Design of Pressure Vessel

DESIGN OF PRESSURE VESSEL

PROJECT REPORT

Submitted by\

MIJO JOSEPH

VIPIN .M VISHNU VIJAY ABSTRACT

This project work deals with a detailed study and design procedure of pressure vessel. A detailed study of various parts of pressure vessels like shell, closure, support, flanges, nozzles etc. Design is carried according to rules of ASME code section VIII, Division I.

The first chapter deals with detailed study of pressure vessel i.e. the various materials used in pressure construction and temperature are mentioned .It also deals with the study of various parts like flanges, support etc. Various methods of fabrication and testing are also included.

The second chapter includes design criteria .This is followed by procedure of design, which include design shell and its components, nozzles, reinforcements etc. LIST OF FIGURES

FIGURE

PAGE

3.9.1

TYPES OF FLANGES

14

3.11.1

TYPES OF SKIRT

17

MODEL OF AMINE ABSORBER 5.1

TOP AND BOTTOM HEAD

29 32

5.2.1

TOP SECTION OF SHELL

35

5.2.2

BOTTOM SECTION OF SHELL

36

5.3

DESIGN OF NOZZLE

59

5.4

DESIGN OF REINFORCEMENT AREAS

72

LIST OF TABLES DESIGN DATA 7

RESULT AND DISCUSSION

28 87

P :

design pressure, kg/cmA2

NOMENCLATURE

T :

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

S :

maximum allowable stress, kg/cmA2

E :

Joint efficiency, %

T :

required the thickness, mm

4

tn :

trn :

fr :

Mt :

minimum thickness provided for the nozzle, mm

selected thickness for the nozzle, mm

strength reduction factor

moment at the skirt to head joint, kg-mm W :

weight of the vessel H : height of center of gravity

Fe :

seismic coefficient

N:

Number of bolts

Ab :

area with in the bolt circles, mmA2

Cb :

circumference of bolt circle, mm

Ba :

required area of one bolt, mm

As :

area within the skirt, mmA2

Cs :

circumference on outer diameter of skirt, mm 5

design pressure, kg/cmA2

P : Dso

outer diameter of the skirt, mm

Dsi:

inside diameter of the skirt, mm

safe bearing load on the concrete, kg/cmA2

Fb

:

I :

width of the base plate, mm

The equations may be written in the following forms t = PRi/(SE-0.6P) = Pro / (SE-0.4P)

Where, t t = minimum required thickness of the shell exclusively of Corrosion allowance

6

P=

design pressure, or maximum an allowable working Pressure welded -joint efficiency

S

maximum stress

Ri

inside radius of the shell

Ro=

allowable

outside radius of the shell

If the thickness of the Shell exceeds 50% of the inside radius, or when the pressure exceeds 0.385SE, the lame equation should be used to calculate the vesselshell thickness. The following forms of the lame equation are given by the code.

With the pressure p known.

t = Ri(Vz-l) = Ro (Vz-1)/Vz

Where, z = S.E+P/(S.E-P) The equation for ellipsoidal head thickness is given by t = PDi/ (2SE+0.2P) =PDo/ (2SE+1.8P) Where, t = minimum required thickness of the ellipsoidal head exclusive of corrosion allowance P = design pressure, or maximum allowable working pressure. E = welded - joint efficiency. S = Maximum allowable stress. 7

Di = inside diameter of the shell Do=outside diameter of the shell CHAPTER

.1

INTRODU

CTION

Chemical engineering involves the application of sciences to the process industries, which are primarily concerned, with the conversion of one material into another by ahemical or physical means. These processes require the handling or storing of large quantities of materials in containers of varied constructions, depending upon the existing state of the material, it's physical and chemical properties and the required operations, which are to be performed. For handling such liquids and gases, a container or vessel is used. It is called a pressure vessel, when they are containers for fluids subjected to pressure. They are leak proof containers. They may be of any shape ranging from types of processing equipment. Most process equipment units may be considered as vessels with various modifications necessary to enable the units to perform certain required functions, e.g. an autoclave may be considered as highpressure vessel equipped with agitation and heating sources.

8

Pressure vessels are in accordance with ASME code. The code gives for thickness and stress of basic components, it is up to the designer to select appropriate analytical as procedure for determining stress due to other loadings. The designer must familiarize himself with the various types of stresses and loadings in order to accurately apply the results of analysis. Designer must also consider some adequate stress or failure theory in order to confine stress and set allowable stress limits.

The methods of design are primarily based on elastic analysis. There are also other criteria such as stresses in plastic region, fatigue, creep, etc. which need consideration in certain cases. Elastic analysis is developed on the assumption that the material is isotropic and homogeneous and that it is loaded in the elastic region. This analysis is not applicable in the plastic range. Under cyclic variation of load causing plastic flow, the material to hardens and the behavior of material becomes purely elastic. This is a phenomenon called shakedown or cessation of plastic deformation under cyclic loading. Elastic analysis is therefore in most important method of designing pressure vessel shells and components beyond the elastic limit, the material yields and the plastic region (spreads with increased value of load. The load for which this occurs is called collapse load rusting pressure.

Limit analysis is concerned with calculating the load or pressure at which flow of jfitructure material occurs due to yielding. However, this method is not usually applied to Resign of pressure vessels. When vessels are subjected to cyclic loading, it is necessary to consider requirements for elastic cycling of the material and the effects of this on component behavior. In the case of a discontinuity of shape, load may give rise 9

to plastic cycling. Under these conditions, shakedown with occur. Maximum shakedown load is twice the first yield load. Therefore, an elastic analysis is valid up to the range of load, under cyclic loading conditions. A factor of safety on the stress or a factor of safety of twenty is applied on the numbers cycles. Design stress is accepted as the lower value. CHAPTER.2 SCOPE

OF THE 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 allinclusive 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 behaviour.

I The ever-increasing use of vessel has given special emphasis to analytical and experimental methods for determining their emphasis to analytical and experimental methods for determining their operating stresses. Of equal importance is the appraising the significance of these stresses. This appraisal entails the means of determining the values and extent of the stresses and strains, establishing the 10

behaviour of the material ■involved, and evaluating the compatibility of these two factors in the media or environment to which they are subjected. Knowledge of material behaviour is required not only to avoid failures, but also equally to permit maximum economy of material choice and amount used. CHAPTER.3 ■

DESIGN CRITERIA

3.1 FACTORS INFLUENCING THE DESIGN

[Regardless of the nature of application of the vessels, a number of factors usually must be considered in designing the unit. The most important consideration often is the selection of the type of vessel that performs the required services in the most satisfactory manner. In developing the design, a number of other criteria must be considered such as the properties of material used, the induced stresses, the elastic stability, and the aesthetic appearance of the unit. The cost of fabricated vessel is also important in relation to its service and useful life.

3.2 DESIGN OF PRESSURE VESSELS TO CODE SPECIFICATION

American, Indian, British, Japanese, German and many other codes are available for design of pressure vessels. However the internationally accepted for design of pressure vessel code is American Society of Mechanical Engineering (ASME).

11

Various codes governing the procedures for the design, fabrication, inspection, testing and operation of pressure vessels have been developed; partly as safety measure. These procedures furnish standards by which, any state can be assured of the safety of pressure vessels installed within its boundaries. The code used for unfired pressure vessels is Section VIII of the ASME boiler and pressure vessel code. It is usually necessary that the pressure vessel equipment be designed to a specific code in order to obtain insurance on the plant in which the vessel is to be used. Regardless of the method of design, pressure vessels with in the limits of the ASME code specification are usually checked against these specifications. 3.3 DEVELOPMENT AND SCOPE OF ASME CODE

In 1911, American Society of Mechanical Engineering established a committee to formulate standard specifications for the construction of steam boilers and other pressure vessels. This committee reviewed the existing Massachusetts and Ohio rules and eonducted an extensive survey among superintendents of inspection departments, Engineers, fabricators, and boiler operators. A number of preliminary reports were issued and revised. A final draft was prepared in 1914 and was approved as a code and copy righted in 1915. The introduction to the code stated that public hearings on the code should be held every two years. In 1918, a revised edition of the ASME code was issued. In 1924, the code was revised with the addition of a new section VIII, which represented a new code for unfired pressure vessels.

3.4 THE API-ASME CODE 12

In 1931, a joint API-ASME committee on unfired pressure vessels was appointed to prepare a code for safe practice in the design, construction, inspection and repair of unfired pressure vessels.

3.5 SELECTION OF THE TYPE OF VESSEL

The first step in the design of any vessel is the selection of the type best suited for the particular service in question. The primary factors influencing this choice are,

. i.

ii. iii. ■ iv.

The operating temperature and pressure.

Function and location of the vessel. Nature of fluid. Necessary volume for storage or capacity for processing.

It is possible to indicate some generalities in the existing uses of the common types of vessels. For storage of fluids at atmospheric pressure, cylindrical tanks with flat bottoms and conical roofs commonly used. Spheres or spheroids are employed for pressure storage where the volume required is large. For smaller volume under pressure, cylindrical tanks with formed heads are more economical.

3.6 TYPES OF PRESSURE VESSELS

3.6.1 OPEN VESSELS 13

Open vessels are commonly used as surge tanks between operations, as vats for batch operations where materials be mixed and blended as setting tanks, decarters, chemical reactors, reservoirs and so on. Obviously, this type of vessels is cheaper than covered or closed vessel of the same capacity and construction. The decision as to whether or not open vessels may be used depends up on the fluid to be handled and the operation.

