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Vertical Unfired Pressure Vessel Components Engineering Essay

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The American Society of Mechanical Engineers was organized in 1880 as an educational and technical society of mechanical engineers. After years of development and public comment, the first edition of the Code, ASME Rules of Construction of Stationary Boilers and for Allowable Working Pressures, was published in 1914 and formally adopted in the spring of 1915. The first Code rules for pressure vessels, entitled Rules for the Construction of Unfired Pressure Vessels, followed in 1925. From this simple beginning the Code has now evolved into the present eleven Section document, with multiple subdivisions, parts, subsections, and mandatory and non-mandatory appendices. Almost all pressure vessels used in the process industry in the United States are designed and constructed in accordance with Section VIII Division 1. In this project, some general concepts criteria related to ASME Code Section VIII are discussed. These include allowable stress, factors of safety, joint efficiency and pressure testing. The objective of this project is to design and analysis Unfired Vertical Pressure Vessel based on ASME Code Section VIII Division 1 and standards. This project only concerned to design main part of pressure vessel like shell, heads, nozzles and supports. The rules in Section VIII Division 1 do not cover all applications and configurations such as designing leg supports. When the rules are not available, another method must be used.

Problem statement

The pressure vessels that not follow any standard codes can be very dangerous. In fact many fatal accidents have occurred in the history of their operation and development. They are many standards and codes that vary from country to country. The common standards and codes that have been used are ASME Boilers and Pressure Vessel Codes, API Standards, PD5500, British Standards, European Codes and Standards and other International Codes. Even though there are computer aided pressure vessel design available in the market, but due to business benefit, the system may not be saleable or pricey. In addition the formulas and concepts applied in the system are always unknown by the users.

Research scope

This project focuses on design and analysis of Unfired Vertical Pressure Vessel based on ASME Code Section VIII Division 1. Based on this code, pressure vessels are application for the containment of internal and external pressure up to 3000 psi. This pressure could be obtained from an external source or by the application of heat from a direct or indirect source or any combination of them. The ASME Code is construction code for pressure vessel and contains requirements, specific prohibitions; and non-mandatory guidance for pressure vessel materials, design, welding and testing. To ensure the objective of this project is achieved, some of the important elements must be consider. There is:

Designing main components of Unfired Vertical Pressure Vessel by refer to ASME Code Section VIII Division 1 and standards.

Analysis of maximum stress value of main components of pressure vessel by finite element using ANSYS software.

Objectives of Project

The purpose of this project is to design and analysis of Vertical Unfired Pressure Vessel based on ASME Code Section VIII Division 1. This researcher points two objectives to be achieved at the end of this research. The objectives are:

1. To design Vertical Unfired Pressure Vessel components based on ASME Code VIII Division 1 and Standards.

2. To analyze maximum equivalent stress (von-Misses), maximum shear stress, maximum deformation and safety factor in shell by finite element using ANSYS software.

Significance of studies

The project will bring a great significant not only for the fertilizer industry but also to the all the manufacturing sector that used a various pressure vessel for daily operation. Nowadays, most the manufacturing industry in Malaysia which used pressure vessel for operational purpose depends on their area of application. As a result, their operation, design, manufacture is regulated by engineering authorities backed up by laws. All pressure vessels are manufactured with the maximum safe operating pressure and temperature. By completing this project, student will gain exposure to the ASME code and standards.

CHAPTER 2.0

LITERATURE REVIEW

2.1 Introduction

The pressure vessels such as cylinder, pipeline or tanks are design and construct to store gas or fluids under pressure. The gas or fluid that being stored may be through change of state inside the pressure vessel, for example case of steam boilers or it might combine with other reagents, such as a chemical plant. The pressure vessels must design with a perfect care because crack of pressure vessels will cause an explosion which may cause of death and loss of property. The material that be used to construct pressure vessels may be ductile such as mild steel or brittle such that cast iron. In generally, pressure vessels and others storage tank such as hydraulic cylinders, gun barrels, pipes, boilers and tanks are important to the chemical, petroleum, petrochemical, nuclear industries and so on. Reactions, separations, and storage of raw materials always occur in this class of equipment. Generally, pressurized equipment is required and been used for a broad range of industrial plant for storage and manufacturing purposes [1].

2.2 Types of Pressure Vessel

The size and geometric form of pressure vessels diverge greatly from the large cylindrical vessels used for high-pressure gas storage to the small size used as hydraulic units for aircraft. Some of the vessels are buried in the ground or deep in the ocean, but most are positioned on ground or supported in platforms. There are mainly two types of pressure vessels usually available in industry:

Spherical Pressure Vessel

This type of pressure vessels are known as thin walled vessels. This forms the most typical application of plane stress. Plane of stress is a class of common engineering problems involving stress in a thin plate. Spherical vessels have the advantage of requiring thinner walls for a given pressure and diameter than the equivalent cylinder. Therefore they are used for large gas or liquid containers, gas-cooled nuclear reactors, containment buildings for nuclear plant, and so on.

