design of pressure vessel 1

7/30/2019 Design of pressure vessel 1

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1. INTRODUCTION

1.1General

A pressure vessel is a closed container designed to hold gases or liquids at a pressure

substantially different from the ambient pressure. They are used to store fluids under pressure.

The pressure vessels are designed with great care because rupture of pressure vessels means an

explosion which may cause loss of life and property. The material of pressure vessels may be

brittle such that cast iron or ductile such as mild steel. And pressure vessels are classified

mainly into two types, (a) According to Dimensions (b) According to end Construction.

The pressure vessels, according to the dimensions are classified as thin and thick

shells. The ratio of internal diameter and wall thickness is the factor which differentiates

between thin and thick shells. If the ratio d/t is more than 10, then it is called thin shell and if

this ratio is less than 10 it is said to be thick shell. The examples of the thin shells are pipes,

boilers and storage tanks while the thick shells are used in pressure cylinders, Gun barrels, etc.

The pressure vessels according to end construction are classified as open end and

closed end. A simple cylinder with a piston is an example of closed end vessel. In case of open

end vessels the circumferential stress is induced in addition to the circumferential stress.

And according to role of process vessels are mainly classified into four types:

(a) Reaction pressure vessel

(b) Heat exchanger pressure vessel

(c) Separation pressure vessel

(d) Storage pressure vessel


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The objective of this project is to design a Horizontal storage pressure vessel which

can store LPG (liquid petroleum gases). In general storage pressure vessels are used to hold

liquid or gaseous materials, storage media or container to balance the pressure from thebuffering effect. In order to achieve better design results, ASME boiler codes were taken into

consideration.

Fig: 1.1 Typical Horizontal LPG Storage Vessel Assembly.

1.2 Objectives and Scope

A Finite element model is developed to analyze the behavior of Horizontal mounded

LPG Vessel subjected to Hydro test condition, and Operation Condition. In Hydro Test

condition, water is used to test the vessel as its density is higher than the LPG. Following are

the loads considered in Hydro Test Condition:


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- Pressure inside the vessel.

o Hydro Test Pressure.

o Pressure Due to Head (100 % head of water).

- External Pressure on the vessel.

- Self weight due to gravity.

In Operating Condition, LPG is used and the vessel is subjected to varying pressure at

different heads as 25%, 50%, 75%, 90%. Following are the loads considered in Operating

Condition:

- Internal Pressure

- Pressure due to head (25%, 50%, 75% and 90%)

- External pressure

- Mound pressure at dome ends

- Mound pressure on cylinder

- Self weight due to gravity.

The Primary Objective of this project is to build a pressure vessel with design

parameters and to validate the results using finite element method.

Scope of designing a pressure vessel covers the design basis for following equipment:

- Vessel

- Stiffeners

- Domes / Hemispherical D-ends

- Sand filling in differential settlements.

- Man-holes


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- Pad-Plates

1.3 Classification of Pressure Vessels:

a) Class 1:Vessels that are to contain lethal or toxic substances.

Vessels designed for the operation below -20C.

Vessels intended for any other operation not stipulated in the code.

b) Class 2:

Vessels which do not fall in the scope of class1 and class 3 are to be termed as class2

vessels. The maximum thickness of shell is limited to 38 mm.

c) Class 3:

There are vessels for relatively light duties having plate thickness not in excess of

16mm, and they are built for working pressures at temperatures not exceeding 250C and

unfired.

Class 3 vessels are not recommended for services at temperature below 0C.

1.4 Categories of Welded Joints:

The term categories specify the location of the joint in a vessel, but not the type of

joint. These categories are intended for specifying the special requirements regarding the joint

type and degree of inspection. IS-2825 specifies 4 categories of welds as shown in the below

figure.

a) Category A: Longitudinal welded joints within the main sheet, communicating

chambers, nozzles and any welded joints within a formed or flat head.

b) Category B : Circumferential welded joints within the main shell, communicating

chambers, nozzles and transitions in diameter including joints between the transitions


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and a cylinder at either the large of small end, circumferential welded joints connecting

from heads to main shells to nozzles and to communicating chambers.

c) Category C: welded joints connecting flanges, tubes sheets and flat heads to mainshells, to formed heads, to nozzles or to communicating chambers and any welded

joints connecting one side plate to another side plate of a flat sided vessel.

d) Category D: welded joints connecting communicating chambers or nozzles to main

shells, to heads and to flat sided vessels and those joints connecting nozzles to

communicating chambers.

Fig: 1.2 Welded Joints- Typical Locations and Categories of Welds

1.5 About ASME Codes:

The American Society of Mechanical Engineers (ASME) set up a committee in 1911

for the purpose of formulating standard rules for the construction of steam boilers and other

pressure vessels. This committee is now called the Boiler and Pressure Vessel Committee. The

Committees function is to establish rules of safety, relating only to pressure integrity,

governing the construction of boilers, pressure vessels, etc.,. This Code contains mandatory


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requirements, specific prohibitions, and non mandatory guidance for construction activities.

The Code is not a handbook and cannot replace education, experience, and the use of

engineering judgment. The phrase engineering judgment refers to technical judgments madeby knowledgeable designers experienced in the application of the Code. Engineering

judgments must be consistent with Code philosophy and such judgments must never be used

to overrule mandatory requirements or specific prohibitions of the Code.

The Boiler and Pressure Vessel Committee deals with the care and inspection of

boilers and pressure vessels in service only to the extent of providing suggested rules of good

practice as an aid to owners and their inspectors.


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2. Design of Pressure Vessel

2.1 Codes and Standards:

Following are the codes and standards had been used to design the Horizontal pressurevessel:

ASME SEC. VIII DIV.1 / For Pressure vessels

ASME SEC. VIII DIV.2 For Pressure vessels (Selectively for high

pressure / high thickness / critical service)

ASME SEC. VIII DIV.2 For Storage Spheres

ASME SEC. VIII DIV.3 For Pressure vessels (Selectively for high pressure)

2.2 Types of Storage Vessels

2.2.1 Selection of storage vessel type:

In addition to mounded storage, other types of vessel may also be used to store LPG

under pressure at ambient temperature.

Spheres

Bullets (above ground)

Submerged storage vessels

The principle of the design of the vessel is that the axi-symmetrical loads are carried

by the shell plates, while bending stresses due to non-symmetrical loads are carried by the

shell plates, while bending stresses due to non-axisymmetrical loads are carried by stiffening

rings (except for small diameter, unstiffened vessels).


