Types Of Pressure Vessel Engineering Essay
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.
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.
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 containers for the containment of internal and external pressure up to 3000 psi. This pressure may 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 mandatory Requirements, specific prohibitions; and non-mandatory guidance for pressure vessel materials, design, welding and testing. To ensure the objective is achieved, some of the important elements must be consideration. 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.
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 stress in shell by finite element using ANSYS software.
Significant 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.
Table : Gantt chart for final year project I and II
CHAPTER 2.0: LITERATURE REVIEW
The pressure vessels such as cylinder, pipeline or tanks are used to store fluids under pressure. The fluid being stored may undergo a change of state inside the pressure vessel as in case of steam boilers or it may combine with other reagents as in a chemical plant. The pressure vessels are designed with great care because rupture of pressure vessels means an explosion which may cause loss of life and property. The material of pressure vessels may be brittle such that cast iron or ductile such as mild steel. Pressure vessels, hydraulic cylinders, gun barrels, pipes, boilers and tanks are, in fact, essential to the chemical, petroleum, petrochemical and nuclear industries. It is in this class of equipment that the reactions, separations, and storage of raw materials occur. Generally, pressurized equipment is required for a wide range of industrial plant for storage and manufacturing purposes.
Types of Pressure Vessel
The size and geometric form of pressure vessels vary greatly from the large cylindrical vessels used for high-pressure gas storage to the small size used as hydraulic units for aircraft. Some 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:
Spherical Pressure Vessel
These pressure vessels are 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. It can also be called as simplified 2D problems. 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:Userszalie87Desktopspherical pressure vessel 2.jpg
Figure 2..1: Spherical Pressure Vessel
Cylindrical Pressure Vessel
This a vessel with a fixed radius and thickness subjected to an internal gage pressure. The 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.
Figure .1.2: Cylindrical (Horizontal & Vertical) Pressure Vessel
Main Components of Pressure Vessel
The main pressure vessel components are as follow:
The shell is the primary component that contains the pressure. Pressure vessel shells are welded together to form a structure that has a common rotational axis. Most pressure vessel shells are cylindrical, spherical, or conical in shape.
All pressure vessel shells must be closed at the ends by heads (or another shell section). Heads are typically curved rather than flat. Curved configurations are stronger and allow the heads to be thinner, lighter, and less expensive than flat heads. Heads can also be used inside a vessel.
Heads are usually categorized by their shapes. Ellipsoidal, hemispherical, torispherical, conical, toriconical and flat are the common types of heads. Figure 2.2.2 shows various types of heads. Ellipsoidal would be the most common type of heads, which is used during the designing of pressure vessels.
Figure 2.2.2: Typical Types of Heads
A nozzle is a cylindrical component that penetrates the shell or heads of a pressure vessel. The nozzle ends are usually flanged to allow for the necessary connections and to permit easy disassembly for maintenance or access. Nozzles are used for the following applications:
Attach piping for flow into or out of the vessel.
Attach instrument connections, (e.g., level gauges, thermowells, or pressure gauges).
Provide access to the vessel interior at manways.
Provide for direct attachment of other equipment items, (e.g., a heat exchanger or mixer).
Nozzles are also sometimes extended into the vessel interior for some applications, such as for inlet flow distribution or to permit the entry of thermowells.
The type of support that is used depends primarily on the size and orientation of the pressure vessel. In all cases, the pressure vessel support must be adequate for the applied weight, wind, and earthquake loads. The design pressure of the vessel is not a consideration in the design of the support since the support is not pressurized. Temperature may be a consideration in support design from the standpoint of material selection and provision for differential thermal expansion.
Typical kinds of supports are as follow:
Tall, vertical, cylindrical pressure vessels are typically supported by skirts. A support skirt is a cylindrical shell section that is welded either to the lower portion of the vessel shell or to the bottom head (for cylindrical vessels). Skirts for spherical vessels are welded to the vessel near the mid-plane of the shell. The skirt is normally long enough to provide enough flexibility so that radial thermal expansion of the shell does not cause high thermal stresses at its junction with the skirt.
Small vertical drums are typically supported on legs that are welded to the lower portion of the shell. The maximum ratio of support leg length to drum diameter is typically 2:1. The number of legs needed depends on the drum size and the loads to be carried. Support legs are also typically used for spherical pressurized storage vessels. The support legs for small vertical drums and spherical pressurized storage vessels may be made from structural steel columns or pipe sections, whichever provides a more efficient design. Cross bracing between the legs is typically used to help absorb wind or earthquake loads.
