VERTICAL PRESSURE VESSEL DESIGN 5.docx

debrebrihan university vertical pressure vessel design by tesema

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COLLEGE OF ENGINEERING AND TECNOLOGY
DEPARTMENT OF MECHANICAL ENGINEERING
COURSE NAME:- MACHINE DESIGN PROJECT 1
COURSE CODE:- MENG-3161
GROUP-5
GROUP MEMBERS ID NUMBER
1,DAGM AWOKE 485/14
2,TIHTNA ASMARE 499/14
3,ABATNEH TESFAW 477/14
4,ASHENAFI DACHEW 481/14
5,BEKALU FEKADE 482/14
6,TESEMA TESHOME 497/14

SUBMITTION DATE .../08/2014 E/C SUBMITTED TO MR. YESEHAK T.

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Table of Contents
NOMENCLATURE ................................................................................................................................4
ACKNOLEDGMEMT .............................................................................................................................5
ABSTRACTIVE .....................................................................................................................................6
CHAPTER ONE ....................................................................................................................................7
1 INTRODUCTION............................................................................................................................7
1.1Definition of Pressure Vessel ...................................................................................................7
1.1.2Classification of pressure vessel ............................................................................................7
1.2 Component of Pressure Vessel ...............................................................................................9
1.3Problem of statement ........................................................................................................... 14
1.4 General objective................................................................................................................. 14
1.5 Specific objective ................................................................................................................. 14
1.6 Scope of Project................................................................................................................... 15
1.7 Limitation the Project........................................................................................................... 15
CHAPTER TWO ................................................................................................................................. 16
BACKGROUND OF RESEARCH ......................................................................................................... 16
2.1 History of pressure vessel And Development ........................................................................ 16
CHAPTER THREEE.............................................................................................................................. 19
LITERATURE REVIEW ..................................................................................................................... 19
3.1 Literature review ................................................................................................................. 19
CHAPTER FOUR................................................................................................................................. 21
SKELETON PROCEDURE ................................................................................................................. 21
4.1 Methodology .............................................................................................................................. 21
CHAPTER FIVE .................................................................................................................................. 25
DESIGN CONSIDERETION ............................................................................................................... 25
5.1Material selection ................................................................................................................. 25
5.2 Operating Temperature and pressure.................................................................................. 26
5.3 Design stress(normal design strength)................................................................................... 27
5.4 Corrosion Allowance ........................................................................................................... 28
5.6 Minimum Thickness ............................................................................................................. 28
5.7 Welded Joint Efficiency ........................................................................................................ 29
5.8 Factor of safety.................................................................................................................... 29

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CHAPTER SIX .................................................................................................................................... 30
DESIGN PROCEDURE ..................................................................................................................... 30
6 DETAILED DESIGN ANALYSIS ................................................................................................... 30
6.4 Design of components.......................................................................................................... 35
Wall thickness of the pipe. ....................................................................................................... 41
6.4.7 Design vessel support ........................................................................................................ 45
6.5 Design of welded joint.......................................................................................................... 54
Factor of safety ......................................................................................................................... 55
CHAPTER SEVEN ........................................................................................................................... 55
CONCLUSION ................................................................................................................................ 55
REFERENCES ..................................................................................................................................... 56

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NOMENCLATURE
Pi…………….…………………internal pressure
Pd………………………………Design pressure
??????………………………………Longitudinal stress inMpa
???h……………………………circumferential (hoop) stress in Mpa
di………………………….…….internal diameter in meter
t…………………………..……..thickness in mm
L………………………………… length of pressure tangent to tangent in mm
FR……………………………… The total resistive force
FT……………………………….The total transverse force
Td………………………Design temperature
Dm………………..Diameter of man hole
Td…………………design temperature
S………………………..allowable stress
Rm……………….radius of man hole
D…………………diameter
A…………………..Area
F………………….force
V…………………volume
Pin…………………………………….internal pressure
Dm………………diameter of man hole
W………………weight

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ACKNOLEDGMEMT

We would like to thank our course Instructor Mr. Yesehak T. to give this project for our group .
This project provides a good knowledge or detail information about design pressure vessel with
choosing of a proper material.

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ABSTRACTIVE



This project work deals with a detail study and design procedure of vertical pressure with a
semi spherical head bracket support with manhole . a detail study of varies parts of pressure
vessel like shell, head ,support ,nozzle the detail study of pressure vessel with properly selected
material from varies materials used in pressure vessel construction is discussed . it also deals
with the study varies parts like head with appropriate shape (semi spherical) and others are
clearly shown.

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CHAPTER ONE
1 INTRODUCTION
1.1Definition of Pressure Vessel
A pressure vessel is a closed container designed to hold gases or liquids at a pressure
substantially different from the ambient pressure. It also defined as a container with a pressure
differential between inside and outside. The inside pressure is usually higher than the outside,
except for some isolated situations. The fluid inside the vessel may undergo a change in state as
in the case of steam boilers, or may combine with other reagents as in the case of a chemical
reactor. Pressure vessels often have a combination of high pressures together with high
temperatures, and in some cases flammable fluids or highly radioactive materials. Because of
such hazards it is imperative that the design be such that no leakage can occur. The legal
definition of pressure vessel varies from country to country, but often involves the maximum
safe pressure (may need to be above half a bar) that a vessel is designed. [1]
1.1.2Classification of pressure vessel
Pressure vessels can be classified in different categories as follows:-
a) According to the end construction
This can be classified in to two groups;
І. Open end construction pressure vessel-In this case, the hoop stress is induced by the fluid
pressure. Example-a simplecylinder with piston, Such as cylinder of press.
ІІ. Closed end construction pressure vessel-In this case, longitudinal stress in addition to hoop
stress are induced.
b) According to dimension
The pressure vessel, according to their dimension, may be classified as:
І .Thin-walled pressure vessel-if the ratio of is less than, then the pressure vessel is called
thin1walled vessel. In this case, the radial stress is small and it can be neglected and the
longitudinal and the circumferential stress are constant.
ІІ. Thick-walled pressure vessel- if the ratio of is greater than or equal to, it is called thick-
walled vessel.in this case, the radial stress can’t be neglected and the other two stress are not
constant. Thick-walled pressure vessels are mostly used in case of high pressure such as guns,
barrels, high pressure cylinder etc.
c). According to geometrical shape
The pressure vessel, according to their geometrical shape, can be classified as follow
i. Cylindrical shape

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ii. Conical shape
iii. Spherical shape with one or two cone
c) According to position arrangement
The pressure vessel, according to their position arrangement, may be classified as follow:
i. Vertical pressure vessel
ii. Horizontal pressure vessel
iii. Spherical pressure vessel
A). Vertical pressure vessel:
Vertical pressure vessels are used for a variety of operational needs, mainly as reservoirs of
compressed air - air chambers, as well as pressurized water tanks or expansion tanks to
compensate for the volume of hot water stations with air or steam cushion - aqua mat, and as
releasers called expanders.
Vertical pressure vessels consist of a cylindrical shell and dished bottoms. They are placed on three
welded legs. The size and positioning of the filler necks is adjusted according to the customer
requirements and accordance with the relevant European standards. The pressure vessels can be
produced from ferrous or austenitic steel.
b) Horizontal pressure vessel:-
The free horizontal pressure vessel in pressure vessel is a container designed to hold gases or
liquids at a pressure substantially different from the ambient pressure.
C) Spherical pressure vessel:-
This type of pressure vessels are known as thin walled vessels. This forms the most typical
application of plane stress. Plane of stress is a class of common engineering problems involving
stress in a thin plate. Spherical vessel have the advantage of requiring thinner walls for a given
pressure and diameter than the equivalent cylinder. Therefore they are used for large gas or liquid
containers, gas cooled nuclear reactors, containment buildings for nuclear plant,
1.1.3Practical use of Pressure Vessel
Pressure vessel is the container for fluid under high pressure. It can store fluid such as liquid
vapor and gas under pressure .pressure vessel had been used in various water treatment
application like mixed bed exchanger, activated carbon filters ,sand filters, dual medical filters
etc. with internal rubber lining carbon steel internal including strainers and screen laterals
,ladders and plate form etc… Pressure vessel is the container for fluid under high pressure.
Pressure vessel has a variety of application. This includes the industry and the private sector.

They appear in this sector as-
*Industrial compressed air reservoirs
* Domestic hot water storage tank

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* Autoclave
* Distillation tower
* Recompression chamber
* Diving cylinder
* Oil refineries and petrochemical plants
* Nuclear reactor vessels
* Submarine and space ship habitats
* Pneumatic reservoirs
* Hydraulic reservoirs under pressure
* Rail vehicle airbrake reservoirs
* Road vehicle airbrake reservoirs
* Storage vessels for liquefied gases such as ammonia, chlorine, propane, butane,
and LPG
N.B -Not only in the above sectors, pressure vessel are used almost in all industries and in
home.

1.1.4Area of application of Pressure Vessel
Pressure vessels are used to store fluids, such as liquids, vapors and gases under pressure
vessels. That means used to for different human activities in modern worlds.
Major uses of pressure vessels are:-
* In brewery and soft drink factories
*power generation industry for fossil and nuclear power
*In pharmaceutical factories
*In oil refineries plant
* In dye factories
*In engine cylinders.
*In medical sterilization system.
* Food production facilities.
* In steam boilers
Generally speaking, pressurized equipment is required for a wide range of industrial plant for
storage and manufacturing purposes

1.2 Component of Pressure Vessel
The major components of pressure vessel are listed below:
1,Head
2,Shell
3,Nozzle
4,Support

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5,man hole
1, Head
All pressure vessel shells must be closed at the ends by heads (or another shell section).Heads is
typically curved rather than flat.
There are four types of head:
A, Flat Plates and Formed Flat Heads.
B, Hemispherical Head
C, Ellipsoidal Heads.
D,Tori-spherical Heads.
Flat plates are used as covers for access ports or manholes, and as the channel covers of heat
exchangers. StandardTori-spherical heads are the most commonly used end closures for vessels up
to operating pressures of 15 bars. They can be used for higher pressures, but above 10 bars, their
cost should be compared with that of an Ellipsoidal head. Above 15 bars, an Ellipsoidal head will
usually prove to be the most economical closure to u Hemispherical head is the strongest shape,
capable of resisting about twice the pressure of a Tori-spherical head of the same thickness. The
cost of forming a hemispherical head will, however, be higher than for a shallow Tori-spherical
head. Hemispherical heads are used for high pressure.
2, Shell
The shell is the primary component that contains the pressure.
Pressure vessel shells are welded together to form a structure that has a common rotational
axis. Most pressure vessel shells are cylindrical, spherical, or conical in shape. Horizontal drums
have cylindrical shells and are fabricated in a wide range of diameters and lengths.


