Training Manual for Pressure Vessel Designers

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Chapter 1: Regulations and Standards 1--1: What are the main regulations and codes that must be followed in pressure vessel design? Answer: 1. Regulations on the Safety Supervision of Special Equipment, issued by the State Council on June 1, 2003. 2. Technical Regulations for the Safety Supervision of Pressure Vessels, issued by the Quality Inspection Bureau on January 1, 2000. 3. Rules for the Licensing and Management of Design Units for Pressure Vessels and Pressure Pipelines, issued by the Quality Inspection Bureau on January 1, 2003. 4. Measures for the Supervision and Management of the Manufacturing of Boilers and Pressure Vessels, issued by the Quality Inspection Bureau on January 1, 2003. 5. GB150 – Steel Pressure Vessels. 6. JB4732 – Steel Pressure Vessels – Analysis and Design Standards. 7. JB/T4735 – Welded Atmospheric Pressure Steel Vessels. 8. GB151 – Shell and Tube Heat Exchangers. 1–2 What are the responsibilities of the pressure vessel design unit? Answer: 1. Be responsible for the accuracy and completeness of the design documents. 2. The design documents for a container shall include at least the design calculations and design drawings. 3. The general layout diagram of the container design shall bear the approval mark from the pressure vessel design unit. What are the scope of application and non-application of GB150-1998 \"Steel Pressure Vessels\" for 1–3 GB? Answer: Scope of application: 1. Steel pressure vessels with a design pressure not exceeding 35 Mpa. 2. The design temperature range is determined based on the allowable operating temperature of the steel. Areas where it does not apply: 1. Containers heated by direct flame. 2. Vessels in nuclear power installations. 3. Containers that are frequently moved. 4. Pressurized vessels that are integral to or serve as components in rotary or reciprocating mechanical devices such as pumps, compressors, turbines, or hydraulic cylinders. 5. Design containers with a design pressure of less than 0.1 Mpa. 6. Containers with a vacuum level below 0.02 Mpa. 7. Containers with an inner diameter of less than 150 mm. 8. Containers that require fatigue analysis. 9. Pressure vessels that are already governed by other industry standards include certain specialized pressure vessels and glass-lined vessels used in industries such as refrigeration, sugar production, papermaking, and beverages. 1–4 What are the scope of application and exclusions of the «Regulations on Safety Supervision of Pressure Vessels»? Answer: Scope of use: (All of the following conditions must be met) 1. Containers with a maximum workpiece pressure (PW) of 0.1 Mpa or higher (excluding liquid pressure). 2. Containers with an inner diameter (for non-circular cross-sections, this refers to the largest dimension of the cross-section) greater than 0.15 m, and a volume V of 0.25 m3 or more ； 3. A vessel whose filling medium is a gas, liquefied gas, or a liquid with a maximum operating temperature equal to or higher than its standard boiling point. Scope of non-applicability: 1. Ultra-high pressure vessels. 2. Various types of gas cylinders. 3. Pressure vessels manufactured from non-metallic materials. 4. Nuclear pressure vessels, auxiliary pressure vessels on ships and railway locomotives, pressure vessels for defense or **equipment, and devices directly heated by flames within the scope of boiler safety supervision (such as flue-type waste heat boilers, etc.). 5. Pressure vessels that operate under normal conditions with a maximum working pressure of less than 0.1 Mpa (including those that need to withstand a pressure of 0.1 Mpa or higher temporarily during feeding or discharging processes, but excluding those that require temporary exposure to a pressure of 0.1 Mpa or higher during processes such as disinfection or cooling). 6. Non-independent pressure-bearing components on the machine (such as compressors, generators, pumps, and the pressure vessels or cylinders of diesel engines, but excluding the drying cylinders in papermaking and textile machinery as well as auxiliary pressure vessels for compressors). 7. Shellless coil exchangers, bellows exchangers, air-cooled exchangers, cooling tubes. 1--5 How do the \"Code for Pressure Vessels\" and GB150-98 define the scope of pressure vessels? Answer: 1. For pressure vessels connected to external pipes and devices: the groove end face of the first circumferential joint where the vessel nozzle is connected to the external pipe ； The first flange sealing surface of the flanged connection ； The end face of the first threaded joint in a threaded connection ； The first sealing surface of a dedicated connection or fitting. 2. Pressure-bearing end caps, flat covers, and their fastening elements for the open sections of pressure vessels. 3. Welded joints of uncompressible elements and compressible elements. 1–6 What are flammable media? Answer: Flammable media refer to gases that, when mixed with air, have an explosion lower limit of less than 10%, or a difference between their explosion upper and lower limits of 20% or more. Such as methylamine, ethane, methane, etc. How is the toxicity level of media 1–7 classified? Answer: Referring to HG’s \"Classification of Toxicity Hazards and Explosion Risks of Chemical Media in Pressure Vessels\": 1. Extremely hazardous: What is the calculation method for 60? When the semi-apex angle a of the conical shell is greater than 60, the thickness of the conical head can be calculated as that of a flat cover. It can also be determined using stress analysis methods (including finite element analysis). 3—34 How is the thickness of the conical shell of a conical head calculated? What does Dc mean? The formula for calculating the thickness of the conical shell is as follows: Here, Dc denotes that GB150-98 permits the conical shell to consist of several sections with different thicknesses at each half-apex angle; the diameter of the conical shell changes from section to section, and therefore the thickness of each section also changes accordingly. The thickness of each conical shell section is calculated using this formula, where Dc represents the inner diameter of the larger end of each conical shell section. 3—35 When reinforcement is required for the larger and smaller ends of a flangeless conical head under internal pressure, what are the requirements for the reinforcement section? How is the thickness of the reinforcement section calculated? What are the limitations? If an increase in thickness is required for reinforcement, the conical reinforcement section and the cylindrical reinforcement section should have the same thickness. The thickness of the thick-end reinforcement section of a flangeless conical head under internal pressure is calculated using the following formula: The value of Q is obtained from Figure 7-12 in GB150-98. The thickness of the reinforced section at the smaller end of a flangeless conical head under internal pressure is calculated using the following formula: The value of Q is obtained from Figure 7-14 in GB150-98. Restriction on the reinforcement section: Under any circumstances, the thickness of the reinforcement section shall not be less than the calculated thickness of the conical shell at the joint. The length of the conical shell reinforcement section, for the larger end, shall be not less than ； For little-endian, it should be no less than. The length of the cylindrical reinforcement section, for the larger end, shall be not less than ； For little-endian, it should be no less than. 3—36 How is the thickness of the large end of a flanged conical head subjected to internal pressure determined in GB150-98? Regarding the determination of the thickness at the larger end of a conical head formed by internal pressure folding, GB150-98 specifies that the thicker of the two values – the thickness of the transition section at the point where it connects to the conical shell and the thickness of the conical shell itself – should be adopted. The transition section is considered as the transition zone of a disc head, and therefore its thickness is calculated based on that of a disc head. The conical shell connected to it has a diameter here that is smaller than the diameter of the cylinder before the transition zone; this diameter can be calculated based on the radius of the transition zone and the half-angle of the conical shell, and then determined using the formula for the thickness of the conical shell. The formula for calculating the thickness of the overfill section at the larger end of the folded conical head in GB150-98: δ=, and the formula for calculating the thickness of the conical shell at the junction with the transition zone: δ=, are both derived based on this mechanism. The f-values of the coefficient K are obtained from Tables 7–4 and 7–5 in GB150–98, respectively. 