3.6.2 CLOSED VESSELS

1

Combustible fluids, fluids emitting toxic or obnoxious fumes and gases must be

stored in closed vessels. Dangerous chemicals, such as acid or caustic, are less hazardous if stored in closed vessels. The combustible nature of petroleum and its products associates the use of closed vessels and tanks throughout the petroleum and petrochemical industries. Tanks used for the storage of crude oils and petroleum products and generally designed and constructed as per API specification for welded oil storage tanks. [3.6.3 CYLINDRICAL VESSELS WITH FLAT BOTTOMS AND CONICAL OR DOMED ROOFS.

The most economical design for a closed vessel operating at atmospheric pressure is the vertical cylindrical tank with a conical roof and a flat bottom resting directly on the bearing soil of a foundation composed of sand, gravel or crushed rock. In cases where it is desirable to use a gravity feed, the tank is raised above the ground, and columns and wooden joints or steel beams support the flat bottoms. 14

3.6.4 CYLINDRICAL VESSELS WITH FORMED ENDS

Closed cylindrical vessels with formed heads on both ends used where the vapour pressure of the stored liquid may dictate a stronger design, codes are developed through the efforts of the American petroleum Institute and the American Society of Mechanical Engineering to govern the design of such vessels. These vessels are usually less than 12 feet in diameter. If a large quantity of liquid is to be stored, a battery of vessels may be used.

3.6.5 SPHERICAL AND MODIFIED SPEHRICAL VESSELS

Storage containers for large volume under moderate pressure are usually fabricated in the shape of a sphere or spheroid. Capacities and pressures used in these types of yessels vary greatly for a given mass; the spherical type of tank is more economical for large volume, low-pressure storage operation.

3.6.6 VERTICAL AND HORIZONTAL VESSELS

In general, functional requirements determine whether the vessel shall be vertical or jjiorizontal. Eg. Distilling columns, a packed tower, which utilizes gravity, require vertical installation.

15

Heat exchanges and storage vessels are either horizontal or vertical. If the vessel to be installed outdoor, wind loads etc, are to be calculated to prevent overturning, thus jhorizontal is more economical. However, floor space, ground area and maintenance requirements should be considered.

3.6.7 VESSELS OPERATING AT LOW TEMPERATURE RANGES

Pressure vessels constructed in such a manner that, a sudden change of section producing a notch effect is present, are usually not recommended for low temperature range operations. The reason is that, they may create a state of stress such that the material will be incapable of relaxing high-localized stresses by plastic deformation, therefore, the materials used for low temperature operations are tested for notch ductility.

Carbon steels can be used down to 60 degree C. Notch ductility is controlled in such as materials through proper composition steel making practice, fabrication practice and heat treatment. They have an increased manganese carbon ratio. Aluminium is usually added to promote fine grain size and improve notch ductility.

Ductility of certain materials including carbon and low alloy steels is considerably diminished when the operating temperature is reduced below certain critical value is usually described as the transition temperature, depends upon the material, method of manufacture, previous treatment and stress system present. Below transition temperature, fracture may take place in a brittle manner with little or no deformation. Whereas, at temperatures above the transition temperature, fracture occurs only after considerable plastic strain or deformation. 3.6.8 VESSELS OPERATING AT ELEVATED TEMPERATURE.

Embrittlement of carbon and alloy steel may occur due to service at elevated temperature. In most instances, brittleness is manifest only when the material is cooled to

jK>om

temperature. This inhibited by addition of molybdenum and also improve tensile and creep properties. Two main criteria in selecting the steel elevated temperature are metallurgical strength and stability. Carbon steels are reduced in their strength properties due to rise in temperature and are liable to creep. Therefore, the use of carbon steel is generally limited to 500dege C.

The SA-283 steels cannot be used in applications with temperatures over 340degreC. The SA-285 steels cannot be used for services with temperature over 482degreC. However, both SA-285 and SA-285 SA-212 steels have very low allowable stress, at higher temperature.

3.7 MATERIAL SPECIFICATION

Plain carbon and low alloy steels plates are usually and where service condition permit because of the lesser cost and greater availability of these steels. Such steels may me fabricated by fusion welding and oxygen cutting if the carbon content does not exceed 0.35%.vessels may be fabricated.

Vessel may be fabricated of plate steels meeting the specification of SA-7, SA-113, Grade A, B, C&D, provided that,

1. operating temperature is between -28degreeC&360degreeC 2. The plate thickness does not exceed 1.5cm 3. The vessels does not contain lethal liquids and gases 4. The steel is manufactured by the electric furnace or open hearth process 5. The material is not used for unfired steam boilers One of the most widely used steel for general purpose in the construction of ■ressure vessel is SA-283, Grade C. This steel has good ductility and forms welds and machines easily. It is also one of the most economical steel suitable for pressure vessels. [However, its use is limited to vessels with plate thickness not exceeding 1.5cm.

For vessels having shells of grater thickness. SA-285 Grade C is most widely used Hi moderate pressure applications. In case of high pressure or large diameter vessels, high strength steel may be used to advantage to reduce the wall thickness. SA-212, GradeB is well suit for such application and requires a shell thickness of only 79% of that required by SA-285, Grade C. This steel also is fabricated but is more expensive than other steels.

Now, many new series of materials like low alloy, high alloy steels, high temperature and low temperature materials are available which can be selected to suit the requirement of every individual need of process industry. The important materials generally accepted for construction of pressure vessels are indicated here. Metals used are generally divided into three groups as.

1. Low cost

Cast iron, Cast carbon and low alloy steel, wrought carbon and

low alloy steel.

2. Medium cost - High alloy steel (12%chromium and above), Aluminum, Nickel, Copper and their alloys, Lead.

3. High cost

-

platinum, Tantalum, Zirconium, Titanium silver.

Materials mentioned (2 & 3) groups are some times used in the form of cladding or bonding for materials in group (1). Also, use non-metallic lining such as rubber, plastics, etc. Vessels with formed heads are commonly fabricated from low carbon steel wherever corrosion and temperature considerations will permit its use because of the low cost, high strength, ease of fabrication and general availability of mild steel. Low and high alloy steel and non-ferrous metals are used for special service. Steels commonly used fall into two general classifications.

1. Steels specified by ASME code.

2. Structural grade steels, some of which permitted by ASME.

3.8 CLOSURES FOR PRESSURE VESSELS

All formed heads are fabricated form single circular flat plate by spinning by drawing with dies in a press. Although the cost of heads formed from flat plates involves additional cost of forming, the use of formed heads as closures usually more economical than the use of flat plates as closures except for small diameters.

A variety of formed heads is used for closing the ends of cylindrical vessels. These include flanged only heads, flanged and shallow dished, torispherical, elliptical, hemispherical and conical shaped heads. For special purposes, flat plates are used to close a vessel opening. However flat heads are rarely used for large vessels.

For pressures not covered by the ASME code, the vessels are often equipped with standard dished heads, whereas vessels that require code construction are usually equipped with either the ASME - dished or elliptical dished heads. The most common shape for the closure of pressure vessels is the elliptical dish. Most chemical and petrochemical processing equipment such as distilling columns, desorbers, absorbers, scrubbers, heat exchangers, pressure surge tanks and separators are essentially cylindrical closed vessels with formed ends of one type or another.

As mentioned above, the most common types of closures for vessels under internal pressure are the elliptical dished head (ellipsoidal head) with a major to minor axis ratio equal to 2.0 : 1.0 and the torispherical head in which the knuckle radius is equal to 6% or more of the inside crown radius (ASME standard dished head). 3.9 FLANGES AND FLANGED FITTINGS A variety of attachments and accessories are essential to vessels. These include flanges for closures, nozzles, manholes and hand holes and flanges for 2- piece vessels, supports platforms, etc,.

Flanges may be used on the shell of a vessel to permit disassembly and removal, for cleaning of internal parts. Flanges are also used for making connections for piping and for nozzle attachments of opening.

A great variety of type and sizes of 'standard' flanges are available for various pressure services. The flanges designated as "American Standards Association (ASME) B 16.5 - 1953" are used for most steel pipelines over 3.8 cm nominal pipe sizes. These flanges are called 'companion flanges', because they are usually used in pairs. Forged steel flanges are manufactured in the following standards types for all pressure ratings.

3.9.1 TYPES OF FLANGES

3.9.1.1 WELDING- NECK FLANGES A sectional view of a welding - neck flange is shown. Welding neck flanges differ from other flanges in that, they have a long, tapered hub, between the flange ring and the welded joint. This hub provides a more gradual transition from the flange ring thickness fo the pipe -wall thickness, thereby decreasing the discontinuity stresses and consequently increasing the strength of the flange. These flanges are recommended for the handling of costly, flammable or explosive fluids, where failure or leakage of the flange joint might disastrous consequences. 3.9.1.2 SLIP-ON FLANGES

The slip-on types of flanges are widely used because of its greater ease of aligned in welding assembly and because of its low initial cost. The strength of this flange as calculated from internal pressure considerations is approximately 2/3rd that of a corresponding weldingneck type of flange. The use of this type of flange should be ' limited to moderate services, where pressure fluctuations, temperature fluctuations, vibrations and shock are not expected to be severing. The fatigue life of this flange is approximately l/3rd that of welding - neck flange.

3.9.1.3 LAP JOINT FLANGES

Lap joint flanges are usually used with a lap-joint stab. These flanges have about the same ability to withstand pressure without leakages as the slip in flange, which is less fhan that of the welding neck flanges. In addition, these flanges have the disadvantages of having only about 10% of the fatigue life of welding neck flanges. For these reasons, these flanges should not be used for connections where, severe bending stresses exist.

