ADDIS ABABA SCIENCE AND TECHNOLOGY

UNIVERSITY
COLLEGE OF ELECTRICAL AND MECHANICAL
ENGINEERING
DEPARTMENT OF MECHANICAL ENGINEERIN

MACHINE DESIGN PROJECT

PROJECT 1 DESIGN OF PRESSURE VESSEL

COMPILED BY
DEJENIE AWOKE ID NO 0383/10
DEMEKE ZEBENE ID NO 0387/10
SUBMITTED TO: Waleligne M. Salilew (MSc.)
6/18/13
GROUP 11

DESIGN OF CYLINDERICAL PRESSURE VESSEL

GIVEN SPECIFICATION
PARAMETERS VALUE UNIT
PRESSURE 1.5 Bar
TEMPRETURE 35 Degree Celsius
VOLUNE 10 Cubic meter
END HEAD SHAPE Hemispherical
SUPPORT TYPE Saddle support
LAYOUT Horizontal
FLUID TYPE Petrol

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 Acknowledgement

We would like to acknowledge and express their appreciation for the co-operation to this project.
First, we would like to thanks our instructor Mr. Waleligne M. Salilew (MSc.) who instruct, teach
and advise us to work on this project properly.
Second, thanks. Our classmates who helped to make this paper possible and for anyone else to
give us a chance to talk with them about this design

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CONTENTS
CHAPTER 1 ..................................................................................................................................................................... 1
BACKGROUND OVERVIEW...................................................................................................................................... 1
HISTORY ....................................................................................................................................................................... 1
1.1 INTRODUCTION ................................................................................................................................................. 2
1.1.1 Types of pressure vessel ..................................................................................................................... 2
1.1.2 Parts of pressure vessels..................................................................................................................... 6
1.2 STATEMENT OF THE PROBLEM .......................................................................................................... 8
1.3 OBJECTIVE ..................................................................................................................................................... 9
1.3.1 General objective.................................................................................................................................. 9
1.3.2 Specific objective .......................................................................................................................................... 9
1.4 SIGNIFICANCE OF THE OF PROJECT .............................................................................................. 10
1.5 METHODOLOGY....................................................................................................................................... 10
1.5.1 Method .......................................................................................................................................................... 10
1.5.2 Procedure..................................................................................................................................................... 11
1.5.3 Material selection for Pressure Vessel Construction..................................................................... 13
CHAPTER 2 .................................................................................................................................................................. 14
LITERATURE REERVIEW ....................................................................................................................................... 14
CHAPTER 3 .................................................................................................................................................................. 16
detail design ..................................................................................................................................................................... 16
3.1 Design of shell .............................................................................................................................................. 17
3.2 Design of hemispherical head.................................................................................................................... 21
3.3 Design of Nozzle .......................................................................................................................................... 21
3.4 Opening Reinforcement ..................................................................................................................................... 23
3.4 Manhole Design ............................................................................................................................................ 24
3.4.1 Manhole opening reinforcement .................................................................................................. 25
3.5 Design of flange............................................................................................................................................ 26
3.6 Design of saddle Support ............................................................................................................................ 28
3.7 Welding design .................................................................................................................................................... 35
CHAPTER 4 .................................................................................................................................................................. 37
summary ........................................................................................................................................................................... 37
4.1 RESULT ......................................................................................................................................................... 37

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 4.2 CONCLUSION ............................................................................................................................................. 37
4.3 RECOMMENDATION ..................................................................................................................................... 37
REFERENCES ................................................................................................................................................................ 38

CONTENTS OF TABLES
TABLE 1: GIVEN DATA FOR A KNOWN PRESSURE ............................................................................ 17
TABLE 2: COMPARISONS OF STRESSES ............................................................................................ 33

CONTENTS OF FIGURES
FIGURE 1: CYLINDRICAL PRESSURE VESSEL...................................................................................... 3
FIGURE 2:SPHERICAL PRESSURE VESSEL........................................................................................... 4
FIGURE 3: CONICAL PRESSURE VESSEL ............................................................................................. 5
FIGURE 4: HEAD OF PRESSURE VESSEL(HEMISPHERICAL) ................................................................. 7
FIGURE 5: MAIN PARTS OF PRESSURE VESSEL ................................................................................... 7
FIGURE 6: CIRCUMFERENTIAL AND LONGITUDINAL STRESSES ........................................................ 19
FIGURE 7: HOOP STRESS ON A SHELL .............................................................................................. 20
FIGURE 8: HEMISPHERICAL HEAD ................................................................................................... 21
FIGURE 9: SECTION OF NOZZLE ....................................................................................................... 22
FIGURE 10: SECTION OF FLANGE .................................................................................................... 27
FIGURE 11: MOMENT APPLIED ON HORIZONTAL PRESSURE VESSEL DUE SADDLE SUPPORT ............. 29
FIGURE 12: DIMENSION OF SADDLE SUPPORT TO BE DESIGN ........................................................... 29
FIGURE 13: SADDLE SUPPORT ANALYSIS ........................................................................................ 34

CONTENTS OF CHARTS
CHART 1: DESIGN PROCEDURE ........................................................................................................ 12

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ACRONYMS
Td: design temperature, °C
c: corrosion allowance, mm
Di: inside diameter of the vessel, mm
Do: outside diameter of the vessel, mm
Ri: inside radius of the vessel, mm
Ro: outside radius of the vessel, mm
t: required the thickness, mm
tn: minimum thickness provided for the nozzle, mm
trn: selected thickness for the nozzle, mm
W: weight of the vessel
L: length of shell
σL: longitudinal stress
σh: hoop stress
σall: allowable stress
τmax: maximum shear stress

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Abstract

The pressure vessel contains high pressurized fluid so that the selection of material and the design
of pressure vessel is needed and most important because since those pressure vessels contain very
high pressure it may blast and hurt many peoples and things if there is not proper selection and
design of material. The material must pass the sequence of hydrostatic test and this test gives
capability of construction to be capable, resist and survive under that high pressure. the design may
be analytical which is by using as per American Society of Mechanical Engineers (ASME) code
section viii division. the dimension and stress which work on the pressure vessel can also be
founded by this very important code. Therefore, in each and every part of the vessel we have tried
to select proper material and calculate the stress which it can resist the high pressure.