3—37 What is the formula for calculating the thickness of a circular flat cover? How was it derived? The formula for calculating the thickness of a circular flat cover is derived on the assumption that a thin circular plate is subjected to a uniformly distributed load, with its edges simply supported or rigidly fixed. Its calculation formula is: 3—38. What formula should be used to check the bending stress in the longitudinal section caused by a constricted-end head? The bending stress acting on the longitudinal section is such that this bending stress must not exceed 0.8 times the allowable stress for the steel used in necked end closures, that is, σm ≤ 0.8t. This is the formula provided in GB150-98 for checking the bending stress acting on the longitudinal section of necked end closures. 3—Under what conditions does GB150-98 permit openings to be made in the walls of pressure vessels without the need for additional reinforcement? Allowance for not requiring additional reinforcement is given under the following conditions: 1. The distance between the centers of two adjacent openings (for curved surfaces, the distance is calculated in terms of arc length) should be no less than twice the sum of the diameters of the two openings. 2. The nominal outer diameter to be taken over is less than or equal to 89 mm. 3. For specifications where the outer diameter and minimum wall thickness of the take over are not reinforced, the options available are: φ25×3.5, φ32×3.5, φ38×3.5, φ45×4, φ57×5, φ65×5, φ76×6, φ89×6 mm. (The corrosion allowance for takeover is 1 mm.) The above provisions apply to vessels with a design pressure of P ≤ 2.5 Mpa. 3—40 How many types are there for reinforcing openings in pressure vessels? What regulations should be followed when using a reinforcing ring structure for reinforcement? Strengthening of openings in pressure vessels can be roughly classified into the following methods based on design approaches: 1. Equal-area strengthening method. 2. Extreme reinforcement method. 3. Stability analysis. 4. Other methods, such as the experimental stress analysis method, and the use of incremental plasticity theory to study container openings and reinforcement, etc. Based on the reinforcement structure, its basic structures can be roughly divided into two categories: 1. Reinforcement ring lap-welded structure. 2. Overall reinforcement structure. When reinforcement is provided using a reinforcing ring structure, the following requirements shall be followed: 1. The standard tensile strength of the steel material used at normal temperature shall be σb ≤ 540 Mpa. 2. The nominal thickness of the shell δn shall be ≤ 38 mm. 3. The thickness of the reinforcement ring shall be less than or equal to 1.5δn. 3—41 Why is it necessary to limit the d/D ratio and the ratio of the major to minor axes of the oblong hole when applying equal-area reinforcement? Holes not only weaken the strength of the container walls but also cause high stress concentrations in the local area around them. Higher local stresses, combined with the stresses generated by various loads on the nozzles, thermal stresses, as well as factors such as the material properties of the container and manufacturing defects, often become the sources of failure for the container. Therefore, for the reinforcement of openings, the stress analysis of the openings should be studied first. Its basic method is to analyze by making small holes in a large plate from the perspective of elasticity theory. I. Small circular holes in a large plate: 1. Stress concentration factor for unidirectional tension: K=3 2. Stress concentration factor for bidirectional tension: K=2.5 II. For holes in a large plate, the stress concentration factor at the edges of elliptical holes can be higher than that of circular holes. Especially when the long axis is perpendicular to the direction of the principal stress, the larger a/b is, the greater the stress concentration factor becomes. III. Small circular holes are made in a cylinder; when the cylinder is flattened, the deformation of these holes is not significant, and they remain approximately circular ； If a large hole is used, it will approximate an elliptical hole after expansion, and the stress concentration factor may increase. Especially when the ratio d/D is relatively large, due to the effect of shell curvature, the edges of the openings will induce additional bending moments, further increasing the stress levels and posing a risk to safety. IV. When the ratio d/D is large, the assumption of \"a small hole in a large plate\" is no longer valid. The calculations used cannot be correct. Therefore, a limit must be imposed on d/D. 3—42 What are the limitations on the maximum diameter of openings in pressure vessel shells? The restrictions are as follows: 1. For cylinders: when their inner diameter Di ≤ 1500 mm, the maximum diameter of the opening d ≤ , and d ≤ 520 mm; when their inner diameter Di > 1500 mm, the maximum diameter of the opening d ≤ , and d ≤ 1000 mm. 2. For convex end caps or spherical shells, the maximum diameter of the opening d ≤ . 3. For conical end caps, the maximum diameter of the opening d ≤ , where Di is the inner diameter of the conical shell at the center of the opening. 3—43 What formula is used to calculate the reinforcement area required for opening reinforcement in pressure vessels? A = dδ + 2δ(δm – C)(1 – fr) is the reinforcement area required for an opening with a flat cover in a pressure vessel; A = 0.5dδp for cases 3–44. What formula is used to calculate the reinforcement area required for openings in pressure vessels subjected to external pressure? A=0.5 3—45 What are the similarities and differences between the equal-area reinforcement method and the pressure area method? The pressure area method is an opening reinforcement technique adopted in the West German codes for pressure vessels and the technical regulations for steam boilers in West Germany, and it is indicated to be applicable to large-opening structures with an opening ratio of up to 0.8. The general formula for this calculation method is: where Ap represents the pressure acting area within the reinforcement range ； Aσ---is the area of the shell, nozzles, and reinforcement metal within the reinforcement range ； P---Design pressure, allowable stress of the material. This formula is based on the balance between the compressed area and the bearing area. The meaning of the equal area method is: to enhance the average strength of the shell by using additional metal in an area equal to that of the opening, thereby compensating for the reduced strength of the shell walls. Their basic starting points are the same. Due to differences in the consideration of the effective range, the results of the overall reinforcement calculation vary. What are the effective reinforcement range and effective reinforcement area for openings in pressure vessels with a mid-diameter of d/2 or D- of 3–46? The effective reinforcement range refers to: 1. Effective width: B=2d 2. B=d+2δn+2δnt; the larger of these two values is taken. The effective height h1 equals the actual external extension height; the smaller of the two values is taken. h2 equals the actual internal extension height; again, the smaller of the two values is used. The effective reinforcement area refers to the metal area within the effective reinforcement zone that can be utilized for reinforcement: A1 represents the excess metal area beyond the design thickness required for the shell to withstand internal or external pressures: A1 = (B-d)(δe-δ) – 2(δnt-C)(δe-δ)(1-fr). A2 represents the excess metal area beyond the design thickness required for the nozzle to withstand internal or external pressures: A2 = 2h1(δnt-δt-C)fr + 2h2(δnt-C-C2)fr. A3 represents the weld area within the reinforcement zone ； A4—Additional reinforcement area within the reinforcement zone. 3—47 What does the design of bolted flange connections include? 