrd

6

10 4-8

l0

t2

1.5

3l

3l

49

,h

,-{l

2t4xJ

i rlt

E,N

e

/9S

z

t?.4

E

II'

Y

ll_______.rl

S.R.

90'

Elbow

16

3.8

3.8

19

3.8 3.8

35

83

4,2

L.R. 90'

Elbow

1E

27

4.r

22

4.1

3l

4.1

45"

lllbow

14

3.4

l6

3.4

73

3.9

1'ee

23 37 41

6

129

ru

",1.{l

3m

+<f

rc

I'langed

Bonnei

Gat€

6.9 7.1

40

4

EO

4.5

190

5

Flanged

Bonnet

Globe or Angle

30

7

64

30

3.8

45

4 4.5

235

Flanged

Bonnet

Check

26

7

5t

3.8

40 60

4.2

300

5.8

Pressure SeaI

Bonnet-Cste

150

Pressure

Seal

Bonnet-Clobe

165

3

o make

up all

joints.

'

16

lt

cu.

ft.

density.



A

. -f

w

fl\

F-:l

-2t"

---i

/-\

-L-t

(--r..}

\.u

z

F

z

J

'

WEIGHTS OF

PIPING

MATERIALS

Appendix

C:

Properties

of Pipe

2o5

2.875'o.D. 2/2"

Ywn

Boldface

type

is seight in

pounds.

Lightfece type

beneai,h

\r'eight is

weight factor

for

insulation.

Insulation thicknesses

and

weights

are

besed

on

everage

conditions and do

not

constitute

a

recommendatioD

for specific

thicknesses

of

materials-

Insula-

tion

weights are

based on

85/6

magnesia

and hydrous

cclcium

silicate

at l1 lbs/cubic foot.

The

listed thicknesses and rveights

of

combination covering lrre the

sums

of the inner laver of

dia-

tomaceous

earth

at

2i

lbs,'cubic

foot and the outer l:r|cr

at

11 lbs/cubic foot.

Insulation

weights

include

al-

lowances

for

wirc,

cemcnt,

can-

vrs,

bends

rnd

print,

but,

not

special surftce linishes.

To

find

the

rveight

of covering

on

flnnges,

valves

or

fittings,

multipiy the \reight

factor

by the

weight

per

foot of

covering

used

on

straiqht

DiDe.

Valve

*eiftrts

are approxi-

mate- When

possible,

obtain

weights fronr

the

manufrcturer.

Oast

iron

valve weiehts ere

for

flanged

end

valves;

stiel

weights

for

*elding

end

valves.

AII flanged

fitting,

flenged

valve

and Iiange

\\eights

include

the

proportionel

iveight of

bolts

or

studs

to

rnake

up all

joints.

z

I

)

z

Temperature Range

'F

Magnesb,

Calcium

Combina-

tion

Fiber-

Sodium

,ffi

9+

i

${lit$

N-ls$

N

/A)

,4"1

,N

g 4

l-{

@

flt'

)

+€

|<IJ

()

z

I

z

.t

*

16

lb

cu.

ft.

density.



rt?

uf

/\

{_0

{l}

L:-I

-{\

fl-\

ri\

{----fr

\iJ

8

z

F

F

z

B

206

Mechanical Design

of Process

Systems

3

tt

"tpr

B.boo" o.D.

WEIGIITS

OF

I'IPING

NIATERIALS

l cnrpentLurc Rcngc

"F

Magnesia

Calcium

(--oDlbi

r-

tron

Fiber-

Sodium

z

F

z

z

z

a

-

ffi

${rn$

Njs

qN

Boldface type

is

weight

in

pounds.

Lightface type beneath

weight is weight Jactot Jor

insuLation.

Insulation ihicknesses and

weights

are based

on

average

conditions and

do not

constitute

a recommendation

for specific

thicknesses

of materials. Insula-

tion

$eights are

based

on 85/p

magnesia

and

hydrous

calcium

silicate

at ll

lbs/cubic foot.

The

listed thicknesses

and weights

of

cornbinetion

covering

are

the

sums

of the inner layer of

dia-

iomrceous

eerth

at 21

lbs/cubic

foot

and the

outer

la]

e. at

11

lbslcubic foot.

Insul{rtion Ncights include

al-

lorvarrces for \\'irc, cenrent, can-

vas,.bands-

and

prlitrt,

but not

speclsL

suf

tace

hnrshes,

-

To

iind the

ueight

of

covering

on

flanges, vs,lves or fittings,

multinl\'

the

weishtfactor

bY

Lhe

weighi

irer

foot 6f

covering'used

on

straight

pipe.

Yalve

weiehts are

aDDroxi-

mete.

Wben-

possible,

-dbtain

weights

from

the ma,nufacturer.

Cs.st

iron

valve weights are

for

flanged end valves; steel

weights

for

rveldinq

end

valves.

All

flanged

fitting,

flanged

valve

and llanee

weiqhts include

the Drooortion;l

weriht

of bolts

or

siudi to

meke

u[

all

joints.

/A

-11

,N

/9N

49 S

t<t

@

0

J{

Fs3

Nom.

Thick.,

In.

*

16

lb

cu. ft.

deDsity.





208

Mechanical Design

of Process Sl

stems

4"

ptpn

4.boo'

o.D.

WEIGHTS

OF PIPING MATERIALS

'li,mtx'nrluro

I

rngo

"I

trlagnesia

Calcium

ComLirur-

iioIl

Iiber-

Sodium

z

k

o

7

F

z

1

/a)

tu

&?

h

:

,t

{l\

tr;:I

tr:JI

/\

\JJ

z

z

3

NrS

{Nj+ln}

N_ts

rx:w

Boldface type

is

weight in

pounds.

Lightface tvpe

bene&th

rveight is

\reight

fsctor Jor

insulation.

Insulation

thicknesses lnd

weights are

based on

average

conditions

and do

not conslitutc

a recommendation

for specific

thicknesses of

mgterials.

Insula-

tion weights

are

based

on

8596

magnesia

and

hydrous

calcium

silicate

&t 11

lbs/cubic foot.

The

Iisted thicknesses

and

\reigllts

of

combinstion covering are

the

sums

of the

inner layer of

dia-

tomaceous

earth at

21 lbs/cubic

foot and the

outer ieter'at

ll

Ibs/cubic fooi,

Insulation

weights

includc

al-

lowances

fo wire,

cement,

can-

vas, bands

and

paint,

but

not

speciel

surlace

6nishes.

-

To find

the

weighl

of

cover;ng

on

flanges,

velves

or

fittings,

multiply

the

weight

frcior

by the

seight

per

foot

of

covcring

uscLl

on

str{righi

pipe.

Vrlve weights are

approri-

mate.

When

possible,

obtoin

$eights

from

thc

mxnufacturer.

Cast

iron

valve Ncights are

for

flanged end

valves

i

steel $eights

for

rveldinq

end

valves.

All flanged fitting,

flrnged

valve

and

flangc wcights

includc

the

proporbionxl

\\eiglrt

of

bolts

or

studs

to mrke

up

all

joinbs.

,.'Nl

/ ,l)

,41

N

/\

F<3

@

fi\

+<l

F<U

\\'stcr-Lhs/l

t

\om.

'l'hick.,

In.

Nom. T)rick.,In.

I

16 li cu.

ft.

density.



Appendix

C: Properties

of Pipe

209

WEIGIITS OF PIPING

MATERIALS

5.56:J"

O.D.

5"

pge

on t5%

rot.

The

righrs oi

of die-

rl er

Di

.lurie

rl-

hut

noi

co\,enng

)r

DJ

rnc

lng

used

approxi-

\reights

flanged

r include

of

bolts

ll

ioints.

Schedule

No.

40

80

120 160

Wall

Designation std.

XS

xxs

Thickness-In.

.258 .500 .62'r

Pipe- Lbs/ Ft, t4.52

20.78

27.M 32.96 38.55

Water-Lbs/Ft

8.66

7

.89 7.09

6.

J3

5.62

ul

,a

g.I/

zf\

F li

E4\

o

f'+

3 4/4-

LJ---.D

{---J--r

\tJ

L.R.,90" Elbow

14.7

1.3

21

1.3

S.R.

C0"

Elbow

9.8

.8

13.7

.8

L.R.

45'

Elborv

7.3

.5

r

0.2

.5

15.6

.5

t7

.7

.5

Tee

t9.E

1.2

26

1.2

39

1.2

43

Laterel

3l

2.5

50

Reducer

6

.4

E.3

.1 .4

t4.2

cop

.7

.7

ll

.7

1l

.7

Tcmpereture ll.enge

"F

00-199 200-20s 300-399 400+s9 500-599 000-699

700,;9c 800-Ec3

900-999

1000-1009

1100,1200

Fiber-

Z Sodium

9

Silicate

Nom.

Thick.,In.

1 1%

2

2ra

2%

3r/t, 3rl 4

4

Lbs/Ft 1.86 2.92 2.92 4.08 6.90 8.41

E.41 10.4

10.4

F-

B

tion

Nom. Thick., In.

2ra

3

3tl

3ti

4

4

Lbs/Ft

7.01 9.30

I1.8 I1.8

14.9 14.9

85%

Iagnesia

Calcium

Nom.

Thick.,In.

I

I

|

1,t;

11,.i

2% 2%

3 3

4

4

Lbs/Ft

2.34 2.34 2.34

3.76

9.31

9.31 14.31

14.37

,ffi

O

-'r-

i

sli19

N-l,Ns

El:::lr$F

Pressure

Rctiltg

psr

Casi,

Stecl

Boldf&ce type is

$eight

in

pounds.

Lightf.rce type

benerth

rvcight

is rveight factor ior

insulrtion.

Insuiation

thicknesses

and

\reights.rre

brsed on

rverage

condirions and do

not

constitute

3 rocommendrtion

for

specihc

thicknesses

of materials.

lnsuh-

tion

wcights

rre

brsed

on t5%

mrgnc-.ia

antl hydrous crlcium

silicatc at

11lbs,/cubic

foot.

The

listed

thickncsses

3nd rreighls oi

combination

covering

lrre

lhe

-cums

of the

inner later

of

dir-

tomrceous

errih

rt

2I

lbs cubic

foot

end the outer

irl er

Dt

ll

lbs/cubic foot.

Insulation

\\0ighis inclurie

rl-

lorvances

for

\\'ire,

cement, crn-

vns,

bends

xnd

p.rint,

hut not

soecial

surfxco

linishes.

1o

hnd

tlrc

$Lrglrt

ol

coverlne

on

llenges, lll\'fs or

littings,

muitiplt thc

\reight

fsctor bl

th.

rveight

per

foot o[

covering

used

on,str.riqht

t,ifo.

\' ive \\'.rqh is rrc

x||ro\l-

mrtc.

\\-hen

possiblc,

oblarn

weiglrts fronr the mllnufrcLurer-

C.rst

iron

vrlve

rvcishis

l|re fol

fluged end

vrlves;steel

\reightl

for

$cLdirrq

end

vlrlvcs.

,\ ll

fianee(l

fittiDs,

flanged

vol\c .rn.l ILrngc

wprgl,ts includ€

tl,c

t,rol,ortionxl

\eight

of bolt

or

siuds to make

up all

joints

125 250

i50

300 400

600 s00

1500

2|rc/..)

Screu

ed

or

Slip-On

20

1.5

32

1.5

l8

1.5

l

5

1.5

73

1.5

100

1.5

162

1.5

259

1.5

'|r,\'elLling

Ncck

22

1.5

49

713

1.5

103

162 293

1.5

Lap

Joini

18

1.5

32 7l

98 168

1.5

Blind I.5

37

t.5 l 5

39

1.5

50

1.5

78

104

1.lr

172

1.5 1.5

0

/a

F ,an

|

/,$

3,\

z

Et\

E

II'

Y

S.R.

90"

Elbo*

58 94 80

t l3

4.3

t23

205

268

4.8

435

5.2

L.R. 90'

Elbow

68 105

9l

t2a

45" Elbow

51

3.3

E3

3.8

66

3.8

98

t23

4

130

350

Tee

90 t45

6.5

ll9

0.4

t72

6.4

179

6.8

304

7

415 665

1-{

-

FdJ

JiLII

;hJ

+<i

rc

Flanged

Bonnct

Cet,e

138 264

7.9

150

4.3

4.9

3r0

455

5.5

615

6

1340

7

Flanged Bonnet

Globe

or

Angle

138 )47

8

ls5

2t5

5

5.2

515 950

6

Flanged Bonnet

Check

llE

7.6

210

E

110

4.3

165

5

1E5

5

350 560

6

1150

7

Pressure

Scal

Bonnet-Cate

350

3.1

520

3.8

865

4.5

Pressure Seal

Bonnet {ilobc

280

4

450

4.5

'

16 lb

cu.

ft. density.

up



2'10

Mechanical

Design

of

Process Systems

6"

pr""

6.625,

o.D.

WEIGHTS

OF PIPING

X{TTERIALS

Tempcraturc

Ilange

"F

Magnesia

Calcium

Combins-

t)on

Fiber*

Sodium

u/

AX

w

{T\

LilI

t---1

\JJ

z

'.

z

F

sq-,$

#r|&

N-S

dISrsS

-Xl

t#

rA

kL

,N

/>

lt'

'{

l-dl

@

ru

1-<i

rc

z

t

z

z

z

z

.|

Boldface

type

is weight in

oounds.

Liehtiace

trpe benea,th

iveight. is

-

tleight

'

iactor

for

Insulation

thichnesses

and

weights are

based

on average

conditions and do

not

constitute

a

recommendation

for

specific

thicknesses

of

materials.

Insula-

tion

weights are

based

on

85%

masnesia

and hvd.ous

calcium

siliate at

11

lbs/cubic foot.

The

listed

thicknesses

and weights

of

combinstion

covering &re the

sums

of the inner

layer of

dia-

tomaceous

es,rth

at

21

lbs/cubic

foot

and

the

outer layer

at

l1

lbs/cubic

foot.

Insulation

$eights include

aI-

Iowrnces

for

\aire,

cement, can-

vas,

bands

and

paint,

but not

special surface

finishes.