C:\Users\zalie87\Desktop\spherical pressure vessel 2.jpg

Figure 2.1 Spherical Pressure Vessel [source: http://communities.ptc.com/thread/39900]

Cylindrical Pressure Vessel

This type of a vessel designed with a fixed radius and thickness subjected to an internal gage pressure. This vessel has an axial symmetry. The cylindrical vessels are generally preferred, since they present simpler manufacturing problems and make better use of the available space. Boiler drums, heat exchangers, chemical reactors, and so on, are generally cylindrical.

 C:\Users\zalie87\Desktop\pressure-vessel-500x500.jpg C:\Users\zalie87\Desktop\vertical_expansion_tank.gif

Figure .2: Cylindrical (Horizontal & Vertical) Pressure Vessel [source: http://www.energyflowsystems.com/pv.htm and http://www.pumpsukltd.com]

2.3 Main Components of Pressure Vessel

The main pressure vessel components are as follow:

2.3.1 Shell

The shell is the main component of any vessels that contains the pressure. Material of shell normally come in plate or rolled steel. Commonly, some pressure vessel shells has a rotational axis and be welded together to form a structure. Most pressure vessel shells are cylindrical, spherical, or conical in shape.

2.3.2 Head

All pressure vessel shells must be closed at the ends by heads. Heads that usually used are typically in curved rather than flat. Configurations of curved shape stronger and let the pressure vessel's heads to be thinner, lighter and less expensive rather than flat heads. Inside a vessel, heads can also be used.

Heads are usually can be categorized by their shapes. Ellipsoidal, hemispherical, torispherical, conical, toriconical and flat are the common types of heads. Figure 2.3 shows various types of heads. Ellipsoidal would be the most common type of heads, which is used during the designing of a new pressure vessel. [11]

Figure 2.3: Typical Types of Heads [source 11]

2.3.3 Nozzles

A nozzle is a cylindrical component that penetrates and mounts whether at the shell or heads of a pressure vessel surface. The nozzle ends are generally flanged. Flanges function is to allow the necessary connections. Flanges also use to permit easy disassembly for routines maintenance or easy access. Nozzles commonly are used for the following applications [11]:

Attach piping for flow inlet or outlet of the vessel.

Attach instrument connections such as level gauges, thermowells, or pressure gauges.

Provide access to the vessel interior at manholes.

Provide for direct attachment of heat exchanger or mixer.

Nozzles sometimes extended into the vessel interior for some applications, such as for inlet flow distribution or to permit the entry of thermowells.

2.3.4 Support

The type of support that is designed and used depends on the orientation of the pressure vessel whether horizontally or vertically. In any situation, the pressure vessel support must be enough to support the applied weight and other loads. Design pressure of the vessel is not being considered in the design of its support because the support is not be pressurized. But, design temperature should be considered for support design. It should be considered from the perspective of material selection and provision for differential thermal expansion.

Several kinds of supports are as follow [11]:

Skirt

This type of support generally been used for tall, vertical, cylindrical pressure vessels. This type of support is a cylindrical shell section which is be weld either to the bottom of the vessel shell or to the bottom head for the cylindrical vessels. Skirt support for spherical vessel is welded to the vessel near the mid plane of the shell. The skirt is normally design long enough to provide flexibility so that radial thermal expansion of the shell does not cause high thermal stresses at its junction with the skirt.

Leg

Small vertical drums are normally supported by legs that are welded to the bottom of the pressure shell. The maximum ratio of support provides for leg length to drum diameter is typically 2:1. The number of legs is designed depends on the drum size and the loads to be carried. Support legs are also usually designed for spherical pressure vessels. The support legs for small vertical vessels and spherical storage vessels normally made from high carbon material such as structural steel columns or pipe sections, which provides a more efficient and perfect design.

Saddle

Horizontal drums are normally supported by saddle. This type of support divides the weight load over a large area of the shell to avoid an unnecessary stress in the shell at two different locations. The width of the saddle is considered by the specific size and design conditions of the pressure vessel. One saddle support is normally fixed or anchored to its foundation. A typical scheme of saddle support is shown on Figure 2.2.4.

Figure 2.4: Typical Scheme of Saddle [source 11]

2.4 Overall Design Procedure of Pressure Vessels

Pressure vessels as components of a complete plant are designed to meet various requirements as determined by the designers and analysts responsible for the overall design. The first step in the design procedure is to select the necessary relevant information, establishing in this way a body of design requirements, as shown in Figure 2.5. Once the design requirements have been established, suitable materials are selected and the specified design code will give an allowable design or nominal stress that is used to dimension the main pressure vessel thickness. Additional code rules cover the design of various vessel components such as nozzles, flanges, and so on. Following these rules an arrangement of the various components are finalized and analyzed for failure. Most of the types of failure relevant to pressure vessel design are stress dependent and therefore it is necessary to ensure the adequacy of the stress distribution and check against different types of postulated failure modes. The proposed design is finally iterated until the most economical and reliable product is obtained. The functional requirements cover the geometrical design parameters such as size and shape, location of the penetrations, and so on. Some of these parameters may have to be fixed in collaboration with the overall design team, but in a majority of situations the pressure vessel designer acts freely on the basis of his or her experience. In the procedure in designing pressure vessels, safety is the main factor that must be consider, especially for the high pressure plant such as nuclear reactor pressure vessels, due the possible impact of a possible severe accident. In general however, the design is a compromise between consideration of economics and safety. The potential risks of a given mode of failure and its consequences are balanced against the effort required for its prevention. The resulting design should achieve an adequate standard of safety at minimum cost. Safety cannot be absolutely assured for these two reasons.