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Fig: 2.1 Bullet (above ground)

The vessels are subjected to internal vapour and hydrostatic liquid pressure (positive

and negative) and to external loads due to the mound (in the axial and radial directions).

Moreover, they may be subjected to live loads and earthquakes, depending on local conditions,

and even to blast waves due to exploding gas clouds.

If a vessel is to be hydrostatically pressure tested on its foundation, longitudinal seams

need to be avoided in the bottom 120 of the circumference. The vessel heads should have a

torispherical or hemispherical shape.

For an optimal vessel design, the thickness of the shell plates should be determined by

the maximum internal design pressure or the external load caused by the mound in

combination with negative internal pressure (if applicable), whichever is governing. For long

vessels with a relatively small diameter, it is necessary to avoid shell plate thickness being

governed by the longitudinal stresses caused by bending, friction etc.Nozzles and manholes should be designed to withstand both internal and external (soil

pressure and piping) loads. Lateral soil pressure on nozzles and manholes can be eliminated by

the use of sleeves.


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Fig: 2.2 Submerged Storage Vessels

The following guidelines are recommended for determining vessel dimensions:

2.2.2 Overall Dimensions

The vessel diameter should not exceed 8 m.

The required volume should not exceed 3500 m3.

2.2.3 Governing load

The vessel length should not exceed 8 times the vessel diameter. In vessels with

a length-to-diameter ratio exceeding this limit, the longitudinal stresses caused by

bending and frictional forces will govern the shell plate thickness, which leads to a less

economical design.

2.2.4 Material and welding constraints:

For vessel cylinder:

- The ratio of plate thickness to vessel diameter should be not less than 1/400.


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- The ratio of external to internal vessel diameter should not exceed 1.3 (this limit is

imposed by the range of applicability which are based on mean diameter rules. For

mounded storage vessels, practical values of the ratio will be well below the limit. For hemispherical domed ends:

- Plate thickness of domed ends should not be less than 12 mm.

- Ratio of plate thickness of domed ends to vessel diameter should not be less

than 0.002 nor exceed 0.16.

For torispherical domed ends:

- The guidelines for hemispherical domed ends apply.

- The inside knuckle radius should not be less than 0.06 times the vessel

diameter.

- The inside knuckle radius should not be less than twice the plate thickness.

- The inside spherical radius should not exceed the vessel diameter.

For stiffening rings:

- For vessels with a diameter of 3.5 m and greater, internal stiffening rings are

normally required.

- The maximum distance between the stiffeners is determining by the

circumferential buckling stress in the shell plates.

- The size of the stiffening rings depends on the load on the vessel and the

distance between the rings. The thickness of the stiffener web and flange

should not exceed the thickness of the shell. For economic reasons, the stiffener


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webs and flanges should be of the same material grade and thickness as the

shell.

- The ratio of the distance between stiffeners to vessel diameter should be of theorder of 1:2 to 1:1. Ratios up to 2:1 may be used, but will lead to less economic

designs.

- Stiffeners placed close to the transition of cylinder to domed end arevery

ineffective.

2.3 DESIGN CRITERIA

Equipment shall be designed in compliance with the latest design code requirements,

and applicable standards/ Specifications.

2.3.1 MINIMUM SHELL/HEAD THICKNESS

Minimum thickness shall be as given below

a)For carbon and low alloy steel vessels- 6mm (Including corrosion allowance not exceeding

3.0mm), but not less than that calculated as per following:

For diameters less than 2400mm

Wall thickness = Dia/1000 +1.5 + Corrosion Allowance

For diameters 2400mm and above

Wall thickness = Dia/1000 +2.5 + Corrosion Allowance

All dimensions are in mm.

b)For stainless steel vessel and high alloy vessels -3 mm, but not less than that calculated as

per following for diameter more than 1500mm.

Wall thickness (mm) = Dia/1000 + 2.5


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Corrosion Allowance, if any shall be added to minimum thickness.

c) Tangent height (H) to Diameter (D) ratio (H/D) greater than 5 shall be considered as

column and designed accordingly.d)For carbon and low alloy steel columns / towers -8mm (including corrosion allowance not

exceeding 3.0mm.

e)For stainless steel and high alloy columns / towers -5mm. Corrosion allowance, if any, shall

be added to minimum thickness.

2.3.2 Vessel sizing

All Stiffeners Based on inside diameter

All Clad/Lined Vessels Based on inside diameter

Vessels (Thickness>50mm) Based on inside diameter

All Other Vessels Based on outside diameter

Tanks & Spheres Based on inside diameter

2.3.3 Vessel End Closures :

- Unless otherwise specified Deep Torispherical Dished End or 2:1 Ellipsoidal Dished

End as per IS - 4049 shall be used for pressure vessels. Seamless dished end shall be

used for specific services whenever specified by process licensor.

- Hemispherical Ends shall be considered when the thickness of shell exceeds 70mm.

- Flat Covers may be used for atmospheric vessels

- Pipe Caps may be used for vessels diameter < 600mm having no internals.

- Flanged Covers shall be used for Vessels /Columns of Diameter < 900mm having

internals.


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- All columns below 900mm shall be provided with intermediate body flanges.

Numbers of Intermediate flanges shall be decided based on column height and type of

internals2.3.4 Pressure

Pressure for each vessel shall be specified in the following manner.

2.3.4.1 Operating Pressure

Maximum pressure likely to occur any time during the lifetime of the vessel

2.3.4.2 Design Pressure

a) When operating pressure is up to 70 Kg./cm2 g , Design pressure shall be equal to operating

pressure plus 10% ( minimum 1Kg./cm2 g ).

b) When operating pressure is over 70 Kg./cm2 g , Design pressure shall be equal to operating

pressure plus 5% ( minimum 7 Kg./cm2g).

c) Design pressure calculated above shall be at the top of vertical vessel or at the highest point

of horizontal vessel.

d) The design pressure at any lower point is to be determined by adding the maximum

operating liquid head and any pressure gradient within the vessel.

e) Vessels operating under vacuum / partial vacuum shall be designed for an external pressure

of 1.055 Kg./cm2 g.

f) Vessels shall be designed for steam out conditions if specified on process data sheet.