To

find

the

weight of covering

on

flanges, valves or

fiLtings,

multiplt; the

weight fxctor

bl the

rveight

per

foot of

covering used

on

straight

pipe.

Valve ueights xre

sppro\i-

mete.

When

possible, obtrin

weights

from ihe

mrnuf&cturer-

Clst

iron

valve

ueights

are

for

flenged end

valvcs;

steel

weights

for

rveidinq end valves.

All flanged litting,

flanged

valvc

3nd

nlnge

Ncrgnts

Incluoe

tLe

DrotJortional

$cieht of

bolts

ot

stud"

to

mrke

up

all

joints.

\\'eter-Ils/Irt

*

16

lb

cu. ft. density.



,qR

Appendix C: Properties of

Pipe

211

WEIGHTS

oF

PIPING

MATERIALS

8.625.

o.D.

8''

"T"e

2

i.

z

B

2

F

-

z

2

F

t'-

r_ j

w

{t}

E:I

,4\"

A

--l-r

\tJ

Temperature

Range

'F

Magnesia

Calcium

Combina,-

tron

Fiber-

Sodium

Boldfnce

tlpe

js

Neiaht in

pounds.

Lighilirce

tvpe bineeth

Neight.

is veight Jactor

lor

Insulation

thicknesscs

cnd

\reights

are based

on average

conditions and do not

constitute

a recommendation

for specific

thicknesses

of materiols. Insula-

tion

rveights

are

based

on

85%

magnesio

and

hJ'drous

calcium

silicate

at 11lbs/cubic foot. The

Iisted

thicknesses

aod

$'eights

of

combinetion covering

are

the

sums of the inner layer of dia-

tomaceous

earth

at 21

lbs/cubic

foot

and

the

outer la]'er at

11

lbs/cubic

foot.

,

Insulation

rveights

include al-

lowances lor

wDe,

cemenl,

c&n-

vas,

bands and

paint,

but

noi

soecial surface finishes.

'

-

To find

the weight

of covering

on

flanges, valves

or

frttings,

multiply

the weight f&ctor by

the

Neight.per folt

of

covering used

on

slrarghl

prpe.

Yalve rveights

are approxi-

mcte. lYhen

possible,

obtcin

lleights from

thc

manufrcturer.

Cast

ilon

valve

weiehts

are

for

flanged

end valves; sGel

\\'eights

Ior

seldinq

end

valvcs.

AII flcneed

fitting,

flanged

valvc

and llangc

rveights

include

tlrc

nroDortioDrl

\eiqht

of bolts

or stu,li

to make ut all

joints.

z

a

7

d

ffi$

ffi

$\

is

d

<,fs$

A

/A

A

gilq

j.43

t4

r\

+<i

FsO

Nom.

Thick.,In.

|

16

lb

cu.

ft.

density.



212 Mechanical

Design of

Process

Systems

10"

prpn

lo.zbo,

o.D.

IVDIGIITS OF

PIPING tr{ATDRIALS

lrmpcr:rturc

lirnge'F

Magnesia

Calaium

Combina-

iion

Fiber-

Sodium

(,

z

(,

z

z

F

P

z

z

.

z

Ih

fl\

L:J

.4'4^

L:-

-l_,

\]J

ffi$

qFl

rr$

N-|s

ryrTqJr

Boldfece t{pe is l\'eight in

pounds.

Lightface

t1'pe

benerth

*eight

is rveight

foctor

ior

insulation.

Insulation thicknesscs and

rveights are based

on average

conditions and do not constit,ute

a

recommcndetion

for specific

thicknesscs

of

materials.

Insula-

tion weights are

based

on 85/o

magnesie and hl drous crlcium

silicate

at

1l lbs/cubic foot. The

listed

thicknesses and weights

of

combination covering are the

sums

of the inner layer of dia-

tomaceous

earth at 2I lbs/cubic

foot and

ihe outer

lsyer

at

11

ibs/cubic

foot.

Insr-rlation Neights

include

al-

lowances

for vire,

cemeni, can-

vas, bands and

B.int,

but not

spacirl surfrce

6nishes.

To find

the

weight

of covering

on ffanges, valves or fittings,

multiplt' the

$eight frctor b

tLe

lieight

t'er

foot of

covering used

on

streight

pipe.

\'rlve \rcights ere approri-

matc.

\Yhen

possiblc,

ol)irirr

rr

ciglrts

from

thc

nrnnufrcturcr.

(lxst

iron

vrlYc

\\'ciglrts

arc

for

lllngcrl

cnd

vrlrcs:

stcoi

teights

fol

lcldilg cnd

vrlves.

-\)l

fl.rngcd fitting,

flnngcd

'rlvc : nd

l]3nge \\'eights include

tlru

prolroriioDxl scislrt

of

l)olts

or

studs

to mrkc

up:rlL

joints.

Ai

/AJ

,-11

,N

/>

lN'

'{

tHt

@

ff1

+<i

f<o

\om.

Thick.,

ln.

\Yelding

Neck

*

16

lb

cu.

ft.

derxity.



WEIGHTS

OF

PIPING

MATERIAI,S

Appendix

C: Propenies

oi PiPe

213

rz.lso'o.D.

12"

Pwr:

30

Schedulc

)io.

|

20

40

60

80

100

120

1{0

Wall

Designation

sid.

XS

Thickness-In.

|

.250

.330

.375

.406

.500

.562

.687

.843

1.000

l-J

Pipe-Lbs/Ft

33.3E

43.8

49.6

53.5

65.4

EE.5

t07

.2

r25.5

139.7

t58

3

Wsier-Lbs/Ft

5l .10

49.7

49.0

48.5

47

.O

46.0

44.0

4

r.6 39.3

{,)

IJJ

f4

.

(_ -f

2n^

F

flIT

Eji1

o

t-

-:

-t

i />"

3tr-

d_l\

L.R. 90'

Elbow

119

3

t51

3

3

S.R.

90"

Elbow

80

2

104

2

L.R.

45" Elbow

60

7E

181

;

1.o

l

Tee

r32

2.5

167

360

2.5

Lateral

1E0

5.4

273

Reducer

33

,|

44

'|

9,1

cap

30

3E

E9

1.5

Temperature

Range

"F

100-199

200-299

300-399

400-49S

500-5s9

600-699

700-799

800-E99

900-999

1000-1099

1100-1:to

Nom.

Thick.,

In.

1)4

114

2

2t/4

3

3

3rlt

4

4

4%

z Calcium

9 silicate

,-

s|

Combina-

;

iron

z

Lbs/Ft

6.04

6.04

8.13

10.5

t2.7

12,1

15.r

17.9

17.9

20.4

20.4

Nom.

Thick.,In.

3

3%

4

4

414

1)i

Lbs/Ft

17.7

21.9

26.7

26.7

31.1

3l,

r

Fiber-

Sodium

Nom.

Thick.,In.

t%

1rt

1r/1,

l1/r,

2\/r,

216

4

4

5

c

Lbs/Ft

14.20

14.20

24.&

4.64

32.&

32.40

,ffi

;+

z #rils

'$$js

{N

Pressure

Rating

psr

Cast

ffi

125

250

150

300

400

600

900

1500

|

2500

Screwed

or

Slip-On

71

1.5

r37

|

72

|

r44

1.5 |

1.5

|

1.5

164

1.5

261

r.o

3EE

820

1611

l.c

Welding

Neck

88

I

163

1.5 I 1.5

212

1.5

434

1.5

843

1.5

1919

1.5

Lap

Joint

72

|

164

1.5

|

1.5

ta7

1.5

286

1.5

433

902

1.5

1573

t.5

Blind

96

177

lllE

r.5 |

1.5

209 261

1.5

341

475 92E

1775

I.D

a4

a

lAl

,

/..4

A,N

I /}

z@

E IP

'{

S.R.

90'

Elbow

265

5

453

5.2

345

5

509

5.2

669

El5

5.8

t474

6.2

L.R.

90' Elbow

6.2

6.2

485

6.2

624

6.2

159E

6.2

45' Elbow

235

4.3

383

2E2

414

4.3

469

4.5

705

4.7

1124

4.8

Tee

403

6E4

7.8

5r3

754

7.4

943

8.3

1361

a.7

r92E

s.3

Flanged

Bonnet

Gate

6E7

7.8

1298

8.5

635

4

1015

5

1420

5.5

215s

7

2770

7.2

4650

8

Fhnqed

Bonnet

Globi

or

Angle

808

9.4

1200

9.5

7t0

5

1410

Flanged

Bounet

Checlc

674

9..1

ll60

9.5

560

720

1410

7.2

2600

8

3370

8

Pressure

SeeI

Bonnet-Gate

1975

2560

6

45r5

7

Pressure

SeaI

Bonnet-Globe

*

16

lb

cu,

fi.

densrty.



214

Mechanical

Design of Process System:

14"

,trr.

14'o.D.

WEIGHTS

OF

PIPING

]IATERIALS

Tcmpcrature Range

'F

Alagnesia

Calcium

Conlbina-

tion

z

|.

z

t

z

F

t

z

1

z

z

{.f

t

/)

fl\

fJJ

-t

c---r---l

\L"J

ffi

S{r-rM

N]s

{N

Boldlace

l\'pc

is

Ncight

in

pounds.

Lightface

tl

pc

l)eneath

*eight

is lYcight

lactor

for

insulation.

Insulation

thicknesses

and

$eights are based

on

lverage

conditions

and

do

not constitutc

a recommendation

for spccific

thicknesses of

rnaterials. Insula-

tion $eights are ba-sed

on E5%

magnesia

and hvdrous

cak.ium

silicate

at 11lbs/cubic

fool. The

listed thicknesses and lreights

of

combination covering Lire

the

sums of

the inner l&\'er

of

dia-

tomaceous

e:irlh at

21

lbs/'cubic

foot and the outer la] er at

11

lbs/cubic foot.

Insulation

\reights include

al-

lorvances for lvire,

cement,

can-

vas, bands and

ptint,

but

not

special su

ace finishes.

To find

the

leight

of covering

on flanges,

valves

or

fittings,

multiplt the

weight

fcclor

b]'the

MeiAht

pcr

foot of covering used

on

strnight

pipe.

Valve

s

eights are

spnro\i-

mate. When

possible,

obtain

weights

from

the

mrnufscturer.

Csst

ilon

velve Neights are

for

flanged end valves:

steel

weights

for

rveldine

end

valves.

All flaneed

fitting,

flanged

valve cnd

flonge

$eights

include

the

nroDortiorrrl

\\'cigl,t

of

holts

or sludi

to

mrkc

up all

joints,

/.4

--ll

,4

,N

i>

0,

.{l

ru

@

0

+<i

FSO

Nom. Thick.,In.

Nom.

Thick.,In.

*

16

lb

cu.

ft.

density



il

tl

Appendix C: Prop€rties of

Pipe

215

WEIGHTS

OF

PIPING

MATERIALS

re"

o.o.

16t'

plpu

Boldfxce

tvDe

is

rveielrt in

pounds.

Lighifirc

tt

pe

benesth

teight

is rveight factor

for

insulation.

Insulrtiod

thicknesses

and

weiqhts

are

bascd

on averase

conditions

and do

not constituie

&

recommend&tiol

for

spccific

thicknesses

of

materials-

Irrsuh-

tion

weights

ere

bosed

on

85%

magnesir

and hydaous

cnlcium

silicate

&t ll lbs/cubic foot. The

listed

thicknesses

&nd \yeights oi

combiortion

covering

are

the

sums oI the

inner layer

oI

dia-

tomaceous

earth

at 2l

lbs/cubic

foot

and

the

outer layer at

rr rDs/cuDLc

ioot.

,

Instrlati<.rn

weights

irclude

al-

low&nces

Io $alrc, cement,

ca

vas,

bands

and

pcint,

but Dot,

specilll

surlace

fi nishes.

To find

the

weight of

coverbg

on

flanges,

v&Ives or

fittings,

multiply

the weight

frctor

by

the

r

eight

per

foot of covering used

on

str&ight

pipe.

Valve Neights are

approxi-

matc. When

possible,

obtrin

weights

from

the

m.nuf&ciurer.

Cllst iron v.rlvc \reights:rre

for

flanged end valves:

steel

$eigh6

Ior

rvelding

end

valves.

All

flcnged fitting,

flanged

vclve

and

flangc wcights include

the

prot)ortionul

Neighi of

lrclr,s

or studs to

make

up 3ll

ioinis.

S$

stjjs

$$l.M

qr\ssF

A

.A

A

B

@

i[I

@

t4

v

z

z

;

z

F

z

z

z

z

1

t

A.

lJj

i\

w

{T\

1-5:I

J,1

E=_:ir

t .+r

fl\

\iJ

Temperature

Ra.nge

'F

I\Iagnesia

Calcium

Combina-

tion

'ih.r-

Sodium

l 100-1200

Flenged

Bonnet

*

16

lb

cu.

ft.

density.




of

Process

Sy:,rems

18"

plpr

18"

o.D.

WEIGI{TS

OF PIPING

MATERIALS

'fcnpcnturc

ll

Dgc

'I,'

Magnesia

Calcium

Combin.r.-

tion

Fiber-

Sodium

z

z

F

z

E

z

t-

Z

F

z

LLl

f^

'4r

fl\

H'

UL,

c.=-=I

IA

\JJ

ffi

ffi

Nl$

si)\r'|\s

Boltlface

tvne

is lcicht in

pounds.

Lig)rifrce

t5

pe

b-enerth

reiglrt. is

\cjght fschor

for

Instrlation

thicknesses

aod

rvciglrts

flrc

l,rsr:d

on r,vcrrge

conditions ltnrl do

not

(oustituta

a

r-ccommcndrtion

for specific

thicknesscs

of matcricls. Insula-

tion

\reights

ore bascd on

85/o

magncsia and

h-Ydrous calcium

silicrte

at 11

lbs/cubic foot.

The

listcd thickncsses

and rveights

of

combination

coveljng

are the

sums

oI the

inncl hver

of

dia-

tomaceous

eorth

at 21

lbs/cubic

foot

and

the

outet laver at

11

lbs/cubic foot.

Insulation s'cights include al-

Iolanr:cs for \rirc,

cemcnt,

con-

ves,

b:rnds

and

print,

but

not

spccial

sur'Iace finishcs.