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 CHAPTER 1
BACKGROUND OVERVIEW

HISTORY

The earliest documented design of pressure vessels was described in 1495 in the book by Leonardo
da Vinci, the Codex Madrid I, in which containers of pressurized air were theorized to lift heavy
weights underwater. However, vessels resembling those used today did not come about until the
1800s, when steam was generated in boilers helping to spur the industrial revolution. However,
with poor material quality and manufacturing techniques along with improper knowledge of
design, operation and maintenance there was a large number of damaging and often fatal
explosions associated with these boilers and pressure vessels, with a death occurring on a nearly
daily basis in the United States.
Local providences and states in the US began enacting rules for constructing these vessels
after some particularly devastating vessel failures occurred killing dozens of people at a time,
which made it difficult for manufacturers to keep up with the varied rules from one location to
another.

The first pressure vessel code was developed

starting in 1911 and released in 1914, starting
the ASME Boiler and Pressure Vessel Code
(BPVC). In an early effort to design a tank
capable of withstanding pressures up to
10,000 psi (69 MPa), a 6-inch (150 mm)
diameter tank was developed in 1919 that
was spirally wound with two layers of high
tensile strength steel wire to prevent sidewall
rupture, and the end caps longitudinally A 10,000 psi (69 MPa) pressure vessel from 1919,
wrapped with high tensile steel banding and steel
reinforced with lengthwise high-tensile rods.
rods to secure the end caps.

The need for high pressure and temperature vessels for petroleum refineries and chemical plants
gave rise to vessels joined with welding instead of rivets (which were unsuitable for the pressures
and temperatures required) and in the 1920s and 1930s the BPVC included welding as an
acceptable means of construction; welding is the main means of joining metal vessels today.

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There have been many advancements in the field of pressure vessel engineering such as advanced
nondestructive examination, phased array ultrasonic testing and radiography, new material grades
with increased corrosion resistance and stronger materials, and new ways to join materials such as
explosion welding, friction stir welding, advanced theories and means of more accurately assessing
the stresses encountered in vessels such as with the use of Finite Element Analysis, allowing the
vessels to be built safer and more efficiently.

1.1 INTRODUCTION

A pressure vessel is considered as any closed vessel that is capable of storing a pressurized fluid,
either internal or external pressure, regardless of their shape and dimensions.
Vessels, tanks, and pipelines that carry, store, or receive fluids are called pressure vessels. Pressure
are used to store and transmit liquids, vapors, and gases under pressure. A pressure vessel is 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.
Pressure vessels are used in a number of industries; for example, the power generation industry for
fossil and nuclear power, the petrochemical industry for storing and processing crude petroleum
oil in tank farms as well as storing gasoline in service stations, and the chemical industry (in
chemical reactors).

1.1.1 Types of pressure vessel

Pressure vessels can be classified by different criteria. Thus, pressure vessels may be classified as
follows:
1. According to size and geometry:

Pressure vessels can theoretically be almost any shape, but shapes made of sections of spheres,

cylinders, and cones are usually employed. A common design is a cylinder with end caps called

heads. Head shapes are frequently either hemispherical or dished (tori spherical).

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More complicated shapes have historically been much harder to analyze for safe operation and are

usually far more difficult to construct.

Based o their size and geometry, most of the time pressure vessels can be:

• Cylindrical

• Spherical and,

• Conical

Cylindrical

Cylinders are widely used for storage due to their being less expensive to produce than spheres.

However, cylinders are not as strong as spheres due to the weak point at each end.

Figure 1: cylindrical pressure vessel

This weakness is reduced by hemispherical or rounded ends being fitted. If the whole cylinder is

manufactured from thicker material than a comparable spherical vessel of similar capacity, storage

pressure can be similar to that of a sphere.

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Spherical

This type of vessel is spherical type of pressure vessel referred for storage of high-pressure fluids.

A sphere is a very strong structure. The even distribution of stresses on the sphere's surfaces, both

internally and externally, generally means that there are no weak points. Spheres however, are

much costlier to manufacture than cylindrical vessels.

Figure 2:spherical pressure vessel

An advantage of spherical storage vessels is, that they have a smaller surface area per unit volume

than any other shape of vessel. This means, that the quantity of heat transferred from warmer

surroundings to the liquid in the sphere, will be less than that for cylindrical or rectangular storage

vessels.

Conical

Conical transitions are really usual in several industrial sectors so that people minimally related to

design or manufacturing of pressure vessels are used to deal with them. The shape or these conical

sections is mainly imposed by the process in which vessel is immersed. In order to satisfy properly

the process needs, the conical part could take different forms.

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The principal classification criterion is established by the external configuration of the cone, which

could be either concentric or eccentric.

Figure 3: conical pressure vessel

2. According to dimension:
The pressure vessels, according to their dimensions, may be classified as thin shell or thick shell.

Thin shell pressure vessel

If the wall thickness of the shell (t) is less than 1/10 of the diameter of the shell (d), then it is

called a thin shell. thin shells are used in boilers, tanks and pipes.

Thick shell pressure vessel

If the wall thickness of the shell (t) is greater than 1/10 of the diameter of the shell (d), then it is

called a thick shell. thick shells are used in high pressure cylinders, tanks, gun barrels etc.

3. According to orientation of shape:

From the above classification, according to size and geometry (cylindrical and conical pressure

vessels) their orientation can be:

• Horizontal

• Vertical

• Slanted

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 4. According to scheme of loading:
Pressure vessels based on their scheme of loading can be:

• Vessels working under internal pressure and

• Vessels working under external pressure

5. According to wall temperature:

Based on wall temperature pressure vessels can be:

• Heated pressure vessel and

• Unheated pressure vessel

6. According to end construction:

The pressure vessels according to the end construction, may be classified as open end or closed

end. A simple cylinder with a piston, such as cylinder of a press is an example of an open-end

vessel, whereas a tank is an example of a closed end vessel. In case of vessels having open ends,

the circumferential or hoop stresses are induced by the fluid pressure, whereas in case of closed

ends, longitudinal stresses in addition to circumferential stresses are induced.

1.1.2 Parts of pressure vessels

Most pressure vessels mainly have four main parts. Those are:
1. Shell
Shell is the primary component contains the pressure. Pressure vessel shells in the form of different
plates are welded together to form a structure that has a common rotational axis. Shells are either
cylindrical, spherical or conical in shape.
2. Head
All the pressure vessels must be closed at the ends by heads (or another shell section). Heads are
typically curved rather than flat.