1. Determine the gasket material, type, and size. 2. Determine the bolt material, specifications, and quantity. 3. Determine the flange material, seal surface type, and structural dimensions. 4. Perform stress verification (all dimensions in the calculations do not include a corrosion allowance). 3—48 What requirements must be met when manufacturing integral necked flanges from steel plates? The following requirements must be met: 1. The steel plate shall undergo ultrasonic testing to ensure there are no delamination defects. 2. The slats should be cut in the direction of steel plate rolling. After bending and butt welding, it becomes a ring, with the surface of the steel plate forming the cylindrical surface of the ring. 3. The butt welds of the rings shall be full-penetration welds. 4. The ring butt welds shall undergo post-weld heat treatment and be inspected by 100% radiographic or ultrasonic testing, with the acceptance criteria specified in the relevant flange standards. 3—49 Under what circumstances should flanges be normalized or fully annealed? Answer: Normalizing or full annealing heat treatment should be carried out in any of the following cases: 1. Carbon steel or low-alloy steel flanges with a flange cross-section greater than 76 mm; 2. Welded integral flanges. 3. Forged flanges. 3—50 What is a narrow-face flange? What is a wide-flange? The contact surface of the gasket lies within the circular area enclosed by the flange bolt holes; this type of flange is known as a narrow-face flange. Flanges whose contact surfaces are distributed on both the inside and outside of the circle centered on the flange bolt holes are called wide-face flanges. 3—How many types are there for the mid-flanges in GB150-98, classified according to their degree of integrity? 51 What are the characteristics of each type? It is divided into three types: 1. Loose flange – The flange does not form a secure connection with the container or pipe fitting; in calculations, it is assumed that the container or pipe fitting does not share the flange’s moment forces. 2. Integral flange: The flange, the flange neck, and the container, or all three of them, can be effectively connected into a single unit structure that together withstands the flange torque. 3. Arbitrary-style flanges are some welded flanges (see Figures (9-1), (h), (I), (J), and (k) in GB150-98); their calculation is carried out as for integral flanges. However, for simplicity, it can also be calculated using the loop flange when the following conditions are met: δ≤15mm, P≤2Mpa, and the operating temperature is less than or equal to 370℃. 3—52 What are the basic conditions for sealing? What is the sealing specific pressure? What is the gasket coefficient? Why is it necessary to check the gasket width? There are two conditions for forced gasket sealing: the pre-sealing condition and the operational sealing condition. The significance of pre-sealing conditions is that, no matter how precisely the sealing surfaces of the flanges are machined, on a microscopic level their surfaces are uneven and contain grooves. These grooves can become leakage paths for the sealing surface. Therefore, softer gaskets must be used; under the force of the pre-tightening bolts, the surface of these gaskets will press into the uneven areas of the flange sealing surface, filling in the grooves and eliminating the aforementioned leakage paths. Sufficient compressive force should be applied to the effective sealing area of the gasket in this unit. The compressive force per unit area is referred to as the gasket sealing specific pressure (unit: Mpa), denoted by y. Different gaskets have different specific pressures. The harder the gasket material, the higher y is. The significance of operating under sealing conditions is that the sealing surfaces, which have achieved a sealed state through pre-tightening, will separate under the influence of internal pressure due to the axial force exerted by this pressure. As a result, the compressive force between the gasket and the sealing surfaces decreases, creating tiny gaps through which the internal pressure medium can leak. To ensure its sealing performance, it is necessary to maintain a sufficient liquid resistance between the gasket and the sealing surface; only when this resistance is greater than the driving force caused by the pressure difference inside and outside the medium can the gasket provide a seal and prevent leakage. Since the fluid resistance between the gasket and the sealing surface is proportional to the gasket compression force. To this end, a sufficiently large compressive force must be applied between the gasket and the sealing surface to ensure that the gap is small enough, thereby giving the liquid sufficient resistance. The ratio of the pressure applied per unit effective area of the gasket to its internal pressure, when a sufficient resistance is maintained between the gasket and the flange sealing surface to prevent leakage, is called the gasket coefficient, denoted by m. Different gaskets have different m values, and m increases as the hardness of the gasket increases. The gasket experiences the greatest compressive force when the bolts are pre-tightened, and it may be compressed into plastic deformation and lose its ability to rebound. Therefore, when the flange is under the pressure of the medium, the separation of the sealing surfaces prevents any rebound that could help to press these surfaces together again; as a result, sufficient contact force (i.e., the compressive force exerted by the gasket) cannot be maintained, leading to leakage. For this reason, the gasket must be compressed during pre-tightening, so that the compressive force per unit of effective sealing surface is not less than the value of y. However, to prevent plastic deformation, this compressive force must not be too high either. For planar sealing, to prevent the gasket from being compressed into plastic deformation, the gasket compression force that should be controlled is approximately 4y. When the gasket is pre-tightened, if the compressive force per unit of effective sealing area is less than y, it will prevent the elimination of \"leakage paths\", thereby failing to meet the sealing requirements. Conversely, when the preload on the gasket is too high (>4y), the gasket loses its elasticity, which again leads to leakage under internal pressure. The verification of the minimum gasket width in gasket calculations serves this purpose. However, this verification allows for experience to be used as a substitute; that is, the minimum width of the gasket can be determined based on experience. 3—53 What is the effective sealing width of a gasket? The width at which the gasket can make contact with the flange sealing surface before pre-tensioning is referred to as the gasket contact width, denoted by N. After the flange bolts are pre-tightened, the deflection of the flange ring causes the flange sealing surface to separate near the inner diameter, resulting in a loss of contact with the gasket at that location; therefore, the gasket can only be compressed near the outer diameter. The width of this compressible portion is called the compression width, denoted by bo. However, compressing the gasket does not mean it provides a seal. Only a gasket of appropriate tightness can provide effective sealing. Therefore, the width of the gasket that can actually provide effective sealing is only a portion of the compression width. That is, the part that is closer to the outer diameter of the gasket. The width of the gasket that actually provides sealing functionality is referred to as the effective sealing width of the gasket, denoted by b. This value is determined as follows: when bo ≤ 6.4 mm, the sealing width b = bo; when bo > 6.4 mm, b = 2.53 × 3–54. How many types of gasket compression forces are there? How to calculate it? 1. Minimum gasket compression force required under pre-tensioned condition: FG = 3.14DGb. 2. Minimum gasket compression force required under operating condition: Fp = 2πDGbmPc. What are the structural features of the reverse flange? 