3, Nozzle
A nozzle is a cylindrical component that penetrates the shell or heads of a pressure Vessel. The
nozzle ends are usually flanged to allow for the necessary connections and to permit easy
disassembly for maintenance or access.
Nozzles are used for the following applications:-
. Attach piping for flow into or out of the vessel.
. Attach instrument connections, (e.g., level gauges, thermos wells, or pressure gauges).
. Provide access to the vessel interior at man-holes.
. Provide for direct attachment of other equipment items, (e.g., a heat exchanger or mixer)

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4, Support
The method used to support a vessel will depend on the size, shape, and weight of the vessel;
the design temperature and pressure; the vessel location and arrangement; and the internal
and external fittings and attachments. Horizontal vessels are usually mounted on two saddle
supports. Skirt supports are used for tall, vertical columns. Brackets, or lugs, are used for all
types of vessel. The supports must be designed to carry the weight of the vessel and contents,
and any superimposed loads, such as wind loads. Supports will impose localized loads on the
vessel wall, and the design must be checked to ensure that the resulting stress concentrations
are below the maximum allowable design stress. Supports should be designed to allow easy
access to the vessel and fittings for inspection and maintenance.
Typical kinds of supports are as follow:-
A. Skirt support
B. Leg
C. Saddle
D. Lug
A. Skirt Support
Tall, vertical, cylindrical pressure vessels are typically supported by skirts. A support skirt is a
cylindrical shell section that is welded either to the lower portion of the vessel shell or to the
bottom head (for cylindrical vessels). Skirts for spherical vessels are welded to the vessel near
the mid-plane of the shell. The skirt is normally long enough to provide enough flexibility so
that radial thermal expansion of the shell does not cause high thermal stresses at its junction
with the skirt.
B. Leg support
Small vertical drums are typically supported on legs that are welded to the lower portion of the
shell. The maximum ratio of support leg length to drum diameter is typically 2:1. The number
of legs needed depends on the drum size and the loads to be carried. Support legs are also
typically used for spherical pressurized storage vessels. The support legs for small vertical
drums and spherical pressurized Storage vessels may be made from structural steel columns or
pipe sections, whichever provides a more efficient design.

C. Saddle support
Horizontal drums are typically supported at two locations by saddle supports. A saddle Support
spreads the weight load over a large area of the shell to prevent an excessive local stress in the
shell at the support points. The width of the saddle, among other design details, is determined
by the specific size and design conditions of the pressure vessel. One saddle support is normally
fixed or anchored to its foundation.
D.Lug support
Lugs that are welded to the pressure vessel shell, which are shown on, may also be used to
support vertical pressure vessels. The use of lugs is typically limited to vessels of small to
medium diameter (1 to 10 ft.) and moderate height-to diameter ratios in the range of 2:1 to 5:1.
Lug supports are often used for vessels of this size that are located above grade within
structural steel. The lugs are typically bolted to horizontal structural members to provide

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stability against overturning loads; however, the bolt holes are often slotted to permit free
radial thermal expansion of the drum.
5,Man hole
It is identical to a nozzle except it does not belt to piping and it has a cover plate which is belted
to the flange.
1.2.1Special Components of Pressure Vessel
1.6.1 Flange
Flanged joints are used for connecting pipes and instruments to vessels, for manhole covers,
and for removable vessel heads when ease of access is required. Flanges may also be used on
the vessel body, when it is necessary to divide the vessel into sections for transport or
maintenance. Flanged joints are also used to connect pipes to other equipment, such as pumps
and valves. Screwed joints are often used for small-diameter pipe connections, below 40 mm.
Flanged joints are also used for connecting pipe sections where ease of assembly and
dismantling is required for maintenance, but pipework will normally be welded to reduce costs.
Flanges range in size from a few millimeters diameter for small pipes, to several meters
diameter for those used as body or head flanges on vessels.
Types of flange, and selectionSeveral different types of flange are used for various
applications. The principal types used inthe process industries are:
1. Welding-neck flanges.
2. Slip-on flanges, hub and plate types.
3. Lap-joint flanges.
4. Screwed flanges.
5. Blank, or blind, flanges.
1,Welding –neck flange have a long tapered hub between the flange ring and the welded joint. This
gradual transition of the section reduces the discontinuity stresses between the flange and branch,
and increases the strength of the flange assembly. Welding-neck flanges are suitable for extreme
service conditions; where the flange is likely to be subjected to temperature, shear and vibration
loads. They will normally be specified for the connections and nozzles on process vessels and
process equipment.
2,Slip-on flanges: slip over the pipe or nozzle and are welded externally, and usually also internally.
The end of the pipe is set back from 0 to 2.0 mm. The strength of a slip-on flange is from one-third
to two-thirds that of the corresponding standard welding-neck flange. Slip-on flanges are cheaper
than welding-neck flanges and are easier to align, but have poor resistance to shock and vibration
loads. Slip-on flanges are generally used for pipe work.
3,Lap-joint flanges: are used for piped work. They are economical when used with expensive alloy
pipe, such as stainless steel, as the flange can be made from inexpensive carbon steel. Usually a
short lapped nozzle is welded to the pipe, but with some schedules of pipe the lap can be formed on
the pipe itself, and this will give a cheap method of pipe assembly. Lap-joint flanges are sometimes
known as “Van-stone flanges”

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4,Screwed-flanges: are used to connect screwed fittings to flanges. They are also sometimes used
for alloy pipe which is difficult to weld satisfactorily.
5,Blind flanges (blank flanges): are flat plates, used to blank off flange connections, and as covers
for manholes and inspection ports.

1.6.2 Gaskets
Gaskets: are used to make a leak-tight joint between two surfaces. It is impractical to machine
flanges to the degree of surface finish that would be required to make a satisfactory seal under
pressure without a gasket.
Gaskets are made from “semi-plastic” materials; which will deform and flow under load to fill the
surface irregularities between the flange faces, yet retain sufficient elasticity to take up the changes
in the flange alignment that occur under load.
A great variety of proprietary gasket materials is used, and reference should be made to the
manufacturers’ catalogues and technical manuals when selecting gaskets for a particular
application. The minimum seating stress y is the force per unit area (pressure) on the gasket that is
required to cause the material to flow and fill the surface irregularities in the gasket face. The
gasket factor m is the ratio of the gasket stress (pressure) under the operating conditions to the
internal pressure in the vessel or pipe.
The internal pressure will force the flanges’ faces apart, so the pressure on the gasket under
operating conditions will be lower than the initial tightening-up pressure. The gasket factor gives
the minimum pressure that must be maintained on the gasket to ensure a satisfactory seal. The
following factors must be considered when selecting a gasket material
The process conditions: pressure, temperature, corrosive nature of the process fluid.
Whether repeated assembly and disassembly of the joint is required.
The type of flange and flange face
Up to pressures of 20 bars, the operating temperature and corrosiveness of the process fluid will be
the controlling factor in gasket selection. Vegetable fiber and synthetic rubber gaskets can be used
at temperatures of up to 100oC. Solid polyfluorocarbon (Teflon) and compressed asbestos gaskets
can be used to a maximum temperature of about 260oC. Metalreinforced gaskets can be used up to
around 450oC. Plain soft metal gaskets are normally used for higher temperatures.
1.6.3 Pressure gages:
Pressure gages are instruments for measuring the condition of a fluid (liquid or gas) that is
specified by the force that the fluid would exert when at rest on a unit area, such as N/c.m2 .

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1.6.4 Temperature gages:
are instruments for measuring the condition of a fluid (liquid or gas) that is specified by the
temperature that the fluid would exert when at rest on a unit area, such as N/c.m2 .

1.3Problem of statement

In human life pressure vessels are the main important devices in different application for our
life in different areas like, house hold, factories, industries &other for storage of fluid and
chemicals. Even if the yare essential in our survival, the occurrence not questionable task. The
failure of pressure vessel is very devastating & serious thing since it leads to several economic
losses, environmental pollution, and danger of life. If its failure leads to these effect, the cause of
the failure is our focus to minimize it as much as possible. So the cause of failure is mostly
design problem which includes: - improper material selection (material defect), not considering
external factors (temperature, pressure, acidic rain etc…), not considering the type of fluid it
store (water, steam, alcohol, benzene or gaseous), and following improper design procedure.
These is not the only cause but corrosion and fatigue effect is also the major problem of failure
Corrosion fatigue is fatigue in a corrosive environment. It is the mechanical degradation of a
material under the joint action of corrosion and cyclic loading. Nearly all engineering structures
experience some form of alternating stress, and are exposed to harmful environments during
their service life. So such things are generally leads to the pressure vessel to fail.
1.4 General objective
The main objective of this project is to design an advanced, problem solving, long lived vertical
pressure vessel by considering the specification

1.5 Specific objective

Specifically, we would like to design vertical positioned pressure vessel with hemispherical
head closure and supported by skirt by considering the effect of internal pressure, temperature
and other real-timed design consideration based on the given design specification by using
currently used design codes and standards of designing pressure vessel.
Specifically, we would like to design support, head, shell and nozzle and it has its own
procedures to design each component and to design the hole assembled of pressure vessels.
The specific objective of the project is to design a vertical pressure vessel subjected to the
following parameters.
Fluid content-------------------------------- any fluid
Maximum tensile stress -------------------80MPa
Internal diameter---------------------------1.4m
Corrosion Allowance-----------------------1.6 mm
Orientation ----------------------------------Vertical
Welding

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1.6 Scope of Project
Pressure vessel used in various field and plays a significant role in our life. So the method of
construction, the serving time length, the factor which distort them and their causes are the
main things that we must observe critically.
We know that mostly pressure vessels fail in short time due to several factors like, corrosion,
design problem, variable environmental condition, and such a like. Considering these problem
we should construct vessel which serve the customer for long time as much as possible. These
can be done designing pressure vessel considering such problems and also giving greater factor
of safety (allowance) to the construction in order to avoid physical damage. Generally our scope
is producing advanced pressure vessel which is resistant to corrosion and failure due to any
other effect like temperature & pressure change by designing properly.
1.7 Limitation the Project
However, we design properly some factors may lead the work un-functional. From these
factors effect of corrosion play the great role due the reason that pressure vessel are all in
contact with water which give great opportunity for rusting, these is impossible prevent
completely. And other extremely variable factors change with time to time and place to place
are difficult to control easily in design. So this things would be the limitation of our.

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CHAPTER TWO
BACKGROUND OF RESEARCH
2.1 History of pressure vessel And Development
Perhaps the earliest reference to the design of pressure vessel was made in about 1495 by
Leonardo da Vinci in his codex Madrid I. Quoting from a translation, Leonardo wrote “we shall
describe how air can be forced under water to left very heavy weights, that is, how to fill skins with
air once they are secured to weights at the bottom of the water. And there will be descriptions of
how to lift weights by tying them to submerged ships full of sand and how to remove the sand from
the ships.” Leonardo’s pressurized bags of air, if implemented did not kill or injure large number of
people and did not force the need for a pressure vessel code.
Numerous boiler explosions took place through the late 1800s and early 1900s. This led to the
enactment of the first code for construction of steam boilers by the Commonwealth of
Massachusetts in 1907. This subsequently resulted in the development and publication of the ASME
Boiler and Pressure Vessel Code in 1914, which sought to standardize the design, manufacturing,
and inspection of boilers and pressure vessels. In 1921 the National Board of Boiler and Pressure
Vessel Inspectors was organized to promote consistent inspection and testing. The publication of
the section on locomotive boilers also appeared in 1921. The ASME and the ASTM (American
Society for Testing and Materials) material specification merged in 1924. The first publication of
Section VIII ‘‘Unfired Pressure Vessels,’’ appeared in 1925. This document was referred to as one of
a theoretical factor of safety of 5. The petroleum industry did not consider it to be adequate for
their purposes and also desired better utilization of available materials. The year 1928 saw the
advent of welded pressure vessels. For higher pressures the welded shells were made thicker than
70 mm. These required nondestructive examination (NDE) before service. In 1934, a joint API–
ASME Committee published the first edition of an unfired pressure vessel code specifically for the
petroleum industry. In 1952 these two separate codes merged into a single code – the ASME
Unfired Pressure Vessel Code, Section VIII. The ASME Pressure Vessel Code, Section VIII Division 2:
‘‘Alternative Rules for Pressure Vessels,’’ was published in 1968 and the original code became
Section VIII Division 1: ‘‘Pressure Vessels.’’ A considerable boost was provided to the understanding
of the basic behavior of pressure vessel components following the development of the nuclear
power program in the U.S. and Europe in the late 1950s and early 1960s. Similar developments can
be found in the British, French, German and Japanese codes, to name but a few. By 1960 the need
for a code for pressure vessels for commercial nuclear plants became imperative. This resulted in
publication of the 1963 Edition, Section III: ‘‘Nuclear Pressure Vessels.’’ This was a design by
analysis code with a theoretical safety factor of 3. After the publication of Section III: ‘‘Nuclear
Pressure Vessels’’ in 1963, it was necessary to modify Section VIII for general pressure vessels.
ASME.[2]