Horizontal drums are typically supported at two locations by saddle supports. A saddle support spreads the weight load over a large area of the shell to prevent an excessive local stress in the shell at the support points. The width of the saddle, among other design details, is determined by the specific size and design conditions of the pressure vessel. One saddle support is normally fixed or anchored to its foundation. The other support is normally free to permit unrestrained longitudinal thermal expansion of the drum. A typical scheme of saddle support is shown on Figure 2.2.4.
Figure 2.2.4: Typical Scheme of Saddle
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 2.3. 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 design of pressure vessels safety is the primary consideration, especially for nuclear reactor pressure vessels, due the potential impact of a possible severe accident. In general however, the design is a compromise between consideration of economics and safety. The possible 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 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 2.3: Design Procedure
CHAPTER 3.0: METHODOLOGY
In this chapter, the selection of pressure vessel is described and the application of selected pressure vessel is been explain. To design of pressure vessel the selection of Code are important as a reference guide to achieve the safety pressure vessel. The selections of ASME CODE Section VIII div 1 are described. The standard of material use 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.
General Design Considerations: Pressure Vessels
Paragraphs UG-4 through UG-15 defines general material requirements. There are several points related to the general material requirements that will be highlighted here.
The main factors that influence material selection are:
Strength is a material’s ability to withstand an imposed force or stress. Strength is a significant factor in the material selection for a particular application.
Strength determines how thick a component must be to withstand the imposed loads.
Corrosion is the deterioration of metals by chemical action. A material’s resistance to corrosion is probably the most important factor that influences its selection for a specific application.
The most common method that is used to address corrosion in pressure vessels is to specify a corrosion allowance. A corrosion allowance is supplemental metal thickness that is added to the minimum thickness that is required to resist the applied loads.
Resistance to Hydrogen Attack
If this hydrogen diffusion continues, pressure can build to high levels within the steel, and the steel can crack.
At elevated temperatures, over approximately 600°F (315,5C), monatomic hydrogen not only causes cracks to form but also attacks the steel. Hydrogen attack differs from corrosion in that damage occurs throughout the thickness of the component, rather than just at its surface, and occurs without any metal loss.
In addition, once hydrogen attack has occurred, the metal cannot be repaired and must be replaced.
Instead, materials are selected such that they are resistant to hydrogen attack at the specified design conditions.
Fracture toughness refers to the ability of a material to withstand conditions that could cause a brittle fracture. The fracture toughness of a material can be determined by the magnitude of the impact energy that is required to fracture a specimen using Charpy V-notch test.
Generally, the fracture toughness of a material decreases as the temperature decreases. The fracture toughness at a given temperature varies with different steels and with different manufacturing and fabrication processes.
Fabricability refers to the ease of construction and to any special fabrication practices that are required to use the material.
Pressure vessels commonly use welded construction. The materials used must be weldable so that individual components can be assembled into the completed vessel.
The pressure vessel design codes and standards include lists of acceptable materials; in accordance with the appropriate material standards.
Design and operating temperature
Paragraph UG-20 provides requirements for establishing the maximum and minimum design temperatures. The maximum design temperature is given as the maximum temperature used in design shall not be less than the mean metal temperature (through the thickness) expected under normal operating conditions for the part considered.
It is common practice to set the maximum design temperature equal to the maximum anticipated fluid temperature with an appropriate allowance; however, the above definition does allow a much less conservative approach to be used. If a pressure vessel is insulated, then the mean metal temperature and process fluid temperature are essentially the same at steady state conditions. However, for non-insulated vessels, the mean metal temperature through the thickness can be considerably different than the process fluid temperature. This can be used to advantage for those applications where the mean metal temperature is significantly cooler than the process fluid.
The operating temperature is fluid temperature that occurs under normal operating conditions. The operating temperature must be set based on the maximum and minimum metal temperatures that the pressure vessel may encounter.