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The reason is that curved configuration is stronger and allow the heads to be thinner, lighter and
less expensive than flat heads. Heads can also be used inside a vessel and are known as
intermediate head. These intermediate heads are separate sections of the pressure vessels to permit
different design conditions.

Figure 4: Head of pressure vessel(hemispherical)

3. Nozzle
Nozzle is a cylindrical component that penetrates into the shell or head of pressure vessel.

They are used for attach piping for flow into or out of the vessel, attach instrument connection
(level gauges, thermo wells, pressure gauges) and to provide access to the vessel interior at man
way or to provide for direct attachment of other equipment items (e.g. heat ex changers).
Nozzles on pressure vessel have following applications:
• Attach piping for flow into or out of the vessel.
• Attach instrument connections, (e.g., level gauges, thermo wells, or pressure
gauges).
• Provide access to the vessel interior at man way.
• Provide for direct attachment of other equipment items, (e.g., a heat ex changer or
mixer).

Figure 5: main parts of pressure vessel

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 4. Support
Support is used to bear all the loads of pressure vessel, earth quack and wind loads. There are
different types of supports, which are used depending upon the size and orientation of the pressure
vessel. It is considered to be the non-pressurized parts of the vessel.

Typical kinds of supports are:

a. Skirt support
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). Tall, vertical, cylindrical pressure vessels are
supported by skirt. For example, tower and reactor.
b. Leg support
Small vertical drums are typically supported on legs that are welded to the lower portion of the
shell. The number of legs needed depends on the drum size and the loads to be carried.
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. The other support is normally free to permit unrestrained
longitudinal thermal expansion of the drum.
d. Lug support
Lugs that are welded to the pressure vessel shell may also be used to support vertical pressure
vessels.
Most of the components are fabricated from plates or sheets. Seamless or welded pipes can also
be used. Parts of vessels formed are connected by welded or riveted joints. In designing these parts
and connections between them, it is essential to consider, the efficiency of joints.

For welded joints, the efficiency may be taken as 100% if the joint is fully checked by a radio
graph and taken as 85%, even if it is checked at only a few points. If the radio graphic test is not
carried out 50 to 80%, efficiency is taken. Efficiency vary between 70 to 85% in the case of riveted
joints. Design procedure is primarily based on fabrication by welding.

1.2 STATEMENT OF THE PROBLEM

Pressure vessel is one of the most devices which used in industries. But there are several defects
created on this device. Such as corrosion, loss of thickness, mechanical & metallurgical failure,
cracking, mechanical deformation etc. and these defects are influenced negatively on the function
of this device. However, these defects can be solved by different mechanisms. Such as by proper
material selection, by proper manufacturing process: like welding process etc.

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From the above defects we would like to mainly focus on corrosion because since the device has
a continuous connection between fluid the materials are greatly affected by corrosion. But due to
different technical and environmental influences we cannot prevent the problem absolutely.
However, as much as possible we would like to design our project (Horizontal layout pressure
vessel) by eradicating these problems to a minimum position. The pressure vessels that not follow
any standard codes can be very dangerous. In fact, many fatal accidents have occurred in the
history of their operation and development. Improper design and construction, irregular testing
and inspection cause safety hazards to pressure vessels. When a substance is stored under pressure,
the potential for rupture and leakage is greater. The risk of damage from a pressure vessel increases
when vessel contents are toxic, flammable or gaseous substances. When a substance is stored under
pressure, the potential for rupture and leakage is greater. Improper design and construction,
irregular testing and inspection causes safety hazards to pressure vessels.
Finally, vessel design and maintenance must be considered carefully as even a small imperfection
increases the risk of pressure vessel failure, posing a serious safety hazard.
1.3 OBJECTIVE

1.3.1 General objective

The general objective of our project is to design a horizontal layout pressure vessel position with
some hemispherical head and with a saddle support by estimating the internal pressure and
temperature and having a medium of petrol at the temperature of 350c and pressure of 1.5 bar.

1.3.2 Specific objective

Specifically, we would like to design each and individual component of pressure vessel such as:
▪ Shell of pressure vessel
▪ Head of pressure vessel
▪ Nozzle of pressure vessel
▪ Support of pressure vessel
When we design as much as possible, we will try to consider each and individual things in order
to full fill a particular need within our overall objective. We will also recognize sub-objective
requirements of the various units that make up the overall design.

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1.4 SIGNIFICANCE OF THE OF PROJECT

In sophisticated pressure vessels encountered in engineering construction; high pressure,

extremes of temperature and severity of functional performance requirements pose exciting design
problems. The word "DESIGN" does not mean only the calculation of the detailed dimensions of
a member, but rather is an all-inclusive term, incorporating:

1. The reasoning that established the most likely mode of damage or failure;
2. The method of stress analysis employed and significance of results;
3. The selection of materials type and its environmental behavior.

The main significance of this project is used to:

• Verify the macro strength part of the pressure vessel.

• Verify the deformation the vessel due to external loading.

• Verify the tightness of the vessel.

• Develop a new product, assurance to the interchangeability and common usability and
is convenient to use and maintain.
1.5 METHODOLOGY

In general, pressure vessels designed in accordance with the ASME Code, Section VIII, Division
1, are designed by rules and do not require a detailed evaluation of all stresses. It is recognized that
high localized and secondary bending stresses may exist but are allowed for by use of a higher
safety factor and design rules for details. It is required, however, that all loading's (the forces
applied to a vessel or its structural attachments) must be considered.
While the Code gives formulas for thickness and stress of basic components, it is up to the designer
to select appropriate analytical procedures for determining stress due to other loadings. The
designer must also select the most probable combination of simultaneous loads for an economical
and safe design.
1.5.1 Method
Our design is horizontal pressure vessel with saddle support. This design is done by using the catia
version 5, auto cad software and if necessary, we use MATLAB to calculate more complex
equations. Before we are going to the software, we have to calculate the diameter, length of the
shell, head, nozzle and support of the horizontal pressure vessel.