3–55 A reverse flange refers to a flat cover that is connected to the cylinder and features large holes with d>1/2Di. For flat covers with holes of d≤1/2Di, they can be designed by using hole reinforcement or by increasing the thickness of the flat cover. For openings with a size of d > 1/2Di, these design methods are no longer applicable; it is advisable to treat the flat cover with the large opening and the cylinder connected to it as reverse flanges, and to design them using the principles of flange design. 3—56 What are the advantages and disadvantages of the sealing surfaces of flat flanges, male-and-female flanges, and tenon-and-socket flanges? The flat flange sealing surface has the advantages of simple structure, easy processing, and ease of applying anti-corrosion linings. Due to the large contact area between this sealing surface and the gasket, if the pre-tightening is not proper, the gasket can easily be pushed out from the sealing surface. It is also difficult to compress, with poor sealing performance; it is suitable for applications with low pressures, generally used at pressures of PN≤2.5Mpa. In flange seals with concave-convex surfaces, the two mating flange surfaces are one concave and the other convex. It is easy to align during installation, effectively preventing the gasket from being pushed out of the sealing surface; its sealing performance is better than that of flat seals. The flange sealing surface of the tenon-and-socket type is formed by the combination of a tenon surface and a socket surface, resulting in a narrower sealing surface. Due to the obstruction by the groove surface, the gasket is not pushed out of the compression surface, and it is less subject to erosion and corrosion by the medium. It is easy to align during installation, the gaskets are evenly stressed, ensuring reliable sealing; it is suitable for use with flammable, explosive, and toxic media. It’s just that the gasket is very narrow, making it difficult to replace. 3—57 What strength criteria are required for flange strength verification? 1. Axial stress: For the overall flange: (except in Figures 9-1(c) and (g)): σH ≤ the smaller of 1.5tf and 2.5nt. For any flange designed as a integral flange, and for the integral flange shown in Figure 9-1(g): σH ≤ the smaller of 1.5tf and 1.5nt. For the integral flange shown in Figure 9-1(c): 1. Hoop stress: σH ≤ 1.5tf. 2. Circumferential stress: σT ≤ tf. 3. Radial stress: σr ≤ tf. 4. Combined stress: ≤ tf. 5. Shear stress: The shear stress under both pre-tensioned and operating conditions shall be less than or equal to 0.8 times the allowable stress of the flange (or cylinder) material at normal temperature and design temperature. 3—58 What are the advantages and disadvantages of two-support and multi-support configurations for horizontal vessels? The mechanical model of a horizontal container is similar to that of a beam. In a multi-support beam, due to the small distance between the supports and the reduced load distributed on each support, the bending moment in the beam is low, and thus the stress is also low. But the fulcrums are required to be at the same level. This is difficult to achieve with large containers. Due to uneven settlement of the foundation, the reaction forces at the multiple supports cannot be distributed evenly. With two supports, there is no issue of reaction forces not being evenly distributed. However, the bending moment across the span is large, and the bending moment at the support sections is also large, resulting in high stresses within the container walls. 3—59 What principles are used to determine the position and fixing method of supports in the design of double-support horizontal vessels? According to the mechanical analysis of a cantilever beam under uniformly distributed loads, when the length A of the cantilever beam is 0.207 times the total length L of the beam, the maximum bending moment across the span is equal to the bending moment at the support section (in terms of absolute value). If the cantilever is made longer, the stress at the supports increases. Therefore, horizontal containers usually require A≤0.2L. Furthermore, since the stiffness of the head is greater than that of the cylinder, the head acts as a reinforcement for the cylinder; if the supports are located near the head, its reinforcing effect can be fully utilized. Therefore, in addition to satisfying A≤0.2L, it is also necessary to satisfy A≤0.5Rm (the average radius of the cylinder). Like vertical vessels, the supports of horizontal vessels should also be fixed to the foundation. However, since the cylinder of a horizontal vessel stretches due to various thermal expansions, if its stretching is prevented by the fixed supports, additional stresses will be generated within the cylinder. Therefore, a horizontal container allows only one support to be fixed, while the bolt holes for the other support are designed as oblong holes to permit sliding. 3—60 What types of loads do tower equipment withstand? What aspects are included in the verification of its strength and stability? The loads to be sustained include: 1. Design pressure. 2. Hydrostatic pressure of the liquid column ； 3. The self-weight of the tower structure (including the internal components and packing), as well as the gravitational load due to the materials contained within it under normal operating conditions or during testing. 4. Weight load of auxiliary equipment, insulation materials, linings, pipes, escalators, and platforms ； 5. Wind load and seismic load. If necessary, the effects of the following loads should also be considered: 1. The forces resulting from the connection of pipes and other components. 2. Forces caused by differences in thermal expansion. 3. Effects of pressure and temperature changes. 4. Forces exerted during transportation or lifting. The above loads generate the following forces: 1. Axial stress caused by internal or external pressure; 2. Axial force caused by gravity during operation or when not in use. 3. Axial force caused by bending moment – Strength or stability verification: 1. Verification of the maximum combined tensile stress in the cylinder: For vessels under internal pressure: σ1–σ2+σ3≤Ktφ; for vessels under external pressure: –σ2+σ3≤Ktφ. 2. Verification of the maximum combined stress in the cylinder: For vessels under internal pressure: σ2+σ3≤Ker; for vessels under internal pressure as well: σ1+σ2+σ3≤Ker. Here, er represents the allowable axial compressive stress for the cylinder, which is determined as follows: er = min(B,t). 3—61. What formula is used to calculate the fundamental natural frequency of vertical vessels with equal diameter and equal thickness? 3—62 How is the basic wind pressure for a particular area determined? According to the GBJ9-87 \"Code for Loads on Building Structures\", the basic wind speed v0 in that area is determined using the following formula: = Where ρ is the air density, which varies with the local altitude and temperature; however, a uniform value of ρ=1.25 kg/m3 is used for calculating the basic wind pressure (this corresponds to the density of dry air at 10°C under 1 atmosphere of pressure). v0 represents the extreme wind speed based on data regarding the local average wind speeds, and this extreme value is closely related to factors such as the average time interval between occurrences and the specified return period. The average wind speed increases with height. In China, the reference value for wind speed is defined as the wind speed at a height of 10 meters, and an average wind speed over 10 minutes is used, with a return period of 30 years. 3—63 Why is it necessary to control the deflection of tower equipment? The ratio of the height to the diameter of the tower equipment is high, and the wall thickness of the equipment’s cylinder is thin; as a result of wind loads, excessive deflection occurs at the top of the tower. This leads to: 1. In the case of plate towers, the trays tilt significantly, resulting in uneven gas-liquid mass transfer and a decrease in the efficiency of the trays, which in turn affects product quality ； 2. The connections between the tower and the pipes are subject to combined stresses of tension, compression, bending, and torsion due to excessive swaying of the tower, which can lead to leaks; this is particularly dangerous for flammable, explosive, and toxic substances ； 3. Excessive deflection at the top of the tower means that the maximum displacement of the tower equipment during swinging results in a large absolute distance from the central axis, which in turn generates significant eccentric bending moments and affects the service life of the equipment. To ensure the proper operation and safe functioning of tower equipment, it is necessary to properly control the deflection at the top of the tower. 