Design and operating pressure
Paragraph UG-21 defines the considerations for establishing the design pressure of a vessel. This paragraph requires the vessel to be designed for at the most severe pressure and temperature that is coincidentally expected in normal operation. This means those conditions such as start-up, shutdown, and any identified upset conditions must be considered when establishing the maximum operating pressure. In an operating system, the set pressure of the pressure relief device must be above the operating pressure by a sufficient amount so that the device does not actuate unintentionally. A vessel must be designed to withstand the maximum pressure to which it is likely to be subjected in operation. Operating pressure is the pressure to be used in operating condition. The operating pressure must be set based on the maximum internal or external pressure that the pressure vessel may encounter.
For vessels under internal pressure, the design pressure is normally taken as the pressure at which the relief device is set. This will normally be 5 to 10 per cent above the normal working pressure, to avoid spurious operation during minor process upsets. When deciding the design pressure, the hydrostatic pressure in the base of the column should be added to the operating pressure, if significant. Vessels subject to external pressure should be designed to resist the maximum differential pressure that is likely to occur in service. Vessels likely to be subjected to vacuum should be designed for a full negative pressure of 1 bar, unless fitted with an effective, and reliable, vacuum breaker.
Design maximum allowable stress (nominal design strength)
Paragraph UG-23 defines the maximum allowable stress for internal and external pressure to be used in the design of Division 1 pressure vessels. The allowable tensile stresses are tabulated in Section II, Part D, of the Boiler & Pressure Vessel Code. There are several points regarding the allowable stress to be used in design that merit discussion. UG-23(a) states that for material identified as meeting more than one material specification or grade, the allowable stress for either specification or grade may be used, provided that all the limitations of the specification is satisfied. UG-23 also provides criteria for the maximum allowable longitudinal compressive stress to be used for cylindrical shells or tubes that are subjected to longitudinal compressive loads. The first criterion is that the maximum allowable longitudinal compressive stress cannot be greater than the maximum allowable tensile stress. The second criterion is based on buckling of the component. Paragraph UG-23(c) states, the wall thickness of a vessel computed by these rules shall be determined such that, for any combination of loadings listed in UG-22 that induce primary stress and are expected to occur simultaneously during normal operation of the vessel, the induced maximum general primary membrane stress does not exceed the maximum allowable stress value in tension.
Typical design stress factors for pressure components are shown in Table 2.
Carbon-manganese, stainless metals
low alloy steels
Austenitic stainless steels
stress or 0.2 per
cent proof stress,
at the design
strength, at room
Mean stress to
at 105 h at the
Table : Design stress factors
Thickness of shell under internal pressure
Information of the minimum required thickness or maximum allowable pressure for a shell under internal pressure are provided in paragraph UG-27. For cylindrical shells, the applicable equations for circumferential stress (the stress acting across the longitudinal seam) are as follows :
Figure 3.2.5: Shell under Internal Pressure
For cylindrical shells for longitudinal stress (the stress acting across the circumferential joints), the applicable equations are
t = minimum required thickness of shell, in. (in the corroded condition)
P = internal design pressure, psi
R = inside radius of shell under consideration, in. (in the corroded condition)
S = maximum allowable stress from the applicable allowable stress table in Section II, Part D
E = the lesser of the joint efficiency for welded joints (as defined in Table UW-12), or the ligament efficiency between openings (when applicable, determined using UG-53).
For spherical shells,
These equations are very straightforward and do not require much discussion. However, there are some pertinent issues that are discussed. These equations are based on thin wall theory, but do include provisions to account for the variation of stress through the wall of the vessel (called the Lame’ effect) that becomes significant for very thick cylinders.
Thickness of shell under external pressure
The method used to design shells and tubes when external pressure is specified as a design load is given in paragraph UG-28. The definitions for various geometries are pictorially shown in Figure 3.2.6a (Fig. UG-28.1). Figure 3.2.6a (a-1) shows a cylindrical vessel with heads having no stiffening rings. Notes 1 and 3 of Figure 3.2.6a provide important requirements for these instances. If the cone-to-cylinder (with or without a knuckle) does not have sufficient moment of inertia, then it may not be considered as a line of support. If the cone-to-cylinder junction does not have sufficient moment of inertia to be considered as a line of support, then the unstiffened length (L) must be taken between lines of support on either side of the joint as shown in Figure 3.2.6a (a-2 and c-2) .