To

find

ihc

\lcight

of covering

on

flanges,

valvcs

oa fittings,

multit)l]'the

xe;ght

factor by the

\eight

pcr

foot of

covering

used

on

stroight

pipe,

Vrlvc

\rriqhts

rre aptrroxi-

mate.

\l'hen possil,le,

-dt,tain

lscights

from

the m$nufacturer.

Cast

iron valve \yciqhts

are

for

flanged

end

velves;

st-eel \\eights

Ior

welding end

valves.

All

flanged

fitting,

flanged

valve

and

flange

scights

include

thc

proDortionrl

\\ci(lrt

of

bolts

or

si,udi

to

meke

up

all

joints.

/'a

IA

rA

,N

/$

4444

@

iln

+<t

rc

a

Il,s

/

Iit

\om.

Thir,k.,

In.

\om. Thitk.,

In.

*

16

lb

cu.

ft.

deDsity.



&Jj

Ih

\-.1-_t

{l\

r'-:

-'t

F4

, ^

*J----

z

F

i:

z

Appendix

C: Propen:*

:: l-

21

1VEIGIITS

OF

]'IPING

}I.\TDRI,\LS

20,,o.D

20" e-,rz

l{t)

3;9

Ll

'17

9

4l.:

'Icmpcraturc

Renge

"F

z

o

F

z

I{agnesia

Calcium

Combina-

tion

Fiber-

Sodium

2a.l

{-1-

1

Roldfrrce

tYpe

is

r..r:

: .:

poun(ls.

Lighthcc

rir)f I

:..,:

\l{ttglrL.

ls \\etglll

Iri-:

: :::

Illsulrtion

thi|krrts...

::

-

vc;ghts

uc

brsc(l

0r ,,. :.

::

corrditiols

urrtl iIr

rror

,,.:.-:.:::-

r

rccommrr{lxti(,n

a,)r

.--

l

tlti<

kncsscs

of mritli,.l:

I:

--.--

tiorr

rveiehts

rLn'

1,,,.t

i :.

:i

I

nNgncsil

rLnd lrr,ir

ru.

-:--

sili(rtc

lri 11 lLs

r

ui,:. :

-

-

listc(l

tLi( knciscs

,t:.

i

,

::

.

, ::-

conrl)in$tion

co\' f:r::

.. ::

sums

of

t))r inncr

-.:.,:

: ,-

tolnxceous

rtLrtlr

:,i

l: .: i

-

:

fooL

oniL

tl)c a';:.:

-

. :

-:

ll

Ibs

r:ulric

fooi

IusulLtion

r,

r::.:.

:: --

loNrurccs

ior r|ir,.

vrLs,

blnrls:i'l:,1

:.:.:

-

:

:

sp(

(

lrLL

:Llr

1t1..

:.:

:,

.

.

\lrgllt

l)ff

iL_'r:

.: '

I

: _.

-

onstfrLigi,:r:

f.

\_rtlvc Li,:::.:.

.:.

-

nlrtr'. \\

'.1:

..

\fi{)its

ir,r:r :].- r..

,

-

(,Lst

ir,:r'. ....

:

--

.- :.-

43.r

fl:LrLgcrl

i r:

i

iot $r:Lli:-ir::- .. :

..

.\ll

:l:,r..r.

:

:

::

:

_' :

vrh-c

rrri i

::.,::r':

r.

::r

-.

tlrc

prorl:l:,:.1-

.::

:

::

or studi i1r ::r:i:

.: ; ,.-

.:

1-1.03

z

g

J

ffi

sm$

N+s

gr(\i.x$

z

F

(,

z

/A

/41

/,+

A\

/>

€4 4

@

fln

J-<f

rc

Pip€'-Lbs./I,t

\\'at.r Lbs/

I,t

300-3c3

100+cc

i00-;9u

1000-6e0

\om.

Thick.,In.

Pressure

Rnting

|

(last.Ir('n

psr

ll25 l2s

l'langed

tsonnet

Globe

or Anglc

*

16

lb

cu. ft.

deDsity,




of

Process

Systems

24"

prpr.

24"

o.D. \T UIGI I1'S

OF

I'IPI\G

}IATEITIALS

Icnrpcrlturc llongc

'F

Magnesia

Calciun

Combinc-

tion

Fiber-

Sodium

Z

F

z

e

z

::

z

ui

f><

w

{T\

trJ-t

-/A

J]\

___-l____-

z

F

p

z

z

ffi

qN

trs

Njs

EN,fr\l

Boldfsce

troe

is weicht

in

pounds.

Liehifl.ce

tvDe

b;neath

ireight.

is

-

rreight'iactor

for

Insulation

thicknesses

and

$'eights

are based

on averaqe

conditions

and do

not, constitule

a

recommendation

lor

specific

thicknesses of

materials.

lnsuh-

tion ucights are

bused

on.85/e

m3gnesla ano nyorous

cstclum

silicate

at ll

lbs/cubic foot.

The

listed

thicknesses

and

lr'eiqhts

of

combinotion covering

arl

the

sums of the inner layer

of

dia-

tomaceous

earth

at 2l lbs/cubic

foot and the

outer

lsyer

at

ll

lbs,/cubic

foot.

d

,N

/D

tt,

.rl

IH

\Y.ltcr-Lbs/It

*

16

lb

cu.

ft.

density.

l=<[J

@

e

++J

rc



{:

Appendix

C:

Properties

of Pipe

219

WEI(IHTS OF PIPING

MATERIALS

za"

o.o

26t'

prps

Llj

/\

Iit

{1\

E--I

J'\

-:I

-I_'

\"J

z

F

F

a

z

t

Temperature

Range

'F

Ilagnesia

z

Calcium

o

brUcate

F

3

combina-

3

tion

3;m::;-

Fiber-

Sodium

z

z

F

ffi

s.{-n$

N-is

fFq.s

Boldface

type is

weight in

pounds.

LiEhtface

type

be-

neath weight is weight

factor

for insulation.

Insulation

thicknesses

and

weights

are

based

on

average

conditions and do

not consti-

tute a recommendation

for

specifrc

thicknesses

of

mat€-

rials. Iosulation

weishts

ate

based

on

85% magndsia

and

hvdrous calcium silicate

at

11

l6s/cubic foot.The

listed

thick-

nesses and weights of

combi-

nation

covering

are the

sums

of the

inner layer of diatoma-

ceous

earth

at

21

lbs/cubic

foot

and the outer layer

at

11 lbs/cubic

foot.

Insulation weights

include

allowances

fo wire. cement,

canvas, bands and

iaint,

but

not special surface ffnishes.

To-find the weiqht of cover-

ing on flanges,

v-alves

or

fit-

tings, multiply

ihe

weight fac-

tor

by

the weight

per.foot of

covetlngused

on siralghl

plpe.

Valve weights are

approxi-

mate.

When

possible,

obtain

weights from manufacturer,

Cast ilon valYe

weights are

for

flansed end

valve€i

steel

weishts Ior weldineend

valves.

A'il flane€d fitting,

flanged

valve and flange weiRhts

in-

clude

the

propo-rtionaf

weight

of

bolts

or

studs to make up

all

joints.

| ,41

AI

/r+

,N

&"f

n'

l

u:-Ji

t<t

@

fi)

+<i

rc

*

16 Ib cu.

ft. deDsitt'.



220

Mechanical Design

of Process

Syslems

28"

prpn

28-

o.D.

WEIGHTS

OF PIPING MATERIALS

Tempelature

Range

"F

nlagnesia

Calcium

Combina-

tion

Fiber-

Sodium

F

B

z

F

F

z

W

/4

{.J-f

Ih

t-+J

{1}

trJ:I

\IJ

ffi$

ffi

ds]-s

iN

,-a

A

tr'

.{

B---Jl

t=<3

@

0

+<i

rc

Boldface type is weight in

pounds.

Lightface

type

be-

neath weight is weight

factor

for insulation.

Insulation

thicknesses

and

weights

are

based

on average

conditions and do not co[sti-

tute a recommendation for

sDecific thicknesses of mate-

rials. Insulation

weights

are

based on 857,

magnesia and

lrydrous

cjrlciuJn silicat€.at 11

lDs/cuorc

root.

I ne lrsteo

[nlck-

nesses

and

weights of combi-

nation covering are

the

sums

of the inner laver of

diatoma-

ceous earth

ai

21

lbs/cubic

foot

and

the

outer layer at

11 lbs/cubic

foot,

Insulation

weights

in€lude

allowances for

wire,

cement,

canvas, bands and

paint,

but

not special surface

finishes,

To find

the weight of

cover-

ing on

flanges, valves

or fft-

tings, multiply

the

weight

fac-

tor

by

the weight

per

foot of

covering

usedon

straight

pipe.

Valve weights are approxi-

mat€. When

possible,

obtain

\reights from manufacturer.

Cast

iron valve weights are

for

flanged end

vslves;

steel

weishts forweldinsend

valves.

A-ll flanged fftting, flanged

valve and flahge weights in-

clude

the proportional weight

of

bolts

or

studs

to

make up

all

joints.

*

16

lb

cu.

ft.

derBity.



F

z

'

t

z

z

u-f

Ih

fl\

E-I

4',q

E::l

L--r-----U

\L/

Appendix

C: Properties

of

Pipe

221

WEIGHTS

0I'

PIPIN(}

MATERIALS

Bo"

o.D.

30"

prpe

Boldface type

is weight

in

pounds.

Lightface

type be-

neath weight

is

weight factor

for insulation.

Insulation

thicknesses

and

weights ale

based

on average

conditions and do

not consti-

tute a

recommendation

for

specilic thicknesses

of

mate-

rials.

Insulation

weights are

based on

859t magnesia

and

hydrous caicium

silicate

at 1l

lbs/cubic foot. The

listed thick-

nesses

and

weights

of combi_

nation covering are

the

sums

of

the

inner layer

of diatoma-

ceous earth at 21

lbs/cubic

foot

and

the

outer

layer

at

11

lbs/cubic foot.

Insulation

rveights include

allorvances

for

wire,

cement,

canvas, banCs

and

paint,

but

not sDecial

surface

finishes.

To_lind the u'eight

of cover-

ing on

flanges,

valves

or

fit-

tinss. multiDl\.the weieht fac-

toibl

the rieight

per-foot of

covering

used on straight

piPe.

Valve weights are approxF

mate.

When

possible,

obtxin

weiqhts

from

manufacturer.

Cist

iron valve

weights

are

for

ffanged end

valves; steei

weights lor

weldingend

valves.

All flanged

fitting,

flanged

valve and flange

weights in-

clude

the

proportional

weight

of

bolts

or

studs

to

make

up

all

joints.

Ilagnesia

Oalcium

Fiber

SodiLtm

tlon

Temperature

Range

'F

G

@

CD+

ffi

E lr-'$

Nls

CI-]-\}

*

16

lb

cu.

ft. density.



222

Mechanical

Design

of

Process

Systems

32"

prcn

sz,

o.D.

WEIGHTS

OF PIPING

MATERIALS

Temperature

Range

.F

Magnesia

Calcium

z

Silicate

{

uomDlna-

5

llon

Fiber-

Sodium

tu?

tg

f\

l_p

{T\

LJJ-

4',4

{-r-,

lr-f-r

\L/

z

7

,@$

3*

3

euls

Boldface

type is

weight in

pounds.

Lightface type be-

neath

weight is weight

factor

for

insulation,

Insulation thicknesses

and

weights are

based

on

average

conditions and do

not

consti-

tute a lecommendation

for

speciflc

thicknesses of mate-

rials. Insulation

weights

are

based on 857.

magnesia

and

hydrous calcium silicat€

at 11

lbs/cubic foot.The listed thick-

nesses

and weights of eombi-

nation covering are

the

sums

of

the

inner

laye of

diatoma-

ceous

earth

at

21

lbs/cubic

foot

and

the outer layer

at

11 lbs/cubic

foot.

Insulation

weiEhts

include

allowances

for

w-ire.

cement.

eanvas, bands and

paint,

but

not special surface

finishes.

To find the weieht of

cover-

ing

on

flanges, valves

or

fit-

tings,

multiply

the weight

fac-

tot

by

the weight

per

foot of

covering

used,on

straight

pipe.

v

alve

wergn s

are approxl-

mate.

When

Dossible. obtain

weights from- manufacturer.

Cast

iron valve weiehts are

for

flanged

end

vatves;

steel

\

eights

f

or

. rrelding

end

valves.

All

flanged

fitting,

flanged

valve and

flange

weights

in-

clude the

orooortional

weieht

of

bolts oi stluds to make-up

all

joints.

,-11

/A

.A

A

4

t"{3

@

m

+<i

t€

z

F

tr

z

lt

fsls

J: i.\\

*

16

lb

cu.

ft.

density.



\

Appendix

C: Properties

of Pipe

223

WEIGHTS OF

PIPING MATERIALS

s4"

o.D.

34"

*trc

G

/.^

u-/

b

/-i\

rT

-r

2,1

c_=_=r

"t\

{---t-r

\IJ

z

F

I

z

F

Temperature

Range

"F

Magnesia

Calcium

Fiber-

Sodium

tion

2

{

z

z

{

z

3

a"

z

3

ffi

ffi

Njis

N

Boldface

type

is

weight

in

pounds.

Lightface

tYPe

be-

ireath

weight

is

weight factor

for insulation.

Insulation

thicknesses

and

weishts

are

based

on

average

conditions

and

do not

consti-

tute a recommendation

for

sDecific

thicknesses

of mate_

rials. Insulation

weights

arq

based

on 857,

magnesis

altd

hvdrous calcium

silicat€

at

11

l5s/cubic f oot. The

listed

thick-

nesses and

weights

of

combi-

nation covering

ale

the sums

of

the

inner

layer of

diatoma-

ceous earth at

21

lbs/cubic

foot

and

the outer

layel

at

11 lbs/cubic

foot.

Insulation

weights

include

allowances

for v/ire, cetnent,

canvas, bands

and

paint,

but

not special surface

frnishes'

To

find

the

weisht

of cover-

ine on ffanees,

v-aives or

fit-

tinles. multi6lv

the weiqht fac-

tor"bi

the iveight

per-foot

of

coverrng

usecl

on

slralghl

plpe.