First, the actual form of loading during service may be more severe than was anticipated at the design stage: abnormal, unpredictable loads inevitably occur during the pressure vessel's lifetime. Second, our knowledge is seldom adequate to provide a qualified answer to the fracture of materials, state of stress under certain conditions, and so on. It is true that although the fundamental mechanism of failure is not sufficiently understood, it is possible to establish preventive measures based on semi empirical methods. Following this line of thinking, the pressure vessels could be classified according to the severity of their operations since this will affect both the possibility of failure and its consequences. These considerations lead to the classification of vessels ranging from nuclear reactor pressure vessels at one end to underground water tanks at the other. The design factor used in the ASME Boiler and Pressure Vessel Code1 is intended to account for unknown factors associated with the design and construction of the equipment. The design formulas and the stress analysis methods are generally approximate and have built-in assumptions. Typically it is assumed that the material is homogeneous and isotropic. In the real world the material has flaws and discontinuities, which tend to deviate from this assumption.

Figure 2.5: Design Procedure

CHAPTER 3.0

METHODOLOGY

3.1 Overview

In this chapter, the information in selection of pressure vessel is described and the application of selected pressure vessel is been discussed. To design of pressure vessel the selection of Code are important as a reference guide to achieve the secure pressure vessel. The selections of ASME Code Section VIII div 1 are described. The standard of material selection used are explains in this chapter. Beside of that, the design and analysis software to obtain the result are introduced. Instead of that, design process methodology is also described.

3.2 General Design Considerations: Pressure Vessels

3.2.1 Materials

General material requirement have been described in paragraphs UG-4 through UG-15. There are some points that must be considered which is related to the general material requirements that will be discussed below. [2]

The main factors of material selection that must be considered are [12]:

Strength

Strength is a material's ability to endure an imposed force or stress applied. Strength is an important factor in the material selection for any particular application.

Strength determines the thick of a component that must be to withstand the forced loads.

Corrosion Resistance

Corrosion defines as the weakening of material by chemical reaction. Material's resistance to corrosion is the most important factor that influences its selection for a specific application.

Specify a corrosion allowance is the common method that used to define corrosion in pressure vessels components.

Fracture Toughness

Fracture toughness defines as the capability of a material to withstand conditions that could cause a brittle fracture. The fracture toughness of a material can be determined by using Charpy V-notch test to define the magnitude of the impact energy and force that is required to fracture a specimen.

Fabricability

Fabricability defines as the ease of construction and to any special fabrication practices that are required to use the material.

Commonly, pressure vessels use welded construction. The materials used must be weldable so that components can be assembled onto the completed pressure vessel.

The pressure vessel design codes and standards include lists of acceptable materials; in accordance with the appropriate material standards.

3.2.2 Design and Operating Temperature

In ASME Code Section VIII Div 1, maximum and minimum design temperatures can be established in Paragraph UG-20. The maximum design temperature can be define as the maximum temperature used in vessel design and it shall not be lesser than the mean metal temperature estimated under normal operating conditions for the part that want to be considered. [3]

The operating temperature is the gas or fluid temperature that occurs under the normal operating conditions. Before designing a vessel, the operating temperature must be set based on the maximum and minimum metal temperatures that the pressure vessel may encounter any situation. [4]

3.2.3 Design and Operating Pressure

Design pressure of the vessel can be established in Paragraph UG-21. In this paragraph, the requirement of the vessel to be designed for any severe pressure and temperature that is coincidentally expected in normal operation has been provided. When establish the maximum operating pressure, all conditions such as start-up, shutdown, and any identified upset conditions can be considered. Set pressure of the pressure relief device in an operating system must be above the operating pressure by a sufficient amount so that the device does not actuate accidentally. A vessel must be designed to withstand the maximum pressure to which it is likely to be subjected in operation condition. Before designing a vessel, the operating pressure must be set based on the maximum internal or external pressure that the pressure vessel may encounter.

