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.
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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:Userszalie87Desktopspherical 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:Userszalie87Desktoppressure-vessel-500×500.jpg C:Userszalie87Desktopvertical_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]
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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
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