Figure 3.2.6a: 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 2007 Code)
Figure 3.2.6b: 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 Code)
When stiffening rings are used to resist external pressure, the provided stiffness has to be continuous around the circumference of the vessel. Gaps are allowed between the ring and the shell; however, the ring has to be continuous and the arc of the gap is limited by Figure 3.2.6b. If the arc of the gap between the ring and shell does not meet the Figure 3.2.6b requirements, then the additional requirements of UG-29(c)(1) through UG-29(c)(4) must be satisfied.
In addition to limits on the number and relative orientation of gaps in adjacent rings, a stiffening ring with a gap that exceeds the value allowed by Figure 3.2.6b may not be considered to be a line of support for determining the unstiffened length of the shell. The stiffening ring may be interrupted if the required stiffness is provided by another component such as a support saddle, tray ring, or internal baffles, provided they are located at the ring and supply the necessary moment of inertia. Stiffening rings may be attached by stitch welds (staggered or in-line), or continuous welds on both sides of the ring, or continuously welded on one side and stitch welded on the opposite side of the ring.
The minimum required thickness of a spherical shell under external pressure, either seamless or of built-up construction with butt joints, shall be determined by the following procedure :
Step 1: Assume a value for t and calculate the value of factor A using the following formula:
Step 2: Using the value of A calculated in Step 1, enter the applicable material chart in Subpart 3 of Section II, Part D for the material under consideration. Move vertically to an intersection with the material /temperature line for the design temperature (see UG-20). Interpolation may be made between lines for intermediate temperatures. If tabular values in Subpart 3 of Section II, Part D are used, linear interpolation or any other rational interpolation method may be used to determine a B value that lies between two adjacent tabular values for a specific temperature. Such interpolation may also be used to determine a B value at an intermediate temperature that lies between two sets of tabular values, after first determining B values for each set of tabular values.
In cases where the value at A falls to the right of the end of the material /temperature line, assume an intersection with the horizontal projection of the upper end of the material /temperature line. If tabular values are used, the last (maximum) tabulated value shall be used. For values at A falling to the left of the material /temperature line, see Step 5.
Step 3: From the intersection obtained in Step 2, move horizontally to the right and read the value of factor B.
Step 4: Using the value of B obtained in Step 3, calculate the value of the maximum allowable external working pressure Pa using the following formula:
Step 5: For values of A falling to the left of the applicable material /temperature line, the value of Pa can be calculated using the following formula:
If tabulated values are used, determine B as in Step 2 and apply it to the equation in Step 4.
Step 6: Compare Pa obtained in Steps 4 or 5 with P. If Pa is smaller than P, select a larger value for t and repeat the design procedure until a value for Pa is obtained that is equal to or greater than P.
Rules for the design of formed heads and sections with the pressure on the concave side of the component are given in paragraph UG-32. The required thickness of ellipsoidal (with the major diameter twice the minor diameter) heads is given by
D = diameter of the ellipse major axis
Figure 3.2.7a: Ellipsoidal head
Other terms are as given for the shell design formulae of UG-27.
Ellipsoidal heads that does not have a major to minor diameter has a ratio of 2:1.The required thickness for a torispherical head with the knuckle radius equal to 6% of the inside crown radius and the inside crown radius equal to the outside diameter of the skirt (straight section of the head attached to the adjacent shell) is given by
Where: L = inside crown radius of the formed head
Figure 3.2.7b: Torispherical head
The above equations for a torispherical head represent only one of many possible combinations of knuckle radius and crown radius.
Openings and Reinforcements
When an opening is made in a pressure vessel, there is a stress intensification resulting from the hole that is formed in the shell. This is analogous to the classical stress concentration effect of a hole in a plate that is loaded in traction. The code reinforcement rules do not consider loads other than pressure. Openings in shells should preferably be round, elliptical, or obround. If the connection is oblique to the surface of the shell, the opening in the shell is elliptical. Obround openings are made by connections formed by parallel sides with semicircular ends. Openings of other shapes shall be provided with a suitable radius to minimize stress concentration effects in the shell. When the strength of vessels with such openings cannot be determined with accuracy or when doubt exists regarding its strength, the proof test provisions of paragraph UG-101 shall be applied.
There is no limit to the size of an opening that may be installed in a pressure vessel. The opening reinforcement rules given in UG-36 through UG-43 apply to openings not exceeding the following: for vessels of 60 in. inside diameter and less, the opening may be as large as one half the vessel diameter, but not to exceed 20 in.; for vessels over 60 in. inside diameter, the opening may be as large as one third the vessel diameter, but not to exceed 40 in.