2.3.4.3 Test Pressure

a) Pressure Vessels shall be hydrostatically tested in the fabricators shop to 1.5 /1.3/ 1.25

(depending on design code) times the design pressure corrected for temperature.


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b) In addition, all vertical vessels / columns shall be designed so as to permit site testing

of the vessel at a pressure of 1.5/ 1.3 / 1.25 (depending on design code) times the

design pressure measured at the top with the vessel in the vertical position andcompletely filled with water. The design shall be based on fully corroded condition.

c) Vessels open to atmosphere shall be tested by filling with water to the top.

d) 1. Pressure Chambers of combination units that have been designed to operate

independently shall be hydrostatically tested to code test pressure as separate vessels i.e.

each chamber shall be tested without pressure in the adjacent chamber.

2.When pressure chambers of combination units have their common elements designed

for maximum differential pressure the common elements shall be subjected to 1.5/ 1.3

times the differential pressure.

3.Coils shall be tested separately to code test pressure.

e)Unless otherwise specified in applicable design code allowable stress during hydro test

in tension shall not exceed 90% of yield point.

f)Storage tanks shall be tested as per applicable code and specifications.

2.3.5 Temperature

Temperature for each vessel shall be specified in the following manner:

2.3.5.1 Operating Temperature

Maximum / minimum temperature likely to occur any during the lifetime of vessel.

2.3.5.2 Design temperature

a) For vessels operating at 0C and over:


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Design temperature shall be equal to maximum operating temperature plus 15 0C.

b) For Vessels operating below 0C:

Design temperature shall be equal to lowest operating temperature.c) Minimum Design Metal Temperature (MDMT) shall be lower of minimum

atmospheric temperature and minimum operating temperature.

2.3.6 Corrosion allowance :

Unless otherwise specified by Process Licensor, minimum corrosion allowance shall

be considered as follows :

- Carbon Steel, low alloy steel column, Vessels, Spheres : 1.5 mm

- Clad / Lined vessel: Nil

- Storage Tank, shell and bottom : 1.5 mm

- Storage tank, Fixed roof / Floating Roof : Nil

For alloy lined or clad vessels, no corrosion allowance is required on the base metal.

The cladding or lining material (in no case less than 1.5 mm thickness) shall be considered for

corrosion allowance. Cladding or lining thickness shall not be included in strength

calculations. Corrosion allowance for flange faces of Girth / Body flanges shall be considered

equal to that specified for vessel.

2.3.7 Capacity

2.3.7.1 Tank

Capacity shall be specified as Nominal capacity and stored capacity, Nominal capacity

for fixed roof tanks be volume of cylindrical shell. Nominal capacity for floating roof tanks


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shall be volume of cylindrical shell minus free board volume. Stored capacity shall be 90% of

Nominal capacity.

2.3.7.2 Sphere

Stored capacity shall be 85% of nominal capacity.

2.3.8 Manholes :

a)Vessels and columns with diameter between 900 and 1000 mm shall be provided

with 450 NB manhole. Vessels and columns with diameter greater than 1000mm shall

be provided with 500 NB manhole. However, if required vessels and columns with

diameter 1200mm and above may be provided with 600NB manhole.

b)For storage tanks minimum number of manholes (Size 500mm) shall be as

follows:

Tank Diameter Shell Roof

Dia. < 8m 1 1

Dia> 8m dia. < 36 dia 2 2

Dia. > 36m 4 2

Floating roofs (pontoon or double deck type) shall be provided with manholes to

inspect the entire interior of the roofs. Size of manhole shall be 500 mm minimum.

2.3.9 Nozzle size :

- Minimum nozzle Size : 40 NB

- Minimum Nozzle Size, Column : 50 NB

- Safety Valve Nozzle : Based on I.D.


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- Self Reinforced Nozzle Neck : Based on I.D.

a) All nozzles and man-ways including self-reinforced type shall be 'set in' type and

attached to vessel with full penetration welds.b)Self reinforced nozzles up to 80mm NB may be 'set on' type.

2.3.10 Vent/Drain Connections:

Vessel shall be provided with one number each, vent/drain connection as per following :

VESSEL VOLUME, m3 VENT SIZE, NB (mm) DRAIN SIZE, NB (mm)

6.0 and smaller 40 40, 6.0 to 17.0 40 50, 17.0 to 71.0 50 80, 71.0 and larger 80 100

2.3.11 Design Theory:

Circumferential or Hoop Stress

A tensile stress acting in a direction tangential to the circumference is called

Circumferential or Hoop Stress. In other words, it is on longitudinal section(or on the cylinder

walls).

Fig: 2.3 Circumferential or Hoop Stresses

Let,

p = Intensity of internal pressure,

d = Internal diameter of the cylinder shell,


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l = length of cylinder,

t = Thickness of the shell, and

t1= hoop stress for the material of the cylinder.Now,

We know that total force on a longitudinal section of the shell

= Intensity of pressure projected Area = p d l ..(i)

and the total resisting force acting on the cylinder wall

= t1 2t l (ii)

From equation (i) and (ii) , we have

t1 2t l = p d l

or

t1 = p x d / 2t

or

t =p x d / 2 t1

2.3.12 Longitudinal Stress

Fig: 2.4 Longitudinal Stress

Let t2 = Longitudinal stress.

In this case, the total force acting on the transverse section


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= Intensity of pressure Cross- sectional Area

= p / 4 (d) 2 . (i)

And total resisting force = t2 d.t (ii)From equation (i) and (ii), we have

t2 d.t = p / 4 (d) 2

t2 = p x d / 4t

2.3.13 Design of shell due to internal pressure:

Fig: 2.5 Shell Internal Pressure

Cylindrical pressure vessel is subjected to tangential (t) and longitudinal(L) stresses.

and

where D= mean diameter=D i + t

The design pressure is taken as 5% to 10% more than internal pressure, where as the

test pressure is taken as 30% more than internal pressure.