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Code Section VIII Division 2: ‘‘Alternate Rules for Pressure Vessels’’ appeared as a result and
provided a theoretical factor of safety of 3. In 1971, Section III: ‘‘Nuclear Power Components’’ were
classified as (a) pumps, (b) valves, and (c) piping. The stress limits for emergency and faulted
conditions were introduced. In addition, the addenda of 1971 added storage tanks. The addenda of
summer 1972 introduced Appendix G on nonductile failure. The Appendix F on evaluation of
faulted conditions was included in the addenda of winter 1972. The design of component supports
and core support structures appeared in the addenda of winter 1973. ASME Section III Division 1 is
devoted entirely to nuclear power components and also contains the rules for the design of nuclear
pumps and valves. The recognition of concrete reactor and containment vessels led to the
publication of the Section II Division 2 code in 1975. Three subsections (NB, NC and ND) of ASME
Section III Division 1 cover the design and construction of equipment of Classes 1, 2, and 3,
respectively. The most stringent is Class 1, which requires design by analysis. Class 2 permits
design by analysis as well as the use of formulas. Class 3 prescribes design by formula, and is
equivalent to Section VIII Division 1. The designer evaluates the safety function of each pressure
vessel and applies the appropriate code class. Design of supports for Section III Division 1 vessels
are not prescribed in the ASME Code. Section III has a subsection NF, which prescribes the design of
supports for Class 1, 2, and 3 pressure vessels. The addenda of winter 1976 changed the
nomenclature of design, normal, upset, testing and faulted conditions to level A, B, C and D service
conditions. In the 1982 addenda, the fatigue curves were extended to 1011 cycles. In the 1996
addenda, the design rules for high-temperature service were incorporated. In 1976, Division 3 was
published which contained rules on transport of irradiated materials. The need for uniform rules
for in-service inspection of nuclear power plants led to the issuance of the 1970 edition of Section
XI: ‘‘Rules for In-service Inspection of Nuclear Plant Components.’’
The organization of the ASME Boiler and Pressure Vessel Code is as follows: 1. Section I: Power
Boilers 2. Section II: Material Specification: i. Ferrous Material Specifications – Part A ii. Non-ferrous
Material Specifications –Part B iii. Specifications for Welding Rods, Electrodes, and Filler Metals –
Part C iv. Properties – Part D 3. Section III Subsection NCA: General Requirements for Division 1 and
Division 2 i. Section III Division 1: a. Subsection NA: General Requirements b. Subsection NB: Class 1
Components Copyright 2005 by CRC Press, Inc. All Rights Reserved. c. Subsection NC: Class 2
Components d. Subsection ND: Class 3 Components e. Subsection NE: Class MC Components f.
Subsection NF: Component Supports g. Subsection NG: Core Support Structures h. Appendices:
Code Case N-47 Class 1: Components in Elevated Temperature Service ii. Section III, Division 2:
Codes for Concrete Reactor Vessel and Containment 4. Section IV: Rules for Construction of Heating
Boilers 5. Section V: Nondestructive Examinations 6. Section VI: Recommended Rules for the Care
and Operation of Heating Boilers 7. Section VII: Recommended Guidelines for Care of Power Boilers
8. Section VIII i. Division 1: Pressure Vessels - Rules for Construction ii. Division 2: Pressure Vessels
– Alternative Rules 9. Section IX: Welding and Brazing Qualifications 10. Section X: Fiberglass-
Reinforced Plastic Pressure Vessels 11. Section XI: Rules for In-Service Inspection of Nuclear Power
Plant Component.
The rules for design, fabrication and inspection of pressure vessels are provided by codes that have
been developed by industry and government in various countries and are indicated in Table 1.1.
The design and construction codes all have established rules of safety governing design, fabrication

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and inspection of boilers, pressure vessels and nuclear components. These codes are intended to
provide reasonable protection of life and property and also provide for margin for deterioration in
service also includes the ASME Boiler and Pressure Vessel Code. Some of the significant features of
the latest version of the ASME Code Section III are: Explicit consideration of thermal stress
Recognition of fatigue as a possible mode of failure The use of plastic limit analysis Reliable
prediction of ductile failure after some plastic action. In addition there is a continuous attempt to
understand all failure modes, and provide rational margins of safety against each type of failure.
These margins are generally consistent with the consequence of the specific mode of failure. A word
or two about the impact of technological advances in pressure vessel design should be mentioned.
The last three decades have seen great strides made in the improvement of digital computations. In
the 1960s the use of computers began to make an impact on design and analysis ofpressure vessels.
The rapid development of finite-element software has remarkably impacted the detailed design of
pressure vessel components. These developments along with continuing increase in computing
speed and storage capacity of the computer have really made the design process extremely quick
and at the same time have led to very accurate design assessment. Initially in the early to mid-
1970s, detailed finite-element analyses were generally performed for confirmatory analyses. Today
these tasks are routinely accomplished in an interactive mode. The three dimensional finite-
element analysis programs using solid elements are rapidly replacing plate, shell, and two-
dimensional programs for routine structural design analysis of pressure vessels. In addition the
concepts of computer-aided design (CAD) and computer-aided manufacturing (CAM) are being
integrated.
In spite of some of the most rigorous, well-conceived safety rules and procedures ever put together,
boiler and pressure vessel accidents continue to occur. In 1980, for example, the National Board of
Boiler and Pressure Vessel Inspectors reported 1972 boiler and pressure vessel accidents, 108
injuries and 22 deaths.2 The pressure vessel explosions are of course rare nowadays and are often
caused by incorrect operation or poorly monitored corrosion. Safety in boiler and pressure vessels
can be achieved by:
* Proper design and construction
* Proper maintenance and inspection
* Proper operator performance and vessel operation
The design and construction cures are dependent upon the formulation and adoption of good
construction and installation codes and standards. Thus the ASME Pressure Vessel Code requires
that all pressure vessels be designed for the most severe coincident pressure and temperature
expected during the intended service. There can be no deviation from this requirement, even if the
severe condition is short term and occurring only occasionally. Bush has presented statistics of
pressure vessels and piping failures in the U.S., Germany and the UK.[3] He has concluded that a 99
percent confidence upper boundary for the probability of disruptive failure to be less than 1 X10^–
5 per vessel year in the U.S. and Germany. According to his study, periodic inspection is believed to
be a significant factor in enhancing pressure vessel reliability, and successful applications of ASME

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Boiler and Pressure Vessel Codes (Sections I and VIII) are responsible for the relatively low re
responsible for in life.
Pierre and Baylac authored an international perspective of the design of pressure vessels in 1992.4
They recommend that the governing authorities be vigilant by constantly monitoring accident
statistics. They also insist that the authorities be prudent and maintain a flexible attitude in
enforcing regulation.[4]












CHAPTER THREEE
LITERATURE REVIEW
3.1 Literature review

M.A Khan et al [6] carried out his research on Stress distribution in horizontal pressure vessel and
saddle supports. Her quarter of the pressure vessel is modelled. After that stress distribution is
carried out for pressure vessel he concluded that highly stressed area is the flange plate of saddle.
M. Javed Hyder et al [7] made research on optimization of location and size of opening in a pressure
vessel cylinder using ANSYS. Analysis is performed for three thick walled cylinder with different
internal diameters. From the research it is concluded that location and size of the hole depends
upon the size of the cylinder. The optimum location is where von Misses stress is, minimum and

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also the whole size should be such that von-Misses stress should be minimum around the vicinity of
the hole.
M. Giglio et al [8] presented his research on Fatigue analysis of different types of pressure vessel
nozzle. He carried out comparison of two different methods for the construction of pressure vessel
nozzle. He conclude that failure of nozzles was carried out by crack passing through their
thickness. Both designs (external and integral reinforcement) give good fatigue life results but
nozzle with external reinforcement is easy to produce than with integral one. Choice of simply
produced nozzle obtained with a cylindrical tube and a reinforcement plate gives good results in
terms of stress and fatigue life.
Avinash R.Kharat et al [9] carried out his research on analysis of stress concentration at opening in
pressure vessel using ANOVA. The motivation for this research is to analyze the stress
concentration occurring at the openings of the pressure vessels and the means to reduce the effect
of the same. Finally conclusion shows that sudden change in strain flow lines causes the strain and
stress to rise abruptly. Strain increases with increasing the opening size in the geometry.
Bandarupalli Praneeth et al [10] carried out research on finite element analysis of pressure vessel
and piping design. The stresses developed in solid layer pressure vessel and multilayer pressure
vessel are analyzed. Here, theoretical and ANSYS results are compared. Finally they conclude that
theoretical calculated values are very close to that of the values obtained from ANSYS is suitable for
multilayer stress concentration fact.
Shyam R. Gupta et al [11] presented their work on the design and analysis of pressure vessel using PV
Elite software. Due to mathematical calculation designing of pressure vessel becomes tedious but by
using software like PV Elite designing of pressure vessel can be done easily. By using this software finally
they conclude that Due to mathematical calculation designing of pressure vessel becomes tedious but
by using software like PVElite designing of pressure vessel can be done easily.
M. Jeyakumar et al [12] researched on “Influence of residual stresses on failure pressure of cylindrical
pressure vessel. Here a main criterion is to study the effect of residual stresses on the pressure vessel.
Conclusion shows that there is a reduction in failure pressure due to unfavorable residual stresses.
Z.ModiAJ,Jadav C.S[13] concluded that the radial stresses in case of hemispherical head pressure vessel
is low compared to other types of head, in this paper author study the comparative structural behavior
of different types of geometry of pressure vessel, the head is under internal uniform pressure, the
analytical and finite element method used for finding stresses in pressure vessel, the aim is finding best
head for specific parameter with finite element analysis of thin cylindrical pressure vessel, here three
types of geometry consider like hemisphere, flat and ellipsoidal and computation result compared with
finite element analysis.
From the literature review it is cleared that study of the different factors that affects the design and
overall life of pressure vessel and they have to be considered in different cases of pressure vessel.
Finite element analysis is an extremely powerful tool for the consideration of pressure vessel. A
structural consideration of the pressure vessel will be implemented. From above literature review -

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Inclination angle, circular cross section with hemispherical ends, fatigue analysis, stress
distribution, fracture analysis, optimization of location and size of opening in a pressure vessel
cylinder are the main criteria among all that have to be considered on the design and analysis of
pressure vessel to avoid any failure of pressure