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1.5.2 Procedure
There are several procedures which are used to design the pressure vessel out of them we have to
use the following one.
Step 1:
Design specification includes the given specification and designer’s specification.
Step 2:
Based on the above specifications we have to select the proper material for our design
depending upon the ASME code or other standard given on our reference books.
Step 3:
Next to the material selection we must check the dimensional analysis for each and every
part based on standard given on our manual.
Step 4:
Design for internal and external application loads on the pressure vessel.
Step 5:
Analysis evaluate and check for hoop/ circumferential, longitudinal and allowable stresses
which developed on the pressure vessel.
Step 6:
Finally, we have to check our design whether it is safe or not by comparing and contrasting
the calculated values with the actual value given in the manual.

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 FUNCTIONAL CHECK CHANGE IN
REQUIREMENT DIMENTION

OPERATIONAL ESTABLISHMENT OF
REQUIREMENT DESIGN CONDITIONS OPTIMUM
S
LIMIT
DESIGN CODES
MATERIAL ANALYZED
SELECTION

PRELIMINARY
LAYOUT
IF DESIGN

FAILS

STRESS ANALYSIS DESIGN CHECK

IF DESIGN

PASSES
FINAL DESIGN

CONSTRACTION,
INSPECTION AND
DELIVERY

chart 1: design procedure

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1.5.3 Material selection for Pressure Vessel Construction

There are parameters that should be satisfied to select economical material for pressure vessels.
These parameters are including the following aspects:
• Strength for design condition
• Strength for desired service life
• Resistance to corrosion in service environment for desired life
• Capabilities for fabrication processes
• Market availability
• Maintenance and repair Cost (first investment and operation cost)
• Ease of maintenance
• Resistance to hydrogen attack
• Fracture toughness
• Ductility
• Temperature treatment

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 CHAPTER 2
LITERATURE REERVIEW

Mr. Mukund Kavekar, Mr. Vinayak H. Khatawate and Mr. Gajendra V. Patil
They have explained about use of composite materials to replace high strength to weight metals
for use of pressure vessels in low weight applications such as aerospace and oil and gas.
N.A. Weil and J.J. Murphy
They have discussed factors determining performance of welded skirt support for vertical pressure
vessels. The importance of thermal effects is emphasized. The performance of the vessel is
assessed on fatigue basis by analysis of local stresses.

Siva Krishna Raparla and T. Seshaiah

In their paper have designed multi-layered high-pressure vessel and compared it with mono block
vessel. Multilayered pressure vessels are built by wrapping a series of sheets over the core tube.
Scope is obtained to select different materials at different layers according to functionality. Inner
layer can be made of anti-corrosive materials while outer layers can be made of material having
high strength. The design is based on ASME Code Section VIII division I. Using muti-layered
pressure vessel results in percentage saving of material of 26.02% reducing overall weight of the
vessel. With help of FEM software, it is found that the stress variation from inner to outer surface
for multilayered vessel is 12.5% and that for solid vessel is 17.35% resulting in more uniform
stress distribution. Thus, the multi-layered pressure vessels are suited for high pressure and high
temperature applications.
Puneet Deolia and Firoz A. Shaikh
They have carried out Finite Element Analysis to estimate burst pressure of mild steel pressure
vessel using Ramberg-Osgood model. Burst pressure is the pressure at which the vessel bursts or
crack and fluid leaks which is undesirable and such pressure must not be exceeded. The burst
pressure can be found out numerically using Ramberg-Osgood material curve. The finite element
method is used to calculate the burst pressure using Ramberg-Osgood equation and then comparing
it with the results obtained from elasto-plastic curve and true stress strain curve. Analyzing the
results Ramberg-Osgood model showed better correlation with the experimental observations as
compared to modified Fuel Formula. Thus, the use of FEM can help save time and cost of actual
testing.

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Sonachalam. M1 and Ranjit Babu. B. G2
have examined optimal ply orientations of symmetrical and asymmetrical shells for maximum
burst pressure. Shells made of Glass Reinforced plastic with-glass-epoxy-fiber. Specimen had 10
layers with different ply orientations. Finite Element Method is used to find the optimum winding
angles. It was found that the maximum hoop stress developed for steel was 212.9 N/mm 2 applied
pressure of 20MPa whereas stress for glass epoxy fiber reinforced plastic was 184 N/mm2 for same
applied pressure.
Simon Sedmaka, Mahdi Algoolb, Aleksandar Sedmakc, Uros Tatica and Emina Dzindoa
They have studied the Elastic-plastic behavior of welded joints during loading and unloading of
pressure vessels. Pressurizing of the model was done in two stages. First the model was loaded to
working pressure then held at a lower pressure for 2 hours. After unloading it was tested at total
working load or water hammer load. They conclude that higher heat input to the weld zone is
better. The HAZ of micro alloyed steel has greater resistance to crack during load variations
compared to WM.
Andrade, Tatiana Lima, Paula, Wagner Andrade de, Junior and Pedro Américo Almeida
Magalhães
They have done a comparative study on methods to analyses stresses in vessel/nozzle due to
external loads. A model of a nozzle without reinforcement is prepared so that comparison can be
done by WRC 107, WRC 297 and FEA method. The WRC (Welding Research Council) Bulletin
107 is a parameterized procedure of stress calculation of nozzle in which the input values are
dimensionless and the stress results are obtained from curves developed based on experimental
data. WRC Bulletin 297 is a supplement to WRC 107 for higher diameter- thickness ratios. The
stress values obtained by the three methods were close and are reliable for pressure vessels and
nozzles that fit in WRC 107 and 297 procedure.
Shyam R. Gupta and Ashish Desai
They have designed a horizontal pressure vessel using PV Elite industrial software. For designing
the vessel very few parameters such as design pressure, design temperature, inside diameter,
volume, material, fluid properties, etc. are required. PV Elite gives the thickness of the shell, head
dimensions as per our selected head, nozzle calculations based on diameter given by us, loading's
on pressure vessels, support design and calculations and all the parameters required to manufacture
a pressure vessel. The stress generated in the vessel due to pressure loading and at discontinuities
is obtained in the report from PV Elite. PV Elite does the calculations as per the code selected by
the user and the stresses are calculated as per Welding Research Council (WRC) 107. The results
obtained from the software are accurate and complying with the standard codes. Use of the
software will help reduce design time and can give precise calculations and required data.