The design pressure is normally taken as the pressure at which the relief device is set for vessel that under internal pressure. To avoid spurious operation during minor process upsets, normally the operation pressure is 5 to 10 per cent above the normal working pressure. The hydrostatic pressure in the base of the column should be added to the operating pressure if deciding the design pressure. [2]

3.2.4 Design Maximum Allowable Stress

Maximum allowable stress that have to be consider in designing a vessel which be used for internal and external pressure has be describe in Paragraph UG-23. The allowable tensile stresses are tabulated in ASME Code Section II, Part D of the Boiler and Pressure Vessel Code. In UG-23(a) indicates that for material that has been identified as meeting more than one material specification, the allowable stress for the specification may be used and provided that all the limitations of the specification is satisfied. In UG-23, criteria for the maximum allowable longitudinal compressive stress to be used for cylindrical shells that are subjected to longitudinal compressive loads also have been provided. The first condition is that the maximum allowable longitudinal compressive stress cannot be greater than the maximum allowable tensile stress. The second condition is based on buckling of the component. In Paragraph UG-23(c), the wall thickness of a pressure vessel shell defined by these rules and it should be determined and the induced maximum membrane stress does not exceed the maximum allowable stress value in tension has been stated. [2]

Typical design stress factors for pressure components are shown in Table 3.1.

Table 3.1: Design stress factors

Property

Material

Carbon

Carbon-manganese, stainless metals

low alloy steels

Austenitic stainless steels

Non-ferrous

metals

Minimum yield

stress or 0.2 percent proof stress, at the design temperature

1.5

1.5

1.5

Minimum tensile strength, at room temperature

2.35

2.5

4.0

Mean stress to

produce rupture

at 105 h at the

design temperature

1.5

1.5

1.0

3.2.5 Thickness of shell under internal pressure

Information and requirement of thickness or maximum allowable pressure for a shell under internal pressure are provided in paragraph UG-27. The equations for circumferential stress which is the stress acting across the longitudinal seam for cylindrical shell are as follows [1]:

or (3.2.5.1)

Figure 3.1: Shell Under Internal Pressure

For cylindrical shells for longitudinal stress which the stress acting across the circumferential joints, the equations are

or (3.2.5.2)

t = minimum required thickness of shell, in. (in the corroded condition)

P = internal design pressure, psi

R = inside radius of shell under consideration, in. (Corroded condition)

S = maximum allowable stress from the applicable allowable stress table in Section II, Part D

E = Joint efficiency for welded joints (Table UW-12), or the ligament efficiency between openings (UG-53).

For spherical shells,

or (3.2.5.3)

These equations are very simple. However, there are some related issues that must be discussed. These two equations are normally based on thin wall theory.

3.2.6 Thickness of shell under external pressure

The information and requirement that used to design shells and tubes under external pressure is given as a design load is given in paragraph UG-28. The definitions for various geometries are graphically shown in Figure 3.2.a (Fig.UG-28.1). [2]

Figure 3.2.a: Diagrammatic Representation of Lines of Support for Design of Cylindrical Vessels Subjected To External Pressure (Source: Fig. UG-28.1 of Section VIII Div. 1 of the ASME 2010 Code)

Figure 3.2.b: Maximum Arc of Shell Left Unsupported Because of Gap in Stiffening Ring of Cylindrical-Shell under External Pressure (Source: Fig.UG-29.2 of Section VIII Div.1 of the ASME 2010 Code)

Stiffness ring that has been provided with continuous around the circumference of the vessel is to resist external pressure. Between the ring and the shell, gaps have been allowed; however, the ring has to be continuous and the arc of the gap is limited by Figure 3.2.b. The additional requirements of UG-29(c) (1) through UG-29(c) (4) should be satisfy when the arc of the gap between the ring and shell does not meet the Figure 3.2.b requirements. [2]

3.2.7 Formed Heads

Information and rules for the design of formed heads are given in paragraph UG-32. The required thickness of ellipsoidal heads formula is given by

or (3.2.7.1)

D = diameter of the ellipse major axis

Figure 3.3: Ellipsoidal head (Source: 7)

Other formulas to design heads are as given in UG-27.Ellipsoidal heads has a ratio of 2:1 if there does not have a major to minor diameter. The torispherical head with the knuckle radius requires a thickness for a equal to 6% of the inside crown radius and the inside crown radius equal to the outside diameter of the is given by [7]

or (3.2.7.2)

Where: L = inside crown radius of the formed head

Figure 3.4: Torispherical head (Source: 7)

3.2.8 Openings and Reinforcements

When designing an opening in a pressure vessel, there is a stress resulting from the hole that is formed on the shell. This is similar to the classical stress concentration effect of a hole in a plate that is loaded in traction. The codes for reinforcement do not consider loads other than pressure. Openings in shells should be round, elliptical, or obround. If the connection is slanting to the surface of the shell, the elliptical opening in the shell will be used. The proof test in Paragraph UG-101should is applied if the strength of vessels with such openings cannot be determined. [2]

There is no limit to the size of an opening that may be designed on a pressure vessel. The opening and reinforcement rules in paragraph UG-36 through UG-43 stated in ASME Code will be apply to openings not exceeding the following vessel size. For example, vessels of 60 inches inside diameter and less, the opening may be as large as one half the vessel diameters, but not to exceed 20 inches. Then, for vessels over 60 inches inside diameter, the opening may be as large as one third the vessel diameter, but not to exceed 40 inches. [2]

Design for Internal Pressure

The total cross sectional or area of reinforcement A in any plane through the opening for a shell or head under internal pressure that has been required shall be not less than

A = dtrF + 2tn trF(1 − fr1 ) (3.2.8.1)