CHAPTER FOUR
SKELETON PROCEDURE
4.1 Methodology
Methodology is a general guide line or procedure which, are used to solve problems to design
our project which is, design of unfired pressure vessel. We will try to follow the following
procedure.
4.1.1 Identification of Need
The information in selection of pressure vessel is described and the application of selected pressure
vessel is been discussed. To design of pressure vessel the selection of Code are important as a
reference guide to achieve the secure pressure vessel. The selections of ASME Code Section VIII div
1 are described. The standard of material selection used are explains in this chapter. Beside of that,
the design and analysis software to obtain the result are introduced. Instead of that, design process
methodology is also described
1) Volume of the fluid to be contain, the total volume or the volume of the fluid within the pressure
vessel either filled or partially filled.
2) Operating pressure, the pressure at which the vessel subjected, this means It is recommended to
design a vessel and its parts for a higher pressure than the operating pressure. A design pressure
higher than the operating pressure with 10 percent, whichever is the greater, will satisfy the
requirement. The pressure of the fluid will also be considering. The maximum allowable working
pressure (MAWP) for a vessel is the permissible pressure at the top of the vessel in its normal
operating position at a specific temperature. This pressure is based on calculations for every

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element of the vessel using nominal thicknesses exclusive of corrosion allowance. It is the basis for
establishing the set pressures of any pressure-relieving devices protecting the vessel.
3) Operating temperature, the temperature of the fluid at which it operates. Design temperature is
the temperature that will be maintained in the metal of the part of the vessel being considered for
the specified operation of the vessel. For most vessels, it is the temperature that corresponds to the
design pressure. However, there is a maximum design temperature and a minimum design
temperature (MDMT) for any given vessel. The MDMT shall be the lowest temperature expected in
service or the lowest allowable temperature as calculated for the individual parts. Design
temperature for vessels under external pressure shall not exceed the maximum temperature
4) Nature of the fluid, can be volatile, reactive, flammable or not to the environment or to the wall of
the pressure vessel.
5) Major fluid connection, the values that connected and acts inlet and outlet.
6) Design of general shape (cylindrical, spherical, and conical) and orientation (if it is vertical or
horizontal).
7) External load: earth quick wind load Failure of the vessel checked by the theory of the pressure
vessel using the maximum shear stress theory, analysis of these variables are design methodology
for pressure vessel.
4.1.2 Definition of problem
We are asked to design vertical pressure vessel which has a capacity to store steam without a great
failures of material or parts.
4.1.3 Analysis and optimization
Selecting suitable material, Geometric, force, cost and stress analysis for each component, Such as:
shell, head, nozzle, flange, bolt and skirt support, Checking the maximum principal and shear stress
with the permissible stress of material.
4.1.4 Selection of material
The selection of material for the design is the first requirement based on the design specification.
There are many consideration to select the material for designing purposes based on considering
the fluid/ medium, the working pressure (internal pressure) and the temperature at the design
pressure vessel. To select the proper material the following criterion should be considered:-

* Availability of the material
*Suitability of the material for working consideration such as pressure, temperature and strength.
*Cost of the material
*performance characteristics (properties)

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* processing (manufacturing) properties
*environmental profile
*economical consideration Material manufacturing process in pressure vessel can use many
types material as designed and customer wants. In our pressure vessel data the working material is
low alloy steel with a strength data at a given working temperature.



Table : Selected material for pressure vessel
Material Tensile strength in Mpa Design of stress at temperature
Low alloy Steel 550 235

Many pressure vessels are made of steel. Some mechanical property of steel achieved by rolling or
forging could be affected by welding unless special precautions are taken.
In addition to mechanical strength current standard dictate the use of steel with high impact
resistance especially for vessel in low temperature. In application where carbon steel is used it
would suffer corrosion so special corrosion resistant material must be used.
Pressure vessel may also be constructed from concrete (PVC) or other material which are weak in
tension Cabling wrapped around the vessel or with in the wall in the vessel itself provide the
necessary tension to resist the internal pressure .there is also a high order of redundancy tank to
the large number of individual cables resisting internal pressure.
Generally pressure vessel may be made from
 Steel on its alloys
 Aluminum and its alloy
 Other metals
 Carbon fiber
4.1.6 Design Parameter
Table design specification
Parameter value unit

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Tensile stress 80 Mpa
Internal diameter 1.4 m
Working fluid Any fluid
Corrosion allowance 1.6 mm
Vessel orientation vertical

4.1.7 Design procedures
* Determine the minimum plate thickness to resist the internal pressure
* Select the size and shape of vessel head (semispherical)
* Increase the basic plate thickness to allow for the bending stress induced by the wind loading at
the base of the vessel and due to wind loading.
* Check the maximum combination stress at the base area with the design stress and check that for
critical buckling stress
* Decide which opening need for compensation
* Use standard flanges
* Design of the support type bracket
* Design the base ring
* Welding

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CHAPTER FIVE
DESIGN CONSIDERETION
5.1Material selection
Several of material have been use in pressure vessel fabrication the selection of material is based on
the appropriateness of the design requirement. All the material used in manufacture of the
receivers shell company with the requirement of the relevant design code and the identification
with mill sheet. The selection of material of the shall take into suitability of the material with the
maximum working pressure and fabrication process .
Thus ,the reason for selecting material can be divided into two generally :-
1, commercial factors of the material such as:-
A, cost
B, availability
C, ease of manufacture
2, Engineering properties of material such as:-
A, strength
B, corrosion resistance
C, fracture toughness
D, fabric ability

A, strength
The strength of the material based on the mechanical properties of the material that used for the
easiest ,fast , strong and low cost of production ,those material abilities to resist some of the
mechanical properties are yield strength ,ultimate tensile strength , creep strength ,
Yield strength
Yield strength is the region which a material change from plastic to elastic deformation , some
materials selected for pressure vessel are shown blow and there yield strength , were the stress
level begin the plastic deformation

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Tensile strength
The ultimate tensile strength (Tensile stress) is a measure the of the basic strength of material ,it is
the maximum stress that material with stand and measure by a standard tensile test.
Creep
Creep is the gradual extension of a material under a steady tensile stress ,over a prolonged period
of time . it is usually only important at high temperature slowly creep under stress loading well
below the level yield point deemed safe in the tensile stress over a prolonged period of time.
B, Corrosion Resistance
Corrosion resistance is partially or completely wearing away, dissolving or softening of any
substance by chemical or electrochemical reaction with it’s environment the term corrosion
specifically applies to the gradual action of natural agent ,such as air or salt water on metals.
C, Fracture Toughness
The ability of the material that absorb energy up to fracture .brittle fracture with out applicable
deformation and by rapid crack propagation the direction of the crack motion is very nearly
perpendicular to the direction of the applied tensile stress and yield stress related with fracture
surface. Brittle fracture is depend on the stress concentration and stress concentration factor.
D, Fabric ability
Based on how it made and cost at production. A guide to the fabrication properties of common
method for metals and alloys.

5.2 Operating Temperature and pressure
5.2.1 Design pressure
A vessel must be design to with the stand the maximum pressure to which it is likely to be
subjected in operation for vessel under internal pressure the design pressure is normally taken as
the pressure at which the device is set this well normally be 3 to 10 above the normal working
pressure to avoid spurious operation during the minor process upset when design pressure the
hydrostatic pressure in the basic of column should be add to the operation pressure it significant.
Vessel subjected to external pressure should be design to resist the maximum differential pressure
that likely to excess in the service vessel likely to be subjected to vacuum should design for equal
negative pressure of the bar. Unless fitted with effective and reliable vacuum breaker
Pd=Pi(5-10%)Pi where Pd= design pressure Pi=internal pressure

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5.2.2 Design temperature
The strength of material decrease with increase temperature , so the maximum temperature ,the
design temperature at which the design stress is evaluated should be taken as the maximum
working temperature of due to allowable for un certainty involving in producing vessel wall
temperature. Decrease in metal strength with a rising temperature ,the lost temperature for metal
design used pressure storage vessel should be taken as 15 degree F . the design temperature for
flange through bolt is usually lower than temperature of operating fluid, unless insulated and it can
be safety assume to be 80% of the vessel temperature.
5.3 Design stress(normal design strength)
For design purpose it is necessary to design a value for the maximum allowable stress (normal de
sign strength) that can be accepted in the material of construction .this is determined by appling a
suitable design stress factor to maximum stress expected to with stand out failure understand test
condition. The design stress factor allow for any un certainty is in the design method the loading the
quality of the material for material not subjected to high temperature the design stress is based on
the proof stress or the tensile strength(ultimate tensile stress) of the material at the design
temperature ,for material subjected to condition at which the creep characteristics of the material.
The average stress to produce rupture after 10^5 hours or the average stress to produce 21%
strain after 10^5 hours at the design temperature.
By considering the given temperature and the selected material the yield stress will be known. We
design by using yield stress of the material. In cylinder there are coordinate systems that describe
the state of stress
The hoop stress circumferential/tangential (σh )
the radial stress
The longitudinal stress ( )
Principal stress are the maximum or minimum order defined.
 Analysis stress in the thine walled vessels; with internal pressure loading
 From thin shell cylinder theory of stress condition.
σh =
PiD
2t
where Pi=internal pressure
B) Longitudinal stress (σL ) sss
σL=
σh
2

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5.4 Corrosion Allowance
The corrosion allowance is the additional thickness of metal added to allow for materials lost by
corrosion scaling . The allowance should be agreed in between customs and manufactures.
Corrosion is a complex phenomena and it is not possible to give rules for the estimation of the
corrosion allowance should be based on experience with the material construction under similar
service condition to those for the proposed design for carbon and low alloy steel, where several
corrosion is not expected minimum allowance of 1.6mm should be used where more conditions
are anti spected. ,
5.5 Design load
A structure must be designed to resist gross plastic deformation and collapse under the condition
loading. the load which a pressure vessel will be subjected to service are listed. they can be
classified as major load that can be consider in design vessel and subsidiary load. The formal stress
analysis to determine the codes and standards where it is not possible to demonstrate the adequate
of the proposed design other means such as compression with the known behavior of existing of
vessel.
Major load
1,design pressure:- include any significant static head of liquid.
2, maximum weight of the vessel and contents under operating list of condition.
3 , maximum weight of the vessel and contents under hydraulic list condition.
4,windload,long vessel with very with small t/r values are subject to distraction from wind
pressure.
5.6 Minimum Thickness
The Rule be a minimum thickness required to insure that any vessel is sufficiently rigid with stand
its own weight and any incident load.
A general guide any vessel thickness should not be less than value given below
Vessel diameter in (m) Maximum thickness
1 5
1 to 2 7
2 to 2.5 9
2.5 to 3 10
3 to 3.5 12

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5.7 Welded Joint Efficiency
The strength of welded joint efficiency will depend on the type of joint and the quality of welding,
The soundness of weld is checked by visual inspection and by non-destructive testing. The possible
lower strength of welded joint compared with plate is usually allowed for design by multiplying the
allowable design stress for the material by welded joint factor. The value of the joint factor used in
design will depend on the type of joint and amount of the radiography required by the design code,
Taking the factor as 1, Implies as the joint is equal as strong as the virgin plate, This is achieved by
radiography.
The complete weld length and cutting out and remarking any defects, the use of lower joints factor
in design, through solving costs on radiography will result in thicker, heavier vessels and the
designer must balance any costs on result solving on inspection fabrication against the increased
cost of the material.
The national code and standards divide vessel construction in to different categories, depending on
the amount of non-destructive testing required the higher categories, requires 100% radiography
of the welds and allows the use of highest value for the weld joint factors. The lower quality
categories required less radiography, but allow only lower joint efficiency factors and plate
restriction on the plate thickness and the type of material that can be used.
The same limitation in the lower construction categories, the three standard specifics in the
construction categories,
Category1Thehigher class requires 100% non-destructive of the welds and that allow the use of all
materials covered by the standard with no restriction on the plate thickness.
Category 2)Requires non-destructive testing but place some limitation on the materials which can
be used and the maximum plate thickness.
Category 3)The lowest cost required only visual inspection of the weld, but restricted to carbon and
carbon manganese steels and austenitic stainless steal and limits are placed on the plate thickness
and the nominal design stress.