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 CHAPTER 3
DETAIL DESIGN

Designers design specification

Before we are going to the detail design we have to select the proper material of the pressure vessel.
On this design we select carbon steel because Carbon steel provides a number of advantages as a
pressure vessel material. In addition to being highly resistant to corrosion, shock, and vibration, it
possesses a high tensile strength - making it ideal for demanding tank applications in a wide range
of industrial processes and it is weldable. It also retains strength at minimal thicknesses, which
reduces the amount of material needed for tank fabrication, thereby lowering fabrication costs.
Carbon steel is also highly recyclable and accounts for more recycled weight annually than
aluminum, plastic, paper, and glass combined. In fact, approximately 50% of its production comes
from reclaimed materials.
• Joint efficiency of longitudinal and circumferential double welded butt joint for this
pressure vessel is 0.85. so, we take J=0.85
• Corrosion allowance c=5mm for carbon steel where sever corrosion is not expected.
• Factor of safety we have selected the factor of safety based on the ASME code section VIII
division 2. because the division 2 vessel requires more complex design procedure as well
as restrictive requirements on design materials and nondestructive examinations, however
higher design stress values are permitted in division 2 vessels and it is possible to design
pressure vessels with safety factor of 3.
We have taken 𝑓𝑜𝑠 = 3
• Design pressure and temperature can be calculated by giving 10% allowance to the
operating pressure and temperature.

10
i.e. 𝑃𝑑 = 𝑃o = 10%𝑃o = 0.15 + 0.15 × 100 = 0.165𝑀𝑃
10
𝑇𝑑 = 𝑇o = 10%𝑇o = 35℃ + 35℃ × 100 = 38.5℃

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 3.1 Design of shell

Pressure and temperature are given to be 1.5𝑏𝑎𝑟 and 35℃ respectively, from the table given for
carbon steel with the respective temperature the design/allowable stress is 135 𝑁⁄𝑚𝑚2 and
ultimate tensile strength is 360 𝑁⁄𝑚𝑚2 .
Length to outside diameter ratio is given by:

Table 1: given data for a known pressure

Pressure in MPa 𝐿 𝐿 𝐷𝑖
𝐷𝑖 𝐷𝑖
0−1.723 3 3 36𝑉
√
5𝜋

1.723−3.445 4 4 3 12𝑉
√
1.5𝜋

>3.445 5 5 33𝑉
√
4𝜋

Since the Pressure is

P=1.65𝑏𝑎𝑟 = 1.65 × 105 𝑏𝑎𝑟 = 0.165𝑀𝑃 and
The given volume is 𝑉 = 10 𝑚3
Hence the pressure is in between 0 −1.723Mpa, therefore, length to diameter ratio is:
𝐿
=3
𝐷𝑖
𝐿 = 3𝐷𝑖 … … … … … … 1

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The length to diameter ratio is known then the internal diameter is given by:

3 6×𝑉
𝐷𝑖 = √
5𝜋

3 6 × 10
𝐷𝑖 = √
5𝜋

𝑫𝒊 = 𝟏. 𝟓𝟔𝟑𝒎 = 𝟏𝟓𝟔𝟑𝒎𝒎
Then the length is:
𝐿 = 3𝐷𝑖
𝐿 = 3 × 1.563𝑚
𝑳 = 𝟓. 𝟔𝟗𝒎 ≈ 𝟓𝒎
Step 1:
A cylinder is considered thin when the ratio of its inner diameter to the wall thickness is more than
15 and When the ratio of the inner diameter of the cylinder to the wall thickness is less than 15,
the cylinder is called a ‘thick-walled’ cylinder or simply ‘thick’ cylinder.
𝐷𝑖 1563
= = 218.6
𝑡 7.15
It is greater than 15 then it is thin cylinder
Next to this we have to find the thickness of the shell. An equation for the minimum thickness of
a sphere can be obtained from,
𝑃𝐷𝑖
𝑡= +𝐶
2𝐽𝜎 − 𝑃
Where:
𝑡 - the thickness of the shell
𝑃 - the internal Pressure of the shell

18 | P a g e
 𝐷𝑖 -the internal diameter
𝜎-the design stresses
Given:
𝑃 = 0.165𝑀𝑃 = 0.165 𝑁⁄𝑚𝑚2
𝐷𝑖 = 1.563𝑚, 𝑅 = 0.7815𝑚
𝜎 = 135 𝑁 ⁄𝑚𝑚2
Solution:
𝑃𝑅𝑖
𝑡= +𝐶
2𝜎 − 0.2𝑃
0.165 𝑁⁄𝑚𝑚2 × (0.7815 × 1000)𝑚𝑚
𝑡= +5
(2 × 135 𝑁⁄𝑚𝑚2 ) − (0.2 × 0.165 𝑁⁄𝑚𝑚2 )
𝒕 = 𝟕. 𝟏𝟓𝒎𝒎 ≈ 𝟖𝒎𝒎
Step 2:
Evaluate the external diameter of the shell by using the value of the thickness,
𝐷o = 𝐷𝑖 + 2𝑡
𝐷o = 1563𝑚𝑚 + (2 × 8)𝑚𝑚
𝑫𝐨 = 𝟏𝟓𝟕𝟗𝒎𝒎
Step 3:
Check for circumferential and longitudinal stresses whether our design is safe or not, by comparing
and contrasting with the allowable stress of the shell material by using the required thickness.
Check for longitudinal stress

Figure 6: circumferential and longitudinal stresses

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 𝑃𝑖 𝐷𝑖
𝜎𝐿 =
4𝑡
0.165 𝑁⁄𝑚𝑚2 × 1563𝑚𝑚
𝜎𝐿 =
4 × 8𝑚𝑚
𝝈𝑳 = 𝟏𝟎. 𝟕𝟓 𝑵⁄𝒎𝒎𝟐
Since, 𝝈𝑳 ≪ 𝝈𝒂𝒍𝒍 our design is safe.
Check for hoop or circumferential stress

Figure 7: hoop stress on a shell

𝑃𝑖 𝐷𝑖
𝜎ℎ =
2𝑡
0.165 𝑁⁄𝑚𝑚2 × 1563𝑚𝑚
𝜎ℎ =
2 × 8𝑚𝑚
𝝈𝒉 = 𝟐𝟏. 𝟒 𝑵⁄𝒎𝒎𝟐
Since, 𝝈𝒉 ≪ 𝝈𝒂𝒍𝒍 our design is safe.
Maximum shear stress is given by:
(𝜎ℎ − 𝜎𝐿 )
𝜏𝑚𝑎𝑥 =
2
(21.4 − 10.7) 𝑁⁄𝑚𝑚2
𝜏𝑚𝑎𝑥 =
2
𝝉𝒎𝒂𝒙 = 𝟓. 𝟒𝟔𝟓 𝑵⁄𝒎𝒎𝟐
Hence, 𝝉𝒎𝒂𝒙 ≪ 𝝈𝒂𝒍𝒍 , our design is safe.