Valve

weights are approxi-

mat€. When

possible,

obtain

weights from

manufacturer.

Cast ilon

valve

weights

are

for

flanged end

valves; steel

weiehts

forweldinsendvalves.

A'il flanged

fitting,

flanged

valve

and flange

weights in-

clude

the

proportional

weight

of

bolts

or

studs

to

make

up

all

joints.

-l)

/A

AI

//

N

/>

+.{

@

m

+<i

rc

*

16

lb

cu. ft.

density.



W

uj

f\

w

{T\

t=l

_/A

--i

A

\iJ

z

F

EI

3

224

Mechanical Design

of Process

Systems

36"

"t"u

s6" o.D.

WEIGHTS OF

PIPING MATERIALS

Temperature

Range'F

Fiber-

Sodirm

I\{agnesia

Ctllcitm,

ffi

6{fliN$

N-S

{raT,s

z

z

F

z

Boldface type is

weight in

pounds.

Lightface type be-

neath

weight

is

weight faetor

for insulation.

Insuiation thicknesses

and

['eights

are

based

on

averag:e

conditiods and do not consti-

tute a

recommendation

lor

specific

thicknesses of mate-

rials. Insulation

weights aae

based on

85% magnesia

and

hydrous calcium silicate

at 11

lbs/cubic foot.

The

listed thick-

nesses

and

weights of

combi-

nation covering are

the

sums

of

the

inner

layer of diatoma-

ceous

earth

at

21

lbs/cubic

foot

and

the

outer

layer

at

11

lbs/cubic foot.

Insulation

weights

include

allowances

for urife, cement,

canvas,

bands

and

paint,

but

not

sDecial

surface

finishes.

To-find

the weight of

cover-

ins

on flanees.

valves or

fit-

tirigs,

multiply

the

weig-ht

fac-

tor

by

the

welgrr

per

lool

or

covering

used on

straight

pipe.

Valve weights

are approxi-

mate.

When

possible,

obtain

weiahts

from

manufacturet.

Cast

iron valve

weights are

for

flanqed

end

valves; steel

weichts iorweldineend

valves,

A-ll flanged

fitting,

flanged

valve and

flange

weights in-

clude

the

proportional

weight

of bolts or

studs

to

make up

all

joints.

,tA

4t

/A

/t\

l|'

tl

6l

.lk{

l-<J

lli'l

+q]

@

Nom.

Thick., In.

Nom.

Thick,,

In.

*

16 ]b cu.

ft. derNity.



Appendix

D

Conversion

Factors

225



226 Mechanical Design of

Process Systems

TO

CONVERT

INTO

Alphabetical Conversion Factors

MULTIPLY

BY

TO CONVERT

tNt0

MULTIPLY

BY

gram-cal/sec

norsepower-hrs

watts

toot-lbs/sec

horsepower

kilowatts

waIls

watts/sq

in.

Cubic Cm,

cu ft

cu In.

cu meters

laters

pecks

pints (dry)

quarts (dry)

Abcoulomb

Acre

acres

acres

actes

acres

acre-feet

acre-feet

amperes/sq cm

amperes/sq

cm

amperes/sq

in.

amperes/sq

In.

amperes/sq

meler

arnperes/sq meter

ampere-hours

arnpere-hours

ampere-turns

ampere-turns/cm

ampere-turn5/cm

ampere-tutns/cm

ampere-turn5/in.

ampere-turns/in.

ampere-turns/ In.

ampere-turns/meter

ampere-turns/meter

ampere-turns/meter

Angstrom

un it

An8stron un

it

Angstrom unit

Ares

ares

Astronomical

LJnit

Atnospheres

atmospneres

atmospheres

atmospneres

atmospheres

atmospneres

atmospheres

atmospheres

Barrels

(U.S.,

dry)

Barrels

(U.S.,

dry)

Barrels

(U.S.,

liquid)

barrels

(oil)

oars

bars

bars

bars

bals

Baryl

Eolt

(US

Cloth)

BTU

8tu

Btu

Btu

Bttr

Btu

t'(U

Btu

8tu

Btu/hr

A

Statcoulombs

Sq. chain

(Gunters)

Rods

Square links

(Gunters)

Hectare

or

sq.

hectometer

sq

feet

sq

mete6

sq mrles

sq

yards

cu feet

gaflons

amps/sq In.

amps/sq

neler

amps/sq cm

amps/sq

meter

amps/sq cm

amps/sq In.

coulombs

faradays

gilberts

amp{urns/in.

amp{urns/neter

gilberts/cm

amp-turns/cm

amp-turns/meter

grlberts/cm

amp/turn5/cm

amp-turns/ in.

gilberts/cm

I ncn

Meter

l\4

icron

or

(i.,lu)

Acre

(US)

sq.

yards

acres

sq meters

Kilometers

Ton/sq. inch

cms

of mercury

ft

of water

(at

4'C)

In.

of mercury

(at

0"C)

kgs/sq

cn

kgs/sq

meter

pounds/sq

jn.

tons/sq

ft

B

cu. tnches

quarts (dry)

8al

tons

gallons (oil)

arrnospnetes

dynes/sq

cm

kgs/sq

meter

pounoS/sq

In.

Dyne/sq.

cm.

Meters

Liter-Atmosphere

ergs

foot-lbs

graln-caloneS

horsepoweahrs

ioules

kjlogram,calories

krlografi-meters

kilowatt-hrs

foot,pounds/sec

2.998 x

10ro

10

160

I x 1Cl'

.4047

43,560.0

4,O47.

1.562

x

10-:

4,840.

43,560.0

3.259 x 1Cl'

6.452

10.

0.1550

1,550.0

I0

I

6.452 x

10-.

3,600.0

0.03731

|.257

2.540

r00.0

t.257

0.3937

39.37

0.4950

0.01

0.0254

0.01257

3937

x

10-'

1x

10-ro

1x 10-.

.0247 |

I19.60

o.o247

|

100.0

1.495 x 101

.007348

76.0

33.90

29.92

1.0333

l0,332.

t4.70

1.058

105.0

31.5

42.0

0.9869

105

..020 x lcr.

2,089.

14.50

1.000

10.409

1.0550 x

10'o

778.3

252.0

3.931

x

l0-l

1,054.8

0.2520

107.5

2.928

x

10-'

o.2t62

Btu

/hr

Btu/hr

Btu

/hr

Btu/min

8tu/min

Btu/man

Btu/min

Btu/sq ftlmin

Bucket

(Br.

dry)

bushels

bushels

bushels

bushe,s

bushels

bushels

bushels

Calories,

gram

(mean)

Candle/sq.

cm

Candle/sq. inch

centares

{centiares)

Centigrade

centiglams

Centiliter

Centiliter

Centiliter

centiliters

centimeters

centimeters

centrmeters

cent,meters

centimeters

centimeters

centimeters

centrmeters

centimeter-dynes

centimeter-dynes

centimeter-dynes

centimeter-grams

centimeter-grams

centimeter-grams

centimeters

of mercury

centimeters

of mercury

centimeters of mercury

centirneters of mercury

centimeters

of rnercury

centimeters/s?c

centameters/sec

centameters/sec

centimeters/sec

centimeters/sec

centlmeters/sec

centimeters/sec

centimeters/sec/sec

centimeters/sec/sec

centarneters/s€c/sec

centimeters/sec/sec

Chain

Chain

Chains

(surveyors'

or

Gunter's)

circular

mils

circular

Inils

Circumference

circular

mils

Cords

Cord

feet

Coulomb

coutomos

c

B.T.U.

{mean)

Lambeats

Lamberts

sq meters

Fahrenheit

glams

Ounce

fluid

(US)

Cubic inch

drams

liters

feet

inches

kilometers

meters

m

es

mallimete6

m ils

yards

cm-grams

meter-xgs

po nd.feet

cm-dynes

rneter-kgs

poundJeet

atmospheres

feet

of water

kgs/sq meter

pounds/sq

tt

pounds/sq

in.

feet

/

min

feet/sec

kilometers/hr

xnotS

mete6/min

miles/

hr

miles

/

rn in

feet/sec/sec

kms/hr/sec

meters/sec/sec

miles/hrlsec

Inches

meters

yards

sq

clns

sq mils

Radians

sq Incnes

cord

feet

cu.

teet

Statcoulombs

faradays

0.0700

3.929 x

l0

'

0.2931

12.96

0.02356

0.01757

t7.57

o.r22r

1.818

x 10'

1.2445

2,150.4

o.03524

4.0

64.0

32.0

3.9685

x

10

I

3.142

.4870

1.0

(C'x9/5)+32

0.01

.6103

2.705

0.01

3-281 x l0-'

0.3937

10-

5

0.01

6.214 x lO-.

10.0

1,094 x

10-I

1.020 x

10-

1.020 x

10-l

7.376 x

10-r

980.7

10-5

7.233

x

10-5

0.01316

0.4461

136.0

27.85

0.1934

1.1969

0.03281

0.036

0.1943

o.02237

3.728 x l0-r

0.03281

0.036

0.01

o.02237

792.00

20.12

22.O0

5.057

r

10-.

0.7854

6.283

7.854 x

10-'

8

l6

2.998

x

10'

1.036

x 10-'



Appendix D: Conversion

Factors

(Continued).

Alphabetical Conversion

Factors

I ULTIPLY 8Y

TO CONVERI

IN?O

MULTIPLY

8Y

227

TO

CONVERT

INTO

coulombs/sq

cm

coulombs/sq

cm

coulombs/sq

in.

cou,ombs/sq

in,

coulombs/sq meter

coulombs/sq meter

cubic

centimeterc

cubic

centirneters

cubic centimeters

cubic

centimete6

cubic centimeters

cubic centimeters

cubic centimeters

cubic centimeters

cubic feet

cubic

feet

cubic

feet

cubic feet

cubic feet

cubic

feet

cubic feet

cubic teet

cubic

feet

cubic feet/min

cubic

teet/min

cubic

teet/min

cubic teet/min

cubic feet/sec

cubic teet/sec

cubic

inches

cubic

inches

cubic

inches

cubic

inches

cubic inches

cubic inches

cubic inches

ctibic

inches

cubic inches

cubic

meters

cubic rneters

cub,c

meters

cubic

meters

cubic

meters

cubac mete6

cubic

meters

cubic

meters

cuDrc meters

cubic

yards

cubic

yards

cubrc

yards

cubic

yards

cubic

yards

cubic

yards

cuorc

yards

cubic

yards

cubrc

yards/min

cubic

yards/fiin

cubic

yards/min

coulombs/sq in,

coulombs/sq meter

coulombs/sq

cm

coulornbs/sq

meter

coulombs/sq

cm

coulombs/sq

in.

cu

feet

cu inches

cu mete6

cu

yards

Sallons

(U.

S. liq.)

liters

pints (U.S.

liq.)

quarts (U.S.

liq.)

bushels

(dry)

cu cms

cu inches

cu meters

cu

yards

gallons

(tJ.S.

iiq.)

liters

pints (U.S.liq.)

quarts (U.S.

liq.)

cu cms/sec

gailons/sec

liters/sec

pounds

of water/min

million

gals/day

gallons/

min

cu cms

cu feet

cu meters

cu

yards

ga

onS

liters

mil-feet

pints (U.S.

liq.)

quarts (U.S.

liq.)

bushels (dry)

cu

cms

cu feet

cu tnches

cu

yards

gallons (U.S.

liq.)

liters

pints

(U.S.

liq.)

quarts (U.S.

liq.)

cu cms

cu feet

cu

rncnes

cu meters

gallons

{U.S.

liq.)

liters

pints

{U.S.

'iq.)

uarts (u.s.

liq.)

cubic

ftlsec

Sallons/sec

liters/sec

0

Gram

seconds

grams

tlers

meters

quadrants

radrans

seconds

64.52

10.

1,550.

l0-.

6.452

x t0-l

3.531

x

10-'

0.06102

10-.

1.308

x 10-'

2.542 x lO-'

0.001

2.113 x l0-

1.057

x

10-

0.8035

-

2A32O.O

t,72A.O

o.02832

0.03704

7.4a0s2

2432

59.84

472.0

0.t247

0.4720

62.43

0.646317

448.831

5.787 x

10-.

1.639 x

10-'

2.143 x

10-5

4.329

x

l0-3

0.01639

1.061x

105

0.03463

0.01732

106

5C.lt

61,023.0

1.308

264.2

1,000.0

2,1r3.0

1,057.

7.646

x

IO'

27.O

46,656.0

0.7646

202.0

764.6

1,615,9

807.9

0.45

3.367

t2.74

oramS

oramS

otams

Dyne/cm

oyne/sq. cm.

Dyne/sq. cm.

Dyne/sq. cm.

dynes

dynes

dynes

dynes

dynes

dynes

oynes/sq

cm

EII

Etl

Em,

Pica

Ern, Pica

*glsec

ergs

ergs

erSs

ergs

ergs

ergs

ergs

ergs

ergs

erg5/sec

ergs/sec

farads

Faraday/sec

faradays

faradays

Fathom

Iathoms

feet

leet

feet

feet

teet

feet

leet

feet

ol water

feet of

water

leet of water

Srams

grains

ounces

Erglsq.

millimeter

Atmospheres

Inch of Mercury at

0'C

Inch of Water at

4'C

grams

JOUTeS/Cm

joules/meter

(newtons)

kilograms

poundals

pounds

bars

Cm.

Inches

Inch

Crn.