3—64 What measures should be taken in case of overpressure occurring during the operation of pressure vessels? An overpressure relief device should be installed. 3—65 GB150-98 Appendix B: How many types of overpressure relief devices are there? For what types of pressure vessels are these overpressure relief devices not suitable? Answer: There are three: 1. 1. Safety valve. 2. Shrapnel device. 3. Combination of safety valves and burst disc devices. Pressure vessels in which a sudden increase in pressure may occur during operation and the reaction rate reaches detonation levels are not suitable for these overpressure relief devices. 3—66 Compare the advantages and disadvantages of safety valves and burst discs respectively 1. A safety valve is an automatic pressure-relief valve that opens under the effect of inlet static pressure; it discharges a certain amount of fluid by relying on the pressure of the medium itself, in order to prevent the pressure inside the container or system from exceeding a predetermined safe level. When the pressure inside the container returns to normal, the valve closes automatically to prevent further flow of the medium. A burst disc device is a non-reclosing pressure relief device. The imported static pressure causes the rupture disc to be pressed and burst, thereby releasing the medium to prevent the pressure inside the container or system from exceeding a predetermined safe level. A new rupture disc must be installed again once the pressure returns to normal. 2. The design pressure of the container is determined separately for different overpressure relief devices. When a safety valve is used, the design pressure of the container is approximately 1.1 times the operating pressure ； For rupture disc devices, the design pressure of the container is 1.1 to 1.7 times the operating pressure. The same operating pressure. Pressure vessels equipped with safety valves have a lower design pressure and thinner wall thickness. 3—67 Under what circumstances must a burst disc device be used? Those that meet one of the following conditions must also be equipped with a burst disc device. 1. Media inside the container that can cause the safety valve to fail. 2. Containers that allow material leakage are not permitted. 3. The pressure inside the container increases too rapidly, so that the safety valve cannot keep up. 4. Other situations in which the safety valve is not suitable. 3—68 What factors should be considered in the structural design of low-temperature pressure vessels? Given that steel transitions from a ductile state to a brittle state as the operating temperature decreases, its impact resistance is reduced ； When unavoidable defects are present, being subjected to stress below the brittle transition temperature will lead to brittle fracture. Therefore, in addition to requiring the steel used in low-temperature containers to have high impact resistance, measures must be taken in the container’s design to prevent brittle fracture; the following issues need to be considered: 1. The structure should be as simple as possible to reduce constraints. 2. Avoid creating excessive temperature gradients. 3. Sudden changes in structural shape should be avoided as much as possible to reduce local high stresses ； The joint end should be polished into a rounded corner to ensure a smooth transition. 4. Welded parts should not be joined using discontinuous spot welds. 5. Shims must be placed under the container’s supports or legs; they must not be welded directly to the shell. 3—69 What are the special requirements for weld inspection of low-temperature pressure vessels? For containers that are required to undergo 100% inspection as specified, their T-joint butt welds and fillet welds must all be subjected to 100% magnetic particle or penetrant testing. The non-compressed components welded to the compressed components shall also be inspected in accordance with the requirements of these provisions. 3—70 What are the requirements for welding low-temperature pressure vessels? 1. Before welding low-temperature pressure vessels, welding procedure qualification tests shall be conducted in accordance with JB4708, including low-temperature Charpy (V-notch) impact tests on the welds and heat-affected zones. The sampling methods and acceptance criteria for impact testing are determined in accordance with the requirements for the base material specified in C2.1. 2. When the base materials on either side of the weld have different testing requirements, the impact test temperature for the weld metal shall be lower than or equal to the higher of the temperatures required for the base materials on those sides. The low-temperature impact energy shall meet the requirements of Table C3 based on the lower value of the tensile strength of the base materials on both sides. The heat-affected zone is determined in accordance with the requirements of the corresponding base material. The tensile and bending properties of the joint shall meet the lower requirement among those of the base materials on both sides. 3. Welding procedure qualification shall be carried out in accordance with JB7408; when welding joints are formed from base materials of different groups, the low-temperature impact test for such welding joints must be re-evaluated. 4. The welding line energy should be strictly controlled. Within the range confirmed by the welding procedure qualification, it is advisable to use a lower welding heat input and perform multiple passes of welding. 5. The surfaces of the butt joints and fillet joints within the welding area must be free from defects such as cracks, pores, and undercuts. There should be no sharp changes in shape; a gradual transition to a circular form is required. 3—71 What is a “low-temperature, low-stress operating condition”? Should containers operating under low-temperature and low-stress conditions be considered as low-temperature pressure vessels? ““A low-temperature and low-stress condition” refers to a situation in which the design temperature of the container or pressure-bearing component is below or equal to -20°C, but its tensile membrane stress is less than or equal to one-sixth of the yield strength of the steel at normal temperature, and not greater than 50 Mpa. When the container or its pressure-bearing components are under \"low-temperature and low-stress conditions\". If its design temperature is 50°C higher than -20°C, the regulations for low-temperature pressure vessels do not apply. 3—72 What are the material selection principles for wave expansion joints? 1. Wave expansion joints made of carbon steel and low-alloy steel are suitable only for temperatures t≤375℃; those made of austenitic stainless steel are suitable for temperatures t≤500℃. 2. For wave-shaped expansion joints made of carbon steel or low-alloy steel, the corrosion margin shall not exceed 1 mm; otherwise, austenitic stainless steel should be used. 3. In the presence of corrosive media such as chlorides, sulfides, acids, and alkalis, or when the operating temperature is high (above 550°C), corrosion-resistant alloys or superalloys should be used to manufacture expansion joints; examples include domestic materials such as FN-2, NS111, and B-315, as well as Incoloy 800 and 825. 3—73 What aspects are included in the strength calculation of wave expansion joints? Answer: The stress calculations are as follows: 1. Circumferential stress in the straight sections of the expansion joint caused by internal pressure: σz = 2. Circumferential membrane stress in the reinforcing rings of the straight sections due to internal pressure: σc = 3. Circumferential membrane stress in the bellows caused by internal pressure: σ1 = 4. Axial membrane stress in the bellows caused by internal pressure: σ2 = 5. Axial bending stress in the bellows caused by internal pressure: σ3 = 6. Axial membrane stress in the bellows resulting from axial displacement: σ4 = 7. Axial bending stress in the bellows resulting from axial displacement: σ5 = 8. Combined stress: σp = σ2 + σ3, σd = σ4 + σ5, σR = 0.7σp + σd 9. Stress verification: a.) σc, σ2, σ1, and σz should each be less than t; b.) σp ≤ 1.5σst; c.) For bellows made of carbon steel or low-alloy steel: σR ≤ 2σst. What kind of verification is required when σR > 2σst for expansion joints made of austenitic stainless steel? 