The principal advantage of these flanges is that the bold holes are easily aligned and this simplifies the erection of vessels of large diameter and usually stiff piping. Theses flanges are also useful in cases where, frequent dismantling for cleaning or inspection is required, or where it is necessary to rotate the pipe by swiveling the flange..

3.9.1.4 SCREWED FLANGES

Screwed flanges can be fastened to the openings by screwing. It can be connected instantly without welding. The only disadvantage is that possibility of leakage. 3.9.1.5 BLIND FLANGES

They are used extensively to blank off pressure vessel openings and hand holes, block off pipes and valves. In this application, a valve followed by blind flange is frequently used at the end of line to permit addition of line while it is 'on stream'.

LAPPED FLANGE

BLIND FLANGE

SLIP ON FLANGES

D-----------~—:--------------~J

Figure: types of flanges

3.10 NOZZLES, OPENINGS AND REINFORCEMENTS Nozzles and openings are necessary components of pressure vessels for the process industries. Openings in a cylindrical shell, conical section or closure may produce stress concentrations, adjacent to the opening and weaken that portion of the vessel. In order to minimize such stress concentrations, it is preferable that the opening be circular in shape. As a second choice the openings may be made elliptical, as a third choice they may be made around. An around opening has two parallel sides and two semicircular ends. Openings of other shapes are permissible if the vessel is tested hydrostatically. If the opening in a closure of cylindrical vessel exceed one-half the inside diameter of shell, the opening and closure should be fabricated. Others require reinforcement. Small sizes of openings welded or brazed to a vessel do not require reinforcement. ■ In the case of shell, opening requiring reinforcement in vessel under internal pressure the metal removed must be replaced by the metal of reinforcement. In addition to providing the area of reinforcement, adequate welds must be provided to attach the metal of reinforcement and the induced stresses must be evaluated.

Materials used for reinforcement shall have an allowable stress value equal to or greater than of the material in this vessel wall except that, when such material is not available, lower strength material may be used; provided, the reinforcement is increased in inversed proportion to the ratio of the allowable stress values of the two materials to the ratio of the two materials

to compensate for the lower allowable stress value of any reinforcement having a higher allowable stress value than that of the vessel wall. 3.11 SUPPORTS FOR VESSELS Cylindrical and other types of vessels have to be supported by different methods. Vertical vessels are supported by brackets, column, skirt, or stool supports, while saddles support horizontal vessels. The choice of type of support depends on the height and diameter of the vessel, available floor space, convenience of location, operating variables, the size of jjhe vessel, the operating temperature and pressure and the materials of construction.

Brackets of lugs offer many advantages over other types of supports. They are inexpensive, can absorb diametrical expansions by sliding over greased or bronze plates, jfcre easily attached to the vessel by minimum amounts of welding, and easily leveled or shimmed in the field. Lug supports are ideal for thick-walled vessels, but in thin-walled vessels, this type of support is not convenient unless the proper reinforcements are used or many lugs are welded to the vessel.

It is also necessary to ensure that, the attachment of the support to the vessel, which is usually by fillet welds should be able to transfer the load safely from vessel to support and that, the support should be strong enough to withstand the load of the vessel.

3.11.1 SKIRT

Vertical vessels are normally supported by means of suitable structure resting on a reinforced concrete foundation. This support structure between the vessel and the j&undation may consist of a cylindrical shell termed as skirt. The skirt is usually welded to the vessel because the skirts are not required to withstand the pressure in the vessel; the selection of material is not limited to codes. The skirt may be welded directly to the bottom dished head, flush with the shell or to the outside of shell. There will be no stress from internal and external pressure for the skirt, unlike for the shell, but the stresses from dead weight and from wind or seismic bending moments will be maximum.

28

3.12ANCHOR BOLT The bottom of skirt of vessel must be securely anchored to the concrete !foundations by means of anchor bolts embedded in the concrete to prevent over turning from bending moments induced by seismic and wind loads.

The concrete foundation is poured with adequate reinforcing steel to carry tensile loads. The anchor bolts may be formed from steel rounds threaded at one end and usually with a curved or hooked end embedded in the concrete will bond to the embedded surface of the steel.

3.13METHODS OF FABRICATION

Process equipment is fabricated by a number of well-established methods such as fcsion welding, casting, forging, machining, brazing and soldering and sheet metal forming. Each method has certain advantages for particular types of equipment. However, fusion welding is the most important method. The size, shape, service and material properties of the equipment all may influence the selection of the fabrication method.

Gray iron casting have been widely used for the mass production of small pipe fittings and are used to a considerable extent for large items such as cast iron pipe, heat exchanger shells and evaporator bodies because of the superior corrosion resistance of cast iron as compared with steel. Large diameter vessels cannot be easily cast, and the strength of gray iron is not reliable for pressure vessels service. Cast steel may be used £>r small diameter thick walled vessels. Further more, because of its higher strength and greater reliability as compared with cast iron; it is more suitable for high-pressure service where metal porosity is

not a problem. The vessel diameter is still limiting because of a problem in casting. Alloy cast steel vessels can be used for high-temperature and high-pressure installation. : Forging is a method of shaping metal that is commonly used for certain vessel jparts such as closures, flanges and fittings. Vessels with wall thickness greater than 10cm ire often forged. Other special methods of shaping metal such as pressing, spinning and rolling of plates are used for forming closures for vessel shells.

Riveting was widely used prior to the improvement of modern welding !fcchniques, for many different kinds of vessels, such as storage tanks, boilers and a verity |tf pressure vessels. It is still used for fabrication of non-ferrous vessels such as copper and aluminium. However, welding techniques have become so advanced, that even these materials are often welded today.

I. Machining is the only method other than cold forming that can be used to exact tecure tolerances. Close tolerances are required for the mating parts of the equipment. Flange faces, bushings, and bearing surfaces are usually machined in order to provide satisfactory alignment. Laboratory and pilot plant equipment for very high-pressure service is sometimes machined for solid stock, pierced ingots and forgings.

3.13.1 FUSION WELDING

It is the most widely used method of fabrication for the construction of steel vessels. This method of construction is virtually unlimited with regard to size and is extensively used for the fabrication and erection of large size product equipment in the field. There are two types of fusion welding that are extensively used for fabrication of welds. These are,

30

1. The gas welding process in which a combustible, mixture of acetylene and oxygen supply the necessary heat for fusion

2. The electric arc welding process, in which the heat of fusion is supplied by an electric arc. Arc welding is preferred because of the reduction of heat in the weld material, reduces the oxidation and better control of deposited weld metal. 3.14 PRESSURE TESTING OF CODE VESSELS

All pressure vessels designed to code specification except those exempted because of small size must be tested hydrostatically, pneumatically or by means of ."PROOF TEST".

In the case of hydrostatic test, the vessel must be subjected to a hydrostatic test pressure at least equal to one and a half times the maximum allowable pressure at the test temperature. Following the application of the test pressure, all joints and connections snust be inspected with the vessel under a pressure not less than 2/3'd of the test pressure. [Although water is used in this test, any non-hazardous liquid may be used below its boiling temperature.

If the vessels are designed so that they camiot safely be filled with water (as in the case of tall vertical towers design to handle vapours pneumatically), testing may be used. The pneumatic test pressure should be at least 1.25 times the maximum allowable pressure at the test temperature. In conducting pneumatic test, the pressure in the vessel should be gradually increased to not more than half the test pressure. There after the test pressure should be raised in increments of 1/10th of test pressure until the test pressure is reached. Following these, the pressure should be reduced to maximum allowable pressure and held there for a sufficient length of time to permit inspection of vessel.

The "proof test" can be used to establish the allowable working pressure in Vessels that have parts, which the stress cannot be computed with satisfactory accuracy. In one procedure

of this test, all areas of probable high stress concentration are painted with a wash of lines or another brittle coating. The pressure is raised and the vessel is inspected for signs of yielding indicated by flaking or strain lines in the wash. The vessel is first observed.

Strain gauge measurements may be used in non destructive testing .In this case the pressure is increased in increments of 1/10th the test pressure, each increment I followed by relaxation of the pressure, until a permanent strain of 0.2% is reached. The

Vessel rating at the test temperature is equal to one-half the pressure producing this jpennanent strain. A modification of the strain gauge measurement procedure is also kennitted by the code. This method involves the use of measuring gauges at diametrically opposed reference points in symmetrical structure.

In another version of the proof test, a sample used is tested to destruction and I identical vessels are rated at the test temperature at l/5th the pressure at which the tested vessels is failed. CHAPTER.4 DESIGN PROCEDURE 4.1 DESIGN

OF SHELL AND ITS COMPONENTS iijhe pressure vessel considered here is a single unit when fabricated. However,

for the

Jonvenience of design, it is divided into the following parts. J (1) Shell; (2)head or cover; (3) nozzles; (4) support; 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.

32

In designing these parts and connections between them, it is essential to take r into account, the efficiency of joints. For welded joints, the efficiency may be taken as 100% if the joint is fully checked by a radiograph and taken as 85%, eve if it is checked at only a few points. If the radiographic test is not carried out 50 to 80%, I efficiency is taken. Efficiencies vary between 70 to 85% in the case of riveted joints. All these are made for pressure vessels operating at pressures less than 200kg/kmA2. Design procedure is primarily based on fabrication by welding.