Considering the joint efficiency, The thickness of shell can be found by following procedure,


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2.3.14 Design of Elliptical Head:

Elliptical heads are suitable for cylinders subjected to pressures over 1.5 MPa. The shallow

forming reduces manufacturing cost. Its thickness can be calculated by the following

equation:

Where,

di = Major axis of the ellipse

W = Stress Intensification Factor

Where,

2.3.15 Design of Man Hole:

Let,


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di = internal dia. Of nozzle

d = di + 2 CA

Where, CA = corrosion Allowance in mmt = Actual thickness of shell in mm

tr = require thickness as per calculation in mm.

tn = Actual thickness of nozzle

trn = Required thickness as per calculation in mm

2.4 Design Calculation

2.4.1 Thickness of cylinder

Internal pressure (P) = 14.5 kg/cm2 or 1.42187 MPa

Internal Diameter (Di) = 7500 mm

Corrosion Allowance (CA) = Nil.Joint Efficiency for shell = 1.

As per the equation:

t

= 32 mm


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2.4.2 Thickness of Hemi-Spherical D-ends:

Where,

k = 2

Generally, k = 2 (however k should not be greater than 2.6)

= 1

Pi = Internal Pressure = 1.42 N/mm2di = Major axis of ellipse = 1500 mm

W = Stress intensification factor = 1

= 18 mm

2.5 Stresses in Pressure Vessels:


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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 allowedfor by use of a higher safety factor and design rules for details. It is required, how ever, that all

loadings ( 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.

The code established allowable stresses by that the maximum general primary

membrane stress must be less than allowable stresses outlined in material sections. Further, it

states that the maximum primary membrane stress plus primary bending stress may not exceed

1.5 times the allowable stress of the material sections. Higher allowable stresses are permitted

if appropriate analysis is made. These higher allowable stresses clearly indicate that different

stress levels for different stress categories are acceptable.

It is general practice when doing more detailed stress analysis to apply higher

allowable stresses. In effect, the detailed evaluation of stresses permits substituting knowledge

of localized stresses and the use of higher allowable in place of the larger factor of safety used

by the Code. This higher safety factor really reflected lack of knowledge about actual stresses.


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A calculated value of stress means little until it is associated with its location and

distribution in the vessel and with the type of loading that produced it. Different types of stress

have different degrees if significance.

The designer must familiarize himself with the various types of stress and loading in

order to accurately apply the results of analysis. The designer must also consider some

adequate stress or failure theory in order to combine stresses and set allowable stress limits. It

is against this failure mode that he must compare and interpret stress values, and define how

the stresses in a component react and contribute to the strength of that part.

2.5.1 Types of Stresses:

Tensile, Compressive, Bending, Axial, Membrane, Principal, Tangential, Strain

Induced, Longitudinal, Shear, Bearing, Discontinuity, Thermal, Circumferential, Radial.

2.5.2 Classes of stress:

- Primary Stress

o General:

- Primary general membrane stress Pm

- Primary general bending stress Pb

o Primary local stress, PL

- Secondary stress:

o Secondary membrane stress. Qm

o Secondary bending stress Qb

- Peak stress. F

Definition and Examples


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2.5.2.1 PRIMARY GENERAL STRESS:

These stress act over a full cross section of the vessel. Primary stress are generally due

to internal or external pressure or produced by sustained external forces and moments.

Primary general stress are divided into membrane and bending stresses. Calculated

value of a primary bending stress may be allowed to go higher than that of a primary

membrane stress.

Primary general membrane stress, Pm

Circumferential and longitudinal stress due to pressure.

Compressive and tensile axial stresses due to wind.

Longitudinal stress due to the bending of the horizontal vessel over the saddles.

Membrane stress in the centre of the flat head.

Membrane stress in the nozzle wall within the area of reinforcement due to pressure or

external loads.

Axial compression due to weight.

Primary general bending stress, Pb.

Bending stress in the centre of a flat head or crown of a dished head.

Bending stress in a shallow conical head.

Bending stress in the ligaments of closely spaced openings.

2.5.2.2 LOCAL PRIMARY MEMBRANE STESS, PL

Pm+ membrane stress at local discontinuities:

o Head-shell juncture


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o Cone-cylinder juncture

o Nozzle-shell juncture

o Shell-flange juncture

o Head-skirt juncture

o Shell-stiffening ring juncture

Pm+ membrane stresses from local sustained loads:

o Support legs

o

Nozzle loads

o Beam supports

o Major attachments

2.5.2.3 SECONDARY STRESS

Secondary membrane stress Qm

Axial stress at the juncture of a flange and the hub of the flange

Thermal stresses.

Membrane stress in the knuckle area of the head.

Membrane stress due to local relenting loads.

Secondary bending stress, Qb

Bending stress at the gross structural discontinuity: nozzle, lugs, etc., (relentingloadings only).

The nonuniform portion of the stress distribution in a thick-walled vessels due to

internal pressure.


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The stress variation of the radial stress due to internal pressure in thick-walled vessels.

Discontinuity stresses at stiffening or support ring.

Peak Stress F

Stress at the corner of discontinuity.

Thermal stress in a wall caused by a sudden change in the surface temperature.

Thermal stresses in cladding or weld overlay.

Stress due to notch effect. (stress concentration).

2.5.3 FAILURE IN PRESSURE VESSELS

2.5.3.1 Categories of Failures:

Material: Improper Selection of materials; defects in material.

Design: Incorrect design data; inaccurate or incorrect design methods inadequate shop

testing.

Fabrication: Poor quality control; improper or insufficient fabrication procedures

including welding; heat treatment or forming methods.

Service: Change of service condition by the user; inexperienced operations or

maintenance personnel; upset conditions. Some types of services which requires

special attention both for selection of materials, design details, and fabrication methods

are as follows:

o Lethal

o Fatigue (cyclic)

o Brittle (low temperature)


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o High Temperature

o High shock or vibration

o Vessel contents

Hydrogen

Ammonia

Compressed air

Caustic

Chlorides

2.5.3.2 TYPES OF FAILURES

Elastic deformation: Elastic instability or elastic buckling, vessel geometry, and

stiffness as well as properties of materials are protecting against buckling.

Brittle fracture: Can occur at low or intermediate temperature. Brittle fractures have

occurred in vessels made of low carbon steel in the 40-50 F range during hydrotest

where minor flaws exist.

Excessive plastic deformation: The primary and secondary stress limits as outlined in

ASME Section VIII, Division 2, are intended to prevent excessive plastic deformation

and incremental collapse.

Stress rupture: Creep deformation as a result of fatigue or cyclic loading, i.e.,

progressive fracture. Creep is a time-dependent phenomenon, whereas fatigue is a

cyclic-dependent phenomenon


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Plastic instability: Incremental collapse; incremental collapse is cyclic strain

accumulation or cumulative cyclic deformation. Cumulative damage leads to

instability of vessel by plastic deformation.