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 3.2 Design of hemispherical head

For no difference in dilation between the two parts (equal diametrical strain) it can be shown that
for steels (Poisson’s ratio D0.3) the ratio of the hemispherical head thickness to cylinder thickness
should be 7/17. However, the stress in the head would then be greater than that in the cylindrical
section; and the optimum thickness ratio is normally taken as 0.6
𝑡 = 0.6𝑡𝑠ℎ𝑒𝑙𝑙
𝑡 = 0.6 × 8𝑚𝑚
𝒕 = 𝟒. 𝟖𝒎𝒎 ≈ 𝟓𝒎𝒎

Figure 8: hemispherical head

Check for the hoop stresses:
𝑃𝑖 𝐷𝑖
𝜎𝐿 =
4𝑡
0.2 × 1563
𝜎𝐿 =
4×5
𝝈𝑳 =15.63𝑵⁄𝒎𝒎𝟐
Hence, 𝝈𝑳 ≪ 𝝈𝒂𝒍𝒍 ,the design is safe.

3.3 Design of Nozzle

The nozzle design is performed According to ASME code section VIII. The nozzle is the part that
have high stress concentration so it must be design safely with reinforcement.
The vessel has to nozzle one center is on the shell above half meter that used as drain the ethanol
out of the vessel and other is on the top the head of vessel to intake or fill the vessel.

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 Figure 9: section of nozzle
Steps for Nozzle Design Calculation
• Minimum shell thickness required is given by
𝑃𝑅𝑖
𝑡= +𝐶
2𝐽𝜎 − 0.2𝑃
0.2 𝑁⁄𝑚𝑚2 × (0.7815 × 1000)𝑚𝑚
𝑡= +5
(2 × 0.85 × 135 𝑁⁄𝑚𝑚2 ) − (0.2 × 0.2 𝑁⁄𝑚𝑚2 )
𝒕 = 𝟕. 𝟏𝟓𝒎𝒎 ≈ 𝟖𝒎𝒎
But from ASME standard the radius of nozzle is not to exceed 70% radius of the vessel i.e.
𝑅𝑛
< 0.7
𝑅𝑣
𝑅𝑛
We have taken; < 0.7
𝑅𝑣

𝑅𝑛 = 0.3 × 0.7815𝑚
𝑅𝑛 = 0.235𝑚
• Minimum required nozzle thickness is given by:
p
𝑡𝑛 = 𝑅𝑛 (exp( ⁄(𝜎 × 𝐽)) − 1)

𝑡𝑛 = 0.235 × 1000(exp(0.2⁄(135 × 0.85)) − 1)

𝑡𝑛 = 0.4𝑚𝑚 + 5𝑚𝑚
𝒕𝒏 = 𝟓. 𝟒𝒎𝒎 ≈ 𝟔𝒎𝒎
The length of the nozzle that insert in the vessel is;
𝐿1 = 2.5𝑡𝑛 = 2.5 × 6𝑚𝑚
𝑳𝟏 = 𝟏𝟓𝒎𝒎

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The length of the nozzle outside the vessel is based on designer preference too much long distance
waste the material and To much short distance taken the nozzle affected by pressure and become
highly stressed due to this reason we have taken,
𝑳𝟏 = 𝟗𝟓𝒎𝒎
The place of nozzle also sealed at low stressed area of vessel is better. we seat the nozzle one at
the top head and nozzle two above (1m to the top) the saddle support.

3.4 Opening Reinforcement

Due to the reduction of some area from the vessel, at that place stress became increased, so we
must reinforce(compensate) that area.
Conditions,
• 𝐴1 + 𝐴2 + 𝐴3 ≥ 𝐴, Opening is adequately reinforced.
• 𝐴1 + 𝐴2 + 𝐴3 < 𝐴, Opening is not adequately reinforced so
reinforced element must be added or thickness increased
where: 𝐴1 - Area of excess thickness in shell or head
𝐴2 - Area of excess thickness in shell or head
𝐴3 - Area available in the nozzle projecting inward
➢ A1=2(J1t-ftr) (t+tn)
Where t- actual thickness of head (without corrosion allowance)
t=8-5=3mm=0.003m
tr- calculated thickness of the shell
tr=7.15mm=0.00715m
tn-actual thickness of nozzle
tn=6mm=0.006m
J1 -the joint efficiency which is equal to 0.85
f- correction factor which is equal to 1
A1=2(0.85*0.003-1*0.00715) (0.003+0.006)
A1=8.28*10-5m2

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➢ A2=2(2.5t) (tn-trn) tr
Where trn- calculated thickness of nozzle
trn=0.00184mm
A2=2(2.5*0.003) (0.006-0.0184)0.00715
A2=0.0442*10-5m2
➢ A3=2(tn-c) tr*h
where h -inside height, h=15mm=0.015m
A3=2(0.006-0.005)0.00715*0.015
A3=0.022*10-5m2
➢ A=dtrf where d- diameter of nozzle
A=2*0.1563*0.00715*1
A=2.2*10-3m2
Therefore
3

 Ai = A1 + A2 + A3 =8.28*10-5+0.0442*10-5+0.022*10-5
i =1

A1+A2+A3=8.3462*10-5
A1+A2+A3<A, Opening is not adequately reinforced so
reinforced element must be added or thickness
increased
Area of reinforcement pad
Aref=A-AT=2.2*10-3-8.3462*10-5
Aref=2.12*10-3m2

3.4 Manhole Design

From the ASME standard of the vessel which have internal diameter exceed 10in must have one
manhole and its size also based on shoulder of man, so
We have taken the diameter 0.5m. The material is carbon steel.