Design for External Pressure

(1) The reinforcement that subject to pressure (external) must be considered for openings in single walled vessels must only 50% of that required in design for internal pressure, where tr is the wall thickness required by the rules for vessels under external pressure and the value of F shall be 1.0 in all external pressure reinforcement calculations. [2]

(2) The reinforcement required for openings in each shell of a multiple walled vessel shall comply with above information when the shell is subject to pressure (external) and with design for pressure (internal) above when the shell is subject to internal pressure, no matter there is a common nozzle secured to more than one shell by strength welds. [2]

3.2.9 Nozzles

The minimum wall thickness of nozzle necks should be determined as given formula below. For access openings and openings used only for inspection [2]:

tUG-45 = ta (3.2.9.1)

For other nozzles:

Determine tb.

tb = min [tb3, max (tb1, tb2)] (3.2.9.2)

tUG-45 = max (ta, tb) (3.2.9.3)

where

ta = minimum neck thickness required for internal and external pressure using UG-27 and UG- 28 (plus corrosion allowance), as applicable. The effects of external forces and moments from supplemental loads (see UG-22) shall be considered. Shear stresses caused by UG-22 loadings shall not exceed 70% of the allowable tensile stress for the nozzle material.

tb1 = for vessels under internal pressure, the thickness (plus corrosion allowance) required for pressure (assuming E p 1.0) for the shell or head at the location where the nozzle neck or other connection attaches to the vessel but in no case less than the minimum thickness specified for the material in UG-16(b).

tb2 = for vessels under external pressure, the thickness (plus corrosion allowance) obtained by using the external design pressure as an equivalent internal design pressure (assuming E p 1.0) in the formula for the shell or head at the location where the nozzle neck or other connection attaches to the vessel but in no case less than the minimum thickness specified for the material in UG-16(b).

tb3 = the thickness given in Table UG-45 plus the thickness added for corrosion allowance.

tUG-45 = minimum wall thickness of nozzle necks

In Paragraph UG-45, the rules for minimum nozzle neck thickness have been provided. A nozzle neck or any other connection shall not be thinner than that required to satisfy the thickness requirements for the loads defined in paragraph UG-22. Except for manhole and other openings that are provided only for access, additional requirements of paragraph UG-45 may require a thicker nozzle neck. [2]

3.2.10 Legs support

Legs supports normally are used to support vertical pressure vessel. Legs support can be made detachable from the vessel. These supports can be bolted or welded to shell plates. Leg supports design method is similar to that for bracket support. If the legs are welded to the shell, then the shear stresses in the weld will be given by [2]:

(3.2.10.1)

Where, tW = Weld Height

LW = Weld Length.

These kinds of supports are suitable only for small and moderate pressure vessels as there is a concentrated local stress at the joint.

Figure 3.5: Leg Support

3.2.11 Joint Efficiency Factors

The strength of a welded joint will depend on the type of joint and the quality of the welding. The soundness of welds is checked by visual inspection and by non-destructive testing (radiography). The possible lower strength of a welded joint compared with the virgin plate is usually allowed for in design by multiplying the allowable design stress for the material by a "welded joint factor" J. The value of the joint factor used in design will depend on the type of joint and amount of radiography required by the design code. Typical values are shown in Table 3. Taking the factor as 1.0 implies that the joint is equally as strong as the virgin plate; this is achieved by radiographing the complete weld length, and cutting out and remaking any defects. The use of lower joint factors in design, though saving costs on radiography, will result in a thicker, heavier, vessel, and the designer must balance any cost savings on inspection and fabrication against the increased cost of materials. [2]

Table .2: Maximum allowable joint efficiency

Type of joint

Degree of radiography

100 %

spot

none

Double-welded butt

or equivalent

1.0

0.85

0.7

Single-weld butt

joint with bonding strips

0.9

0.80

0.65

In ASME Code Section VIII Division 1, joint efficiency factors influence the level of examination of joints on pressure vessel. The degree of examination influences the required thickness through the use of Joint Efficiency Factors, E. This factor is sometimes referred to as Quality Factors or weld efficiencies serve as stress multipliers applied to vessel components when some of the joints are not fully radiographed. Basically, ASME Code Section VIII Division 1 vessels have variable factors of safety and it depending on the radiographic examination of the main vessel components joints. For this project, fully radiographed longitudinal butt-well joints in cylindrical shell use a Joint Efficiency Factor, E of 1.0. There are four joint categories require that have been identified in ASME Code Section VIII Division 1. They are categories A, B, C and D as shown in figure below. [2]

Figure 3.6: Welded Joint Categories (Source: 2010 ASME VIII Div1)

3.2.12 Corrosion allowance

The corrosion allowance is the additional thickness of metal added to allow for material lost by corrosion and erosion, or scaling. The allowance to be used should be agreed between the customer and manufacturer. Corrosion is a complex phenomenon, and it is not possible to give specific rules for the estimation of the corrosion allowance required for all circumstances. The allowance should be based on experience with the material of construction under similar service conditions to those for the proposed design. For carbon and low-alloy steels, where severe corrosion is not expected, a minimum allowance of 2.0 mm should be used; where more severe conditions are anticipated this should be increased to 4.0 mm. Most of design codes and standards available specify a minimum allowance of 1.0 mm. [2]