Dyne

-

cm/sec

Btu

dyne-centimeters

foot'pounds

Srarn-calo

es

Sram-cm5

Joules

Kg-carofles

Kg-melers

kilowatFhrs

watt-houts

Btu/min

ft-lbs/min

ft-lbs/sec

kg-calories/min

kilowatts

microfarads

Ampere {absolute}

ampere-hours

coulombs

l{eter

feet

centimeters

krlometers

meters

rniles

(naut.)

miles

(stat.)

millimeters

mr

ls

armospnere5

an. of mercury

Kgs/sq

cm

0.01745

0.1667

2.778 x

10

1

r0.0

10.0

10.0

0.r371429

0.125

3.6967

1.7714

27.3437

0.0625

.01

9.869 x

10-'

2.953 x

l0-'

4.015 x

10-'

1.020

x

10

I

10-'

10-

'

1.020

x

10

6

7.233 r 10-5

2.248 x 10-'

10-6

114.30

45

.4233

1.000

9.480 x l0-rr

1.0

7.367 x

10-l

0.2389 x

10

'

1.020

x

10

3.7250 x

10-r'

10-

'

2.389 x

l0

-rl

1,020

x

10-'

O.277ax

I0

t3

0.2778

x

10

-ro

5,688 x 10-,

4.427

x

lO-'

7.3756

x 10-l

1.341 x

l0-ro

1.433

x

l0-'

10-,0

10

9.6500

x

lcr

26.4O

9.649

x

l0

1.828804

6.0

30.48

3.048

x 10

'

0.3048

1.645 x l0-.

1.894 x

10

.

304.8

1.2 x

lg

0.02950

0.8826

0.03048

degrees/sec fadians/sec

degrees/sec revolutaons/min

degrees/sec

revolltions/sec

oeKa8rams

gtams

dekaliters

liters

dekamete6

meters

Drams

(apothecaries'

or troy)

ounces

(avoidupois)

Drams

(apothecarieS'

or

troy)

ounces

(troy)

Drams

(U,S.,

fluid or apoth.)

cubic cm.

Dalton

days

decrgrams

deciliters

oecrmelers

degrees

(angle)

degrees

(angle)

degrees

{angle)

1.650 x l0-1.

86,400.0

0.1

0.1

0.1

0.01111

0.01745

3,600.0




of

Process

Systems

TO

CONVERT INTO

(Continued).

Alphabetical

Conversion

Factors

MULTIPLY

BY

TO

CONVERT

INTO

MULTIPLY 8Y

teet of water

feet

of

water

feet of water

teet/m

in

feet/ min

feet/ min

feet/ min

feet/min

feet/sec

feet/sec

feet/sec

feet/sec

feet/sec

feet/sec

teet/sec/sec

feet/sec/sec

feet/sec/sec

feet/sec/sec

feet/

100

feet

Foot

-

candle

foo pounds

foot-pounds

loot.pounds

foot-pounds

foot-pounds

foot-pounds

foot-pounds

foo pounds

foot-pounds/min

foot-pounds/min

loot-pounds/mjn

loot-pounds/m,n

foot-pounds/min

toot-pounds/sec

foot-pounds/sec

foot-pounds/sec

toot-pounds/sec

foot-pounds/sec

Furlongs

turlongs

furlonBs

Sallons

garrons

galrons

Sallons

gallons

gallons

gallons

(liq.

Br. lmP,)

gallons

(U.S.)

gallons

of watef

gallons/min

gallons/min

gallons/min

gausses

Sausses

Sausses

gausses

gilberts

gilberts/cm

gilberts/cm

gilberts/cm

cills

(British)

gills

Sills

Grade

Grains

kgs/sq

meter

pounds/sq

ft

Pounds/sq

in.

cms/sec

teet/sec

kms/hr

meters/min

miles/hr

crns/sec

kms/hr

knots

meters/min

miles/hr

males/

rn

in

cms/

sec/sec

kms/hr/sec

meters/sec/sec

miles/

hrlsec

per

cenl

graoe

Lumen/sq. meter

Btu

ergs

grarl1-calofles

np-nrs

JOules

kg'calories

kg-meters

kilowatt-hrs

Btu/min

foot-pounds/sec

hotsepowel

kg-calories/min

kilowatts

Btu/hr

Btu/min

horsepower

kg-calories/min

kilowatts

miles

(u.S.)

rooS

feet

cu cms

cu feet

cu

Inches

cu meters

cu

yards

liters

gallons

(U.S.

liq.)

eallons

(lmp.)

pounds

of water

cu

ftlsec

liters/sec

cu ft/hr

lanes/sq in.

weDers/sq

cm

webers/sq in.

webers/sq meter

ampere-turns

amp-turns/cm

amp-turns/in

amp-turns/meter

cubic cm.

liters

pints

(liq.)

Radian

drarns

(avoirdupois)

304.8

62.43

0.4335

0.5080

0.01667

0.01829

0.3048

0.01136

30.48

1.097

0.5921

18.29

0.6818

0.01136

30.48

1.097

0.3048

0.6818

1.0

10.764

1.286 x 10-3

1.356 x 10'

0.3238

5.050 x l0-'

1.356

3.24

x 1.0

.

0.r383

3.766 x l0-'

1.286 x l0-3

0.01667

3.030

x 10

-5

3.24

x

lO-.

2.260 x l0-5

o.o77 17

1.818 x

l0-'

0.01945

1.356 x 10-'

o.125

40.0

660.0

3,785.0

23i.0

3.785

x 10-'

4.951 x

10-t

3.785

1.20095

o.83267

8.3453

2.22a

x

l'-t

0.06308

8.0208

6.452

l0-l

6.452 x

10-,

10-.

0.7958

0.7958

2.02r

79.58

142.O7

0.1183

0.25

.01571

0.03557143

grains

(troy)

grains

(troy)

Srains

(troy)

giains

(troy)

Srains/U.S.

gal

grains/U,S.

8al

grains/lmp.8al

Srams

Srams

Sralns

Srams

Srams

grams

grams

grams

grams

g,ams

grams/cm

Slams/cu

cm

gr-arns/cu

cm

Srams/cu

cm

grams/

liter

grams/

liter

grams/liter

grams/liter

grams/sq

cm

gram-calones

gram-calories

Sram-catones

Stam-catofles

Sram-calories

gram-calones

gram-caloraes/sec

gram-centimeters

gram-centimeters

gram-centrmeters

gram'centametels

grafi-centimeters

Hand

nectares

nectares

neclograms

hectoliters

hectometers

hectowatts

henries

Hogsheads

(British)

Hogsheads

(U.S.)

Hogsheads

(U.S.)

hoasepower

holsepower

horsepower

horsepower

(met.ic)

(542.5

ft

lb/sec)

horsepower

(550it

lb/sec)

horsepower

ho15epower

horsepower

horsepower

(boiler)

horsepo',ver

(boiler)

horsepower-hrs

horsepower-hrs

horsepower-hrs

horsepower-hts

norsepower-nrs

grains (avdp)

grams

ounces

(avdp)

pennyweight

(troy)

parts/million

pounds/million

gal

parts/rnillion

oynes

Slarns

joules/cm

joules/meter

(newtons)

kilograms

milligrams

ounces

{avdp)

ouhces

(troy)

pounoals

pounds

pounds/inch

pounds/cu

ft

pounds/cu

in

pounds/mil-toot

grains/gal

pounds/

gal

pounds/cu

ft

parts/nillion

pounds/sq

ft

6tu

foot-pounds

horsepowet-hrs

kilowatt-hrs

watt-hr9

Btu/hr

Btu

ergs

joules

kg-cal

xg-meters

Cm.

acres

sq feet

grams

liters

meters

watts

millihenries

cubac ft.

cubic

ft.

gallons

(U.S.)

Btu/min

foot-lbs/min

foot-lbs/sec

horsepowet

(550

ft lb/sec)

horsepower

(metric)

(542.5

ft lb/sec)

kg.calories/min

kilowatts

watts

Btu/hr

kilowatts

Btu

ergs

footl

bs

gram.calol|es

JOU

leS

1.0

0.06480

2.0833

x

10-t

0.04167

17.118

142.56

14.286

980.7

15.43

9.807 x

lo-t

9.807 x

10-

0.001

1,000.

0.03527

0.03215

0.07093

2.205 x l0-'

5.600

x

l0-r

0.03613

3.405

x

l0-t

58.417

8.345

o.062427

1,000.0

2.0481

3-9683

x

10-t

4.1868

x l0'

3.0880

1.5596 x l0-.

1.1630 x l0-.

1.1630 x 10-3

14.286

9,297 x lO-.

980.7

9.807 x

l0-5

2,343 x 10-3

10

-'

10.15

2.471

1.076 x 103

100.0

100.0

100.0

100.0

1,000.0

10.114

8.42184

42.44

33,000.

550.0

0.9863

1.014

10.68

0.7

457

7 45.7

33.479

9.803

2,547.

2.6845

x 10u

1.98

x l0'

641,190.

2.684

r l0'



Appendix D:

Conversion

Factors

(Continued),

Alphabetical

Conversion

Factors

TO

CONVERT

229

TO COI{VERT

tt{To

ilIULTIPLY

BY

INTO

HULTIPLY BY

ho.sepower-hrs

horsepower-hrs

horsepower-hrs

nours

houls

HundredweiShts

(long)

Hundredweights

(long)

Hundredweights

(short)

Hundredweights

(shortl

Hundredweights

(short)

Hundredweights

(short)

inches

inches

inches

inches

Inches

inches

inches of

mercury

atmospheres

inches

of

mercury

feet of water

inches

of

mercury

kgs/sq cm

inches

of

mercury

kgs/sq meter

inches

ot

mercury

pounds/sq

tt

inches of mercury

pounds/sq

an.

inches of water

(at

4'C) atmospheres

inches of watet

(at

4'C)

inches

of

mercury

inches of water

(at

4'C) kgs/sq cm

inches of

water

(at

4'C)

ounces/sq

in.

inches

of water

(at

4'C)

pounds/sq

ft

inches

of water

(at

4'C)

pounds/sq

in.

International

Ampere

Ampere(absolute)

InternationalVolt

volts(absolut€)

Inte.nationalvolt

Joules(absolute)

lniernational

volt Joules

kg.calories

l(g-meters

kilowatt-hrs

qays

pounds

tons

(long)

ounces

(avoirdupois)

pounos

tons

(metric)

tons

(long)

I

centimeters

meIels

miles

millimeters

mils

yaros

641.1

2.7X7 x lU

o.7457

4.167 x

10-r

5.952

x l0-r

t12

0.0s

t600

100

0.0453592

o.0446429

2.540

2.540x

10-t

1.578 x 10-5

25.40

1,000.0

2.77a x

rO-'

0.03342

0.03453

345.3

70.73

o.4912

2.458 x

10-

0.07355

2.540

x

l0-1

0.5781

5.204

0.03613

.9998

1.0003

l-593

x

10-''

9.654

x

l0'

kilograrns/sq

cm

kilograms/sq

cm

kilograms/sq cm

kilograms/sq rneter

kilograms/sq

meter

kilograms/sq meter

kilograms/sq meter

kilograms/sq

meter

kalograms/sq meter

kilograms/sq mm

kilogram-calories

kilogram-calories

kilogram-calories

kilogram-caloraes

kilogram.caloaies

kilogram-calories

kilogram-calories

kilogram meter9

kilogram

meters

kilogram meters

kilogram meters

kilogram meters

kilogram meters

kilolines

kiloliters

kilometers

kilometers

kilometers

kilometers

kilometers

kilometers

kilometers

kilometers/hr

kilometers/hr

kilometers/hr

kilometers/hr

kilometers/hr

kilometers/hr

kilometers/hrlsec

kilometers/hrlsec

kilometers/hrlsec

kilometers/hrlsec

kilowatts

kilowatts

kilowatts

kilowatts

kilowatts

kilowatts

kilowatt-hrs

kilowatt-hrs

kilowatt-hrs

kilowatt-hrs

kiiowatt-hrs

kilowatt-hrs

kilowatt-hrs

kilowatt-hrs

kilowatt-hrs

kilowatt-hrs

knots

knots

l(nots

knots

inches of mercury

pounds/sq

lt

pounos/sq

In.

atmospheres

oars

teet ot water

inches ot mercury

pounds/sq

ft

pounds/sq

in.

kgs/sq meter

Btu

foot-pounds

hp-h.s

joules

kg-meters

kilojoules

kilowatt-hrs

Btu

foo pounds

JOUIeS

kg-calories

kilowatt-hrs

liters

centimetels

{eet

inches

meterS

miles

millimeters

yards

cms/sec

feet/min

teet/sec

knots

meters/nin

miles/hr

cms/sec/sec

ft

/sec/sec

meters/sec/sec

miles/hrlsec

Btu/min

foot-lbs/min

foot-lbs/sec

norsepower

kg-calories/min

Btu

foot-lbs

24.

2,O44.

14.22

9.678

x

10-'

98.07

x

l0-.

3.281

x l0-:

2.896

x 10-l

0.2044

1.422

x

l0-'

10.

3,088.

1.560 x 10-1

4,186.

426.9

4.186

1.153

x

l0-3

9.294 x

10-r

9.804 x 10'

9.804

2.342

x

lO''

2.723 \ 1O-'

1,000.0

1,000.0

10,

3,281.

3.937

x lO

1,000.0

0.6214

lCl'

1,094.

27.74

54.68

0.9113

0.6214

27.74

0.9113

0.2774

0.6214

4.426

\ W

737.6

1.341

14.34

1,000.0

3,413.

3.600

x 10r'

2.655 x 10.

JOUIeS

joules

joules

ioules

joules

joules

joules/cm

ioules/cm

joules/cm

.loules/cm

ioules/cm

Btu

9.480

x

10-'

ergs

107

footpounds

0.7376

kg-calories 2.389

x

l0-'

kg-meters 0.1020

watlhrs 2.77Ax lO-'

grams

1.020

x 10.

dynes

10'

joules/meter(newtons)

100.0

poundals

723.3

pounds

22,44

kilograms

kilograms

kilograms

kilograms

kilograms

kilograms

kilograms

kilograms

kilograms/cu meter

kilograms/cu

meter

kilograms/cu

fieter

kilograms/cu

meter

kilograms/meter

Kilogram/sq.

cm.

kilograrns/sq

cm

kilograms/sq crn

K

dynes 980,665.

grams

1,000.0

joules/cm

0.09807

joules/meter(newtons)

9.807

poundals

70.93

pounds

2205

tons

(lond

9,842 x 10-'

tons

(short)

1.102 x

10-r

grams/cu

cm 0.001

pounds/cu

tt 0.06243

pounds/cu

in, 3,613 x 10-5

pounds/mil-foot

3.405 x 10-'o

pounds/ft

0,6720

oynes

980,665

atmospheres

0.9678

feet

of

water 32.81

gram-calories

859,850.

horsepower-hrc

1,341

joules

3.6

x

lcl.

xg.carofles

5bu.5

k8-meters

3.671

x

10'

pounds

ot water

evaporated from and

at212'F.