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 Thickness analysis
PR
tm= +c
J - 0.6P
0.2 * 0.7815
tm= + 0.005
135 * 0.85 − 0.6 − 0.2
tm=0.0064m=6.4mm
Approach to the standard tm=7mm
Inward height of the manhole

L=2.5tm
L=2.5*7
L=17.5mm
Out ward height of the manhole we assume L=95mm like the nozzle height. The Position of the
manhole at the lower cross section of the vessel shell for the safety.
The Manhole opening also needs reinforcement so we must be reinforcing it.
The excess area available on the components is enough to reinforce without any wedding of pad.
3.4.1 Manhole opening reinforcement

➢ Excess area in shell

A1=2(Jt-ftr) (t+tm)
Where:
t-actual thickness of shell
t=8-5=3mm
tr-calculated thickness of shell
tr=8mm
tm-actual thickness of manhole
tm=5mm
A1=2(0.85*3-1*8) (3+6)
A1=24.8mm2

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➢ Area excess in outward manhole
A2=2(2.5*tm) (tm-trm) tr
where
trm -calculated thickness of man hole
trm=5.4mm
A2=2(2.5*5) (6-5.4)8
A2=120mm2
Area excess in inward manhole
A3=2(tm-c) tr× ℎ where ℎ is inside height. (h=0.0235m)
A3=2(5-2) *8*0.0235
A3=1.128m2
The required area is;
A=dtrf=0.1563*5*1=0.7815m2
❖ A1+A2+A3=24.8*10-6+120*10-6+1.128=1.128m2
A1+A2+A3>A
This indicate area excess available in the component is enough for reinforcement.

3.5 Design of flange

Flanges can be designed by detailed calculation or by using tabulated data tables. For our design
flanges are designed by using tables that are formulated by ASME for different material since it
leads to better accuracy of as compared to detailed calculation.
We designed flange for nozzles and manhole. We have to select carbon steel for all flanges and
the rating is 300.
The flange type is weld neck flange because it manufactured easily and easily assembled.
Slip-on flanges, Figure: 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. Figure shows a forged
flange with a hub; for light duties slip-on flanges can be cut from plate.

26 | P a g e
 Figure 10: section of flange
➢ Flange design for nozzles
Size of flange (D)=R*2=0.235*2m=0.47m=18in
From table,
Outside diameter of flange O= 711.2mm
Thickness of flange Q=60.3mm
slip on screwed socket Ys=88.9mm
lap joint YL=130.2mm
Welding neck Yw=158.8mm
Diameter of hub at bevel H=457.2mm
Diameter of hub at face X=533.4mm
Diameter of raised face R=533.4mm
Welding neck Bw=438.2mm
Slip on socket Bs=462.3mm
Lap joint BL=462.3mm
Diameter of bolt circle C=628.7mm
➢ Bolt design for connecting flange of nozzle
We can determine all the dimension of bolt using table. And we used this table for our design
The material for bolt we select carbon steel we have bolt hole diameter
Diameter of bolt hole d=34.9mm
Diameter of bolt=31.8mm
Number of bolt N=20
minimum thread length-69.9mm

27 | P a g e
➢ Flange design for manhole
Size of flange (D)=0.5m=20in
Outside diameter of flange O= 774.7mm
Thickness of flange Q=63.5mm
slip on screwed socket Ys=95.3mm
lap joint YL=139.7mm
Welding neck Yw=136.5mm
Diameter of hub at bevel H=508mm
Diameter of hub at face X=587.4mm
Diameter of raised face R=584.2mm
Welding neck Bw=489mm
Slip on socket Bs=513.1mm
Lap joint BL=514.4mm
Diameter of bolt circle C=685.8mm
➢ Bolt design for connecting flange of manhole
We can determine all the dimension of bolt using table. And we used this table for our design
The material for bolt we selects carbon steel we have bolt hole diameter
Diameter of bolt hole d=34.9mm
Diameter of bolt=31.8mm
Number of bolt N=20
minimum thread length-73mm
3.6 Design of saddle 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; 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.

28 | P a g e
 Though saddles are the most commonly used support for horizontal cylindrical vessels, legs can
be used for small vessels. A horizontal vessel will normally be supported at two cross-sections; if
more than two saddles are used the distribution of the loading is uncertain.
A vessel supported on two saddles can be considered as a simply supported beam, with an
essentially uniform load, and the distribution of longitudinal axial bending moment will be as
shown in Figure. Maxima occur at the supports and at mid-span.

Figure 11: moment applied on horizontal pressure vessel due saddle support

The dimension of the saddle of the vessel is as shown below,

Figure 12: dimension of saddle support to be design

The cylindrical shell acts as a beam over the two supports to resist bending by the uniform load of
the vessel and its contents. The total weight of the vessel and its contents is equal to 2Q. If the
vessel is composed of a cylindrical shell with a formed head (i.e. tori spherical, elliptical, or
hemispherical) at each end that is supported by two saddle supports equally spaced and with A
≤ 0.25 L, then the moment at the saddle, M1,the moment at the center of the vessel, M2, and the
shear force at the saddle, T, may be computed using the following equations,

29 | P a g e
Firstly, we have to find the net weight of the vessel
➢ Weight of the shell:
Ws=0.898*D*L*t
Ws=0.898*1.579*5*0.008*1000m
Ws=56.7kg
➢ Weight of the head
Wh=0.307*D2*t
Wh=0.307*1.5792*0.005*1000
Wh=3.8kg
➢ Weight of two heads WH=2*Wh
WH=2*3.8
WH=7.6kg

➢ Weight of the nozzle

W1=0.898*d*Li*t
W1=0.898*0.47*0.015*0.006
W1=0.000038*105kg

W2=0.898*d*Lo*t
W2=0.898*0.47*0.095*0.006
W2=0.00024*105kg

W3=0.898*D*Li*t
W3=0.898*0.482*0.015*0.006
W3=0.000041*105kg

W4=0.898*D*Lo*t
W4=0.898*0.482*0.095*0.006
W4=0.00026*105kg

30 | P a g e
 Wref=0.898*d*l*t
Wref=0.898*2*0.026*0.015*0.006
Wref=0.000023*105kg

➢ Total weight of the nozzle

Wn=W1+W2+W3+W4+Wref
Wn=0.000038+0.00024+0.000041+0.00026+0.000023
Wn=0.000602*105kg
Wn=60.2kg
Calculating Weight of Fluid
Wf=ρ*v density of petroleum is 881kg/m3
Wf=881*10
Wf=8810kg
The vessel also another additional component like flange and manhole. To get the weight of this
material we use the formula, table or assumption but we assume the weight of these material for
our design,
Weight of another components Wo=20kg
Operating weight of vessel
Q=Ws+Wh+WHWn+Wf+Wo
Q=56.7+3.8+7.6+60.2+8810+20
Q=8958.3kg*9.81m/s2
Q=87880.9N