3.3 Finite Element Analysis by ANSYS

This project is set out to verify finite element analysis, FEA when applied to pressure vessel design. Finite Element Analysis is a simulation technique. Function of this technique is to evaluate the behavior of components, equipment and structures for various loading conditions including applied forces, pressures and temperatures. There are many complex engineering problems with non-standard shape and geometry can be solved using this analysis [5]. Results that can be achieve by this analysis such as the stress distribution, displacements and reaction loads at supports for any model. There are number of scenarios can be done such as design optimization, material weight minimization, shape optimization, code compliance and more by using this analysis[10]. The finite elements analysis was performed using ANSYS software. ANSYS widely used in the computer-aided engineering (CAE) field in many industries [10]. ANSYS software helps engineers and designers to construct computer models of structures, machine components or systems by applying operating loads and other design criteria and to study physical responses such as stress levels, temperature distributions, pressure and more. It permits an evaluation of a design without having to build and destroy multiple prototypes in testing. In this project, the analysis will be test on cylindrical shell of the unfired vertical pressure vessel to see the maximum deformation, maximum equivalent (von-Misses) and maximum shear stress of the shell's material.

Figure 3.1: Example of ANSYS analysis; Maximum shear stress of Elliptical Head [source 1].

CHAPTER 4.0

RESULT AND ANALYSIS

4.1 Design Data and Calculation

Table 4.1: Pressure Vessel Design Data

Design code

: ASME Section VIII Division 1

Type of vessel

: Vertical

Inside diameter

: 1300.0 mm

Temperature

Design

: 70.0 °C

Operating

: 30.0 °C

Pressure

Design

: 44 BarG

Operating

: 24.9 BarG

Corrosion allowance

: 3 mm

Type of fluid

: Natural gas

Max. Liquid level

: Not applicable

Radiography

: Full

Joint efficiency

: 1.0

Type of head

: 2:1 Ellipsoidal

Weight

Empty

: 4791 kg

Operating

: 4850 kg (approximate)

4.1.1 Material

For selecting material for construction these pressure vessel components, there are several rules should be consider that available in paragraphs UG-4 through UG-15. For this project, material that will be use is in carbon and low alloy steel's class which is SA-516-70. This type of material has been choosing based on design pressure and design temperature because it is suitable for moderate and lower temperature service applications. [2]

4.1.1.1 Properties of Material

Table 4.2: Properties of Material

Material

SA-516 Gr 70

Form

Plate

Composition

C-Mn-Si

Tensile strength

552 MPa

Yield point

260 MPa

Density

7.85 g/cm3

Melting Point

1510 °C (2750 °F)

4.1.2 Design Pressure

Refer to ASME code in paragraph UG 21, the design pressure is a pressure that is used to design a pressure containing system or piece of equipment. With the design pressure, it is recommended for engineer to design a vessel and its components. Design pressure must 5-10 % higher than operating pressure, whichever is the higher, will fulfill this requirement. The pressure of the fluid and other contents of the pressure vessel are also considered. For this project, design pressure is 44.0 BarG. [2]

4.1.3 Operating Pressure

Operating pressure is a pressure that less than the maximum allowable working pressure at which the pressure vessel is normally operated. Recommended value is 30% below maximum allowable working pressure. [2]

4.1.4 Maximum Allowable Stress Value

Refer to ASME code in paragraph UG 23, the maximum allowable stress value that the maximum stress allowed in material that used to design pressure vessel components under this rules. The allowable stress value for most material at design temperature is the lower 2/7 the minimum effective tensile strength or 2/3 the minimum yield stress of the material. For this project, the allowable stress value is obtained from table in ASME Code Section II; Part D. Below is allowable stress value that simplified from the table in section II, Part D. [2]

Material

Metal temperature not exceeding deg, F

Maximum Allowable Stress, psi

SA-516 Gr 70

-20 to 650

17500

Table 4.3: Maximum Allowable Stress Value

4.1.5 Thickness of Shells under Internal Pressure

t = PR per UG 27 (c)

(SE- 0.6P)

= (44 x 10^5) (653 mm)

(1206.58 Bar) - 0.6(44 Bar)

= 24.35 mm

24.35 mm + corrosion allowance, 3 mm = 27.35 mm

So, use t = 28 mm

Maximum Allowance Working Pressure, MAWP

P = SEt per UG 27 (c)

R + 0.6t

= (1206.58 x 10^5) (1) (28 mm)

650 mm + 0.6 (28 mm)

= 51.98 Bar

Stress, σhoop = P (R + 0.6t)

Et

= (4.4 x 10^6) (0.650m + 0.6(0.028 m)

(1)0.028 m

= 105.25MPa

Stress, σlong = P (R - 0.4t)

2Et

= (4.4 x 10^6) (0.650 m - 0.4(0.028 m)