5.8 Factor of safety
Factor of safety or design stress factor allows for any un certainty in the design method the
loading ,quality of material. For material it is not subjected to high temperature the yield stress
(proof stress) or the tensile strength (ultimate tensile stress) of the material at the design
temperature.
For effective design the factor of safety is grater than 5.

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CHAPTER SIX
DESIGN PROCEDURE
6 DETAILED DESIGN ANALYSIS
The given and assumed Specification and their corresponding
values of our design
Parameter value
Design pressure …………………… …………………………….3.5MPa
Design temperature ………………..…. …………………………300
0c
Vessel orientation ……………………………………………….vertical
Cover shape ……………………………......................................semi spherical
Support type …………………………………………………. bracket
Total inside diameter…………………. ………………. …. …….1.4m
Corrosion allowance ……………… …………………………….1.6mm
We have to give recommendation about the design of vessel. State the problems during
designing of the pressure vessel.
6.1 Design pressures
For vessels under internal pressure, the design pressure is normally taken as the pressure
at which the relief device is set. This will normally be 5 to 10 per cent above the normal
working pressure, to avoid spurious operation during minor process upsets. When deciding
the design pressure.
PD=Pi + Pi× PD=
100?????? + 10??????
=100PD=110Pi 100 Pi=
100??????-----------------------from the given
parameter PD=3.5MPa

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100×3.5MPa=110×PiMPa=350MPa
Pi= = 3.18????????? where Pi=Internal pressure
Pi=3.18Mpa
If internal is greater than 300psi or
PD > 0.385×SE then the wall of the cylinder is thick, otherwise
PD< 0.385×SE or less than 300psi, the wall of the cylinder is thin.
1????????? = 6894.8??????
??? = 3.5?????????
??? > 300psi, therefore thick walled pressure vessel is employed.
6894.8??????
6.2 Design temperature
The strength of metals decreases with increasing temperature so the maximum
allowable design stress will depend on the material temperature. The design temperature
at which the design stress is evaluated should be taken as the maximum working
temperature of the material, with due allowance for any uncertainty involved in predicting
vessel wall temperatures.
Then at T=300℃
From typical design stress table find the stress for low alloy steel:-
Design stress: - 235 N/mm
2
Tensile strength: - 550 N/mm
2
6.3 Design stress
For design purposes it is necessary to decide a value for the maximum allowable stress
(nominal design strength) that can be accepted in the material of construction. The design
stress factor allows for any uncertainty in the design methods, the loading, the quality of
the materials, and the workman ship.
For materials not subject to high temperatures the design stress is based on the yield stress
(or proof stress), or the tensile strength (ultimate tensile stress) of the material at the
design temperature.
6.3.1Welded joint efficiency, and construction categories
The strength of a welded joint will depend on the type of joint and the quality of the
welding.

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The soundness of welds is checked by visual inspection and by non-destructive testing
(Radiography).The possible lower strength of a welded joint compared with the virgin plate
is usually allowed for in design by multiplying the allowable design stress for the material
by a "Welded joint factor" J. The value of the joint factor used in design will depend on the
type of joint and amount of radiography required by the design code.


Table 1: welded joint factor [1]
In our design we use joint efficiency value E=1
The above table is above the maximum allowable joint efficiency .taking the factor as 1.0
implies that joint is equally as strong as a virgin plate. This is achieved by radio graphing
complete weld length. Cutting out and remarking any defects. The use of lower joint factor
factors in design through saving cost in radiography will result in any cost saving on
inspection and fabrication against the increased cost of materials.
6.3.2 Corrosion allowance
The additional thickness of metal added to allow for material lost by heat, corrosion and
erosion or scaling. The allowance to be used should be agreed between the customer and
manufacturer’s corrosion is complex phenomena, and it is not possible to give specific rules
for the estimation of corrosion allowance required for circumstances. The allowance should
be based on the experience with the material construction under similar surface conditions
to those for the purposed design. For carbon and low alloy steel with several condition are
anticipated, this should be increased to 4mm most design codes and standard specify a
minimum allowance of 1mm.
For carbon and low alloy steel where sever corrosion is not expected the minimum
allowance given that 2mm should be used. I.e. CA=2mm
Two procedures are method 1 and 2 that determining the diameter and the length of vessel.
Method 1:
Where; Pi=internal pressure=3.18MPa
CA=corrosion allowance=2mm

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E = Joint efficiency=1
S=allowable stress =235MPa

F1= Pi/CSEF1&F2=vessel ratios

Method 2:
F2=C [SE/PD – 0.6]



Table 2


Diameter for different
??? ratios
???

???

???
D
3
??? ??????
√
??????
4
??? ?????????
√
?????????
5

??? ??????
√
??????

We can select method is preferable then method one because it gives more accurate
diameter and length

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Figure 1 [19]


2
&#3627408447;
&#3627408439;

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Figure 2[20]
6.4 Design of components
6.4.1 Design of shell
According to geometrical shape of shell classified in to three thus are:-
 Cylindrical shell
 Conical wall
 Spherical vessel with one or two cones
We are Hemispherical shape because these vessels are:-

• Easy to manufacture and install economical to be maintain
• The cylindrical shall stress will be greater than that of the two principal stress
To calculate diameter, length and thickness of the shell use the following data and formulas
Table 3 L/D approximation table
Pressure(psi) L/D ratio
0-250 3

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250-500 4
>500 5

To determine L/Di ratio we use both vessel ratio and total inside volume reading from
chart:
From the table our design pressure is greater than 500psi we choose L/Di=5
Then L=5Di=5x1.4m=7m
L=7m

Where V = Volume of vessel in ???
3
Di = Internal diameter of shell in m
L = Length in m

Vs=&#3627409163;&#3627408439;??????
2
L/4
vs=3.14 x 1.4x 1.4 x7/4
Vs=10.77m
3

Where Vs=volume of shall
D = Internal diameter of shell in m
L = Length in m
Vh=
2
3
&#3627409163;&#3627408453;??????
3

Vh=
2
3
x3.14x(0.7x0.7x0.7) m
3
Vh=0.7m
3
Vt=2Vh+Vs=2x0.7m
3 + 10.77m
3
Vt=12.17m
3
Where Vh =volume of head
Vt=volume of total

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Stress analysis of semi spherical shell

The design thickness is the minimum required thickness plus an allowance for shrinkage.
Now, we can calculate the thickness of the plate by considering the welding efficiency and allowable
stress
We use double butt joint with weld efficiency E=1

From ASME standard formula of thickness for cylindrical shell is calculated for circumferential and
longitudinal cases and larger is taken as follows;

t=
&#3627408451;&#3627408453;??????
2&#3627408440;&#3627408454;+0.4&#3627408451;??????
…………………. for longitudinal

table 4 relation of thickness


Thickness, t Thickness,
T
Pressure, P
Part

Internal
diameter
Outer
diameter
Internal
diameter
Outer
diameter
Stress, ??????
Formula
Shell
For
circumferential

??????&#3627408453;??????
&#3627408454;&#3627408440;−0.6??????

??????&#3627408453;&#3627408476;
&#3627408454;&#3627408440;+0.4??????


&#3627408454;&#3627408440;&#3627408481;
&#3627408453;??????+0.6&#3627408481;

&#3627408454;&#3627408440;&#3627408481;
&#3627408453;&#3627408476;−0.4&#3627408481;
??????ℎ=
??????&#3627408453;&#3627408474;
&#3627408481;

Shell for
longitudinal
??????&#3627408453;??????
2&#3627408454;&#3627408440;+0.4??????

??????&#3627408453;&#3627408476;
2&#3627408454;&#3627408440;+1.4??????


2&#3627408454;&#3627408440;&#3627408481;
&#3627408453;??????−0.4&#3627408481;

2&#3627408454;&#3627408440;&#3627408481;
&#3627408453;&#3627408476;−1.4&#3627408481;
??????&#3627408473;=
??????&#3627408453;&#3627408474;
0.2&#3627408481;

Head ??????&#3627408453;??????
2&#3627408454;&#3627408440;−0.2??????

??????&#3627408453;&#3627408476;
2&#3627408454;&#3627408440;+0.8??????


2&#3627408454;&#3627408440;&#3627408481;
&#3627408453;??????+0.2&#3627408481;


2&#3627408454;&#3627408440;&#3627408481;
&#3627408453;&#3627408476;−0.8&#3627408481;


σl= σh=
PRm
2t

t=
&#3627408451;??????&#3627408453;??????
&#3627408454;&#3627408440;−0.6&#3627408451;??????
…………………….for circumferential
where
Pi= internal pressure= 3.18Mpa
S= allowable stress for selected material= 250Mpa
E= welding efficiency=1
R= radius of the shell = 0.7m
D=inner diameter=1.4m

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MACHINE DESIGN Page 38

Then t=
&#3627408451;??????&#3627408453;??????
2&#3627408440;&#3627408454;+0.4&#3627408451;??????
…………………. for longitudinal
t=
3.18Mpax0.7m
235Mpax1+0.4∗3.18Mpa

t= 9.42mm…. for longitudinal
t=
&#3627408503;??????&#3627408505;??????
&#3627408506;??????−&#3627409358;.&#3627409364;&#3627408503;??????
…………………….for circumferential
t=
3.18Mpax0.7mm
235Mpax1−0.6x3.18Mpa

t=9.54mm………….. for circumferential
so, we take one the thickness t=9.54mm
Then corrosion allowance for steel pipes of compressed air is 1.6mm
tshall= t +CA
tshell = 9.54mm+1.6mm= 11.14mm ……. thickness of shell
the outer diameter of shell is given by
Do=Di+2t
=1.4m+2x9.54mm
Do=1.4191m
from the thickness diameter ratio, we have
if
t
D
>
1
10
…..it is called thick cylinder
if
t
D
<
1
10
…… is called thin cylinder
there fore
t
D
=
11.14mm
1.4m
=0.0079<0.1, because of this the shell is thin.
It is true, so we can conclude that, the pressure vessel is thin shell pressure vessel
 From thin shell cylinder theory of stress condition.
σh =
PiD
2t
=
3.18Mpax1.4m
2(11.14mm)
=199.82Mpa
Hoop stress = 199.82Mpa< 235Mpa a Design is safe
σh=2 σL
σL=
σh
2
=
199.82
2
=99.91Mpa
Longitudinal stress = 99.91Mpa< 235Mpa Design is safe
τmax=
??????ℎ+??????&#3627408447;
2
=
199.82&#3627408448;&#3627408477;&#3627408462;+99.91&#3627408448;&#3627408477;&#3627408462;
2

τMAX=149.865Mpa
then the design is safe because maximum shear stress is less than 235Mpa

39. Vertical pressure vessel

MACHINE DESIGN Page 39

6.4.2 DESIGN OF HEAD
All pressure vessel shell must be closed at the end by heads. The ends of cylindrical vessel
are closed by head various shapes. This are:-
 Flat plates head
 Hemispherical head
 Ellipsoidal head etc.
our design head is hemi spherical(semi spherical) head

Our given head has hemispherical shape .so to design it we start from calculating its
thickness by using the following steps:

In these section the main design parameters are the following
 Thickness of the head
 Stress checking at different point
 Design pressure determination
thead =
&#3627408503;??????&#3627408505;&#3627408528;
&#3627409360;&#3627408506;??????+&#3627409358;.&#3627409366;&#3627408503;??????
………… for hemispherical head Ro=Do/2=0.7096m

thead=
&#3627409361;.&#3627409359;&#3627409366;??????&#3627408529;?????? &#3627408537; &#3627409358;.&#3627409365;&#3627409358;&#3627409367;&#3627409364;&#3627408526;
&#3627409360;(&#3627409360;&#3627409361;&#3627409363;??????&#3627408529;??????∗&#3627409359;)+&#3627409358;.&#3627409366;(&#3627409361;.&#3627409359;&#3627409366;??????&#3627408529;??????)
=4.78mm
thead =4.78mm
where Pi= internal pressure
Ro= outer radius of pressure vessel
S= allowable stress
E= welding efficiency
Do = outer diameter
Then

now we have to determine the maximum pressure that the hemispherical head can carry and we
use their pressure in other component. The outer pressure must become;
Po = 3.18Mpa.
Now we will check pressure;
 Next, we will check the longitudinal and circumferential stress at different position. With the
allowable stress of the material for failure of the head due to the internal stress.
 At the center of the head
Po=
2&#3627408454;&#3627408440;&#3627408481;
&#3627408453;&#3627408476;−0.8&#3627408481;
=
2??????235??????4.78
0.7096−0.8×3.35
=3.178Mpa≅3.18&#3627408448;&#3627408477;&#3627408462; Mpa=Pi
Then when our pressure is equal I.e. Pi=Po= 3.18Mpa,
Therefore our design is safe.