3–74 Fatigue life verification is required: 1. Calculation of the number of cycles at fatigue failure: N= 2. Determination of the allowable number of cycles: = nf≥15 3–75 From a design perspective, what are the various failure criteria for pressure vessels? What are their respective views on this? The failure mentioned here is a design perspective, a commonly accepted criterion. Mainly: 1. Elastic failure: This theory of failure holds that when the metal on the inner wall of the container reaches its actual yield stress, it loses its purely elastic properties and enters a plastic state, at which point the container fails. This view holds that plastic deformation of the material alters the quality of the metal and leads to corrosion; therefore, containers are required to operate in an elastic state without allowing plastic deformation. 2. Plastic failure: This theory holds that when the material on the inner surface of the container undergoes plastic deformation, the deformation is confined to a very small range due to the restraint provided by the elastic layer on the outside, so the container does not reach a dangerous state. Only when the plasticity propagates from the inside to the outer wall does the container volume expand significantly, leading to instability; it is at this point that the load-bearing limit is reached. This perspective takes the overall yield of the container walls as the criterion for container failure. 3. Elastoplastic failure: This criterion is applicable to situations involving repeated loading. It assumes that the stresses in different parts of the vessel play different roles in causing its failure. For example, even when the stress in the vessel is much lower than the material’s yield point, local areas where the cylinder and the head meet the nozzles may have reached the yield point and thus experience plastic deformation, while adjacent areas remain elastic. Under repeated loading, such local plastic deformation does not necessarily lead to vessel failure; damage only begins to accumulate once the \"stability\" limit is exceeded, but failure does not occur immediately. 4. Blasting failure: For an ideal plastic material, once the container as a whole yields, plastic deformation continues to increase even if the pressure no longer rises, the wall thickness keeps decreasing, ultimately leading to the destruction of the container. The failure due to blasting theory holds that materials are not ideally plastic; due to strain hardening, the container will not break down if the pressure does not continue to increase. It is only when the pressure reaches a certain level that the container fails and bursts ； In the design, a safety factor is applied to the blasting pressure based on the operating pressure. This criterion is generally used in the design of ultra-high pressure vessels. In addition to the above four failure criteria, there are also creep failure, fracture failure, etc. 3—76 No What is the explosion limit? When combustible gases, vapors of combustible liquids, or combustible dusts mix with air to reach a certain concentration, an explosion occurs upon encountering a source of fire. The concentration range of an air mixture at which explosion can occur is known as the explosive limit. This limit is expressed as the volume percentage of flammable gases, vapors of flammable liquids, or flammable dusts in the air. The lowest concentration is called the \"lower explosive limit\", while the highest concentration is called the \"upper explosive limit\". 3-77 What are the ignition point and flash point? Answer: The ignition point refers to the heat of combustion released when a combustible substance is heated and ignited; this heat is sufficient to cause the substance to emit enough combustible vapor to sustain the combustion. The lowest temperature required to heat this substance at this point is known as its “ignition point”. Also known as the ignition point. The lower the ignition point of a substance, the easier it is to burn. The flash point is the lowest temperature at which the vapor emitted by a flammable liquid, when mixed with air, can catch fire upon contact with an ignition source. The flash point is different from the ignition point; the flash point is slightly lower than the ignition point. 3-78 How are flammable and combustible liquids classified? Answer: They are generally divided into four grades and two categories: Grade 1 has a flash point of 45°C to ≤120°C; Grade 4 has a flash point >120°C. Liquids of grades 1 and 2 are referred to as flammable liquids, while liquids of grades 3 and 4 are known as combustible liquids. 3-79 What are chemical hazardous substances? Answer: Any substance that possesses various degrees of hazards such as flammability, explosiveness, toxicity, corrosivity, or radioactivity, and which can, as a result of external factors such as friction, impact, vibration, exposure to fire, sunlight, moisture from water, temperature changes, or contact with other substances with conflicting properties, cause combustion, explosion, poisoning, burns, and other forms of injury to people or lead to damage to property, is considered a chemical hazardous substance. 3-80 What are the types of mechanical property tests for the weld joints of 3-80 welding test plates? Answer: 1. Tensile test 2. Bend test 3. Impact test. 3-81 After the bend specimen is cold-bent to the specified angle as required, what kind of cracks or defects are allowed on its tensile surface? Answer: There shall be no cracks or defects on the tensile surface that are longer than 3 mm in length in any direction. Cracks at the edges and corners of the specimen are not taken into account, but the length of cracks at those edges and corners caused by slag inclusions or welding defects shall be included in the evaluation. 3-82: What are the acceptance criteria for the impact test of product welding test plates??? Answer: The acceptance criteria: The specified value for impact energy at room temperature is as given in the drawings or relevant technical documents, but it must not be less than 27 J (the impact energy for three standard test pieces). The specified value for low-temperature impact work is in accordance with the relevant provisions of Appendix C (the appendix to the standard). At the test temperature, the average impact energy of the three specimens must not be lower than the specified value; one of the specimens may have an impact energy that is lower than the specified value, but it must still be at least 70% of that specified value. 3-83 If the tensile, bending, and impact tests on the welding test plates of pressure vessels fail, how should retesting be carried out? Answer: 1. If the tensile and bending tests on the welding test plates fail, retesting is permitted. For the non-compliant items, double samples shall be taken for re-inspection, and the compliance criteria shall meet the requirements of clauses E3.4 and E4.4 in GBl50---98 respectively. 2. If the results of the impact test do not meet the requirements of clause E5.4 in GBl50--98, another set of (3) specimens can be taken for testing. The qualifying criterion is that the average impact energy of the 6 specimens in each of the two groups must not be lower than the specified value; it is allowed for two specimens to have values below this specified value, but only one specimen may have a value that is 70% below the specified value. ‘ 3-84 Under what circumstances should pressure vessels and their pressure-bearing components be heat-treated? Answer: Vessels and their pressure-bearing components shall undergo post-weld heat treatment if they meet one of the following conditions. 