3.53

pounds

ot water raised

frcm62"

to 212"

F.

22.75

feet/hr 6,080.

kilometers/hr

1.8532

nautical

miles/hr

1.0

statute

miles/hr

1.151



Mechanical Design

of

Process

Systems

(Continued).

Alphebetical Conversion

Factors

TO

CONVERT INTO

MULTIPLY

BY

TO

COI{VERT

INTO

IT.IULTIPLY BY

230

knols

knots

leaSue

Light

year

Light

Year

lines/sq

cm

lines/sq

in.

lines/sq

in.

lines/sq

in.

lines/sq

in.

links

(engineer's)

links

{surveyor's)

liters

liters

liters

liters

liters

liters

liters

liters

liters

liters/min

liters/min

lumens/sq

ft

Lumen

Lumen

Lumen/sq.

ft.

tux

maxwells

megaltnes

megohms

megohms

fieters

meters

metets

meters

metels

meters

meters

meters

metets

meters/m,n

meters/man

meters/mrn

meters/min

metels/min

meters/min

meters/sec

mete6/sec

meters/sec

meters/sec

mere6/sec

metels/sec

meters/sec/sec

meters/sec/sec

mete6/sec/sec

mete6/sec/sec

meterkilograms

meteFkilograms

meteFkilograms

microfarad

micrcgrams

micrchms

yards/hr

feet/5ec

L

miles

(approx.)

Miles

Kilometers

gausses

Sausses

weDers/sq cm

w€bers/sq

in.

webers/5q

meter

inches

inches

bushels

(U.S.

dry)

cu cm

cu

feet

cu

tnches

cu

mete6

cu yards

eallons

(u.S.

liq.)

pints

(U.S.

liq.)

quads

(U.S.liq.)

cu

ft/sec

gars/sec

foot-candles

Spherical

candle

power

Watt

Lumen/sq.

meter

foot-candles

tl

kilolines

webels

maxwells

microhms

ohms

centimeters

leet

anches

kilometers

miles

(naut.)

miles

(stat.)

millimeters

yards

cms/sec

feet/min

teet

/sec

kms/hr

knots

miles/hr

feet/ m in

feet/sec

kilomete15/hr

kilometers/min

miles/hr

miles/min

ft/sec

/sec

kms/hrlsec

rniles/hrlsec

cm-dynes

cm-grams

pound-feet

farads

glams

megohIns

2,027.

1.589

5.U

5.9

x

10rr

9.46091

x

10"

1.0

0.1550

1.550

x

l0-'

l0-l

1.550 x

10-r

t2.o

0.02838

1,000.0

0.03531

61.02

0.001

1.308 x

10-r

0.2642

2.r13

1.057

5.886

x

l0-'

4.403 x 10-'

1.0

.07958

.001496

10.75

0.0929

0.001

10-l

1Cl.

10u

10.

100.0

39.37

0.001

5.396

10-1

6.214

x 10-'

1,000.0

1.094

1.179

1.567

3.281

0.05458

0.06

0.03238

0.03728

195.8

3.281

5,O

0.06

0.03728

r00.0

3.281

9.807 x 19

lCr'

l0-.

10-.

10-rl

microhms

micrcliters

Microns

miles

(naut.)

miles

(naut.)

miles

(naut.)

miles

(naut.)

miles

(naut.)

miles

(statute)

miles

(statute)

miles

(statute)

miles

(statute)

miles

(statute)

miles

(statute)

miles

(statute)

miles/hr

rniles/h.

miles/hr

miles/hr

miles/hr

miles/hr

miles/hr

miles/h.

miles/hr/sec

miles/hrlsec

miles/hrlsec

miles/hr/sec

miles/min

miles/min

miles/min

miles/min

miles/min

mil-feet

milliers

Millimicrons

Milligrams

milligrams

milliSrams/litet

millihenrie5

milliliters

millimeters

millimeters

millimeters

millimeters

millimeters

millimete6

millimeters

millimelers

million

gals/day

mils

mils

mils

mils

mils

miner's

incheg

Minims

(British)

Minims

(U.S.,

flu;d)

minutes

(angles)

minutes

(angles)

minutes

(angles)

minutes

(angles)

myriagrams

myriameters

myriawatts

ohms

liters

mererc

feet

kilometers

meters

miles

(statute)

yards

centametels

feet

inches

kilometers

meterc

miles

(naut.)

yards

cms/sec

leet/man

feet/sec

kms/ht

kms/min

knots

meters/min

miles/min

cms

/

sec/sec

feet

/

sec

/sec

kms/hr/sec

meters/sec/sec

cms/sec

teet/sec

kms/min

knots/min

miles/hr

cu

inches

kiloSrams

meters

g|a

Ins

grams

parts/million

henries

liters

centimetels

feet

inches

kilometers

meters

miles

mrls

yards

cu

ftlsec

centimeters

feet

anches

kilorneters

yaros

cu

ft/min

cubic

cm.

cubac cm.

oeSrees

quadrants

radians

seconds

kilograms

kilometers

kilowatts

10-.

10-.

I

x

10-'

6,04O.27

1..'J5

1,853.

1.1516

2,027.

1.509

)(

1Cl'

5,280.

6.336

x 10

r.609

1,509.

0.8684

1,760.

M.70

8&

t,467

1.609

o.o26a2

0.8684

26.42

0.1667

44.70

L.467

1.509

0.4470

2,642.

88.

0.8584

60.0

9.425

x

10-'

1,000.

I

x

lo-t

0.01543235

0.001

1.0

0.001

0.001

0.1

3.281

x 10-t

10-.

0.001

6.214

x 10-'

1.094

x l0-'

1.54723

2.540 x

10-t

8.333

x

10-

0.001

2.540

x

10-'

2.77Ax

lO-'

t.5

0.059192

0.0516r2

0.01667

1.852 x 10-'

2.909

x l0-.

60.0

10.0

10.0

10.0

8.686

1x105

N

decibels

Dynes






of

Process

Systems

TO CONVERT

INTO

(Continued).

Alphabetical

Conversion

Factors

MUI.TIPLY

BY

TO

COI{VERT

INTO

MULTIPLY

BY

revolutions/min/min

revolutions/min/min

revolutions/min/min

revolutions/sec

revolutions/sec

revolutions/sec

revo,utions/sec/sec

revolutions/sec

/sec

revolutions/sec/sec

Rod

xoo

Rods

(Surveyors'

meas.)

rods

Scruples

seconds

{angle)

seconds

(angle)

seconds (angle)

seconds

(angle)

Slug

Slug

Sphere

square centimeters

square cent|melerS

square centimeters

square centrmeters

square

cen melers

square

centimeters

square

centimeters

square feet

square

Inches

square

Inches

square

Inches

square

Inches

square

inches

square

k'lometers

square

kilofleters

square

kilometers

square

kilorneters

square

kilometers

square

kilometers

square

kalometers

square

meters

square

meters

square melers

square

meters

square

meters

square

meters

square

meters

square

miles

square

miles

square

miles

square

mrles

square

millimeters

square

millimeters

square

millimeters

square

millimeters

square

rn ils

radians/sec/sec

revs/sec/sec

oegrees/sec

radians/sec

radians/sec

/sec

revs/min/min

revs/man/sec

Chaan

(Gunters)

Meters

yards

feet

s

gra,ns

minutes

quaoranls

radians

Kilogram

Pounds

Steradians

circular

lnils

sq feet

sq rnches

sq miles

sq

millimeters

sq

yards

acres

circular

mils

sq

cms

sq inches

sq mrles

sq millimeters

sq

yaros

circu lar

mils

sq cms

sq teet

sq millimeters

sq mils

sq

yards

acreS

sq

cm5

sq

ft

sq mrles

sq

yards

sq cns

sq

feet

sq miles

sq millimeters

sq

yards

actes

sq feet

sq xms

sq meters

sq

yards

circular mils

sq

cms

sq feet

sq inches

circular mils

1.745

x

10-r

0.01667

2.778 x 10-.

360.0

6.283

50.0

3,600.0

60.0

.25

5.029

20

2,778\ lO

.

0.01667

3.087 x 10-6

4.848 x l0-l

14.59

32.17

1.973 x 10'

1.076 x l0-3

0.1550

0.0001

3.861 x

10-'r

r00.0

1.196 x 10-.

2.296 x 10-,

1.833 x l0o

929.O

144.0

0.09290

3.587 x

l0-r

9.290 x lCr

0.1111

1,273

x

106

6.452

6.944 x

l0-3

10.

7.716 x

10-.

247.1

10x

10.76 x 106

1.550 x 10'

106

0.3861

1.196 x

106

2.471 x lO-.

10.

10.76

1,550.

3.861 x 10-'

1Cp

1.196

640.0

27.88 x 10.

2.590

2.590 x 10d

3.098

x 106

1,973.

0.01

1.076 x 10-r

1.550

x 10-

1.273

square mrls

squate nrl5

square

yards

square

yards

square

yards

square

yards

square

yards

square

yards

square

yards

6.452 x 10-6

10-6

2.066 x 10-.

8,361.

9.0

't

,296.

0.8361

3.228

x 1O-,

8.361

x l0'

.39370

.003336

temperature

("F)-32

temperature

('C)

5/9

tons

(long)

kilog€ms

1,016.

tons

(long) pounds

2,240.

tons

{long)

tons

(short)

1,120

tons

(metric)

kilograms

1,000.

tons

(metric) pounds

2,205.

tons

(short)

kilograms

907.1848

tons

(short)

ounces 32,000.

tons

(short)

ounces

(troy)

29,166.65

tons

(short) pounds

2,000.

tons

(short)

pounds

(troy)

2,430.56

tons

(short)

tons

(long)

0.89287

tons

(short)

tons

(metric)

0.9078

tons

(short)/sq

ft

kgs/sq

meter

9,765.

tons

(short)/sq

ft

pounds/sq

in.

2,000,

tons

of

water/24

hrs

pounds

of

water/hr

83.333

tons

of

water/24

hrs

gallons/min

0.16643

tons of water/24

hrs

cu

ltlhr

1.3349

tempemture

("c)

+273

temperature

('c)

+

r7.78

temperalure

("F)

+460

Volt/

inch

Volt

(absolute)

watts

walls

watts

watts

Watts

(Abs.)

Watts

(Abs.)

watt'hours

watt-hours

watt-hours

watt'hours

watt-hours

watt-hours

watt-hou.5

watt-hours

Watt

(lnternational)

sq

cns

sq Inches

acres

sq cms

sq inches

sq meters

sq males

sq millimeters

T

absolute

temperature

('C)

1.0

temperature

('F)

1.8


('F)

1.0

v

Volt/cm.

Statvolts

w

Btu/hr

Btu/min

eags/sec

foot-lbs/min

toot'lbs/sec

horsepower

horsepower

(metric)

kg-calories/min

kilowatts

B.T,U.

(mean)/man.

joules/sec.

Btu

erSs

foofpounds

gram-caloneS

horsepolver-hrs

kilogram-calories

kalogram-meters

kilowatt-hrs

Watt

(absolute)

kilolines

3.4129

0.05688

107.

44.27

0.7374

1.341

x

l0-1

1.360

x

10-

0.0t 433

0.001

0.056884

I

3.413

3.60

x

10'o

2,656.

859.85

1.341

x l0-1

0.8605

0.001

1.0002

1Cp

10,



Appendix

D:

Conversion Factors

2Sg

Synchronous

Speeds

Frcqusncy r

120

syncnronou3

sPc.o

- T;;Ei;;-

FIEQUEiICY

60.ycle

50

.y.lc

50 Gycl.

142.9

136.4

130.4

r25

t20

5.a

||t.l

t

o7.l

103.5

100

96.8

93.7

90.9

88.2

85.7

83.3

8r.l

78

-9

76.9

75

6

8

l0

l2

II

t6

t8

2l

26

28

30

32

31

36

38

10

3600

r800

1200

900

600

5r 4.3

150

400

360

327

.2

300

276.9

257

.1

210

225

2n.8

200

t89.5

r80

3000

t

500

1000

750

600

500

124.6

375

300

272.7

250

230.8

211.3

200

187.5

175.5

166.7

157

-9

150

| 500

375

300

250

214.3

187.5

166.7

t50

136.4

lt5.a

t 07. t

100

93.7

88.2

83.3

78,9

75

12

11

a6

18

56

60

62

61

66

58

72

71

76

80

171.1

|

63.6

|

56.5

l50

111

t38.5

133.3

128.6

t21.1

120

rr6.t

2.5

t0t. I

r

05.9

102

-9

t00

97 .3

91.7

92.3

?0

Courtely

Ingersoll-Rand

Co.



234

Mechanical Design

of

Process

Systems

Temperature

Conversion

Thc G.nter .oluh'l

of nu|'b.t

in

boldfo..

.efeB to the teDperotur. in desreei, either

Cenrig.odc

or

Fohrenh.ir, whidr

it ir

d. ir.d

to conv.rt inlo

lh.

lf.o.v.rtins kom fohr€nhcil

lo

Ccntis.ode

degr€e . the equivolent tempe.oiure will

bc

found

in

lh.lefi

col'r6n, whileil

convc.li.s

lron d.s.c€i

to

d.gr..r

fobr.nhi.t,

thc

oniy€r

$ll

b.

fo'rnd in

the

column

on

thc

right.

C.ntigrodc Fohrenhcil

Centigrode

enlisrod.

C.ntlgrod.