High Strain: Low cyclic fatigue is strain-governed and occurs mainly in lower

strength/ high-ductile materials.

Stress corrosion: It is well know that chlorides cause stress corrosion cracking in

stainless steels; likewise caustic service can cause stress corrosion cracking in carbon

steel. Materials selection is critical in these services.

Corrosion fatigue: Occurs when corrosive and fatigue effects occur simultaneously.

Corrosion can reduce fatigue life by pitting the surface and propagating cracks.

Material selection and fatigue properties are the major considerations.

2.5.4 Special Problems:

Thick walled pressure vessels

Mono-bloc-Solid Vessel Wall.

Multilayer Begins with a core about in. Thick and successive layers are applied.

Each layer is vented (except the core) and welded individually with no overlapping

welds.

Multi-Wall Begins with a core about in. to 2 in. thick. Outer layers about the same

thickness are successive Shrunk fit over the core. This creates compressive stress in

the core, which is relaxed during pressurization. The process of compressing layers is

called auto-frettage from the French word meaning Self-Hooping.


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Multilayer auto-frettage Begins with a core about in. thick. Bands or forged rings

are slipped outside and then the core is expanded hydraulically. The core is stressed

into plastic range but below ultimate strength. The outer rings are maintained at a

margin below yield strength. The elastic deformation residual in the outer bands

induces compressive stress in the core, which is relaxed during pressurization.

Wire wrapped vessels: Begins with inner core of thickness less than required for

pressure. Core is wrapped with steel cables in tension until the desired auto frottage is

achieved.

Coil warpped vessels: Begin with a core that is subsequently wrapped or coiled with a

thin steel sheet until the desired thickness is obtained. Only two longitudinal welds are

used, one attaching the sheet to the core and the final closures weld. Vessels 5 to 6 ft in

diameter for pressure up to 5000psi have been made in this manner.

2.5.5 THERMAL STRESS

Whenever the expansion or contraction that would occur normally as a result of

heating or cooling an object is prevented, thermal stresses are developed. The stress is

always caused by some form of mechanical restrain.

Thermal stresses are secondary stresses because they are self-limiting. Thermal

stresses will not cause failure by rupture. They can however, cause failure due to

excessive deformations.

2.5.6 DISCONTINUITY STRESSES

Vessel sections of different thickness, material, diameter and change in directions

would all have different displacements if allowed to expand freely. However, since they are


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connected in a continuous structure, they must deflect and rotate together. The stresses in the

respective parts at or near the juncture are called discontinuity stresses.

Discontinuity stresses are secondary stresses and are self-limiting.Discontinuity stresses do because an important factor is fatigue design where cyclic

loading is a consideration.

2.6 Loadings:

Loadings are forces are the causes of stress in pressure vessels. Loadings may be

applied over a large portion (general area) of the vessel or over a local area of the vessel.

General and local loads can produce membrane and bending stresses. These stresses are

additive and define the overall state of stress in the vessel or component.

The stresses applied more or less continuously and uniformly across an entire section of

the vessel are primary stresses.

The stresses due to pressure and wind are primary membrane stresses.

On the other hand, the stresses from the inward radial load could be either a primary

local stress or secondary stress. It is primary local stress if it is produced from an

unrelenting load or a secondary stress if produced by a relenting load.

If it is a primary stress, the stress will be redistributed; if it is a secondary stress, the load

will relax once slight deformation occurs.

Basically each combination of stresses (stress categories will have different allowable,

i.e.,

o Primary Stress: Pm < SE

o Primary membrane local (PL):


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PL = Pm + PL < 1.5SE

PL = Pm + Qm < 1.5SE

o Primary membrane + secondary (Q):

o Pm + Q < 3SE

Loading can be outlined as follows:

Categories of landings

a) General loads Applied more or less continuously across a vessel section.

Pressure loads : Internal or external pressure (design, Operating,

hydrotest, and hydrostatic head of liquid).

Moment loads : Due to wind, seismic, erection, transportation.

Compressive / Tensile Loads : Due to dead weight, installed equipment,

ladders, platforms, piping and vessel contents.

Thermal loads : Hot box design of skirt-head attachment.

b) Local Loads : Due to reactions from supports, internal, attached

Piping, attached equipment, i.e., platforms, mixers, etc.

a) Radial Load: Inward or Outward

b) Shear Load : Longitudinal or circumferential

c) Torsional Load

d) Tangential Load

e) Moment Load: Longitudinal or circumferential

f) Thermal Load

Types of Loadings:


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Steady Loads: Long term duration, continuous.

a) Internal/ external pressure

b) Dead weight

c) Vessel Contents

d) Loading due to attached piping and equipment

e) Loadings to and from vessel supports

f) Thermal loads

g) Wind loads

Non-Steady Loads- Short term duration, Variable.

a) Shop and field hydro-test

b) Earthquake

c) Erection

d) Transportation

e) Upset, emergency

f) Thermal Loads

g) Startup, Shut down

2.7 CAD Details:

Modelling is done using CATIA/Hypermesh.


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Fig: 2.6 Pressure Vessel CAD Details.

2.7.1 Design of Vessel:


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Fig: 2.7 Vessel Design Details

2.7.2 Design of Domes:

Fig: 2.8 Hemi Spherical Dome Ends Design Details.

2.7.3 Design of Stiffeners:


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Fig: 2.9 Stiffeners Design Details

2.7.4 Design of top Domes:

Fig: 2.10 Domes A / Dome B on both ends of vessel.