(2)0.028 m

= 50.19 MPa

Factor of safety = σyield

σhoop

= 120.658 Mpa

105.25 MPa

= 1.14

4.1.6 2:1 Ellipsoidal Head thickness

t = PD per UG 27 (d)

(2SE-0.2P)

= (44 x 10^5) (1303 mm)

2 (1206.58 x 10^5) (1) - 0.2 (44 x 10^5)

= 23.85 mm

23.85 mm + corrosion allowance, 3 mm = 26.85 mm

So, use t = 28 mm

h = D

4

= 1300

4

= 325 mm

Maximum Allowance Working Pressure, MAWP

P = 2SEt per UG 27 (d)

D + 0.2t

= 2 (1206.58 x 10^5) (1) (23.85 mm)

1303 mm + 0.2 (23.85 mm)

= 44 Bar

Stress, σ = P (D + 0.2t)

2 t

= (4.4 x 10^6) (1.303 m + 0.2(0.024 m)

2(1)(0.024 m)

= 119.88 MPa

4.1.7 2:1 Nozzle and Flanges

4.1.7.1 (Inlet and Outlet)

t = PR per UG 45

(SE- 0.6P)

= (44 x 10^5) (152.4 mm)

(1206.58 x 10^5) - 0.6(44 x 10^5)

= 5.68 mm ~ 6 mm

6 mm + corrosion allowance, 3 mm = 9 mm

So, use t = 9 mm

Length of pipe 12" = 211.85 mm

Flanges

Based on slip-on Flanges - ANSI B16.5 300lbs

Table 4.4: Slip-On Flanges - ANSI B16.5 300lbs for 12 Inch

Nominal pipe size

Outside diameter

Overal diameter

Inside diameter

Flanges thickness

Overall length

Hub diameter

Face diameter

No.of holes

Bolt hole

Diameter of circle of holes

12"

323.8

520.7

327.1

50.80

73.15

374.6

381.0

16

31.70

450.8

4.1.7.2 (Manhole)

t = PR per UG 45

(SE- 0.6P)

= (44 x 10^5) (254 mm)

(1206.58 x 10^5) - 0.6(44 x 10^5)

= 9.47 mm ~ 9.5 mm

9.5 mm + corrosion allowance, 3 mm = 12.5 mm

So, use t = 12.5 mm

Length of pipe 20" = 252 mm

Flanges

Based on slip-on Flanges - ANSI B16.5 300lbs

Nominal pipe size

Outside diameter

Overal diameter

Inside diameter

Flanges thickness

Overall length

Hub diameter

Face diameter

No.of holes

Bolt hole

Diameter of circle of holes

20"

508

774.7

513.1

63.50

95.20

587.2

584.2

24

35

685.8

Table 4.5: Slip-On Flanges - ANSI B16.5 300lbs for 20 Inch

4.1.7.3 (Liquide Outlet)

t = PR per UG 45

(SE- 0.6P)

= (44 x 10^5) (25.4 mm)

(1206.58 x 10^5) - 0.6(44 x 10^5)

= 0.95 mm ~ 1 mm

1 mm + corrosion allowance, 3 mm = 4 mm

So, use t = 4 mm

Length of pipe 20" = 271.8 mm

Flanges

Based on slip-on Flanges - ANSI B16.5 300lbs

Nominal pipe size

Outside diameter

Overal diameter

Inside diameter

Flanges thickness

Overall length

Hub diameter

Face diameter

No.of holes

Bolt hole

Diameter of circle of holes

2"

60.3

165.1

62

22.30

33.20

84

91.90

8

19.10

127.0

Table 4.6: Slip-On Flanges - ANSI B16.5 300lbs for 2 Inch

4.1.8 Leg support

For designing leg support, there are no specific rules or codes that describes in ASME Code Section VIII Div 1. So, in this project, the leg supports was designed based on available support that be designed for knock out drum by Petronas Fertilizer Sdn. Bhd.

4.2 Details drawing by Catia

Figure 4.1: Unfired Vertical Pressure Vessel

[Please refer Appendix 1]

Figure 4.2: Shell

[Please refer Appendix 2]

Figure 4.3: Top Ellipsoidal Head

Figure 4.4: Bottom Ellipsoidal Head

[Please refer Appendix 3 & 4]

Figure 4.5: Leg Support

[Please refer Appendix 5]

4.3 Static Structural Analysis Result and Discussion

From the finite element analysis for all load cases by static structural analysis using ANSYS software, there are result are obtained.

4.3.1 Static Structural Analysis of Shell with Nozzles

C:\Users\zalie87\AppData\Roaming\Ansys\v140\preview.png

Figure 4.6a: Total Deformation of Shell with Nozzles

The figure above shown the total deformation of the shell with nozzle attached. From the results of analysis, it was observed that the maximum deformation occurred at the junction of pressure vessel's shell and the nozzle. The maximum deformation was 0.52119 mm.