5t.t

60.0

62.8

65.6

68.3

71.1

73.9

76.7

79.1

s2.2

85.0

82.8

90.6

93.3

96.1

93.9

100.0

t02

t04

t07

I

l0

ll3

I t6

I

l8

l2l

121

127

129

132

135

138

t{

143

lt6

l,a9

154

t60

t66

t71

177

182

t88

r93

t99

201

210

216

221

266

2e1

293

3ll

t20

-273.17

-a59.f

-268 -r50

-262

-aao

-257

-{30

-25t

-420

-216 -aro

-210

-,100

-231

-390

-229

-3r0

-223

-370

-2t8

-360

-212

-350

-207

-3{0

-20t

-330

-t96

-310

-190 -3ro

-t81

-300

-179

-290

-173

-2t0

-f69

-2f3

-168 -rro

-t62

-260

-157

-250

-tsf

-rao

-116

-230

-f10

-220

-t34

-210

-129

-2oO

-123 -r90

-I8

-tlo

-l

12

-tto

-107

-r50

-tot

-t50

-96

-lao

-90

-r30

-8{

-120

-79

-ll0

-73.3

-I00

-67

-S

-90

-62.2

-r0

-59.r

-56.7

-53

.9

-51

.l

-18.3

-15.6

-/2.9

-,40.0

-31.1

-3t

.7

-28

.9

26.1

formulor

ol

lh. .isht hoy oho

bc ured

(ony€rling

Ceotigrcdc

or

foh.enhcil

125.6

127.1

12r.2

r3l .0

r

32.8

| 34.6

| 36.4

| 38.2

l{0.0

l4r.8

t 13.6

I t5.4

117.2

149.0

t

50.8

152

.6

154.4

|

56.2

158.0

159.8

161 .5

163.4

165.2

167

.O

168.8

170.6

t72.1

171.2

't76.O

177

.8

t79.6

l8l .,a

I83.2

185

.0

186.8

|

88.6

|

90

.,4

192.2

194.0

| 95.8

197

.6

r

99.4

201 .2

203.0

204.8

206,6

208.4

21o.2

212

.0

221

230

239

218

257

Dcqree' Fohr.,

'F

=

fc

+

.ot

-.0

esr€e'

cent,'.=|et

+

,ot

-ro

=

|

et-r'r

Degreca KeMn,'K:'C

+

273.2

-75

-ro

-55

-50

- l

-50

a5

-40

-35

-25

-20

-ll

-t0

-159.1

-151

-136

-4t8

-100

-382

-361

-316

328

-292

-271

-256

-238

-220

-202

-181

-t66

-148.0

-t

30.0

-t

l2 .0

-t

03.0

91.0

-65.0

-75,0

-67

.0

-58

.0

-49.0

10.0

-31.0

-22.0

-t3.0

-4.0

.4.0

-20.6

-16.7

-t6.1

-15.0

-14.1

-13.t

-r3.3

-t2.8

-t2.2

-

.l

-10.6

-10.0

-8

.9

-8

.3

6.7

-6.1

-1.1

-3

.9

1.7

0.0

0.6

LI

1.7

2.2

3.9

4.1

5.0

5-6

6.1

6.7

8.3

8.9

9.1

10.0

l0 .6

23

-O

32.0

35.6

37.1

39 .2

al.0

t2.a

lt.6

16.4

18.2

50.0

5r.8

57 .2

59.0

60

.8

62.6

61.1

66.2

d8.0

69.8

71.6

73.1

75.2

77 .O

80.6

8?

.,{

81.2

86.0

82.8

89.6

9t.,(

93.2

95.0

96.8

98.6

100.,{

102.2

104

.0

105.8

to7

.6

109.1

l]1.2

113.0

114.8

| 16.6

118.,(

120.2

122.O

t23.9

.l

11.7

12.2

I2.8

13.3

t3.9

14.1

15.0

l5 .6

16.

r

t6.7

t7 .2

l, .8

|

8.3

l8 .9

I9..{

20.0

20.6

2t.r

2t .7

22.2

22.8

23.9

21.1

25.0

25.6

26.1

26.7

27

.2

27 .S

28.3

28.9

29

.1

30.0

30.6

3t.l

31

.7

32.2

32.8

33.3

33.9

31.1

35.0

35.6

36.7

37

.2

40 .6

43.3

,(6.l

18.9

5t

.7

52

53

5a

55

55

57

5t

59

60

6l

52

53

61

55

55

57

5l

59

70

fl

72

,3

7a

75

76

7a

,9

o

8l

a2

83

4

E5

86

a,

E8

89

90

9l

92

93

9a

95

95

9f

9A

99

100

t05

0

||5

t20

t25

t30

t35

ta0

l,l5

150

I55

t50

65

t70

t75

ta0

l15

t90

195

200

205

210

212

215

220

225

230

235

240

7.45

250

233

260

255

270

215

280

2t5

290

495

300

310

320

330

3:10

350

360

370

3t0

390

a0o

at0

420

r30

-5

0

I

2

3

1

5

7

I

9

t0

u

t3

l4

l5

t5

t7

t8

t9

20

2l

23

24

25

25

27

2.4

a9

30

3l

32

34

35

36

3f

3l

39

40

/tt

12

{3

14

15

a6

17

/t8

a9

50

5l

227

238

213

219

251

260

329

338

317

356

371

383

392

,a0r

110

ttt

119

t2a

137

116

155

461

173

ta2

191

500

509

518

536

545

554

s72

590

608

626

611

662

680

69S

,16

731

v0

748

805

L0

gza

aso

812

150 860

170

678

{r0 896

ago 914

500

932

i.lo thc

othcr

3cal.r.

c

+32

Degrccr

Rcrftlne,

ol

:oF+459.7

9



Altitude

and Atmospheric

Pressures

Hs

Ab3.

Aooendix

D:

Conversion

Factors

235

Kelrq

H9 Abr.

PSIA

-5000

,{500

,{000

-3500

-3000

2500

-2000

-1500

-1000

t0,000

t5,000

20,000

25,000

30,000

35,000

40,000

15,000

50,000

55,000

60p00

20.000

80,000

90,ooo

t00,000

120,000

t,(0,000

160,000

180,000

200p00

220,000

2,{0,000

260,000

280,000

300,000

,(00,000

500,000

600,000

900,000

tp0o,ooo

1.200,000

1,400,000

1,600,000

l,8oo,ooo

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236

Mechanical Design

of Process

Systems

I

9

o

A



Index

American

Society

of

Mechanical

Engineers. See

ASME.

API,

degrees for hydrometer,

conversions, tables

of, 92

defined,8T-88

ASME

Section

VIII

Division

I

joint

reliability

factor,

l13-l14

joint

types

for tubesheets. I

l5

maximum tube

joint

force,

ll3,

157

tube

joint

load criteria,

113

vessel code,

99,

101

Axial

flow

compressors

aircraft,

for,

59

airfoil

blades

for

pitch,

58

size,58

applications of, 44,

58-59

characteristic curve

for,

59

operating range

of, 49

surge limit

of, 59

Beams,

boundary

conditions for,

continuous beams,

142

Bins

arching

(rathole,

l-2,

6)

critical

dimension

for,

3,

12

critical flow factor

for,

4

critical hooper

dimensions, 6

dead storage,

1-2

degradation

flow

condition,

1

design of, reasons

for inefficiency,

1

flow,

erratic,

I

flushing of, 1

funnel

flow in,

1,

6,

8

hoop

pressure

in,

rnaximum,

6

hooper

angle,

3

mass flow

in,

1, 3-4,

6,

8,

11

piping,3

angle of internal friction, 3-4, 6-7

angle of friction, effective, 6*7

critical

flow factor

for,

7

piping

factor, 304

pneumatic gases

in, 7

pressure

vessels,

differences from,

1

segregation, 1

shear stress,

1

solid flow,

pressure

distribution for, 8

steady flow, consolidating

pressure

for, 3

structural

design,

conical

portions,

rectangular,

17

frame

detail,

20

stiffener design, 14-16

hoop

force,

16

stresses in, 13- 14

truss design, 18-20

wall friction

angle,

4-5

Blowers

and fans, 59

Bulk

solid

properties

bins, in, 1, 6

bulk density,

3,

6

critical dimensions

of,

3

pressure

of,

consolidating,

4, 6-7

stresses,

hooper wall, on, 3

solids,

in, 3

typical values oi 7

yield

strength, solid

material,

3

Centrifu

gal

compressors

actual, or inlet,

flow

rate,

80

advantages

of, 43-44

affinity laws, 50

237



Mechanical Design

of Process Systems

anti-surge devices for, 52

diagram

of,

53

applications of, 49

compressibility curves

for,

81

compressibility factor,

significance of, 83

compression

process,

diagram of, 50

compression

ratio

of,

50,

80-81

discharge temperature

average,80

dependence on ratio of

specific heats, 83

frame data, typical,

80

gas,

cyclic

vibration

of,

50-51

noise

induced by, 50-51

gas

inlet conditions,

50

impeller, 49

types

of,

52, 52

inlet

parameters,

effect

of

varying.

52

intercoolers,

sizing of, 50

mechanical

losses of, 82

percentage

of

power

required,

83

mixtures

compressibility factors

for, 79-81

specific

heats

for,

79

nncratinc

'arlo"

44

performance

curves, typical,

51

polytropic

head,

81

maximum

per

stage,

82-83

significance

of,

83

polytropic

relations

for,

46-50

pressure

versus capacity

for

constant speed compressor, 52

rpm, required, 82

selection

of, 79-83

shaft

power,

required,

expression

for,

82

single

stage, 49-50

specific

heat

ratio

significance

of, 83

stages,

required number of,

82

standard

cubic feet,

use of, 52

surge,50

control

of,

52

surge

limits, 50,

52

temperature,

discharge, 49-50

temperature

ratio for,

81

volumetric

flow,

expression for, 80

Centrifugal

pumps

advantages

of,

31

API

hydrometer,

conversion factors, table of,

92

defined,

ST-88

bearings, 34

outboard type, 34

brake horsepower, 34, 36, 70, 9l

required,96

shut-off, at,

36

by-pass for, 34, 36

casrngs,

horizontally

split,

32

vertically

split, 32

advantages

of,

32

components

of,

33

efficiency

of, 70

head, total, 36

heat

dissipation

in,

34, 36

intercooler for,

37

Hydraulic Institute,

68, 71-72

hydraulic requirements

of, 34, 36-37

impeller,

axial

flow

pump,

for,

32

mixed

flow

pump,

for,

32

vanes

of,

32

radial type, 32

volute of, 32

net

positive

suction

head

(NPSH)

definition of, 34

pressure pads

for,

91

Newtonian

fluids,

68

non-Newtonian fluids, 68, 79

packtng,

32

performance

curves

for,

34

typical, 69, 75, 95

pressure

drop

discharge

line, for,

67

-68,

9l, 95-96

friction factor for,

66-67

,

89-91, 93, 95-96

suction line for, 65-66,

90-91,93,95,97

viscosity, effects

of,

68,

70-72

seals,32

double seals

criteria for use, 32

types of, 35

seal

flush,

34

single seals

types

of,

35

versus double seals, 32

selection

of,

70

total dynamic

head, application of, 70, 74

types of, 31, 34-35

vaporization of

pumped

liquid,

causes of, 34

viscous liquids,

pumping

of, 37

correction-factor curves,

37,

38-39

criteria for, 37

equivalent water-performance

of, 37

water horsepower, 34, 36

defined, 36

Compression, ideal

gas

compressibility

factor

discharge, at,

45



mean,

45

suction,

at,

45

isentropic

(reversible adiabatic),

46-49

adiabatic

efficiencY,

46

energy,

isentroPic, 46

polytropic

efficiencY,

46

principles

of, ff

44-48

real

gas.

compressibility

factor.

44

Compressors

acfm,59-60

advantages

of,

59-60

conversion

to, standard

volumetric

flow,

60

actual

volumetric

flow.

See

acfm'

flow conditions,

sPecifYing,

59

actual,

or inlet

flow,

59

mass flow,

59

standard

volumetric

flow,

59-60

mass

flow,

conversion

to

standard

volumetric

flow,

60

principles

of

comPression,

44-48

scfm,

59-60

specifing

flow

conditions,

59

acfm,

exPression

for,

60

actual,

or

inlet

flow,

59

mass

flow,

59

specific

volume,

exPression

for,

60

standard

volumetric

f1ow, 59-60

standard

volumetric

flow

compressibilitY

factor,

59

conversion

to

actual

or mass

flow, 60

disadvantages

of,

60

specific

volume,

exPression

for,

59

'ttandard"

condition,

defined,

59-60

comparisons

of

various

forms,

60

volume

flow,

equation

for,

59

types

of,

43

volume

flow,

exPression

for,

59

External

loading

on

shell structures

applications

of

,

l7Q-17

5

"critical

value,"

170

shell

thickness,

170

Fans

and

blowers,

59

Flow

of solids,

problems of, 1-3

Gas

compressibility

tactor,

44

general

gas

law,

44

specific

heat

ratio

for,

44

universal

gas

constant,

44,

59

Gear

pumps, 37,

40

Heat

transfer,

convection

of,

air

normal

to cylindeq

126

Hydraulic

Institute,

37

Hydraulics

API

hydrometer

conversion

factors,

table

of,

92

defined,8T-88

Internal

pressure,

stress

concentration

factor,

169

lsentropic

comPression

brake

horsepower,

48

discharge

temperatue,

48

head,

adiabatic,

46

heat, mechanical

equivalent

of,

45

horsepower,

ratio

of

isentroPic,

45

horsepower

input

for single

stage,

45

ideal

eas,

45

adia--batic

efficiencY,

45

horsepower, isentropic,

45

mechanical

efficiencY,

45

overall

adiabatic

efficiencY,

45

multistage,46

perfect

gas,

formulations

for,

44

real

gas,

formulations

for,

45

isentropic

exPonent

for,

45-46

relations,

basic

versus

polytropic compression,

47

reversible,48

Jenike

and Johanson

method,

1-8

Lifting

lug

design, 170-175

choker

angle

for,

175

standard

designs

for,

171

L'Hospital's

rule,

165

Ingarithmic

mean

temperature

difference.

See

LMTD.