2.7.5 General Notes:


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a) All Dimensions are in MM unless otherwise specified.

b) All flange bolt holes to straddle vessel principal centre lines unless otherwise specified.

c) Protect all machined surfaces and threaded connections with rust preventive

immediately after machining.

d) Install wood or steel protectors for fittings immediately after testing.

e) No welding is to be done on the vessel after PW heat treatement/SR unless otherwise

permitted by CODE/Specification

f) Assembly Instruction: orientation angles 0, 90, 180, 270 of all components shall match

with that of assly.

g) All pressure resistant butt welds are full peneration welds with backchipping and

welding. wherever backchiping is not possible, root run shall be carried by GTAW.

h) All sharp corners/ edges are to be rounded off

i) The vessel shall have a slope of 1:200 between the ends with respective to true

horizontal, however manway/ nozzle falnge faces except the liquid receipt nozzle N1

shall be perpendicular to horizontal.

j) Centerline of all nozzles, manholes and dome shells shall coinside with true vertical

axis and shall be chekced with plumb


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k) The out of roundness of the cylinder sections i.e, the difference between max & min

internal diameters measuered at any cross section shall not exceed 0.5% of nominal

diameter.

l) For circumferential joints, the misalignment of center lines of plates shall not exceed

10% of thickness of thinner plate or 3mm. Whichever is smaller.

m) Prior of hydrotest, all weld spatter, metal dust etc. shall bee removed from the tank

during hydrotest, tank is to be supported on sand bed.

n) During hydro test settlement monitoring of the equipment supported on its foundation

should be per-formed. The settlement should be monitored at 0,25,75, & 100 % filling

and after 48 hrs with the vessel completely fillet. The settlement rate during the testing

period must dimnish with the time. Otherwise the vessel must be partially emptied and

corrective action on the foundation shall be taken.

o) Radiography shall be carried out before and after post weld heat treatement and wet

fluorescent magnetic particle testing of weld shall be carried out after post weld heat

treatment.

p) All forgings aand nozzle flanges shall be MP/DP tested. After machining.

q) Nozzle necks fabricated from plate to be fully radiographed

r) Hydrotest of vessel shall be carried out only once after stress relieving.

s) NDT of weld joint after hydrotest is as follows:


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a) 100% UT exam of all the T joints from inside of the vessel followed by DP Test.

b) MP of 1/3rd length of fillet of inner shell to stiffener web as 0,90,180,270deg

c) after final inspection, before gassing up and commissioning of LPG bullet an

ultrasonic shell thickness testing shall be carriedout on the internal walls of bullet at

the points designated for 5 years.

t) Gasket seating surface of all flanges shall have smooth finishing to 125 AARH.

u) surface preparation & painting:

a) all external surfaces steel structure painting council: surface preparation

specification SSPC - SP10, painting: polyurethane protective coating as per techincal

specification.

v) Flange dimensions for nozzles shall be as per ANSI B16.5 upt 24 size/ANSI B16.47

series-A above 24 size.

w) All nozzles shall be provided with insulating gasket, insulating bolts and insulating

washer under backup washers and nuts.

x) Bullet is to be stress relievedand hardness controlled at 200 BHN after PWHT.

y) bullet is to be designed for the following foundation settlement values

a) immedciate settlement value 5mm

b) longterm settlement value 15mm


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c) maximum settlement value between center and end of the bullet (at empty,

operation, testcase) 10mm

z) RF pads shall be pnematically testedfor tightness to 1.0KG/sq.cm(g) with soap solution

on all attachmetn welds.

aa) RF pad is to be welded to shell such that one tell tale hole will be at the bottom of the

pad in the erected position of the vessel.

bb) Tell tale/Vent holes shall not be plugged and shall be filled with hard grease after

PWHT.

cc) wherever RF pad crosses the weld seam the weld seam is to be ground flush.

dd) calibration: volumetric capacity calibration shall be done for the LPG bulletby

statutory authorities

ee) Work at site:

a) each bullet will be fabricated in 4 sections in vertical position, stiffener rings shall

be welded in this position, These modules will be turned horizontal dn each module

placed on temporary saddles maintaing slope as per drawing assembly and welding of

all nozzles, doems and D'ends shall be completed on the individual modules.

b) after completion of welding of all attachments,the modules shall be individually

stress relieved.


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c) all external surfaces of modulesshall be blast cleaned and painted. the modules will

be placedon foundation bed maintaining slope and C-Seams between modules shall be

completed. local SR will be carried out for these C-Seams followed by Hydrotest ofbullet.

d) the PH of fresh water used for hydrotestshall be between 6 & 7.The vessel shall be

completely drained, cleaned and dried with hot air after Hydrotest.

e) Internal surfaces of bottom 90 section of bullet shall be properly cleaned and painted

with amine cured epoxy paint. Frames shall be used for lifting/handling the shell

courses over the foundation during different stage of fabrication, inspection and

testing.

2.8 Design Data for CAD:

a) Process fluid : LPG (commerical grade)

b) Design pressure - internal : 14.5 Kg/cm2(g)

c) Design Pressure External : 1.856 Kg/cm2(g)

d) Design Temperature : -27 to +55 C

e) Hydro Test Pressure : 19.75 Kg/cm2(g)

f) Operating temperature : Amb C

g) Water Capacity : 2165 Cu.m

h) Storage Capacity of LPG (working) : 1000 MT


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i) Position : Horizontal

j) Dished Ends : Hemispherical

k) Class of Hazard : Flammable

l) No. of bullets : 3

m) Liquid flow rate (feed) : 330 Cu.m/hr

n) Liquid flow rate (loading) : 200 Cu.m/hr

o) Boing Point : Range >-40 C

p) Density of liquid water : 1000 Kg/m3

q) Desnity of LPG : 550 Kg/m3

r) Physical Condition : liquid, Vapour

s) Vapour pressure : 8.5 Bar @ 20C

t) Flash Point : -104 C

u) COmposition : propane -60%, Butene-40%

v) Physical state : Gas at 15C at one ATM

w) Design Code : PD 5500

x) Radiography : 100 % Before and after PWHT

y) Weld joint efficiency : 1.0


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z) Vapour Density :1.5

aa) Corrosion allowance : 1.5mm

bb) Length of Vessel : 44000mm

cc) Diameter of vessel : 7500mm

dd) Empty Weight : 319Tonnes

ee) Hydro Test Weight : 2505 Tonnes

ff) Operating Weight : 1379 Tonnes

gg) Painting -External : Polyurethane Coating

hh) Painting Internal : Surface preparation specification SSPC SP10

2.9 Materials Used in Design:


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Fig: 2.11 Material Details

Inorder to meet the requirements of ASME Boiler and Pressure Vessel Codes, SA 537

CL- 1 is a grade of Carbon Manganeses-Silicon Steel is used in this design. Here are the

compositions used in this material:

Pmax = 0.035

Smax = 0.040

Cu max = 0.035

Ni max = 0.25Cr max = 0.25

Mo max = 0.08

Heat Treatment = normalized, ultrasonically tested, impact tested


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Tensile Strength = 70 90 Ksi / 485 620 MPa

Yield Strength = 50 Ksi / 345 MPa

Other Characteristics of this material:- The plates shall be free of scales and rolled in the direction of length specification and

shall be supplied in the normalized condition. Accelerated cooling by liquid quenching

or other means is not permitted.