4.1.1 DESIGN OF CYLINDRICAL SHELLS UNDER PRESSURE The equation for determining the thickness of cylindrical shells of vessels under internal pressure are based upon a modified membrane-theory equation. The modification empirically shifts the thin wall equation to approximate the "Lame" equation for thickwalled vessel's shown above. '4.2 WELDING STANDARDS r

The success of fabrication by welding is dependent upon the control of the welding variables such as experience and training of the welder, the use of proper materials, and welding procedures. An inexperienced welder or welder using inferior materials, incorrect procedures can fabricate a vessel that has a good appearance but has unsound joints, which may fail in service. Thus, it is essential that the welding variables be controlled in order to produce sound joints in the equipment. A number of codes and standards have been published for the puipose. The American welding society (AWS) established the basic standards for quantifying operators and procedures. These standards of qualification form the basis of most of the

standards in various codes. For practical purposes, therefore the rules for qualifying welders and welding procedures are essentially the same in the various codes and standards. Each fabrication shop should establish welding procedures best suited to its need and its equipment. To meet the welding standards previously mentioned, it is not necessary that, regardless of the procedures used, the welded joints must pass the qualification tests for welding procedures. To meet welding standards, welded joints must be tested to determine tensile strength, ductility, and soundness. The required tests for the welding procedures specified by API standard 12C involved the following. A. For Groove Welds:

1. reduced section -section-tension test ( for tensile strength ) 2. Free bend test ( for ductility)

I

3. root bend test( For soundness )

4. face bend test( for soundness) 5. Side bend test ( for soundness)

i

B. For fillet welds: 1. transverse -shear test( for shear strength)

I

2. free bend test( for ductility) 3. Fillet weld - soundness test

The minimum results required by the tests such as those listed above are described in detail in the various codes. A few representative requirements are: a. The tensile strength in the reduced section tension test shall not be less than 95% of the minimum tensile strength of the material being welded 34

b. The minimum permissible elongation in the free bend test is 20% c. The shearing strength of the welds in the transverse shear test shall not be less than 87% if the minimum tensile strength of the material being welded. d. In the various soundness tests, the convex surface of the specimen is examined for the appearance of cracks or other defects. If any cracks exceeds 0.3cm, in any direction, the joint is considered to have failed.

H2.1 TYPES OF WELDED JOINTS A variety of welded joints are used in the fabrication of vessels. The selection of the type of joint depends upon the service, the thickness of the metal fabrication procedure and code requirements. The following figure is a diagram from the API-ASME code for unified pressure vessels which illustrates some of the types of welded joints used in the welding of steel plates for the fabrication of pressure vessels.

4.2.2 WELDED JOINTS EFFICIENCIES The use of welded joints may result in reduction in the strength of the part at or near the world. This may be result of metallurgical discontinuities and residual stress. The code rules make allowance for these factors by specifying joint efficiencies for various types of welds with and with out stress relief and radiographing. The designs are permitted some option in the selection of the kind of weld joint to be and in whether or not, the welded joints must be radio graphed. All vessel shells having a thickness greater than 1 1/4 inch or greater than ld+50)120 {where, d=inside diameter or 20 inches which ever is greater: must be thennally stress received.

f Vessels of any thickness can be fabricated from the following steels must be stress relieved, SA-301.Grade B; SA-302; SA-270, Grades WC-5; SA-357; SA-387;Grades B,C,D and E and chrome-molybdenum steel having a chrome content greater than 0.7%. In addition, vessels having a shell thickness greater than 1.4 cm must be thermally stress relieved if they are fabricated of the following steels: SA-202; SA-203; SA-204; SA-225; SA-299; SA-301, Grade A, and any steel having specified molybdenum content of 0.4 to 0, 65% and a chrome content not greater thanvO.7%. In addition, steel greater than 2.5 cm in thickness must be stress relieved if they meet the Verification of the following: SA-212; SA-105, Grade II: SA181, Grade II; SA-266, Grade II SA-94; & SA-216, Grade WCB.

1

36

If high alloy steels are used, stress relieving is not required in the case of austenitic chromium nickel steels. The increase in joint efficiency may be used if these steels are heat treated at over 480 x C. if the vessel are constructed of ferrites chromium stainless steels, stress relieving is required in al vessel thickness except in the case of type 415 welded with electrodes, a process producing austenitic weld. The code gives the temperature and describes the procedures to be used in thermal stress relieving.

Radiography examination is required for double welded butt joints. If the plastic thickness is greater than 2.5 cm complete radio graphing of each welded joints is required, if the vessel is fabricated of SA-202; SA-203; SA-204; SA-225; SA-299; SA-301 or SA-302. Vessels of thickness that are fabricated if SA-353, SA-357 or SA-387. must be radio graphed. Also vessels constructed of high alloy steels such as type 405 jwelded with straight chromium electrodes and type -410 &430 welded with any electrodes must be radio graphed in all thickness except when carbon content does not exceed 0,08%, the plate thickness does not exceeded 3.8 cm and austenitic welds are used. k.23 JOINT EFFICIENCIES AND CORROSION ALLOWANCES In vessels for atmospheric storage, the welded joints are seldom stress relieved or radio graphed. The welded seams may not be as strong as the adjacent rolled steel plate ■ the shell. It has been found from experience that, an allowance may be made for such Weakness by introducing a "joint efficiency factor E" in the equations. This factor is llways less than unity and is specified for a given type of welded construction in the prarious codes.

The thickness of the metal, C allowed for any anticipated corrosion is then added to the calculated required thickness, and the final thickness value rounded off to the nearest nominal plate size of equal or greater thickness.

4.3 DESIGN TEMPERATURE The temperature used in design shall not be less than the mean metal temperature except operating conditions for parts considered. If necessary, the metal temperature shall be determined by computation using accepted heat transfer procedures or by measurement from 37

equipment in service, under equivalent operating conditions. In no case, shall the temperature of surface of metal exceed the maximum temperature listed in the stress tables for materials not exceed the maximum temperature limitation specified elsewhere in ASME section VIII. div 1.

4.4 NOZZLE NECK THICKNESS

Except for opening for inspection only, the wall thickness of a nozzle neck or other connection shall not be less than the greater of the following.

1. The thickness computed for the applicable loadings in UG -22 plus the thickness added for corrosion allowance in the connections.

2. Smaller of the following. For vessels under internal pressure only, the thickness required for pressure for the shell at the location. Where the nozzle neck attached to the vessel, but on no case less than the minimum thickness specified for the material in UG - 16.

4.5 DESIGN OF NOZZLE REINFORCEMENT As per UG -37 of ASME sec. VIII, div 1. Reinforcement is provided in amount and distribution such that the area requirements for reinforcement are satisfied for all planes through the centre of opening and normal to vessel surface. For a circular opening in a cylindrical shell, the plane containing the axis of shell is the plane of greatest loading due to pressure. Not less than half the required reinforcement shall be on each side of the centre line of single openings. Reinforcement is provided for openings having diameter greater than 50mm.

4.6 DESIGN OF SKIRT SUPPORT

1

38

Skirt is the most frequently used and the most satisfactory support for vertical

:

vessels. It is attached by continuous welding to the head and usually the required size of | this welding determines the thickness of the skirt.

In calculation of the required weld size, the value of the joint efficiency is given by the code UW -12 may be used. DESIGN DATA

[Code

ASME SECTION VIII DIVISION 1, 2004 DEA 1.051 30 Kg/cm2 40°C 35 Kg/cm2 70°C 1 3 mm 45.5 Kg/cm2 N/A

Fluid handled Specific gravity of fuel Operating pressure Operating temperature Design pressure Design temperature Joint efficiency Corrosion allowance Hydraulic test pressure Pneumatic test pressure

39

I.JHUHJ 40

U90KD

CHAPTER.5

DESIGN CALCULATIONS

5.1 DESIGN OF HEAD 5.1.1 TOP HEAD Equipment

Top Head

Code

ASME Section VIII Div 1

Material

SA516 Gr60 118 MPa

Max Allowable Stress, S

3.434 MPa 343 K

Design Pressure, P Design Temperature, T | A 2:1 ellipsoidal top head is selected. According to UG 31 of ASME Section VIII Div 1, Minimum thickness required, tr =

P.D 2SE - 0.2 P

where D =intemal diameter E= joint efficiency=1.0

C= corrosion allowance= 3mm tr:

3.434 x 1000 2xH8xl-0.2x 3.434

41

+C,

Nominal thickness,

t = 20mm

Height of ellipsoidal head, h = D/4= 250mm

42

t,= 17.59mm

43

Ll.2 BOTTOM HEAD fcquipment

Bottom Head

Code

ASME Section VIII Div 1

Material

SA516Gr 60

Max Allowable Stress, S

118MPa

besign Pressure, P Design

3.455 MPa

Temperature, T

343 K

A 2:1 ellipsoidal bottom head is selected. According to UG 31 of ASME Section VIII Div 1, Minimum thickness required, tr =

P.D 2SE-0.2P

E= joint efficiency=1.0 C= corrosion allowance= 3mm tr = 3.455x1000 2xH8xl0.2x3.455

t = 17.68mm 44

Nominal thickness,

t = 20mm

Height of ellipsoidal head, h = D/4= 250mm

45

Fig: top head and bottom head 46

20 rrirn THICKNESS

5.2 DESIGN OF SHELL

5.2.1 TOP SECTION (ABOVE LT1) Equipment

Shell

Code

ASME Section VIII Div

Material

SA516Gr 60 118 MPa

Max Allowable Stress, S

3.434 MPa 343 K

Design Pressure, P Design Temperature, T

A 2:1 ellipsoidal top head is selected. According to UG 27 of ASME Section VIII Div 1, Minimum thickness required, tr =