40. Vertical pressure vessel

MACHINE DESIGN Page 40

Stress Head analysis

σl= σh=
PRm
2t
, but Rm =
Ri + Ro
2
=
0.7&#3627408474;+0.7096&#3627408474;
2
=0.7048&#3627408474;&#3627408474;, Rm=704.8mm
Where Rm= mean radius
σh= σl=
3.18&#3627408448;&#3627408477;&#3627408462; ??????704.8&#3627408474;&#3627408474;
2(4.78&#3627408474;&#3627408474;)
=234.44&#3627408448;&#3627408477;&#3627408462;
τmax=
??????ℎ+??????&#3627408447;
2
= 234.44Mpa < 235Mpa
so, it is less than allowable stress, then it is safe at the center due to tangential load.
σl= σh = 234.44Mpa <235MPa

but, when we check for all values of σl= σh not more than the material allowable stress; therefore,
our design is safe.
6.4.3DESIGN OF PIPES
Inside diameter of the pipe.
The inside diameter of the pipe depends upon the quantity of fluid to be delivered.
Let Dp = Inside diameter of the pipe,
Q=Area x velocity=&#3627409163;&#3627408439;&#3627408477;
2
/4 x v
Dp=1.13√(
&#3627408452;
&#3627408483;
)
v = Velocity of fluid flowing per minute, and
Q = Quantity of fluid carried per minute.
We know that the quantity of fluid flowing per minute,
&#3627409164;&#3627408484;&#3627408462;&#3627408481;&#3627408466;&#3627408479;=1000&#3627408472;&#3627408468;/&#3627408474;
3
where t= time taken in minute
M=&#3627409164;&#3627408457;t
M=1000*12.17(
&#3627408446;??????
&#3627408474;
3
∗&#3627408474;
3)

M= 12170kg
a=velocity/time= L/t
2
F= ma =
&#3627408474;&#3627408473;
&#3627408481;
2
, but F=????????????∗&#3627408436;=
??????&#3627408439;
2
4
x Pi=
3.14∗0.15
2
∗3.18∗1000000
4
=56,166.75N
&#3627408481;
2
=
12170kgx7m
56166.75N
=1.517&#3627408480;&#3627408466;&#3627408464;&#3627408476;&#3627408475;&#3627408465;

41. Vertical pressure vessel

MACHINE DESIGN Page 41

&#3627408483;=
7&#3627408474;
0.0205 &#3627408474;??????&#3627408475;&#3627408482;&#3627408481;&#3627408466;
=341.46
&#3627408474;
&#3627408474;??????&#3627408475;&#3627408482;&#3627408481;&#3627408466;
then Q = A*v =
??????&#3627408439;
2
4
∗341.46=6.031
Dp = 1.13√
&#3627408504;
&#3627408509;
= 1.13*0.1328m = 0.1501m=150.1mm
Wall thickness of the pipe.
After deciding upon the inside diameter of the pipe, the thickness of the wall (t) in order to
withstand the internal fluid pressure (p) may be obtained by using thin cylindrical or thick
cylindrical formula. The thin cylindrical formula may be applied when
(a) the stress across the section of the pipe is uniform,
(b) the internal diameter of the pipe (D) is more
than twenty times its wall thickness (t), i.e.
D/t > 20, and
(c) the allowable stress (σ??????) is more than six
times the pressure inside the pipe (p),
i.e. σ&#3627408481;/p > 6. ??????t=
&#3627408451;&#3627408439;
2&#3627408481;
,&#3627408481;=
&#3627408451;&#3627408439;&#3627408477;
2??????&#3627408481;
, t=
3.18&#3627408448;&#3627408477;&#3627408462;×150.1&#3627408474;&#3627408474;
2×235
= 1.016mm
twall=1.016mm
6.4.4 DESIGN OF MANHOLE
Manhole is an open area that the person can enters through it for the purpose of working in the
interior part of the vessel. Why because when the vessel needs operation the man easily enters and
he also do his practical work that the vessel needs.
It is the assumption hole that designed by considering the man’s size.
Components of manhole to be designed;
 Thickness
 Hoop stress
 Nominal diameter
Where tmh = thickness of manhole
rm = radius of manhole
dm =diameter of manhole
??????&#3627408439; = &#3627408465;&#3627408466;&#3627408480;??????&#3627408468;&#3627408475; &#3627408477;&#3627408479;&#3627408466;&#3627408480;&#3627408480;&#3627408482;&#3627408479;&#3627408466;
Allowable stress, S
Joint efficiency, E = 1

42. Vertical pressure vessel

MACHINE DESIGN Page 42

&#3627408481;&#3627408474;ℎ =
??????&#3627408465;∗&#3627408465;&#3627408474;
2&#3627408454;&#3627408440;−1.6??????&#3627408465;

??????ℎ =
&#3627408451;&#3627408439;
2&#3627408455;&#3627408474;
, but tmh =
3.5∗&#3627408439;&#3627408474;
2(235∗1)−1.6(3.5)
= 0.007536 dm
??????ℎ=
3.5&#3627408448;&#3627408477;&#3627408462;∗&#3627408465;&#3627408474;
2∗0.007536dm
=232.197&#3627408448;&#3627408477;&#3627408462;< 235Mpa which is safe. Then we assume the diameter of the
manhole is 700mm=70cm, tmh=0.007536Dm = 0.007536*700 = 5.3mm
A=&#3627409163; dm*tmh = &#3627409163;∗700∗5.3&#3627408474;&#3627408474;
2
=11649.4&#3627408474;&#3627408474;
2
=0.0116494&#3627408474;
2

6.4.5 DESIGN OF BOTTOM DISCHARGE
Pipe loads are the net forces and moments exerted on equipment.
notations;
Design pressure, PD=3.5MPa
nominal diameter of pipe, Dp=150.1mm
joint efficiency, E=1
thickness of bottom discharge, &#3627408481;&#3627408463;=unknown
nominal radius bottom discharge, Rb= 75.05mm
&#3627408481;&#3627408463; =
&#3627408451;&#3627408439; &#3627408453;&#3627408463;
2&#3627408454;&#3627408440;−0.8&#3627408451;&#3627408439;
+&#3627408438;&#3627408436;=
3.5&#3627408448;&#3627408477;&#3627408462;(75.05&#3627408474;&#3627408474;)
2(235∗1)−0.8(3.5&#3627408448;&#3627408477;&#3627408462;)
+&#3627408438;&#3627408436;=0.562&#3627408474;&#3627408474;
&#3627408481;&#3627408463; = 0.562mm, though our thickness of nozzle is 0.562mm, we must be adding some corrosion used
as shrinkage resistance. Let c= 1.6mm
&#3627408481;&#3627408463; = 0.562mm add 1.6mm of shrinkage resistance
&#3627408481;&#3627408463; = 1.6mm+0.562mm =2.162mm
therefore, the outside diameter of bottom discharge pipe is given;
Dbd=Dp+2tb = 150.1mm+2(2.162) mm=154.424mm
then find mass of the bottom discharge water by taking the length of bottom discharge pipe
assumption Let assume the length of pipe is 85mm
&#3627408474;&#3627408463;=v×&#3627409164; where: &#3627408474;&#3627408463;=mass of bottom discharge
&#3627408474;&#3627408463; = (A×&#3627408447;)&#3627409164;&#3627408484;&#3627408462;&#3627408481;&#3627408466;&#3627408479;
&#3627408474;&#3627408463; = &#3627409163;
&#3627408439;p
2
4
×85&#3627408474;&#3627408474;×
1000&#3627408472;??????
&#3627408474;
3
=1.503&#3627408472;&#3627408468;/&#3627408474;
3

according specification, the discharge pipe is attached on the bottom of vertical pressure vessel, the
pipe attached on the shell of cylinder thickness of

43. Vertical pressure vessel

MACHINE DESIGN Page 43

Where
Z –section modulus
Z=
1
6
×&#3627408447;×&#3627408481;&#3627408463;
2
=
1
6
×85×(2.162)²= 66.2&#3627408474;&#3627408474;
3

Where &#3627408458;&#3627408463; = force exerted on the discharge
Mb = bending moment
&#3627408458;&#3627408463; =Pi×&#3627408436; = 3.18Mpa×
??????&#3627408439;&#3627408477;
2
4
=3.18&#3627408448;??????&#3627408462;×
??????×150.1&#3627408474;&#3627408474;
2
4
=56241.66&#3627408449;
Mb =
??????&#3627408463;??????&#3627408447;
4
=
56241.66×85&#3627408474;&#3627408474;
4
=1195135.275&#3627408449;&#3627408474;&#3627408474;=1195.14&#3627408449;&#3627408474;
let as calculate the extension ?????? of nozzle inside the vessel: -
??????=2.5 tb
?????? =2.5*2.162mm=5.405mm
The hoop stress induced in the pipe: -
where
??????ℎ=
&#3627408451;??????&#3627408439;&#3627408463;
2&#3627408481;&#3627408463;
=
3.18&#3627408448;&#3627408477;&#3627408462;×150.1&#3627408474;&#3627408474;
2(2.162&#3627408474;&#3627408474;)
=110.389&#3627408448;&#3627408477;&#3627408462;
??????ℎ is less than ??????&#3627408481;=235&#3627408448;&#3627408477;&#3627408462; Then the design is safe

6.4.6 DESIGN OF NOZZLE
In this section the main design parameters are: -
Finding required thickness of nozzle and Finding the required area
Nozzle design for 150mm nominal diameter (&#3627408439;&#3627408475;) opening pipes on the head (two longitudinal
pipes), We select 10 % for more safety purpose or from the standard design of pressure vessel.
Because our nozzle or pipe may be affected by shrinkage. ??????&#3627408465;= Pi+ 10%Pi+0.1(3.18Mpa) =3.5Mpa
Where &#3627408439;&#3627408475; = nominal diameter of nozzle
&#3627408481;&#3627408475; =nozzle thickness
??????&#3627408465; = design pressure
&#3627408453;&#3627408475; = nozzle radius
σℎ=
&#3627408451;&#3627408465;&#3627408439;&#3627408475;
2&#3627408481;ℎ
=
3.5&#3627408448;&#3627408477;&#3627408462;×150&#3627408474;&#3627408474;
2×4.78&#3627408474;&#3627408474;
=54.92&#3627408448;&#3627408477;&#3627408462; ……… its less than design stress so our design is safe