Design for Internal Pressure
The total cross sectional area of reinforcement A required in any given plane through the opening for a shell or formed head under internal pressure shall be not less than
A = dtrF + 2tn trF(1 âˆ’ fr1 ) (22.214.171.124)
Design for External Pressure
(1) The reinforcement required for openings in single- walled vessels subject to external pressure need be only 50% of that required in (c) above, 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) The reinforcement required for openings in each shell of a multiple-walled vessel shall comply with (1) above when the shell is subject to external pressure, and with design for internal pressure above when the shell is subject to internal pressure, regardless of whether or not there is a common nozzle secured to more than one shell by strength welds.
The minimum wall thickness of nozzle necks shall be determined as given below. For access openings and openings used only for inspection:
tUG-45 = ta (126.96.36.199)
For other nozzles:
tb = min [tb3, max (tb1, tb2)] (188.8.131.52)
tUG-45 = max (ta, tb) (184.108.40.206)
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
Paragraph UG-45 provides rules for minimum nozzle neck thickness. 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 manways and other openings that are provided only for access, additional requirements of paragraph UG-45 may require a thicker nozzle neck.
In certain cases, legs can be made detachable to the vessel. These legs can be bolted to plates. The design for leg supports 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:
Where, tW = Weld Height
LW = Weld Length.
These types of supports are suitable only for small vessels as there is a concentrated local stress at the joint.
Figure 3.2.10: Leg support
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.
Type of joint
Degree of radiography
joint with bonding strips
Table : Maximum allowable joint efficiency
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.
Figure 3.2.11: Welded Joint Categories (ASME VIII Div1)
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 design codes and standards specify a minimum allowance of 1.0 mm.
Finite Element Analysis by ANSYS
This project is set out to verify finite element analysis, or FEA, when applied to pressure vessel design. Finite Element Analysis is a simulation technique which evaluates the behavior of components, equipment and structures for various loading conditions including applied forces, pressures and temperatures. Thus, a complex engineering problem with non-standard shape and geometry can be solved using finite element analysis where a closed form solution is not available. The finite element analysis methods result in the stress distribution, displacements and reaction loads at supports for the model. Finite element analysis techniques can be used for a number of scenarios, e.g. design optimization, material weight minimization, shape optimization, code compliance and more. The Finite elements analysis was performed using FEA software ANSYS. ANSYS is a finite element analysis (FEA) code widely used in the computer-aided engineering (CAE) field. ANSYS software allows engineers to construct computer models of structures, machine components or systems; apply operating loads and other design criteria; and 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 vertical unfired pressure vessel. All result from the analysis using ANSYS will be compared to the calculation.
Figure 3.3: Example of ANSYS analysis; Maximum shear stress of Elliptical Head
CHAPTER 4.0: RESULT AND ANALYSIS
4.1 Design Data and Calculation
: ASME Section VIII Division 1
Type of vessel
: 1300.0 mm
: 70.0 °C
: 30.0 °C
: 44 BarG
: 24.9 BarG
: 3 mm
Type of fluid
: Natural gas
Max. Liquid level
: Not applicable
Type of head
: 2:1 Ellipsoidal
: 4791 kg
: 4850 kg (approximate)
Table : Pressure Vessel Design Data
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.
220.127.116.11 Properties of Material
SA-516 Gr 70
Table : Properties of Material
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.
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.
4.1.4 Maximum Allowable Stress Value
Refer to ASME code in paragraph UG 23, the maximum allowable stress value is the maximum unit 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.