P.R

SE-0.6Pwhere R =internal radius E= joint efficiency= 1.0 C= corrosion allowance= 3mm

tr

3.434x500 118x1-0.6x3.434

47

+3

t,= 17.81mm

Nominal thickness, t

20mm

48

5.2.2 BOTTOM SECTION (BELOW LT1) Equipment Code Material

Shell

Max Allowable Stress, S

ASME Section VIII Div 1

Design Pressure, P

SA516 Gr60 118 MPa

Design Temperature, T

3.452 MPa 343 K

[A 2:1 ellipsoidal top head is selected. : According to UG 27 of ASME Section VIII Div 1, Minimum thickness required, tr =

P.R

+

SE-0.6P where R =internal radius E= joint efficiency=1.0 C= corrosion allowance= 3mm

3.432x500 .118xl-.6x3.432

t, = 17.89mm

49

Nominal thickness,

t = 20mm

50

LTI

.20 mm thick

LT 2 1000 i

53

Figure: top shell section and bottom head section

54

5.3 DESIGN OF NOZZLES

Nozzle Mark

:V

Equipment

: Vent with valve : 50 mm

SizeNB

: ASME Section VIII

Code

Div : SA 106 GrB : 118

Material

MPa : 3.434 MPa : 343 K

Max Allowable Stress, S Design Pressure, P Design Temperature, T

From ANSI 3.36.10, Outside diameter D0 = 60.3 mm Ro = 30.15 mm

A'

P.Ro

SE + 0.4P

+C

3.434x30.15 118x1+0.4x35

3.87 mm

B' = head thickness + C

- 2 0 + 3

= 23 mm

+3

From pipe tables, Standard wall thickness = 3.91 mm C =

(std. wall thickness x .875) + C

= (3.91 x .875)+ 3

= 6.421 mm D' = lesser of (B' and C) =

6.421 mm E' = greater of (A' and D') = 6.421

mm

Required thickness, tr =

E'-C .

+C

875 From pipe tables, Standard wall thickness = 11.91 mm

57

6.421-3 .

+3

875 -

6.91 mm

Selected thickness, t = 7.14 mm

Schedule 160

58

Nozzle Mark

Ml

Equipment Size NB

Manhole

(Code Material

600 mm

Max Allowable Stress, S

ASME Section VIII D

Design Pressure, P

SA 106 G r B 118 MPa

Design Temperature, T

3.452 MPa 343 K

From ANSI 3.36.10, Outside diameter, D0 = 609.6 mm Ro = 304.8 mm

A'

P.Ro

+ C ■= 11.81 mm

L§E + 0.4P

3.452x304.8 118x1+0.4x3.452

= 11.81mm

B' = shell thickness + C

20 + 3

+3

23 mm C = (std. wall thickness x .875) + C = (11.91x.875) +3 = 13.42

mm

D' = lesser of (B' and C) = 13.42 mm

E' = greater of (A' and Dv) = 13.42 mm

Required thickness, t, =

+C

E' - C . 875

13.42-3 .875

14.91 mm

Selected thickness, t = 15.88 mm

+3

Nozzle Mark Equipment

M2

Size NB Code Material

Manhole

Max Allowable Stress, S

600 mm

Design Pressure, P

ASME Section VIII Div 1

Design Temperature, T

SA 106 G r B 118 MPa 3.434 MPa 343 K

From ANSI 3.36.10, Outside diameter, D0 = 609.6 mm Ro = 304.8 mm

A'

P.Ro .

+C

SE + 0.4P-

3.434x304.8 118x1+0.4x3.434

11.76 mm

B' = shell thickness + C

20 + 3

+3

23 mm

C = (std. wall thickness x .875) + C - = (ll.91x.875) +3 I; F 13.42

mm

D' = lesser of (B' and C) = 13.42 mm E' = greater of (A' and D') =

13.42 mm

iRequired thickness, tr =

E'-C .

+C

875

13.42-3 . 875

= 14.91 mm

Selected thickness, t = 15.88 mm

+3

[Nozzle Mark

: LT1

Equipment Size NB

Level Transmitter 50 mm

[Code

ASME Section VIII Div 1

! Material

SA 106 G r B 118 MPa

Max Allowable Stress, S

3.434 MPa 343 K

Design Pressure, P Design Temperature, T From ANSI 3.36.10, Outside diameter, D0 = 60.3mm Ro = 30.15mm

A'

P.Ro

+C

SE + 0.4P3.434 x 30.15

+3

118x1+0.4x3.434

From pipe tables, Standard wall thickness = 3.91 mm

65

- 3.87 mm

B' = shell thickness + C

= 20 + 3

= 23 mm

66

fc' = (std. wall thickness x .875) + C

(3.91x.875) + 3

6.421 mm

W = lesser of (B' and C) = 6.421 mm

p' = greater of (A' and D') = 6.421 mm

From pipe tables, Standard wall thickness = 3.91 mm

67

Required thickness, t,

E' - C .

+C

875

6.421 -3 . 875

+3

.

6.91 mm

Selected thickness, t = 7.14 mm

Schedule 160

68

[Nozzle Mark

Design Temperature, T

Equipment

From ANSI 3.36.10,

Size NB

I

LT2

Code

Level Transmitter 50 mm

Material

ASME Section VIII Div

Max Allowable Stress, S

SA 106 G r B 118 MPa

Design Pressure, P

3.452 MPa 343 K Outside diameter, D0 = 60.3 mm R0 = 30.15 mm

A'

P.Ro

-C

SE + 0.4P

3.452 x 30.15 18x1+0.4x3.452

3.87 mm

B' = shell thickness + C

= 20 + 3

+3

23 mm C = (std. wall thickness x .875) + C . =(3.91x.875) + 3

= 6.421 m m . M = lesser of (B' and C) = 6.421 mm

I' = greater of (A' and D') = 6.421 mm

Required thickness, t, =

E' - C .

+C

875

6.421-3 .

+3

875 .

= 6.91 mm

Selected thickness, t = 7.14 mm

Schedule 160

Nozzle Mark

SP1

Equipment Size NB

Standpipe

Code Material

50 mm

Max Allowable

ASME Section VIII Div 1

Stress, S Design

SA 106 G r B .118 MPa

Pressure, P Design

3.452 MPa 343 K

Temperature, T From ANSI 3.36.10,

Outside diameter, D0 = 60.3 mm R0 = 30.15 mm

A'

P.Ro

+C

SE + 0.4P

3.452 x 30.15

+3

118x1+0.4x3.452

3.87 mm

B' = shell thickness + C

= 20 + 3

= 23 mm

From pipe tables, Standard wall thickness = 3.91 mm = (3.91*.875) + 3 f =6.421 mm D' = lesser of (B' and C) = 6.421 mm I E' =

greater of (A' and D') = 6.421 mm

C = (std. wall thickness x .875) + C 73

[Required thickness, tr =

E'-C .

+C

875

6.421-3 . 875

+3

.

6.91 mm

Selected thickness, t = 7.14 mm Schedule 160 SP2 Nozzle Mark

Standpipe

Equipment Size NB

50 mm

Code . Material

ASME Section VIII Div 1 SA 106 G r B 118 MPa 3.434 MPa 343 K

Max Allowable Stress, S Design Pressure, P Design Temperature, T iFrom ANSI 3.36.10, Outside diameter, D0 = 60.3 mm R0

30.15 mm

=

A'=

P-Rn

+C

LSE + 0.4P.

' 3.434 x 30.15

+3

118x1+0.4x3.452

= 3.87 mm

B' = shell thickness + C

20 + 3

23 mm

From pipe tables, Standard wall thickness = 3.91 mm = (3.91x.875) + 3 = 6.421 mm D' = lesser of (B' and C) =

6.421 mm I

E' = greater of (A' and D') = 6.421 mm

76

Required thickness, tr =

E'-C .

+C

875

6.421-

+3

3 .875

6.91 mm

Selected thickness, t = 7.14 mm

Schedule 160

Nozzle Mark

B2

Equipment Size NB

Treated gas outlet 100 mm

Code Material

ASME Section VIII Div 1

Max Allowable Stress,

SA 106 G r B 118 MPa

S Design Pressure, P

3.434 MPa 343 K

Design Temperature, T From ANSI 3.36.10,

Outside diameter, D0 = 114.3 mm R0 = 57.15 mm

A'

P.R0

+ C =4.64 mm

SE + 0.4P

3.434x57.15

+3

118x1+0.4x3.434

4.64 mm

B' = head thickness + C

= 20 + 3

= 23 mm

From pipe tables, Standard wall thickness = 4.78 mm = (4.78*.875) + 3 5

= 7 . 1 8 mm

D' = lesser of (B' and C ) = 7.18 mm

1

E' =

greater of (A' and D') = 7.18 mm ■

C = (std. wall thickness x .875) + C 79

Required thickness, tr =

E' - C .

+C

875

7.18-3 .

-3

875

7.78 mm

Selected thickness, t = 7.92 mm

Schedule 40

Nozzle Mark Equipment

Al

Size NB Code Material

Recycle gas inlet 100

Max Allowable Stress,

mm

S Design Pressure, P

ASME Section VIII D

Design Temperature, T

SA 106 G r B 118 MPa

From ANSI 3.36.10,

3.434 MPa 343 K

Outside diameter, D0 = 114.3 mm Ro = 57.15 mm

A' =

P-RQ

+C

SE + 0.4P-

3.434x57.15 U

+3

18x1+0.4x3.434-

= 4.64 mm

B' = head thickness + C

20 + 3

= 23 mm

From pipe tables, Standard wall thickness = 4.78 mm (4.78x.875) + 3

7.18 mm

D' = lesser of (B' and C ) = 7.18 mm E' = greater of (A' and

D') = 7.18 mm *

82

[Required thickness, tr =

E'-C .