1. The steel thickness δs shall meet the following conditions: a) For carbon steel and 07MnCrMoVR, the thickness must be greater than 32 mm (or greater than 38 mm when preheating before welding to over 100°C) ； b) 16MnR and 16Mn with a thickness greater than 30 mm (or greater than 34 mm when preheating before welding to over 100°C) ； c) For 15MnVR and 15MnV, when the thickness is greater than 28 mm (or greater than 32 mm when preheating before welding at over 100°C) ； d) 15MnVNR, 18MnMoNbR, 13MnNiMoNbR, 15CrMoR, 14CrlMoR, 12Cr2MolR, 20MnMo, 20MnMoNb, 15CrMo, 12CrlMoV, 12Cr2Mol, and 1Cr5Mo steels of any thickness ； e) For welded joints with different steel thicknesses δs, the aforementioned thickness shall be considered as the thinner one ； For welded joints formed by joining dissimilar steel grades, it shall be determined according to the more stringent heat treatment requirements ； 2. Cylinder thickness δn for cold forming and medium-temperature forming. Those that meet the following condition: for carbon steel and 16MnR, δn is not less than 3% of the designed inner diameter Di ； The nominal thickness of other low-alloy steels shall be not less than 2.5% of the design inner diameter Di. 3. Cold-formed heads shall be heat-treated. This restriction does not apply when the manufacturing unit ensures that the properties of the material after cold forming meet the design and usage requirements. Unless otherwise specified in the drawings, cold-formed austenitic stainless steel heads and welded joints of austenitic stainless steel may not require heat treatment. 4. Vessels marked with stress corrosion in the diagram. 5. The diagram indicates containers used to hold media with extreme or high toxicity. For the post-weld heat treatment of 3-85 pressure vessels, if the equipment is too long and the treatment must be carried out in stages within the furnace, what conditions must be met? During segmented processing, the length of the repeated heat treatment section should be no less than 1500 mm, and insulation measures should be taken for the portion outside the furnace to prevent temperature gradients from affecting the structure and properties of the material. 3—86 What conditions must be met and what measures should be taken when local heat treatment is applied to welds? Local heat treatment is permitted for welding joints of types B, C, and D, for type A welding joints where a spherical head is connected to a cylinder, and for areas that require repair due to defects. During local heat treatment, the heating width on each side of the weld should be no less than twice the thickness of the steel material ； When welding the joint to the shell, the heating width must be no less than 6 times the thickness of the steel material. Insulation measures should be taken in the area near the heating zone to prevent the temperature gradient from affecting the structure and properties of the material. 3--87 Under what circumstances should pressure vessels and their components be subjected to acid washing and passivation? The surfaces of vessels made of austenitic stainless steel or composite steel plates that require corrosion protection should be treated with acid washing and passivation. Austenitic stainless steel and composite steel components that require corrosion resistance are heat-treated in accordance with the specifications given in the drawings; they also need to undergo acid washing and passivation. According to standards GB150—98, under what circumstances should Class A and Class B welds of pressure vessels be subjected to X-ray or ultrasonic testing? Answer: a) For carbon steel and 16MnR with a steel thickness δs > 30 mm ； b) 15MnVR, 15MnV, 20MnMo, and austenitic stainless steels with a steel thickness δs > 25 mm ； c) Steel with a minimum standard tensile strength value of σb > 540 MPa ； d) 12CrMo, 15CrMoR, 15CrMo with steel thickness δs > 16 mm; Cr-Mo low-alloy steels of any other thickness ； e) Containers subjected to pressure testing ； f) The diagram indicates containers holding media with extreme or high hazard toxicity ； , g) Containers specified in the pattern to be inspected 100% ； h) Class A welded joints for the inner cylinder of multi-layer wrapped pressure vessels ； i) Class A welded joints of each single-layer cylinder in the heat-shielded pressure vessel ； J) Full inspection of the weld intersection areas and areas below them: (a) First, assemble the plates, and then inspect all the joints on the convex head ； (b) Welded joints that are covered by gussets, supports, shims, internal components, etc ； (c) The welded joint enclosed within a circle with the center of the opening as its center and 1.5 times the diameter of the opening as its radius ； (d) Welded joints where the embedded fitting is connected to a cylinder or head; (e) Welded joints where fittings with a nominal diameter of 250 mm or more are connected to long-neck flanges, or where one fitting is connected to another. 3—89 According to the ‘Code for Pressure Vessels’, under which circumstances must the butt welds of butt joints in pressure vessels undergo full radiographic or ultrasonic testing? Answer: Full radiographic or ultrasonic testing is required in the following circumstances: 1. When full radiographic or ultrasonic testing is specified in GBl50 and GBl51; 2. For Class III pressure vessels ； ’ 3. Those with a design pressure of 5 MPa or greater ； 4. Reaction pressure vessels and storage pressure vessels containing flammable media among the second category of pressure vessels ； 5. Shell-and-tube waste heat boilers with a design pressure of 0.6 MPa or higher ； 6. Pressure vessels designed using fatigue analysis: 7. The weld coefficient of 1.0 is adopted in the design (except for seamless tube shells) ； 8. Pressure vessels that cannot be subjected to internal or external inspections or pressure tests after use ； 9. Pressure vessels that utilize electroslag welding ； 10. Pressure vessels made of aluminum, copper, nickel, titanium, and their alloys that meet one of the following conditions: (1) The medium is flammable or presents an extremely high, high, or moderate level of hazard. (2) Those that use pneumatic testing. (3) Those with a design pressure of 1.6 MPa or higher. 3-90 Under what circumstances should magnetic particle or penetrant testing be performed on the surface of welds? Answer: Welded joints that meet one of the following conditions require magnetic particle or penetrant testing of their surface in accordance with the methods specified in the drawings. a) Class C and Class D welded joints on containers as specified in items c) and d) of 10.8.2.1 ； b) Welded joints of category C for laminated pressure vessels made of laminated materials with a minimum standard tensile strength of σ>540 MPa: c) Surfacing welds: d) Welded joints of the composite layer in composite steel plates ； e) Surfaces of grooves on materials with a minimum standard tensile strength of σ>540 MPa and Cr-Mo low-alloy steels after flame cutting, as well as the surfaces of areas where defects in such containers have been ground out or rewelded, and the weld surfaces at locations where clamps and pulling aids have been removed. f) In the case of 10.8.2. 1 Welded joints of nozzles with a nominal diameter of less than 250 mm on the container, connected to long-neck flanges, as well as connections between nozzles. 3—91 What are the requirements set by the Code of Practice regarding the selection of testing methods for welded joints of pressure vessels? Answer: 1. Radiographic inspection shall be used when the wall thickness of the pressure vessel is less than or equal to 38 mm ； When radiographic testing cannot be used due to structural reasons or other factors, recordable ultrasonic testing can be an alternative. For steel pressure vessels with a standard tensile strength of 540 MPa or higher and a shell thickness of more than 20 mm, in addition to radiographic testing, local ultrasonic testing shall be carried out on each butt weld: 2. For butt joints of pressure vessels with a wall thickness of 38 mm or more, if radiographic inspection is chosen ； In such cases, each weld shall also undergo local ultrasonic testing; the additional local testing shall cover all weld intersections, with a testing ratio of 20% of the original non-destructive testing ratio specified in Article 84 of the Acceptance Criteria. 3. For fillet joints and T-joints that require flaw detection, when radiographic or ultrasonic testing is not possible, 100% visual inspection shall be carried out. 4. For butt joints of pressure vessels made of non-ferrous metals, radiographic inspection should be preferred as much as possible. 