Metal temperature not exceeding deg, F
Maximum Allowable Stress, psi
SA-516 Gr 70
-20 to 650
Table : Maximum Allowable Stress Value
4.1.5 Thickness of Shells under Internal Pressure
t = PR per UG 27 (c)
= (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, Ïƒhoop = P (R + 0.6t)
= (4.4 x 10^6) (0.650m + 0.6(0.028 m)
= 104.78 MPa
Stress, Ïƒlong = P (R – 0.4t)
= (4.4 x 10^6) (0.650 m – 0.4(0.028 m)
= 50.19 MPa
Factor of safety = Ïƒhoop
= 104.78 MPa
4.1.6 2:1 Ellipsoidal Head thickness
t = PD per UG 27 (d)
= (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
= 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, Ïƒ = P (D + 0.2t)
= (4.4 x 10^6) (1.303 m + 0.2(0.024 m)
= 119.88 MPa
4.1.7 2:1 Nozzle and Flanges
18.104.22.168 (Inlet and Outlet)
t = PR per UG 45
= (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
Based on slip-on Flanges – ANSI B16.5 300lbs
Nominal pipe size
Diameter of circle of holes
Table : Slip-On Flanges – ANSI B16.5 300lbs for 12 Inch
t = PR per UG 45
= (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
Based on slip-on Flanges – ANSI B16.5 300lbs
Nominal pipe size
Diameter of circle of holes
Table : Slip-On Flanges – ANSI B16.5 300lbs for 20 Inch
22.214.171.124 (Liquide Outlet)
t = PR per UG 45
= (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
Based on slip-on Flanges – ANSI B16.5 300lbs
Nominal pipe size
Diameter of circle of holes
Table : 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
Refer to appendices
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
Figure 4.3.1a: Total Deformation of Shell with Nozzles
The figure above show the total deformation of the shell with nozzle attached. From the results of analysis, it can be observed that the maximum deformation occurs at the junction of pressure vessel’s shell and the nozzle. High stress concentration is developed at this location due to abrupt change in the geometry and the consequent change in stress flow. The maximum deformation is 1.2754 mm.
Figure 4.3.1b: Equivalent (von-Mises) Stress
Based on figure above, the maximum stresses occur at the nozzle neck. The minimum stress occurs on the shell surface is 92.662 KPa and the maximum stress value obtained is 343.04 MPa.
Figure 4.3.1c: Maximum Shear Stress
The figure above represents 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 is lowest and the red color indicate the maximum stress occur while applying the pressure.
Figure 4.3.1d: Safety Factor
From the analysis of shell with nozzles attached, the minimum value of factor safety obtained is 1.2883. Because of some lack, the value of safety factor is quite low but still satisfied since it not under the value of 1.
4.3.2 Static Structural Analysis of Shell without Nozzles
Figure 4.3.2a: 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 0.33066 mm.
C:Userszalie87AppDataRoamingAnsysv140preview.pngFigure 4.3.2b: 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 127.45 MPa. The maximum stress occurs at the bottom of the shell.
Figure 4.3.2c: Maximum Shear Stress
Based on figure above, the maximum stresses occur at the bottom of shell. The minimum stress occurs on the top shell surface is 9.5236 MPa and the maximum stress value obtained is 66.101 MPa which is represented with red color.
Figure 4.3.2d: Safety Factor
From the analysis of shell without nozzles attached, the minimum value of factor safety obtained is 2.04. The value of safety factor is approximately close to the value from the calculation which is 2.09. Compared this two values has prove the calculation and analysis is tally.
CHAPTER 4.0 CONCLUSION
As the project is completed, it can be concluded that the objectives of this project are successfully done. This project has lead to several conclusions. However, major conclusions are as below:
The design of pressure vessel is initialized with the specification requirements in terms of standard technical specifications along with numerous requirements that lay hidden from the market. The design of a pressure vessel is more of a selection procedure, selection of its components to be more precise rather designing each and every component. Regarding storage of fluid for a pressure vessel system should be preferred due to its simplicity, better sensitivity, higher reliability, low maintenance, compactness for the same capacity. The storage of fluid at high pressure in the pressure vessel is at the heart of its performance and is the first step towards the design.
The pressure vessel components are merely selected, but the selection is very critical, a slight change in selection will lead to a different pressure vessel altogether from what is aimed to be designed. It is observed that all the pressure vessel components are selected on basis of available ASME standards and the manufactures also follow the ASME standards while manufacturing the components. So that leaves the designer free from designing the components. This aspect of design greatly reduces the development time for a new pressure vessel. It also allows the designer the freedom to play with multiple prototypes for the pressure vessel before finalizing the decision. The pressure vessel selection procedure after determining the inputs is a simplified process and can be automated to shorten the design cycle.
For pressure vessels, finite element analysis provides an additional tool for use in analysis. However, it must be compared to other available data, not taken as being correct just because it looks right. Used with this understanding, finite element analysis offers great insight into the complex interactions found in pressure vessel design. With the help of finite element analysis, we can study the actual maximum stress distributions in the different components of pressure vessel and the actual behavior of pressure vessel.