- The plates shall be supplied with gas/sheared edges with tolerances as per SA 20 latest.

Manual gas cutting is not acceptable. Tolerance on thickness shall be positive only.

- All the plates shall be supplied in normalized condition.

- The plates shall be free from injurious defects and shall have work-man like finish.

Reconditioning/ repair of plates by welding is not permitted.

- The carbon content for plates shall not exceed 0.23%.

o Additionally, one of the following requirements for carbon equivalent based on

heat analysis, shall be also satisfied:

Ceq = C + 0.42

3. Analysis Theory

3.1 Loading Conditions

3.1.1 Loads due to the Mound:

The weight of the mound above the vessel is transmitted by the vessel to the

foundation and will result in radial loads on the vessel.


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If a vessel is installed in an excavation, it is possible that after backfilling the sides of

the excavation exert a horizontal supporting pressure on the vessel. This beneficial effect

depends very much on the quality and compaction of the fill material. For vessels installedabove grade level only horizontal support pressure from the mound is to be taken into account.

Passive and active soil pressures at the heads of the vessels are to be taken into account

in the design.

Mounded storage vessels ought not to be exposed to traffic or similar live loads on top

of the mound. If this cannot be avoided, a suitable live load shall be taken into account in the

design. The weights of pipelines, pipe bridges, etc. are also to be taken into account, if

applicable.

3.1.2 Loads due to uneven support by the foundation:

Uneven support of the vessel may occur due to:

- Variation in subsoil characteristics along the length of the vessel, which may cause a

variation in support, as the vessel with settle as a rigid body.

- Construction tolerances of the vessel and its foundation, due to which the vessel with

normally not sit perfectly on the foundation.

To allow for the effect of tolerances and uneven support, a maximum and a minimum

soil reaction are assumed. Both soil reaction distributions are to be taken into account. Based

on the soil investigation data, calculations of bending moments, shear forces etc., taking into

account the long-term bedding/sub-grade modulus of the subsoil, should be carried out using a

beam on elastic foundation method.

3.1.3 Loads due to Temperature and Internal Pressure Variations


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Variations in vessel temperature and internal pressure will cause variations in the

length and radius of the vessel. With regard to variations of the vessel length, the following

effects need to be taken into account in the vessel design: Increase of the soil pressure on the domed ends of the vessel.

Longitudinal stresses in the vessel caused by friction and restrained expansion.

The magnitude of these effects depends on the length and the diameter of the vessel,

the thickness of the soil cover and the configuration of the soil cover over the domed ends.

3.1.4 Loads due to Earthquake

If the vessel has to be designed against earthquakes, the earthquake load is usually

introduced into the calculation as a horizontal load on the vessel. A vertical component may

also be taken into account, although its overall effect is usually not significant.

The required design earthquake load is usually specified in local building codes,

expressed in terms of horizontal acceleration and sometimes also vertical acceleration. For

large vessels the effect of liquid sloshing should be considered.

The consequences of liquefaction of the subsoil, if relevant, also need to be considered.

3.1.5 Loads due to an External Explosion:

An explosion of a gas cloud in the vicinity of a mounded storage vessel would result in

a pressure increase on the soil cover. This increase in pressure can be taken into account in the

analysis as an increase of the weight of the soil cover.

The magnitude of the overpressure will be stated by the purchaser. It may, typically,

vary between 0 and 0.15 bar. The overpressure should be multiplied by a suitable reflection


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coefficient. Because of the damping effect of the mound, a reflection coefficient of 1.5 should

be used instead of the more commonly used (in codes) reflection coefficient of 2.

3.1.6 Load Combinations:

In the design of the vessel, the loads previously described need to be combined, as

applicable, with the hydrostatic test pressure, normal operating loads, retest pressures and

those caused by emergencies. Explosion and earthquake loads may be assumed not to occur at

the same time, and not during pressure testing.

3.1.7 Dead Weight (Load 1):

The dead weight Q1 is the weight of the shell over the centre-to-centre distance

between two stiffeners, together with the estimated weight of one stiffening ring. This weight

is allocated to one stiffener and is:

Q1 = 2R x t x m x L + W (kN)

Where m is the weight of vessel material per m3 (kN/m3)

3.1.8 Weight of Liquid Fill (Load 2)

The weight of the maximum volume of liquid Q2 to be allocated to one stiffener is:

Q2 = R2 x 1 x L (kN)

Where 1 is the weight of contents per m3 (kN/m3)

The formula is correct for test conditions. However, in normal operating conditions

there is always a gas pocket, giving a slight reduction of liquid load and thus of stress in the

vessel wall. National regulations may allow the operator of an installation to guarantee a

maximum filling level below 100% and to have this taken into account in the design of the

vessel, leading to a thinner wall than would otherwise have been the case.

3.1.9 Internal Design Pressure (Load 3)


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With regard to the stiffener design, the internal pressure only affects the normal force

N3 in the stiffening ring, not the bending moment of the shear force. Part of the shell plates

will act together with the ring to carry the loads to which the rings are subjected. This part, theworking width w, is:

w = 2 x 0.78 x (R x t) (m)

The internal pressure, acting on the working width, will cause a normal tensile load N3

which is carried by the combination of stiffening ring and working width of the shell plate.

N3 = p3 x w x R (kN)

Where p3 = the internal design pressure

3.1.10 Negative Internal Design Pressure (Load 4)

Internal pressure may become negative is extreme conditions, e.g. very high outward

flow rates, or rupture of a bottom outlet pipe resulting in gas freely bleeding into the open air.

The buoyancy effect of the negative internal pressure adds to the external soil pressure exerted

by the mound.

The negative internal pressure only affects the normal force N4 in the stiffener, not the

bending moment or the shear force. The negative internal pressure p4 will cause a

compressive normal load N4 in the stiffening ring and the plates of the working width.

3.2 Stress Analysis of Mounded Storage Vessels

The difference between the stress analysis for a mounded storage vessel and that for a

conventional pressure vessel is due to the mound and the method of support. The mound and

the support cause bending moments, normal forces and shear forces in the wall of the cylinder.