44. Vertical pressure vessel

MACHINE DESIGN Page 44

then, the required thickness will be
&#3627408481;&#3627408475;=
&#3627408451;&#3627408439;&#3627408453;&#3627408475;
&#3627408454;&#3627408440;−0.8&#3627408451;&#3627408439;
+CA=
3.5&#3627408448;&#3627408477;&#3627408462;×150&#3627408474;&#3627408474;
235&#3627408448;&#3627408477;&#3627408462;×1−0.8×3.5&#3627408448;&#3627408477;&#3627408462;
+1.6mm = 3.86mm


Required Area of the nozzle(An)
An= ??????Dn×tn=??????(&#3627409359;&#3627409363;&#3627409358;mm×3.86mm)=1818.06mm
2 =0.00181806m
2

Reinforcement of opening
Around the opening the vessel must be reinforce with an equal amount of metal which has been cut
of the opening. The reinforce may be an integral part of the vessel and nozzle or may be an
additional reinforcement.
Areas of reinforcement required
For vessel under internal pressure the total cross-sectional area required for reinforcement of
opening shell not be less than;
The reinforcement scheme is shown in fig below the reinforcement area required is
Ar=Dn*t n
=150mm*3.86mm
Ar= 579mm
2
The reinforcement area available in the shell (up to a distance), A1 is given by
A1= (2Dn-Dn) (ts- tsl) = Dn (ts-tsl)
= 150mm (11.14mm-9.42mm) =258mm
2
The reinforcement area available in nozzle wall is available in to two parts A21and A22
A21=2(2.5ts) (ts – tn)=2*2.5*(11.14mm(11.14-3.86mm))=405.5mm
2
A22=2(2.5ts) (ts)=2*2.5*(11.14mm*11.14mm)=620.5mm
2
The total area available for reinforcement is AT is given by
AT=A1+A21+A22
= 258mm
2+405.5mm
2+620.5mm
2

45. Vertical pressure vessel

MACHINE DESIGN Page 45

AT=1284mm
2
Ar=AT+2Dntp where Dn=diameter of nozzle
AT=total area of reinforcement
tp=thickness of additional metal
Ar=reinforcement area
tp=Ar-AT/2Dn
tp=2.35mm
Area of add
Aadd=Dnx tp
=150mmx2.35mm
Aadd=352.5mm
2
6.4.7 Design vessel support
design procedure of support
1, Calculate total weights
That means WT=&#3627408458;&#3627408480;ℎ&#3627408466;&#3627408473;&#3627408473;+&#3627408458;ℎ&#3627408466;&#3627408462;&#3627408465;+&#3627408458;&#3627408475;&#3627408476;&#3627408487;&#3627408487;&#3627408473;&#3627408466;+&#3627408458;&#3627408474;&#3627408466;&#3627408465;??????&#3627408482;&#3627408474;
A) Weight of shell
І) The weight of shell consists or the sum of the weight of shell because of it’s material and
the fluid or the gas stored
І.1) Shell material
&#3627408458;&#3627408480;ℎ&#3627408466;&#3627408473;&#3627408473;= &#3627409163; *&#3627408439;&#3627408474;∗t*L*ρ*g Where : &#3627408439;&#3627408474;=mean diameter of the shell
t = thickness of the shell
L = length of the shell
ρ = density of the shell material
g = Gravity takes 9.81m/s²
- The volume of the shell is calculated as :-
V=Π/4*di²*L Where : Di= internal diameter of the shell
L= length of the shell
Di = 1.4m , L= 7m, &#3627408481; &#3627408480;ℎ&#3627408466;&#3627408473;&#3627408473;=11.14mm,
V= &#3627409163;&#3627408485;
D??????
2
4
*L=3.14x(1.4m)
2x7m
V=10.77m
3

46. Vertical pressure vessel

MACHINE DESIGN Page 46

Do=Di+2t
Do =1400mm+2(11.14mm)
Do = 1422.28 =1.42228m
 For the shell material
V = &#3627409163; /4(Do²- Di²)*L
V = &#3627409163; /4[(1.42228)²-(1.4)²]*7m
V= 0.345 m ³
&#3627408458;&#3627408480;ℎ&#3627408466;&#3627408473;&#3627408473; =&#3627409163;∗&#3627408439;o*&#3627408481;shell *L*&#3627409164;*g
= &#3627409163;∗1.42228&#3627408474;∗0.0114&#3627408474;∗7&#3627408474;∗7850
&#3627408446;??????
&#3627408474;
3
∗
9.81&#3627408474;
&#3627408480;
2

&#3627408458;&#3627408480;ℎ&#3627408466;&#3627408473;&#3627408473;=2797.61 N
since density of low carbon alloy steel is 7850
&#3627408446;??????
&#3627408474;
3

B, Mass of internal fluid
In this particular case, the fluid medium is water product
Take the density of the fluid is =1000 Kg/m³ it is density of water.
&#3627408448;&#3627408462;&#3627408480;&#3627408480; &#3627408476;&#3627408467; &#3627408484;&#3627408462;&#3627408481;&#3627408466;&#3627408479; (&#3627408448;&#3627408484;)=&#3627408481;ℎ&#3627408466; &#3627408481;&#3627408476;&#3627408481;&#3627408462;&#3627408473; &#3627408457;&#3627408476;&#3627408473;&#3627408482;&#3627408474;&#3627408466; &#3627408476;&#3627408467; &#3627408477;&#3627408479;&#3627408466;&#3627408480;&#3627408480;&#3627408482;&#3627408479;&#3627408466; &#3627408483;&#3627408466;&#3627408480;&#3627408480;&#3627408466;&#3627408473; &#3627408485; &#3627409164;&#3627408458;
&#3627408448;&#3627408484;=12.17 m ³ ×
1000&#3627408472;&#3627408468;
m ³
=12170&#3627408472;&#3627408468;
Weight of H2o=Mw×&#3627408468;
=12170&#3627408472;&#3627408468;×9.81&#3627408474;/&#3627408480;
2=119387.1N
Weight of H2o=119387.1N
C) Weight of head

&#3627408458;ℎ=&#3627408474;&#3627408462;&#3627408480;&#3627408480; &#3627408476;&#3627408467; ℎ&#3627408466;&#3627408462;&#3627408465;∗&#3627408468;&#3627408479;&#3627408462;&#3627408483;??????&#3627408481;&#3627408486;
Take g=9.81m/s
2
&#3627408458;
ℎ=&#3627408474;
ℎ×&#3627408468;
&#3627408474;
ℎ=&#3627408483;
ℎ×&#3627409164;

47. Vertical pressure vessel

MACHINE DESIGN Page 47

&#3627408474;
ℎ=
4&#3627409163;
3
×(&#3627408453;&#3627408476;
3
−&#3627408453;??????
3
)×&#3627409164;
&#3627408474;
ℎ=
4&#3627409163;
3
×((0.7114&#3627408474;)
3
−(0.7&#3627408474;)
3
)×7850
&#3627408446;&#3627408468;
&#3627408474;
3

&#3627408448;
ℎ=559.78&#3627408446;&#3627408468;
Therefore &#3627408458;
ℎ=559.78&#3627408446;&#3627408468;×9.81
&#3627408474;
&#3627408480;
2

Wh= 5491.4N
D Weight of nozzle Notations;
&#3627408458;&#3627408475; = weight of nozzle
&#3627408448;&#3627408475;=mass of nozzle
g= gravity = take 9.81 m/s
2
&#3627409164;&#3627408484;&#3627408462;&#3627408481;&#3627408466;&#3627408479;=1000
&#3627408472;&#3627408468;
&#3627408474;
3

Wn=mass of nozzle*gravity
&#3627408458;
&#3627408475;=&#3627408474;
&#3627408475;×&#3627408468;
&#3627408474;
&#3627408475;=&#3627408483;×&#3627409164; , V=A*L= &#3627409163;(&#3627408453;
2
)&#3627408447;×&#3627409164; ,but R
2 =&#3627408453;
&#3627408475;&#3627408476;
2

&#3627408474;
&#3627408475;=&#3627409163;(&#3627408453;
&#3627408475;&#3627408476;
2
−&#3627408453;
&#3627408475;??????
2
)&#3627408447;×&#3627409164;
Dno=2tn + Dni =2×3.86mm +150mm=157.72mm =0.15772m
Where Dno=outer nominal diameter
Dni=inner nominal diameter = 0.15m
&#3627408474;
&#3627408475;=&#3627409163;((0.15772/2)
2
−(0.15&#3627408474;/2)
2
))×0.085&#3627408474;×1000
&#3627408446;&#3627408468;
&#3627408474;
3

M
&#3627408475;=0.159kg
Therefore
&#3627408458;
&#3627408475;=0.159&#3627408446;&#3627408468;×9.81
&#3627408474;
&#3627408480;
2

&#3627408458;
&#3627408475;=1.56&#3627408449;

48. Vertical pressure vessel

MACHINE DESIGN Page 48

Mass of longitudinal and lateral pipe
Now let as assume the length of pipe is 85mm
&#3627408447;&#3627408466;&#3627408481; &#3627408481;&#3627408462;&#3627408472;&#3627408466; &#3627408465;&#3627408466;&#3627408475;&#3627408480;??????&#3627408481;&#3627408486; &#3627408476;&#3627408467; &#3627408474;&#3627408462;&#3627408481;&#3627408466;&#3627408479;??????&#3627408462;&#3627408473; &#3627408462;&#3627408481; [7850&#3627408446;&#3627408468;/&#3627408474;3]
&#3627408474;&#3627408477;= &#3627408457;&#3627408477;× &#3627409164;&#3627408474;&#3627408462;&#3627408481;&#3627408466;&#3627408479;??????&#3627408462;&#3627408473;
=&#3627408436;&#3627408477;×&#3627408447;&#3627408477; &#3627408485; &#3627409164;&#3627408454;&#3627408454;
&#3627408474;&#3627408477; =&#3627409163;(&#3627408453;&#3627408477;ₒ^2− &#3627408453;&#3627408477;
2
)&#3627408477;×(&#3627408447;&#3627408477;)&#3627408485; &#3627409164;&#3627408473;&#3627408476;&#3627408484; &#3627408464;&#3627408462;&#3627408479;&#3627408463;&#3627408476;&#3627408475; &#3627408480;&#3627408481;&#3627408466;&#3627408466;&#3627408473;
&#3627408448;&#3627408477;=&#3627409163;(
76.066
1000
)
2
−(75.05/1000)^2) &#3627408485; 0.085&#3627408474; &#3627408485; 7850&#3627408472;&#3627408468;/ &#3627408474;³)
Mp=0.322kg

&#3627408454;&#3627408476; &#3627408467;&#3627408476;&#3627408479; &#3627408481;ℎ&#3627408479;&#3627408466;&#3627408466; &#3627408477;??????&#3627408477;&#3627408466;&#3627408480; (??????.&#3627408466;.,&#3627408481;&#3627408484;&#3627408476; &#3627408473;&#3627408476;&#3627408475;&#3627408468;??????&#3627408481;&#3627408482;&#3627408465;??????&#3627408475;&#3627408462;&#3627408473; &#3627408462;&#3627408475;&#3627408465; &#3627408476;&#3627408475;&#3627408466; &#3627408473;&#3627408462;&#3627408481;&#3627408466;&#3627408479;&#3627408462;&#3627408473; &#3627408477;??????&#3627408477;&#3627408466;&#3627408480;)
3×0.322&#3627408472;&#3627408468;=0.965&#3627408472;&#3627408468;
&#3627408484;&#3627408466;??????&#3627408468;ℎ&#3627408481;&#3627408476;&#3627408467; &#3627408477;??????&#3627408477;&#3627408466;=mp×&#3627408468;
=0.965&#3627408472;&#3627408468;×9.81&#3627408474;/&#3627408480;2=9.46N
&#3627408458;&#3627408477;=9.46&#3627408449;
Total mass of pressure vessels
The total mass of the pressure vessel including the mass of the fluid medium, the Cylindrical shell
section, the above and bottom head end closures is
Total weight = weight of Shell + weight of Head + weight of water + weight of pipes + weight of
nozzle