+C

875

7.18-3 .

+3

875

= 7.78 m m Selected thickness, t = 7.92 mm Bl Schedule 40 RDEA outlet 80 mm Nozzle Mark Equipment

ASME Section VIII Di

Size NB Code Material

SA 106 G r B 118 MPa

j Max Allowable Stress, S

3.455 MPa 343 K

Design Pressure, P Design Temperature, T [From ANSI 3.36.10, Outside diameter, D0 = 88.9 mm R0 = 44.45 mm

:A' =

P-PvQ

+C

SE + 0.4P

3,455x44.45

+3

118x1+0.4x3.455

4.29 mm

,B' = head thickness + C =

20 + 3 = 23 mm

From pipe tables, Standard wall thickness = 4.37 mm C =

(std. wall thickness x .875) + C = (4.37*.875) + 3

= 6.82 mm

D' = lesser of (B' and C ) = 6.82 mm

E' = greater of (A' and D ' ) = 6.82 mm

85

Required thickness, tr

F -C .

+C

875

6.82-3

+3

- .875

7.37 mm

Selected thickness, t = 7.62 A2 mm Schedule 80 LDEA outlet 80 mm Nozzle Mark ASME Section VIII Div Equipment Size NB

SA 106 G r B 118 MPa

Code Material

3.434 MPa 343 K 86

Max Allowable Stress, S Design Pressure, P Design Temperature, T From ANSI 3.36.10, Outside diameter, D0 = 88.9 mm R0 = 44.45 mm

C = (std. wall thickness x .875) + C 87

A'

P-RQ

SE

+C

+ 0.4P.

3.434x44.45 ll

+3

18x1+0.4x3.434-

= 4.28 mm

B' = head thickness + C

20 + 3

23 mm

From pipe tables, Standard wall thickness = 4.37 mm (4.37x.875) +3

6.82 mm

D' = lesser of (B' and C) = 6.82 mm E' =

greater of (A' and D') = 6.82 mm

88

Required thickness, tr =

E' - C .

+C

875

6.82-3

+3

.875

7.37 mm

Selected thickness, t — 7.62 mm

Schedule 80

5.4

DESIGN OF NOZZLE REINFORCEMENT

According to UG 39 of ASME Section VIII Div 1, reinforcement is a necessity in the following cases: (i)

Nozzles having thickness greater than 89 mm, for a shell thickness not greater than 10 mm.

(ii)

Nozzles having thickness greater than 60 mm, for a shell thickness not greater than 20 mm.

Hence, considering the case (ii), reinforcement is provided for the following nozzles: M1/M2 (Size NB: 600 mm) A1/B2 (SizeNB: 100 mm) B1/A2 (Size NB: 80 mm) Limits of reinforcement: (i)

Parallel to vessel walls:

91

Greater of [(nozzle inside diameter) & (nozzle inside radius+ shell thickness + nozzle thickness)] (ii)

Normal to vessel walls: Lesser of [(2.5 times shell thickness) & {(2.5 times nozzle thickness) + (reinforcement plate thickness)}]

92

Nozzle Mark

Ml

Size NB

600 mm

Code

ASME Section VIII Div

Materials

Shell

SA516Gr

Nozzle

60 SA 106

Reinforcement 118 MPa

GrB SA516

3.452 MPa

Gr 60

Max. Allowable Stresses Design Pressure

Outside diameter of nozzle, do

= 609.6 mm

Inside diameter of nozzle, di

= (do - 2 x nozzle thickness) = (609.6-2x15.88) = 577.84 mm

Inside radius of nozzle, ri Required

= 288.92 mm

thickness of seamless nozzle, trn Nominal

= 14.91 mm

thickness of nozzle, tn Required thickness

= 15.88 mm

of shell, tr Nominal thickness of vessel wall,

= 17.89 mm

t Thickness of reinforcement plate, tp Inside

= 20 mm

radius of vessel, Ri

= 20 mm (assumed) = 301.5 mm

Limits: X = Parallel to vessel wall: 2>f greater of [di and (ri +t + tn)] = 2 x [577.84] = 1155.68 mm

(i)

Y = Normal to vessel wall: lesser of [2.5t and (2.5 tn + tp)] = 50 mm 93

A l = (X-do) tp x (Sp/S) = 10921.6 mm2 = (1155.68-577.84) x20x (118/118).

94

= 11565.6 mm2 Area of shell

available, AT = (t-tr) di = (20-17.89) x 577.84 = 1219.24 m m 2

A2" = 2(t- tr) x (t+tn) =151.41 mm2 = 2* (20-17.89)x(20+15.88) = 151.41 mm2

A2= greater of (A2' and A2") = 1219.24 mm2

Area of nozzle available, A3 = 2(tn-trn) Y (Sn/S) = 2x(15.88-14.91)x50x( l 18/118) = 97

mm2

Total available area, AA= A l +A2+A3= 12237.84 mm2 = 11565.6+1219.24+97 = 12881.84 mm2

Required areas for reinforcement: A'= di x tr = 577.84 x 17.89 = 10337.56 mm2

A"=2xtrx(ro-ri)(l-Sn/S) = 2xl7.89x(304.8-288.92)x( l - ( l 18/118) = 0

mm2

RA = A'+A" = 10337.56 mm2 Available area (AA) is greater than the required area (RA); hence selected thickness, tp=20 mm is secure.

Nozzle Mark

M2

Size NB

600 mm

Code

ASME Section VIII Div 1

Materials

Shell

SA516Gr

Nozzle

60 SA 106 Gr

Reinforcement : 118 MPa : Max. allowable Stresses

3.434 MPa

B S A 5 1 6 Gr 60

Design Pressure = 609.6 mm = (do - 2

Outside diameter of nozzle, do Inside diameter of nozzle, di

x nozzle thickness) = (609.62x15.88) = 577.84 mm = 288.92 mm = 14.91 mm =

Inside radius of nozzle, ri Required

15.88 mm = 17.81 mm = 20

thickness of seamless nozzle, trn Nominal

mm

thickness of nozzle, tn Required thickness of

= 20 mm (assumed) = 301.5

shell, tr Nominal thickness of vessel wall, t

mm

Thickness of reinforcement plate, tp Inside radius of vessel, Ri

Limits: X = parallel to vessel wall: 2 = 2 x [577.84] = 1155.68 mm

greater of [di and (ri +t + tn)]

(ii)

Y = Normal to vessel wall: lesser of [2.5t and (2.5 tn + tp)] = 50 mm

A l = (X-do)tp x (Sp/S) = 10921.6 mm2 = (1155.68-577.84)x20x(l 18/118) - 11565.6 mm2

Area of shell available, A2' = (t-tr) x di = (20-17.81) x 577.84 = 1265.47 mm2

A2" = 2(t- tr) x (t+tn) = 2 x (20-17.81) x (20+15.88) = 157.15 mm2 A2= greater of (A2' and A2") = 1265.47 mm2 Area of nozzle available, A3 = 2 x (tn-trn) xYx (Sn/S) = 2 x (15.88-14.91) x50x (118/11 = 97 mm2

Total available area, AA= A1+A2+A3 = 11565.6+1265.47+97 = 12928.07 mm2

Required areas for reinforcement: A'= di x tr = 577.84x17.81 = 10291.33 mm2

A"=2xtrx(ro-ri)(l-Sn/S)

= 2 x 17.81 x(304.8-288.92)x(l-( 118/118)) = 0 mm2

RA = A'+A" = 10291.33 mm2

Available area (AA) is greater than the required area (RA); hence selected thickness, tp=20 mm is secure. Nozzle Mark

100 mm

Size NB

ASME Section VIII Div 1

Code

Shell

Materials

Nozzle

SA516Gr

Reinforcement 118 MPa

60 SA 106

3.452 MPa

GrB

Max. allowable Stresses

SA516Gr

Design Pressure

60

A1/B2

Outside diameter of nozzle, do

114.3 mm

Inside diameter of nozzle, di

(do - 2 x nozzle thickness) (114.3-2x7.92) 98.46 mm

Inside radius of nozzle, ri Required

49.23 mm

thickness of seamless nozzle, tm Nominal

7.78 mm

thickness of nozzle, tn Required thickness of

7.92 mm

shell, tr Nominal thickness of vessel wall, t

17.81 mm

Thickness of reinforcement plate, tp Inside

20 mm

radius of vessel, Ri

20 mm (assumed) 301.5 mm

Limits:

101

X = parallel to vessel wall: 2

greater of [di and (ri +t + tn)]

= 2x98.46 = 196.92

mm

Y = Normal to vessel wall: lesser of [2.5t and (2.5 tn + tp)] = 39.8 mm

Al=(X-do) tpx(Sp/S) = (196.92-114.3) x20x (118/118)' = 1652.4 mm2

Area of shell available, A2' = (t-tr) di = (20-17.81) x 98.46 = 215.63 mm2

A2" =2(t- tr) x (t+tn) = 2x (20-17.81)x(20+7.92) = 122.29 mm2

A2= greater of (A2' and A2") = 215.63 mm2

Area of nozzle available, A3 = 2(tn-trn) Y (Sn/S) = 2x (7.92-7.78)x39.8x( l 18/118) =11.14 mm2