C:\Users\zalie87\AppData\Roaming\Ansys\v140\preview.png

Figure 4.6b: Equivalent (von-Mises) Stress

Based on figure above, the maximum stresses occurred at the nozzle neck. The maximum stress value obtained is 141.28 MPa. The maximum equivalent stress obtained from the analysis was large than maximum allowable stress because of attached nozzle neck due to sudden change in the shell geometry and the resulting of change in stress flow.

C:\Users\zalie87\AppData\Roaming\Ansys\v140\preview.png

Figure 4.6c: Maximum Shear Stress

The figure above represented the maximum shear stress that occurs on the shell. There are colors that represent the level of stress that occur on the shell surface. The blue color indicate the area which the stress was lowest and the red color indicated the maximum stress occur while the pressure has been applied.

C:\Users\zalie87\AppData\Roaming\Ansys\v140\preview.png

Figure 4.6d: Safety Factor

From the analysis of shell with nozzles attached, the minimum value of factor safety obtained is 0.85406. Because of some lack, the value of safety factor is quite low compared to theoretical value. It is because the maximum equivalent stress that been obtained was large than maximum allowable stress.

4.3.2 Static Structural Analysis of Shell without Nozzles

C:\Users\zalie87\AppData\Roaming\Ansys\v140\preview.png

Figure 4.7a: Total Deformation of Shell without Nozzles

For the analysis of shell without nozzle attached, the result has been shown above. Compared with the previous analysis on shell with the nozzles, the value of maximum deformation is less which is only 0.33246 mm.

C:\Users\zalie87\AppData\Roaming\Ansys\v140\preview.pngFigure 4.7b: Equivalent (von-Misses) Stress

The figure shown above is the result of equivalent (von-Misses) stress that occurs on the shell surface at about design pressure of 4.4 MPa. The red color represents the maximum stress which is 116.67 MPa. The maximum stress occurs at the bottom of the shell. The maximum allowable stress for this shell is 120.658 MPa. So, the value obtained in this analysis was below than maximum allowable stress. It can be said that this shell was safe.

C:\Users\zalie87\AppData\Roaming\Ansys\v140\preview.png

Figure 4.7c: Maximum Shear Stress

Based on figure above, the maximum stresses occur on the surface inside shell. The minimum shear stress occurs on the top shell surface 16.272 MPa and the maximum shear stress value obtained is 61.08 MPa which is represented with red color.

C:\Users\zalie87\AppData\Roaming\Ansys\v140\preview.png

Figure 4.7d: Safety Factor

From the analysis of shell without nozzles attached, the minimum value of factor safety obtained by calculation is 1.14. The value of safety that obtained by this analysis is 1.03 because the maximum equivalent stress that be obtained is less than hoop stress in manual calculation. So the percentage of factor safety between value from calculation and analysis is about 9.6 % and it's acceptable.

CHAPTER 5.0

SUMMARY

5.1 Conclusion

As the project is completed, it can be concluded that the objectives of this project are successfully done. This project had lead to several conclusions. However, major conclusions are as below:

From overall study of ASME Code Section VII Division 1 in designing vertical pressure vessel, it be said that the main requirement that used to design this type of vessel was be studied properly. Because of some lack of information is ASME Code such as rules for designing leg support, the component had been designed just based on available designed that widely used in industry. This project only focused on design requirements in ASME Code, so the rule for fabrication and inspection did not be involved.

From the analysis of shell with attached nozzle, the maximum deformation of the shell has been obtained. The deformation value was below the allowable deforming for the shell material. Then, the maximum equivalent stress (von-Misses) also has been obtained over the maximum allowable stress. This was because of the geometry of the shell has been changed during nozzle attached. This problem occurred also because of the change is stress flow during the loads has been applied. Because of over maximum equivalent stress obtained, the value safety factor had been affected. The value of safety factor obtained was less than 1; it might be not good enough but it still can be considered.

From the analysis of shell without nozzle, the maximum deformation is less than deformation in shell with attached nozzle analysis. Then, the value of maximum equivalent stress (von-Misses) obtained was less than maximum allowable stress. The value was approximated to the calculation value at about 3.3 %. So, the value of safety factor obtained also close to the calculation value in term of maximum allowable stress per maximum equivalent stress. Hence, the shell was in safe condition when the operating pressure been applied.

However, although the code for design a pressure vessel had been studied properly, some of information was not described in details. So, this design was not too safe and good enough for fabrication. Many requirements still had to be considered to make this design perfect. There were many codes and rules should be studied and understood properly. By the way, as been stated earlier, this project has achieved the objectives and fulfills the requirement of Final Year Project II.

5.2 Recommendation

Apparently, in term of design rules, there are many aspects to looking further improvement to have a complete and perfects vertical pressure vessel. The design codes and standards must be appropriately revised to make sure the design is safe enough.

Because of lack of information from the ASME Codes Section VIII Division 1 in designing this pressure vessel, some of the criteria required cannot be applied. Some of the information in ASME Code is confidential and need to ask for their permission before used it. Sometimes engineer, designer or organization needs to buy their codes and standards which are very expensive.

There are others codes and standard in designing pressure vessel available. There also has software to design pressure in the market. Maybe by using others codes and standard or software may improve the procedure in designing pressure vessel


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