LMTD,

application

of,

148-149,

154, 160, 162'

t65

correction

factot

F,

117

-l2l

multipass

exchangers,

variance

in, 117

variance in

shell

and tube

heat

exchangers,

117

zero

LMTD

exchanger,

165

Multi-stage

reciprocating

compressors,

58

Non-Newtonian

fluids,

162

Nozzle

reinforcing

pads

disadvantage

of

pads,

170

pad

width,

maximum,

170

Nusselt

number,

125-126,

156

Petroleum

fractions

API hydrometer

for, 87-88

Plate-fin

heat

exchangers

advantages

of,

147



24O

Mechanical Design of Process

Systems

applications

of,

99

disadvantages

of, 147

illustrated, 149

Kays

and

London

coefficients,

148

thermal shock

and fatigue, 148

uses of, 147- 148

vacuum brazing of, 148

Polytropic

compression

efficiency

overall

polytropic,

48

polytropic

vs. isentropic, 46-47

gas

horsepower,

47

head,

adiabatic,

47

horsepower, compressor

(polytropic

head), 48

perfect

gas,

for, 47

polytropic

exponent,

46

polytropic

head

(compressor

horsepower), 48

real

gas,

for,

47

relations, basic versus isothermal compression, 47

Positive-displacement

pumps

applications

of,

31

brake horsepower, 77

definition

of, 31

efficiency

of,

77

pump

selection, use

in,

77

gear pumps,

37

,

40, 78

heat dissipation

in,

43

intercooler,

43

temperature switch,

43

net

positive

suction

head.

See

Pumps.

performance curves

for rotary

gear pumps,

79

pressure

drop

suction

line, 74

velocity

heads,

74

pressure protection

for, 42-43

priming

of,

79

reciprocating

pumps

diaphragm

pumps,

3l

piston pumps,

31

nlrrnocr nrrmnc 1l

rotary

pumps

cam

pumps,

31

gear

pumps,

31

lobe

pumps,

31

screw

pumps,31

types

of, 37

vane

pumps,

31

screw

pumps,

40-41

vane

pumps,

37

Prandtl

number,

125,152, 156,

164

Pulsation

response spectra

compression

bottles,

64,

65

typical,65

methods

of

predicting,

64

orifice

plates,

application of,

65

piping

system

excited by,

65

pulsation

bottles. See Compressor bottles.

pulsation

dampener.

See

Compressor bottles.

reciprocating equipment, induced by, 62,

&-65

Southwest

Research

Institute,

64

Structural Dynamics

Research

Corporation,

(scRc),

64

surge drums. See Compressor bottles.

Pumps

API degrees, defined, 87-88

calculation sheet

for, 36, 70,

77

flow

capacities

of,

34

head,

friction,

40

static discharge,

40

static suction, 40

total discharge,

40

total dynamic, 34, 40

total static, 40

total suction,

40

Hydraulic

Institute, 68,

7

|

-72

inline, nozzle loadings for,

61

lift

static suction,

40, 42

for water

maximum recommended, 43,

77

theoretical,

43,

77

total

suction,

40, 42

motors,

NEMA

frame dimensions,

73

NPSH

definition

of, 34

pressure

pads

for,

91

priming

of,

79

pump

Hydraulic

Design,

calculation

sheet,

36,

70,77, 93,95-96

pump

selection

guide,

32

types of, 3l

uses

of, 31

velocity

heads,

effect on

pumps,

40

Reciprocating

compressors

adiabatic

compression, work required

for,

58

adiabatic exponent,

53

adiabatic expressions

for, 44-46,

53

adiabatic

process,

57

applications

of, 43, 84-86

clearance capacity,

effect

of, 55

clearance

pockets,

43

stop

valve,

53

volumetric efficiency, effects on,

56

compressibility

factors

discharge,

58

inlet, 58



fr

lnder

API

618, 61

API

criteria,

61-62

NEMA.

See

Nozzle

Loadings.

nozzle

loadings

on, 61-62

allowable,

defined,

61-62

NEMA,61_62

applications

for,

61

options

to, basic,

62

steam

turbines,

ideal

expansion

joint,

64

turbo-expanders,

reasonable

values

for,

63

typical

for in-line

pumPs,

61

piping

systems

for,

60-65

pulsation bottles.

Se? Pulsation

response spectra.

steam

turbines,

piping

to,

62

surge

drums.

'gee

Pulsation

response

spectra

Rotary

pumps,

types

of, 37

Screw

pumps,

40-41

Shell-and-tube heat exchangers

advantages

of,

99

ASME Section

VIII

Division

I

Code, 99,

101

ASME

tube

joint

load

criteria,

1 13- 1

15

joint

reliability

factor.

I

l3-l14

maximum

tube

joint

force,

113

tube

joint

load, 113

baffle

cuts,

111

baffle details,

111

baffle

lanes,

channel

and

head, 128

baffle

plates,

99

baffle windows,

139

various schemes,

139

baffles

annular

orifices,

110

doughnut

and disc tYPes,

110

flow

direction,

used

for,

107

horizontally

cut,

107, 109

longitudinal,

109

structural

supports,

as,

107

verticaliy

cut, 107

vibration

dampers,

as,

107

baffle

windows,

Ill

basic components

of,

107

-112

caloric

temperature,

117

,

122-123,

158

Kern

relationships

for,

I22

caloric

versus arithmetic

rnean, 122

chlorine

superheater

design,

154- 160

chiller,

101

condenser,

101

deflexion

or

ligament efficiency,

158

design

classifications

of,

101

final condenser,

101

fixed tubesheet,

102-1O4

fixed tubesheet

design,

100

floating

heat exchangers

211

compressor

horsepower,

factors

affecting,

53

compression

ratio,

58,

84

compressor

bottles.

See Pulsation

response spectra.

cylinders,

size

of, 86

cylinder displacement,

86

diatomic

gases,

57

discharge

temperature,

85

efficiency,

volumetric,

86

Neerken

equation

for,

86

gas

temperature,

exPression

for,

58

horsepower,

theoretical,

58

parameters

affecting,

58

horsepower

per

million

curves,

85

correction

factors

for,

85

intercoolers

for, 84

multiple

staging

of, 58

advantages

of,

58

compression

ratio for,

84

cylinder

size,

58

cylinders,

number

of,

58

flywheei,

effect

on, 58

torque,

effect

on,

58

operating

range,

44

piston rod diameter,

86

polytropic

exponent,

57

Chlumsky

recommendations

for, 57

pressure-volume diagram,

56

ratios

of clearance

volume

to

volume

swept

by

piston,57

reciprocating

compressor

cycle,

53,

55

re-expansion

process,

57

schematic

of, 87

volumetric

efficiency

curves

for

determining,

57

expression

for,

53, 57

for a

perfect

gas,

57

parameters

that

affect,

53, 57

theoretical,53

Regenerated

gas

exchanger

design of,

148- 153

vibration

check, 153-

154

Reinforcing

pads

(external

loadings)

pad

width, maximum,

170

disadvantage

of

pads,

170

Reynolds

number, 9,

66-67,

7

4,

89

-91,

93, 95

-96,

t25-127,

140, 141,

l5l-152,

156-157,

1U

non-Newtonian

fluid,

Metzner-Reed,

162-163

versus

drag coefficients

for

long circular

cylinders,

r42

Rotating equipment

APr 611,61

APr

612,

61

API

617, 61



242

Mechanical Design

of

Process

Systems

internal floating

head design, 103-104

advantages of, 104

outside-packed floating head

design, 103-104

operating

range,

104

packed

latern

ring design,

103-1M

operating range,

1M

pull-through

bundle

design, 103- 104

limitations

of, 104

types

of, 103- 104

forced circulation reboiler,

101

fouling

resistances,

recommended

minimum,

125

friction

factors

for, shell-side

surfaces,

140

heat transfer

bulk temperature of fluid,

125

continuity

equation,

128

convection, basic

expressions

for, 115

factor

jH,

129,138,

152,

157

film

coefficients,

shell-side, 128

Kern

correlation, 128

fouling factors,

124

bare tubes

versus

finned

tubes,

124

definition of, 124

versus thermal

conductance, 124

fouling

resistance,

124

Fourier's law of

heat

conduction, 116

Grimson equation,

for

film

coefficient,

126

inside film coefficient, 122, 151

laminar,

125

turbuient,

125

laminar boundary

layer. 125

modes of, 115

McAdams correlation,

125

outside film coefficient,

lZZ,

126, 1,29

overall

heat

transfer

coefficient,

152

caloric, 117, 122, 152,

157,

158

parameter

jH,

129,

138

effective diameters

for, 129

versus Reynolds number,

138

shell-side

film

coefficient, 151-152,

156,

163-t64

tube-side film

coefficient,

151,

i54-156

tube

wall

resistance,

124

turbulent

boundary

layer,

125

impingement

baffles,

i28

latent heat, I 16- 117

ligament or deflexion

efficiency, 158

LMTD

correction factor

R

117- 121

multipass exchangers, variance in, 117

variance

of, 117

overall

heat

transfer coefficient, 122

caloric, 117, 122,

152

partial

condenser, 101

process

evaluation of,

115-140

reboiler,

99, l0l

kettle

type,

99

regenerated

gas

exchanger

design, 148-153

sensible

heat,

116-

117

shell-side,

defined, 99

shell-side

equivalent,

tube diameter, 129, 152,

156, 164

shell-side

pressure

drop, \39, 152-153,157,

164-165

expression for, 139, 152

shell-side mass

density, 151

shell-side mass flow

rate, G,, 139, 152-154, 156,

I O-l

Sieder-Thte correlation,

laminar

flow, for,

125,

162

turbulent flow, for,

125

steam

generator,

101

TEMA

class

B

exchanger,

99, lO4

class

C

exchanger,

99,

104

class R exchanger,

99,

104

comparisons

of

types, 105

mode constants

for

tubes,

112

natural frequencies

of

straight

tubes,

I 12- I l3

natural frequencies of U-tubes,

113

nomenclature of, 102

TEMA specification

sheet, 150, i55

tubes, stress, allowable compressive,

l12

tubesheets, compressive stress induced

OD,

lll

thermosyphon

reboiler, 101

tie rods

TEMA

recommendations for, 110

uses of,

110

tube arrangements,

pros

and cons of,

129

tube

bundle, 99, 126, 128

flow area

of, 152

Keys

and

London

constants

foq 129

tube

bundle cross-flow arca, 128

staggered

inline,

for,

128

triangular

layouts, for, 128

tube count tables,

130-

137

tube

geometry

angtlar

pitch,

126-127

diamond-square

pitch,

126

-

127

inJine square

pitch,

126-127

inJine triangular

pitch,

126-127

tubes

bare, 107

bend radii,

minimum, 109

boundary

layer,

125

laminar,

125

turbulent,

125

buckling

of



{

2rl:t

Euler

columl

formula,

114

exchanger

tubes, 113

Johnson short

column

equation,

1i4

finned,

107

foreign

deposits,

124

inside

film

coefficient,

122

outside

film

coefficient,

122

pitch,

nominal,

114

stress

factors for,

159- 160

tabulated

properties

of, 108

tubesheets, 99

double

tubesheets, 110

uses of,

110

maximum

radial

stresses

in, 159

single

tubesheets,

110

tubesheet-tube

connections,

typical,

I 1

1

tubesheet

layouts

staggered

in{ine,

for,

128

triangular

layouts,

for, 128

typical,

128

tube-side

defined,

99

tube-side mass

flow rate,

151,

162

tube vibrations.

See Tube

vibrations.

tube wall

temperature,

117,122,

124

U-tube

exchangers

kettle

type

reboiler,

100

tubesheet

fot

103

vaporizer,

101

vapor-liquid

equilibrium

calculations,

I

17

vertical

gas-gas

exchanger,

151

Silos.

See

Bins.

Specific

diameter,

48

versus

specific

speed,

49

Specific

speed,

48

versus

specific

diameter,

49

Stack

design

anchor

bolt torque,

26-27

base

support

detail

for,

27

carbon

precipitation

in,

8

buckling

stress

allowable,

22

deflection,

dynamic,

26

deflection,

static,

26

excitation,

flexural,

9

flexural

frequency,

9

lining

of,

8

effect

of,

8

gunite,8

modulus

of

elasticity

of,

8

Michell

and

Love

equation,

9,

28

ovaling,8-9

flexural

modes

of,

9

in-plane,

9

out-of-plane,9

modes of,

9

ovaling

frequency.

See

Flexural

frequencl

-

ovaling rings,

9,

26

natural

frequency

of,

9,

26

reasons

for,

9

section

modulus

of, required,

9

pressure

vessels,

vertical

differences

bef$'een.

8

seismic response

spectra,

8

vibration,

cantilev

er, 25

-26

vortex

shedding

frequency,

9, 26

vortex

strakes, 9-11,

27

-28

clearances

for,

11

critical wind velocities

for, 10

fabrication

detail

of, 11

fabrication,

method

of,

11

helix

angle

of, 10

length of, 10

Morgan equation,

10, 28

radius

of

curvature

of, l0

strake height,

10

range

for,

10

wind

design

anchor

bolt design

for,

23

bearing

pressure

for,

23

base

plate,

Brownell

and Young

method, 24

chair design,

Brownell

and

Young

method,

24-25

compression

rings, gusset plate

thickness,

required,25

effective diameters

for,

20

weld,

skirt-to-base

ring,

25

wind

load,

2l-22

wind

moment,

21-22

wind

pressure,

21

wind

response

spectra,

8

Steam

turbines

piping

of,

62

Strouhal

number,

9

Suction lift,

IOr WAIe\

+5, I

I

TEMA

class B

exchanger,

99,

104

class

C

exchanger,

99,

104

class

R

exchanger,

99,

104

heat exchanger

specification

sheet,

150-161

mode

constants

for

tubes, 112

natural

frequencies

of,

straight

tubes,

112- 113

U-tubes,

113

nomenclature

for

shell-and+ube

heat

exchangers.

102

standard,

TEMA,

99,

104

TEMA

types, composition

of, 105

tie rods,



244

Mechanical

Design

of

Process

Systems

recommendations

for,

1

10

uses

of,

110

tube

joint

load

formulations,

113

tubes,

minimum

bend

radii,

109

stress,

allowable

compressive, I

12

tubesheets,

turbulence

deflection,

root-mean-square,

145

joint

efficiency,

145

pressure

distribution for,

144-

145

response

spectra,

145

Wambsganss

and

Chen relation, 146

Venturi

effect,

144

Documents

Mechanical Design of Process System Volume-2(Shell and Tube Rotating Equipment)-Keith Escoe