Total weight =&#3627408458;&#3627408480; +&#3627408458;ℎ +&#3627408458;&#3627408484; +&#3627408458;&#3627408477;&#3627408480; +&#3627408458;&#3627408475;
=2797.61&#3627408449;+5491.4&#3627408449;+119387.1&#3627408449;+9.46&#3627408449;+1.56&#3627408449;
=127687.13&#3627408449;

6.4.7.1 Design on support
Design of Bracket support
Notation;

49. Vertical pressure vessel

MACHINE DESIGN Page 49

W=weight of vessel
N= number of bracket
Q=
&#3627408510;
&#3627408527;
, load of one bracket
R= radius of head
H= lever arm of load
2A, 2B = dimensions of wear plate
S = stress th = wall thickness of head
K= factors, see from chart
C=√&#3627408488;&#3627408489; = radius of circular wear plate
D= 1.82
&#3627408490;
&#3627408505;
√
&#3627408505;
&#3627408507;

DESIGNING DATA
In the design of bracket some parameters are.
W=127687.13N
n = 4
Q=
&#3627408536;
&#3627408527;
=31921.78N
Ri = 0.7m
H = 0.145m
2A = &#3627409358;.&#3627409363;&#3627408526;
2B = 0.5m
t = 0.00942m
Pi = 3.18MPa
S=235MPa
Yield point =530MPa, E = 1
??????=&#3627409361;&#3627409358;°,????????????s??????=0.866
C =√&#3627408436;&#3627408437;=√0.25(0.25)=0.25 &#3627408474;

50. Vertical pressure vessel

MACHINE DESIGN Page 50

D = 1.82
&#3627408464;
&#3627408453;
√
&#3627408453;??????
&#3627408455;
=1.82(
0.25&#3627408474;
0.7&#3627408474;
)√
0.7&#3627408474;
0.00942
=5.603&#3627408474;
Values of k can be taken from the charts shown below
Stress in vessel on bracket support
NOTES:
Positive values denote tensile stresses and negative values denote compression.
Computing the maximum tensile stresses, in formulas for S1 and S2, K, K3, K5 and
K, denote negative factors and K2, Kq, KG and K8 denote positive factors.
Computing the maximum compression stresses, in formulas for S1 and S2, K1, K2,
K3, K4, K5, KG, K, and K8 denote negative factors. The maximum tensile stresses S1 and S2,
respectively, PIUS the tensile stress due to Internal pressure shall not exceed the allowable tensile
stress value of head material. The maximum compression stresses S1 and S2, respectively, plus the
tensile stress
due to internal pressure shall not exceed the allowable compression stress value of head material.

51. Vertical pressure vessel

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52. Vertical pressure vessel

MACHINE DESIGN Page 52


K1=0.025 k4=0.002 k7=0.004
K2=0.008 k5=0.005 k8=0.003
K3=0.026 k6=0.07

Longitudinal stress
Maximum tensile stress:
S1=
&#3627408452;
&#3627408481;
2
{&#3627408464;&#3627408476;&#3627408480;??????(−&#3627408472;1+6&#3627408472;2)+
??????
&#3627408453;
√
&#3627408453;
&#3627408455;
(−&#3627408472;3+6(&#3627408472;4))}
S1=
31921.78&#3627408449;
(0.00942&#3627408474;)
2
{cos30(−0.025+6(0.008))+
0.145&#3627408474;
0.7&#3627408474;
√
0.7&#3627408474;
0.00942&#3627408474;
(−0.0026+ 6(0.002)}
S1=-80.22Mpa
The stress due to internal pressure
??????=
??????&#3627408453;
2&#3627408481;
=
(3.14&#3627408448;&#3627408477;&#3627408462;)(0.7&#3627408474;)
2(0.00942&#3627408474;)
=116.67&#3627408448;&#3627408477;&#3627408462;
The sum of tensional stress =116.67&#3627408448;&#3627408477;&#3627408462;−80.22&#3627408448;&#3627408477;&#3627408462;=36.45Mpa
Maximum compressive stress

53. Vertical pressure vessel

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S1=
&#3627408452;
&#3627408481;
2
{&#3627408464;&#3627408476;&#3627408480;??????(−&#3627408472;1−6&#3627408472;2)+
??????
&#3627408453;
√
&#3627408453;
&#3627408481;
(−&#3627408472;3−6(&#3627408472;4))}
S1 =
31921.78
(0.00942)2
{cos30(−0.025−6(0.008))+
0.145
0.7
√
0.7
0.00942
(−0.026− 6(0.002)}
S1=-148.21MPa
The stress due to internal pressure =116.67MPa
The sum of stresses = (116.7-148.21)Mpa
??????T=31.51MPa compression
Circumferential stress
Maximum tensile stress
S2 =
&#3627408452;
&#3627408481;
2
[ cosᾳ(-k5+6k6)+
??????
&#3627408453;
√
&#3627408453;
&#3627408481;
(-k7 - 6k8 )]
S2=
31921.78&#3627408449;
0.00954
2
[cos30(-0.005+6(0.007)) +
0.145&#3627408474;
0.7&#3627408474;
√
0.7&#3627408474;
0.00954&#3627408474;
(-0.004 - 6(0.003)]
S2=-103.66MPa
The stress due to internal pressure = 116.7MPa
The sum of stress = (116.7-103.66) Mpa
?????? T = 13.04MPa
Maximum compression stress
S2 =
&#3627408452;
&#3627408481;
2
[ cosᾳ(-k5-6(k6)) +
??????
&#3627408453;
√
&#3627408453;
&#3627408481;
(-k7 - (6)k8 )]
S2=
31921.78&#3627408449;
0.00954
2
[cos30(-0.005-6(0.007)) +
0.145
0.7
√
0.7
0.00954
(-0.004 - 6(0.003)]
S2=-106.78Mpa
The stress due to internal pressure =116.7Mpa
The sum of compressive stress =(116.7Mpa-106.7Mpa)=9.92Mpa
Both longitudinal and circumferential stress are less than 235Mpa
Therefore the design is safe.

54. Vertical pressure vessel

MACHINE DESIGN Page 54

6.5 Design of welded joint
A welded joint is a permanent joint which is obtained by the fusion of edge of two parts to
be joined together with or without the application of pressure and filler material. Electric
arc welding is extensively used because of greater speed of welding
The main considerations involved in the selection of weld type are:
 The shape of welded component required
 The thickness of the plates to be welded
 The direction of force to be applied
We Select butt joint especially square butt joint for the connection of both head and shell
but for the connection of nozzle to the shell part select corner joint.
Stresses for Welded Joints:
The stresses in welded joints are difficult to determine because of the variable and
unpredictable Parameters like homogeneity of the weld metal, thermal stresses in the
welds, changes of physical Properties due to high rate of cooling etc. The stresses are
obtained, on the following assumptions :
 The load is distributed uniformly along the entire length of the weld, and
 The stress is spread uniformly over its effective section.
The following table shows the stresses for welded joints for joining ferrous metals with
mild steel electrode under steady and fatigue or reversed load.
Table 4.6 standard for weld design

Table 5
Since the stress in the weld material is to be 81Mpa our electrode is mild steel electrode.
The recommended minimum size of weld is 14mm size of weld between 25mm- 55mm
F=π(do
2-di
2) ??????&#3627408474;/4
F=3.14(157.72
2-150
2) ??????&#3627408474;/4
F=1864.84 ??????&#3627408474;

55. Vertical pressure vessel

MACHINE DESIGN Page 55

And FR=Pd x π x di/4
FR=3.18x3.14x150
2/4
FR=56166.75N
By equating F and FR calculate ??????&#3627408474;
??????&#3627408474;= 56166.75N/1864.84mm
2
??????&#3627408474;=30.19N/mm
2
where: - d???= outside diameter of the nozzle
di= inside diameter of the nozzle

Factor of safety
A factor of safety is given by Fs=maximum tensile strength/ultimate tensile stress
Fs=550Mpa/80mpa=6.87 which is greater than 5 so our design is safe.

CHAPTER SEVEN
CONCLUSION
We conclude from this pressure vessel design the pressure with the internal diameter of
1.4m assume which is used to store steam at 3.5Mpa and 300℃ designed above is designed
to satisfy all the required requirements to be able to work under the above parameters. A
corrosion allowance of 1.6mm is applied during the design to prevent failure due to
corrosion and fatigue conditions. Our pressure vessel is made with the material called low
alloy steel for all part and carbon steel for bracket support design due to have higher than
to those of the plain carbon steels for given applications. This means it has higher strength,
hardness, hot hardness, wear resistance, toughness, and more desirable combinations of
these properties. .
The design of pressure vessel in initialized with the specification requirements in terms of
standard technical specifications along with numerous requirements that lay hidden from
the market.
The storage of fluid at higher pressure in the pressure vessel is at the heart of its
performance and it is the first step towards the design. The pressure vessel components are
merely selected, but the selection is very critical. A slight change in selection will lead to a
different pressure vessel altogether from what is aimed to be designed.

56. Vertical pressure vessel

MACHINE DESIGN Page 56

It is observed that all the pressure vessel components are selected on basis of available
ASME standards and the manufactures also follow the ASME standards while
manufacturing the components.
Selection of pressure vessels components should be according to standards rather than
customizing the design:-the standards lead to;
 A universal approach
 Less time consumption
 Easy replacement
 So less overall cost.






REFERENCES
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[ 4.] Pierre, D., and Baylac, G., French pressure vessel regulations within the European context, J.
Pressure Technol., 114, 486–488, 19
[5]. Pressure vessel hand book -10th edition.
[6]. M.A Khan, “Stress distribution in horizontal pressure vessel and saddle supports”, International
Journal of Pressure Vessels and Piping 87 (2010).
[7]. M. Javed Hyder, K, Asif,” “Optimization of location and size of opening in a pressure vessel
cylinder using ANSYS”, engineering Failure Analysis 15 (2008).
[8].M. Giglio, “Fatigue analysis of different types of pressure vessel nozzle”, revised 19 November
2002; accepted 20 November 2002
[9]. Avinash R.Kharat, V. V. Kulkarni,” Analysis of stress concentration at opening in pressure vessel
using ANOVA”, IJRET: International Journal of Research in Engineering and Technology, eISSN:
2319-1163

57. Vertical pressure vessel

MACHINE DESIGN Page 57

[10]. Bandrupalli Praneeth, T.B.S.Rao, “Finite Element Analysis of Pressure Vessel and Piping
Design “, Inte
[11]. Shyam R. Gupta, Ashish Desai, Chetan P. Vora, “Optimize nozzle location for minimization of stress
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[14].Pressure vessel design manual ,Dennis Moss third edition.
[15]. Pressure vessel hand book ,Henry,H-Bendar second edition.
[16].Brownell, Lloyd E. and Edwin H. Young, Process Equipment Design: 1959
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[18].Mahajan, Kanti K., Design of Process Equipment: Second Editon
[19] K.Abakan’s, Hydrocarbon Processing and Petroleum Refiner, June 1963
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