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3.2.1 Circumferential Bending:

Several loads, such as dead weight, weight of liquid, mound loads, etc. cause

circumferential bending moment in the wall of the cylinder (in addition to the in-planestresses). The relatively thin shell plates of the cylinder cannot carry these bending moments.

3.2.2 Unstiffened Cylinders

In vessels without stiffening rings, the circumferential bending moments will be

transmitted to the domed ends by shear stresses in the shell. The domed ends are, in

comparison with the cylindrical shell, quite rigid and may be capable of carrying these

bending moments provided the length of the vessel is relatively short.

For diameters over approximately 3.5 meters, the required plate thickness of an

unstiffened cylinder tends to become uneconomically heavy (due to negative internal pressure

and/or external pressure) and it is therefore usually more economic to use cylinders provided

with stiffening rings at regular intervals.

3.2.3 Stiffened Cylinders

When stiffening rings are used, the circumferential bending moments in the shell plates

are transmitted to the stiffening rings by shear forces.

If the distance between two stiffeners is L, the bending moments due to all the loads on

a part of the cylinder with a length L need to be allocated to one stiffening ring.

The domed ends are very rigid in a radial direction when welded to the cylindrical

shell, and will also act as stiffeners. Torispherical heads are more rigid stiffeners than

hemispherical heads. A stiffening ring just next to a torispherical head is therefore

unnecessary, and the first stiffener need not be located close to the transition from dome to


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shell. However, irrespective of the chosen spacing between stiffeners, it is recommended that

the stiffeners nearest to the domed ends be located at a distance from the domedhead-to-shell

weld of half the diameter of the vessel.The plate thickness of the domed ends needs to be increased by the relevant percentage

as shown in above table in order not to exceed the maximum design stress. This may add to

the cost of the vessel, but because of uneven external soil pressure the plate thickness of

hemispherical ends is usually more than required for internal pressure only.

3.2.4 Normal forces and shear forces:

The shear forces acting in the plane of the stiffening rings will cause bending moments

and also normal and shear stresses in the ring. It should be noted that the presence of the

stiffening rings, whilst greatly reducing the bending moments in the shell plates, does not

eliminate the normal tensile and compressive stresses in the shell plates.

3.3 Stress and Stability Criteria:

Irrespective of the formulae used in the various applicable design codes (ASME, BSI,

CODAP, etc.) for establishing the dimensions of vessel components (shell thickness, stiffener

dimensions), vessels for use in the European Economic Area must conform to the

requirements of the European Pressure Equipment Directive (PED).

For a particular load case, the calculated stress in a vessel component as a proportion

of allowable stress will depend on the code applied. Allowable material stresses for 'non-

credible' events (see below) may in some codes be higher than prescribed by PD 5500 to

which the present document refers in the main formulae for establishing the dimensions of

vessel components for 'normal' operating conditions.


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A non-credible event is a load combination with a chance of occurrence of less than

10-5. An example of such a non-credible event is illustrated in Figure A.11. The vessel is at

maximum fill and at maximum temperature. Liquid bleeds to open air via a broken outlet pipeor valve causing the internal pressure to drop to zero. The vessel bends due to the internal

liquid load and external loads, including normal forces and moments from the mound, causing

high tensile and compressive stresses at midspan.

Fig: 3.1 Tensile / Compressive Stresses in Vessel due to differential settlements.

The high compressive stresses might result in buckling of the tank shell but the

contents will be maintained within the vessel.


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In view of the very low chance of occurrence (< 10-5 in the lifetime of the vessel) as

well as the fact that the vessel will deform but will still contain the product, it is

recommended, based on international research, to verify the acceptability of these stressesagainst higher allowable stress instead of against the allowable material stress.

Designing the vessel according to the above philosophy will result in a thinner shell

thickness for the non-credible event. PD 5500 only gives the interaction formula below for

combined external pressure and longitudinal compression:

(z / z allow) + (pex / pex allow) 1

Where:

z = design longitudinal compressive stress

z allow = allowable pressure stress

pex = design external pressure

pex allow = allowable external pressure

However, the formula is valid only for longitudinal general membrane stress. The

bending phenomenon described here is not incorporated. The formula does not therefore

provide the required gain in the higher allowable material stress, where 35% increase is

allowed for the non-credible event.

3.3.1 Shell Plates:

Special criteria for allowable longitudinal compressive stress may be defined in the

design code, in which the risk of buckling is covered. Buckling of the shell plates in the

circumferential direction due to external pressure. For mounded storage vessels, the buckling

strength in one direction does not have to be reduced because of compressive stresses in a

direction perpendicular to it. This is different from ultimate strength considerations, where


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longitudinal stresses have an influence on the allowable circumferential stresses and vice

versa.

3.3.2 Domed Ends

It should be realized that the external pressure due to Pressure due to mound, Axial

loads due to changes in vessel length, Seismic loads, External pressure caused by explosion of

gas clouds and supporting pressure by the foundation is not equally distributed, which has an

unfavourable effect on the buckling behavior. Therefore a sufficient safety margin should be

maintained. For a hemispherical end, which is the most economical ( and therefore preferred)

solution for vessels with diameters over 3.5 to 4 m, the wall thickness should not be less than

12mm. For these cases, a stiffening ring should be provided at a distance from the domed

head-to-shell weld of approximately half the diameter of the vessel. If such a ring is not

provided, the analysis of the domed end will be more complicated.

3.3.3 Stiffening Rings

The calculated stresses due to the bending moments, normal forces and shear forces

need to satisfy conventional requirements for steel structures. These include the stresses in the

welds with which the stiffening rings are attached to the shell plates.

If very heavy stiffening rings would have to be provided in order to maintain

acceptable stresses, consideration may be given to installing columns between the top and

bottom of the rings. This will reduce the bending moments in the rings considerably. The

columns should be able to transmit compressive loads only, not tensile loads. This will avoid

secondary bending moments due to internal pressure. The analysis of the stiffening rings,

however, becomes much more complicated. Application of a finite element method will be

required. The deformations of stiffening rings are calculated by the equation:


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(m)

Where

y = vertical deflection of top of ring (m)

R = diameter of ring (m)

E = modulus of elasticity (kN/m2)

I = moment of inertia (second moment of area) of ring (m4)

M = bending moment for circumferential angle (kNm)

( = 0 at top and = at bottom.

design of pressure vessel 1

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