Total available area, AA= A1+A2+A3 = 1652.4+215.63+ 11.14 = 1879.17 nun2

Required areas for reinforcement: A'= di x tr = 98.46x17.81 = 1753.57 mm2

102

RA = A'+A" = 10291.33 mm2

Available area (AA) is greater than the required area (RA); hence selected thickness, tp=20 mm is secure. A"= 2(trxro-ri) (1-Sn/S) = 2xl7.81x (57.15-49.23) x (1-(118/118)) = 0 mm2

103

Design Pressure

Nozzle Mark

A2/B1

Size NB

80 mm

Code

ASME Section VIII Div 1

Materials

Shell Nozzle

0 SA 106 Gr

Reinforcement : 118 MPa : 3.452 MPa

Max. allowable Stresses

SA516Gr6

B SA516Gr 60

Outside diameter of nozzle, do

88.9 mm

Inside diameter of nozzle, di

(do - 2 x nozzle thickness) 88.9-(2x7.62) 73.66 mm

Inside radius of nozzle, ri Required

36.83 mm

thickness of seamless nozzle, trn Nominal

7.37 mm

thickness of nozzle, tn Required thickness

7.62 mm

of shell, tr Nominal thickness of vessel wall,

17.89 mm

t Thickness of reinforcement plate, tp Inside

20 mm

radius of vessel, Ri

20 mm (assumed) 301.5 mm

Limits:

_ X = Parallel to vessel wall: 2 x greater of [di and (ri +t +tn)] = 2x73.66 = 147.32 mm

105

RA = A'+A" = 1753.57 mm2

Available area (AA) is greater than the required area (RA); hence selected thickness,. tp=20 mm is secure. Y = Normal to vessel wall: lesser of [2.5t and (2.5 tn + tp)] = 39.05 mm Al=(X-do)tpx(Sp/S) = (147.32-88.9) x20x (118/118) = 1168.4 mm2 Area of shell available, A2' = (t-tr) di = (20-17.89) X73.66 155.43 mm2 A2" =2(t- tr) x (t+tn) = 2x (20-17.89) x (20+7.62) = 116.56 mm2

A2= greater of (A2' and A2") 155.43 mm2

Area of nozzle available, A3 = 2(tn-trn) Y (Sn/S) = 2 x (7.62-7.37) x39.05x (118/118) 19.53 mm2

Total available area, AA= A1+A2+A3 \

= 1168.4+155.43+19.53 = 1324.36 mm2

106

Required areas for reinforcement: A'= di x tr = 73.66x17.89 = 1317.78 mm2

A"= 2(trxro-ri)(l-Sn/S) = 2xl7.89x(44.45-36.83)x( l -(118/118) = 0 mm2

107

RA = A'+A" = 1317.78 mm2

109

fig limits of nozzle reinforcement

111

5.5 ESTIMATION OF LOADS ERECTION Shell

7i/4[(Do

Top Head

2

-Di2) x h x p] = 5030.946 kg 1/3 x

4/3 x 7i [(Ro3-Ri3) x h] = 44.478 kg

Bottom Head Ladder

1/3 x 4/3 x 7i [(Ro -Rf) x h] = 44.478 kg 1250 kg (assumed) 35 kg (assumed)

Demister Nozzles: i) Nozzle Mark: V Pipe mass: 7i/4[(do2-di2) x h x p] = 2.012 kg Type of flange: 300# Mass of flange: 4.1 kg Total mass = 6.112 kg

ii) Nozzle Mark: M1/M2 Pipe mass: 7c/4[(do2-di2) x h x p ] = 192.95 kg Type of flange: 300# Mass of flange: 247 kg Total mass = 439.95 kg x 2 = 879.91 kg

iii) Nozzle Mark: LT1/LT2 Pipe mass: 7t/4[(do2-di2) x h x p] = 6.741 kg

Type of flange: 300# Mass of flange: 4.49 kg Total mass = 11.23 kg x 2 = 22.46 kg

113

iv) Nozzle Mark: SP1/SP2 Pipe mass: 7i/4[(do2-di2) x h x p] = 6.741 kg Type of flange: 300# Mass of flange: 4.49 kg Total mass = 11.23 kg x 2 = 22.46 kg

v) Nozzle Mark: A1/B2 Pipe mass: 7i/4[(do2-di2) x h x p] = 4.215 kg Type of flange: 300# Mass of flange: 11.82 kg Total mass = 16.04 kg x 2 = 32.08 kg

vi) Nozzle Mark: A2/B1 Pipe mass: T:/4[(do2-di2) x h x p] = 3.293 kg Type of flange: 300# Mass of flange: 7.12 kg Total mass = 10.413 kg x 2 = 20.83 kg

Total mass of nozzles

: 983.843 kg

Trays

: 10 x 0.8 x n/4 x Di2 x h x p = 246.614kg

Total Weight

7635.36 kg

HYDRO TEST

Mass of water

: (Volume of water) x p = 7866.164 kg :

Total mass

15523.524 kg

OPERATION

Mass of water column

: (Volume of water up to LT1) x p = 1610.06 kg

Mass of liquid hold up on trays Total mass

: 10 x 0.8 x TI/4 x Di2 x h x p = 125.66 k t : 7635.36 kg + 1610.06 kg + 125.66 kg = 9371.082 kg

SHUT DOWN

anen

7c/4[(Do

Top Head

(Ri+2C)3) x h] = 38.238 kg 1/3 x 4/3 x 7i [(Ro3-(Ri+2C)3) x h] =

Bottom

38.238 kg

2

-(Di+2C)2) x h x p] = 4288.88 kg 1/3 x 4/3 x 7i [(Ro3-

Head Nozzles: i)

Nozzle Mark: V

Pipe mass: 7t/4[(do2-(di+2C)2) x h x p] = 1.632 kg Type of flange: 300# Mass of flange: 4.1 kg Total mass = 5.73 kg

ii) Nozzle Mark: M1/M2 Pipe mass: 7t/4[(do2-(di+2C)2) x h x p] = 78.162 kg Type of flange: 300# Mass of flange: 247 kg

117

Total mass = 325.162 kg x 2 = 650.32 kg iii) Nozzle Mark: LT1/LT2 Pipe mass: 7i/4[(do2-(di+2C)2) x h x p] = 3.66 kg Type of flange: 300# Mass of flange: 4.49 kg Total mass = 8.15 kg x 2 = 16.3 kg iv) Nozzle Mark: SP1/SP2 Pipe mass: 7t/4[(do2-(di+2C)2) x h x p] = 6.741 kg Type of flange: 300# Mass of flange: 4.49 kg Total mass = 8.15 kg x 2 = 16.3 kg

v) Nozzle Mark: A1/B2 Pipe mass: 7t/4[(do2-(di+2C)2) x h x p] = 3.465 kg Type of flange: 300# Mass of flange: 11.82 kg Total mass = 15.28 kg x 2 = 30.56 kg

vi) Nozzle Mark: A2/B1 Pipe mass: 7t/4[(do2-(di+2C)2) x h x p] = 2.69 kg Type of flange: 300# Mass of flange: 7.12 kg Total mass = 9.81 kg x 2 = 19.62 kg 118

Total mass of nozzles

: 738.83 kg

Total Weight

:5104.196 kg

119

5.6 CALCULATION OF MOMENT

DUE TO WIND LOAD

Basic wind speed, Vb = 39 m/s

from IS 875 Part 3

Design wind speed, Vz = Vb x k l x k2 x k3, kl

k2

k3

= risk factor = 1.062 = terrain roughness factor = 0.8

from IS 875 Part 3

= topography factor = 1.0

Vz = 3 9 x 1.062 x . 8 x 1

33.134 m/s

Design wind pressure, Pz =0.6 x Vz2/9.81 = 67.148 kg/m2

Wind Force, Fw = Cf x Pz x Ae x Kp = 80.133 kgf = 786.104 N

Moment due to wind load at W.L. = Fw x h i = 786.104 x 10.2

IS 1893 120

= 8.018 kNm

Moment due to wind load at T.L = Fw x h2 = 786.104 x 10.25 = 8.057 kNm

Moment due to wind load at base. = Fw x h3 = 786.104 x 11.25 = 8.843 kNm

121

DUE TO SEISMIC LOAD Fundamental time period, T = 0.085 H 075 = 0.085 x (11.25)075 = 0.522 s T < 0.7 therefore, force due to vibration = 0 N

ERECTION

Base shear force, V = (ZIC/Rw) x W kg = 16.23 kN, where, Z = seismic zone factor = 0.3 I = occupancy importance coefficient = 1.0 C = 1 . 2 5 s/ (T)

(2/3)

=2.89, s = 1.5 Rw =

Numerical Coefficient = 4 W = 7635.359 kg

Moment due to seismic load, Ms = V x 2H/3 = 16.23 x 1 0 0 0 x 2 x 11.25/3 = 1.218 x l 0 5 N m

OPERATION

Mass below LT1, W l =3237.481 kg Mass above LT1, W2 =6007.938 kg HI = 2x2.8/3 = 1.87m H2 = 2 x 8.45/3 + 2.8 = 5.633 + 2.8 = 8.43 m H = (Wl x H I + W2 x H2)/(W1 + W2) = 6.13 m Base shear force, V = (ZIC/Rw) x W kg = 19.93 kN, where, Z = seismic zone factor = 0.3

—

IS 1893

I = occupancy importance coefficient = 1 . 0 C =1.25 s/(T)(2/3) =2.89, s = 1.5 Rw = Numerical Coefficient = 4 W = 9371.082 kg Moment due to seismic load, Ms = V x H = 19.93 x 1000 x 6.13 = 1.221 x 105 Nm