A=0.25L=0.25*5m=1.25m
H- diameter of shell divided by two
H=D/2=1.563/2=0.7815m
R-outer radius of the shell
R=1.579/2=0.7895m

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 1.25 (0.78952 − 0.78152 )
(1 − + )
5 2 ∗ 1.25 ∗ 5
𝑀1 = −87880.9 ∗ 1.25(1 − )
4 ∗ 0.7815
(1 + )
3∗5
M1=-41580Nm

2∗(0.78952 −0.78152 )
87880.9∗5 1+ 1.25
52
M2= ∗( 4∗0.7815 −4∗ )
4 1+ 5
3∗5

M2=69027Nm
The longitudinal membrane plus bending stresses in the cylindrical shell between the supports are
given by the following equations. Longitudinal stress at top of the shell:
𝑃 × 𝑅𝑚 𝑀2
σ1 = − 2 ×𝑡
2×𝑡 𝜋 × 𝑅𝑚
0.2 × 1571 69027
σ1 = +
4×8 𝜋 × 15712 × 8
σ1 =9.82kgf/mm2
Longitudinal stress at the bottom of the shell:
𝑃 × 𝑅𝑚 𝑀2
σ2 = + 2 ×𝑡
2×𝑡 𝜋 × 𝑅𝑚
0.2 × 1571 69027
σ2 = −
4×8 𝜋 × 15712 × 8
σ2 =9.82 𝑘𝑔𝑓/𝑚𝑚2

32 | P a g e
Longitudinal stress at top of Shell at supports
𝑃 × 𝑅𝑚 𝑀1
σ3 = − 2 ×𝑡
2×𝑡 𝜋 × 𝑅𝑚

0.2 × 1571 41580

σ3 = +
4×8 𝜋 × 15712 × 8
σ3 = 9.82𝑘𝑔𝑓/𝑚𝑚2
Longitudinal stress at the bottom of the shell at support
𝑃 × 𝑅𝑚 𝑀1
σ4 = + 2 ×𝑡
2×𝑡 𝜋 × 𝑅𝑚
0.2 × 1571 41580
σ4 = −
4×8 𝜋 × 15712 × 8
σ4 = 9.8𝑘𝑔𝑓/𝑚𝑚2
T is the total shear force induced on the saddle support and it is determined by the following
equation:
Maximum shear force in the saddle
𝑄 × (𝐿 − 2 × 𝐴)
𝑇=
4
𝐿 + 3𝐻

87880.9 × (5 − 2 × 1.25)
𝑇=
4
1.25 + 3 0.7815

𝑇 = 95856.13𝑁
Table 2: comparisons of stresses
Sr. Stress Location Manual Allowable
No and Type Result Stress
(N/mm2) (N/mm2)
1 σ1 96.22 135
2 σ2 96.22 135
3 σ3 96.33 135
4 σ4 96.44 135

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We conclude that the calculated stress is less than the design or allowable stress so that the design
is safe.

Figure 13: saddle support analysis

To get the above dimension we use the table by determining the vessel diameter therefore the
vessel diameter is 1.579m approximately 1.6m then:
Maximum weight =330kN
V=0.98m
Y=0.2m
C=1.41m
E=0.62m
J=0.35m
G=0.14m
t2=12mm
t2=10mm
bolt diameter d=24mm
bolt hole h=30mm
By using the above dimension, we have to design the saddle support of the horizontal pressure
vessel.

34 | P a g e
3.7 Welding design

L=Length of the weld= 75mm=2.925in

t=Throat area= 0.707*L=53.025mm=1.69in
Ʈ =Direct shear stress due to weight of the vessel
Ʈ =W/A, ........formula where w is total operating weight.
Z= section modules.
Allowable stress=the stress of the support material which is saddle =18500psi
σ=maximum normal stress.
τ =maximum shear stress
A=π *L*t =3.14*2.925*1.69
A= 394.462mm
Therefore
Ʈ =3338.968/15.53.
Ʈ =1.482 N/mm2
σb=Bending stress = m/z
m=37.15*2.925
108.663psi.
Z =Section modules
Z= (π *L2)/4
Z= (3.14*2.925*2.925)/4
Z=170.688mm
Therefore
Bending stress =
σb =M/z=108.663/6.72=16.17psi
σ=Then the maximum normal stress is
σ=1/2 σb +1/2(σb 2+ Ʈ2)1/2
σ=0.5*16.17+0.5(16.172+2152)0.5

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 σ=0.799 N/mm2
The design is safe.
τ=The maximum shear stress is calculated as
τ=1/2(σb 2+ Ʈ2)1/2
τ=0.74325484N/mm2
From the above result we can conclude that the maximum shear and normal stress are less than the
material allowable stress therefore the design is safe.

36 | P a g e
 CHAPTER 4
SUMMARY

4.1 RESULT

All parts of the vessel have designed according to ASME code in the given parameters.
We have calculated the yield strength and other major stress which are taken place on the selected
material.

4.2 CONCLUSION

From our design we can conclude that the design of pressure vessel initialized with the
specification requirements based on the standards given on reference books. Our design of pressure
vessel mainly considers on the basic procedures of each and every component design.
Regarding the storage capacity of fluid for pressure vessels system should be preferred due to its
simplicity, better sensitivity, higher reliability, low maintenance, compactness for the same and
similar capacity. The selection of the materials is based on the ASME standards.

4.3 RECOMMENDATION

➢ When we go over through this project, we have done many calculations. Hence, we made
approximations and assumptions so results may vary.
➢ On this so many sources are available, anyone can get access.

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REFERENCES
Design of Machine Elements | 4th Edition Paperback – 1 July 2017
by V B Bhandari (Author)

Casti Guidebook to ASME Section VIII Division 1 - Pressure Vessels, 2nd Edition Paperback –
Import, 16 August 2000
by N/A Casti Publishing (Author)

Coulson and Richardson's Chemical Engineering: Chemical Engineering Design v. 6 (Coulson &
Richardson's Chemical Engineering) Paperback – Import, 31 May 1996
Pressure Vessel Design Manual Paperback – 1 January 2005
by Moss (Author)

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