SHUT DOWN

Mass below LT1, W l =1366.063 kg Mass above LT1, W2 =3738.133 kg H I = 2 x 2 . 8 / 3 = 1.87 m H2 = 2 x 8.45/3 + 2.8 = 5.633 + 2.8 = 8.43 m H = (Wl x H I + W2 x m)l (Wl + W2) = 6.67 m Base shear force, V = (ZIC/Rw) x W kg = 10.85 kN,

where, Z = seismic zone factor = 0.3 I = occupancy importance coefficient = 1 . 0

IS 1893 C = 1 . 2 5 s / ( T ) (2/3) =2.89,

s = 1.5 Rw = Numerical Coefficient = 4 W = 5104.196 kg Moment due to seismic load, Ms = V x H = 10.85 x 1000 x 6.67 = 0723 x 105Nm

5.7 DESIGN OF SKIRT SUP] Equipment

Skirt Support (Flared type)

Code

ASME Section VIII Div 1

Material

S A 5 1 6 G r 7 0 118 MPa

Max Allowable Stress, S Design

343 K

Temperature, T 1400 mm Inside diameter of skirt, D

1.0

Efficiency of butt weld joint, E

1.2 2 1 x l 0 5 N r a 9371.082

Moment acting on the skirt ,Ms

Kg

Weight during operation, W

Required thickness,

t=

12Ms

+

W 124

7tR2SE

TtDSE

= 12xl.221xl05 + 9371.082 7tx7002xll8xl

7txl400xll8xl = 8.21

mm

125

5.8 DESIGN OF ANCHOR BOLTS

FOR SEISMIC LOAD Equipment

Anchor Bolt

Code

ASME Section VIII Div 1

Material

SA 193 Gr B 7 130 MPa

Max Allowable Stress, S

343 K

Design Temperature, T

No. of bolts, N = 12 Therefore, diameter of bolt circle, Db = 69 in = 1752.6 mm Area within bolt circle, A = 7t/4(Db2) = 24.112 mm2

Circumference of bolt circle, C = n Db = 5505.95 mm

Selected base block

: Type A

Max tensile force, T = (12Ms/A) - (W/C) = 607.497 x 105

,W= 9371.082 kg in

operation phase Required area of one bolt = (T.CV(S.N) = 214.35 mm2 = 0.332 in2 which corresponds to 0.75" bolt

Adding a corrosion allowance of .12", bolt selected is of .87"

Hence, an M22 bolt is selected.

FOR WIND LOAD Equipment

Anchor Bolt

Code

ASME Section VIII Div 1

Material

S A 1 9 3 Gr B 7 130 MPa

Max Allowable Stress, S

343 K

Design Temperature, T

No. o f b o l t s , N = 12 Therefore, diameter of bolt circle, Db = 69 in = 1752.6 mm Area within bolt circle, A = 7t/4(Db2) = 24.112 mm2

Circumference of bolt circle, C = n Db = 5505.95 mm

Selected base block

: Type A

Max tensile force, T = (12Mw/A) - (W/C) = 43.842 x 105

, W= 9371.082 kg

operation phase Required area of one bolt = (T.C)/(S.N) = 15.47 mm2 which corresponds to 5/8" bolt 128

Hence, an M20 bolt is selected.

Considering both cases, selected bolt

: M22 CHAPTER.6 ANALYSIS

6.1 WIND ANALYSIS ERECTION

Stress due to wind load, Sw = Mw x (Y/I) = 551811.975 N/m2 = .552 MPa, Where Y=Do/2 = . 52 m 1 =7t/64(Do4-Di4) = 8.337 mm4

Stress due to internal pressure, Sp = Pd/4t = 42918750 N/m2 = 42.92 MPa

Stress due to erection load, SI = W/7tDi = 1.192 MPa

Combined stress:

(a) Windward side: S= Sw +Sp - SI = 44.664 MPa

Combined stress: 129

(b) Leeward side: S=Sp-Sw-Sl = 41.176 MPa OPERATION Stress due to wind load, Sw = Mw x (Y/I) = 551811.975 N/m2 = .552 MPa, where Y=Do/2 = . 52 m, I =?i/64(Do4-Di4) = 8.337 mm4

Stress due to internal pressure, Sp = Pd/4t = 42918750 N/m2 = 42.92 MPa

Stress due to erection load, SI = W/7tDi = 1.463 MPa

(a) Windward side: S= Sw +Sp - SI = 42.009 MPa (b) Leeward side: S=Sp-Sw-Sl = 40.905 MPa SHUT DOWN Stress due to wind load, Sw = Mw x (Y/I) = 551811.975 N/m2 = .552 MPa, where Y=Do/2 = . 5m I =7t/64(Do4-Di4) = 8.337 mm4

Stress due to internal pressure, Sp = Pd/4t = 42918750 N/m2 = 42.92 MPa 130

Stress due to erection load, SI = WAiDi = 0.797 MPa

Combined stress:

(a) Windward side: S= Sw +Sp - SI = 42.675 MPa (b) Leeward side: S=Sp-Sw-Sl = 41.571 MPa S < Allowed stress, hence the shell thickness selected is secure.

Combined stress: 131

6.2 SEISMIC ANALYSIS

ERECTION

Stress due to seismic load, Ss = Mw x (Y/I) = 74.52MPa, where Y=Do/2 = .52 m, I =7t/64(Do4-Di4) = 8.337 mm4

Stress due to internal pressure, Sp = Pd/4t = 42918750 N/m2 = 42.92 MPa

Stress due to erection load, SI = W/nDi = 1 . 1 9 2 MPa

Combined stress:

(c) Windward side: S= Ss +Sp - SI = 116.25 MPa (d) Leeward side: S=Sp-Ss-Sl = -32.792 MPa

OPERATION

Stress due to seismic load, Ss = Mw x (Y/I) = 74.77MPa, where Y=Do/2 = .52 m, I =7t/64(Do4-Di4) = 8.337 mm4

Stress due to internal pressure, Sp = Pd/4t = 42918750 N/m2 = 42.92 MPa

132

Stress due to erection load, SI = W/nDi = 1.463 MPa

Combined stress: (c) Windward side: S= Ss +Sp - SI = 116.23 MPa (d) Leeward side: S=Sp-Ss-Sl = -33.313 MPa

SHUT DOWN

Stress due to wind load, Sw = Mw x (Y/I) = 44.24MPa, where Y=Do/2 = .52 m, I =7r/64(Do4-Di4) = 8.337 mm4 Stress due to internal pressure, Sp = Pd/4t = 42918750 N/m2 = 42.92 MPa

Stress due to erection load, SI = W/rcDi = 0.797 MPa

Combined stress:

(c) Windward side: S= Sw +Sp - SI = 86.363 MPa (d) Leeward side: S=Sp-Sw-Sl = -2.301 MPa

S < allowable stress, hence the shell thickness selected is secure

133

The nozzles that require reinforcement are M l , M2, B2, A l , A2 and B l . The nozzle reinforcement thickness obtained from calculation is 20mm. CHAPTER.7 RESULTS AND

DISCUSSION

7.1 ELLIPSOIDAL HEADS SL NO NAME OF PART 1 Top Head 2 Bottom Head 7.2 CYLINDRICAL SHELLS SL NO NAME OF PART 1 2

Top Section Bottom Section

THICKNESS mm - 20 20

DIAMETER mm 1000 1000

HEIGHT mm 250 250

THICKNESS mm 20 20

DIAMETER mm 1000 1000

HEIGHT mm 6800 1800

7.3 NOZZLES SL NO

NOZZLE MARK

1 Vent with valve (V) 2 Manhole (MI/M2) 3 Level Transmitter (LT1/LT2) 4 Stand Pipe (SP1/SP2) 5 Treated gas outlet (B2) 6 RDEA outlet ( B l ) 7 LDEA inlet (A2) 8 Recycle gas inlet ( A l ) 7.4 NOZZLE REINFORCEMENT

DIAMETER (Do) mm 60.3 609.6 60.3 60.3 114.3 88.9 88.9 114.3

134

THICKNESS mm 7.14 15.88 7.14 7.14 7.92 7.62 7.62 7.92

7.5 SUPPORT Skirt support is selected for this pressure vessel. The thickness of the skirt support obtained is 8.21mm.

7.6 ANCHOR BOLTS Number of bolts required = 12 Bolt circle diameter obtained = 1752.6mm By considering both wind load and seismic load selected bolt is M22.

7.7 ANALYSIS

In both wind analysis and seismic analysis, the combined stress of stress due to seismic load, wind load, internal pressure and erection load is less than that of the allowable stress, so selected thickness of shell is secure. 8.

CONCLUSION .

The column is successfully designed so that it with stand all the mechanical stresses acting on it. The column is analyzed under various conditions of operation. All forces are carried according to ASME codebook. The column also with stand the internal pressure of fluid at working conditions. The various forces analyzed are pressure exerted by gas on the shell, weight of the fluid, wind force, seismic force. The stresses in above-mentioned conditions are found out and thickness of various parts is selected such that the stresses produced in each member are with in the maximum allowable range. All the selected have been successfully verified and hence the design of column is safe.

135

9.

REFERENCES ■

1. ASME CODE BOOK 2004 2. K. Linghaiah: Machine Design Handbook (Vol .1), 3. Harvey John.H: Pressure Vessel Design 4. Chuse Robert: Pressure Vessel ASME CODE Simplified 5. R.W Nichole: Pressure Vessel Engineering Technology 136

6. Pressure Vessel Designing: Henry.H. Betner 7. Dennis. R. Mon: Pressure Design Manual

137