5. Magnetic particle testing should be given priority for surface inspection of ferromagnetic material pressure vessels. 3-92 What are the acceptance criteria for radiographic or ultrasonic testing of welds in pressure vessels? Answer: Radiographic testing of welds is carried out in accordance with JB4730-94; for Class A and Class B welds that are tested 100%, a grade of II is considered acceptable ； For Class A and Class B welds subjected to local testing, grade III is considered acceptable. Ultrasonic testing of welds is carried out in accordance with JB4730-94. For Class A and Class B welds that are tested 100%, a grade of I indicates compliance ； For Class A and Class B welds subjected to local testing, grade II is considered acceptable. 3-93 What should be done if unacceptable defects are detected through radiographic or ultrasonic testing? Answer: For welds that show unacceptable defects as identified by radiographic or ultrasonic testing, repair welding should be carried out after those defects are completely removed, and the affected area should be rechecked using the same testing methods until it meets the required standards. For welds that are subject to local inspection, if any unacceptable defects are found, the inspection length shall be increased at both ends of such defects by an amount equal to 10% of the length of that weld, with this increased length being no less than 250 mm. If any unacceptable defects remain, 100% inspection of that weld is performed. 3-94 What are the requirements when installing inspection holes in pressure vessels? Answer: 1. When the inner diameter of the pressure vessel is 1000 mm or greater, at least one manhole should be provided ； For pressure vessels with an inner diameter of 500 mm or more and less than 1000 mm, one manhole or two handholes shall be provided ； For pressure vessels with an inner diameter of 300 mm or more and less than 500 mm, at least two manholes shall be provided. 2. The diameter of circular manholes should be no less than 400 mm, while the dimensions of elliptical manholes should be no less than 400X300 mm ； The diameter of the circular hand hole should be no less than 100 mm ； The size of the oval hand hole should be no less than 75X50 mm. 3. Pressure vessels shall be equipped with removable end caps (covers, etc.) or other lids that can be opened and closed; wherever such elements serve as manholes or access holes, there is no need to provide additional manholes or access holes. However, its size shall not be less than the specified size of the manhole or handhole it replaces. 4. If a threaded plug inspection hole is provided on the pressure vessel, it is not necessary to provide a manual access hole ； The nominal pipe diameter of the threaded plug should be no less than 50 mm. 5. The location of the inspection holes should be reasonable and appropriate to facilitate internal cleaning. Handholes or threaded pipe plug inspection holes should be provided separately on the end closures or on the cylinder near the closures. 6. The manhole of spherical pressure vessels should be located on the pole band. 3-95 Pressure vessels shall not be equipped with removable heads or covers to replace manholes or handholes, nor shall they have inspection holes ； What requirements must be met? · Answer: The following requirements must be satisfied simultaneously: 1. The medium used in operating the pressure vessel is non-corrosive or only slightly corrosive; designers who do not require internal inspection and cleaning should indicate this in the overall design drawings and technical documents ； 2. All welds shall be subject to complete non-destructive testing ； 3. The designer shall indicate the calculated thickness on the design drawings ； —So as to conduct thickness inspections during use. 3-96 Under what pressure conditions should oil-resistant rubber asbestos sheets be used as gaskets for containers containing hydrocarbon compounds or organic solvents? Why? Answer: Not more than 2.5 MPa; they are suitable for butt-welded flanges, and the compressive force applied is not high enough to damage them. Furthermore, benzene, as an organic solvent, has a dissolving effect on nitrile rubber in oil-resistant rubber asbestos gaskets, which is also one of the reasons for limiting the pressure to no more than 2.5 MPa. 3-97 What are the concepts of primary stress, secondary stress, and peak stress? Answer: Primary stress is the normal or shear stress required to balance pressure and other mechanical loads. Primary stress is divided into the following three categories: 1. Primary overall film stress is the primary film stress whose influence extends throughout the entire structure. During the plastic flow process, the overall film stress does not redistribute; it will directly lead to structural failure. 2. Primary local film stress: A stress level that is greater than the primary overall film stress, but whose influence is limited to a local area of the structure. When local plastic flow occurs in the structure, such stresses will be redistributed. If left uncontrolled, when load is transferred from a high-stress area of the structure to a low-stress area, excessive plastic deformation will occur, leading to failure. 3. Bending stress per unit length: The bending stress that is distributed linearly across the thickness of the cross-section, required to counteract pressure or other mechanical loads. Secondary stress is the normal or shear stress required to satisfy external constraint conditions or the requirement for continuous deformation of the structure itself. The basic characteristic of secondary stress is its self-limiting nature; that is, local yielding and a small amount of deformation are sufficient to satisfy the constraints or the requirements for continuous deformation, thereby preventing further increase in deformation. As long as it is not loaded repeatedly, secondary stress will not cause structural failure. Peak stress is the stress increment added to the primary and secondary stresses, caused by local structural discontinuities or the effect of local thermal stresses. 3-98 What are the basic characteristics of peak stress? Under what circumstances is it necessary to limit peak stress? Answer: Its characteristics are self-limiting and localized; it does not cause significant deformation ； Its hazard lies in the fact that it may cause fatigue cracks or brittle fracture. Frequent alternating loads or temperature changes can easily lead to fatigue; in such cases, the peak stress should be controlled. 3-99 What are the advantages and disadvantages of ultrasonic testing compared to radiographic testing? Answer: 1. Compared to radiographic testing, ultrasonic testing has the following advantages: a. It has high sensitivity in detecting hazardous defects such as cracks and lack of fusion ； b. It can detect materials with thicknesses of several meters, whereas X-rays can generally only detect thicknesses of 40–60 mm; only by using a 9MoV linear accelerator is it possible to detect materials with a thickness of 400 mm ； c. Testing can be performed from either side of the material, and it is possible to test and monitor containers in use ； d. Fast flaw detection speed, capable of determining the depth and location of defects ； e. The equipment is simple, and the testing costs are low ； f. It is harmless to the human body. 2. In comparison, ultrasonic testing has the following disadvantages: a. Damage assessment is not intuitive, and qualitative evaluation is relatively difficult ； b. There are no original records of the test results ； c. The test results are highly affected by human factors. 3-100 Why is it required to perform ultrasonic testing on ESW welds after normalizing? Answer: The welds formed by ESW develop coarse columnar crystals, which increases ultrasonic attenuation; moreover, grain boundary reflections occur, making it difficult to detect defects. After normalizing, the grain structure is refined, enabling defects to be detected. Therefore, ultrasonic testing of electroslag welds must be carried out after normalizing. 3-101 Why should important steel plates be inspected on both sides? Answer: In the detection using a straight probe with multiple reflections, the echo generated by defects near the bottom surface of the steel plate can easily overlap with the bottom wave, leading to missed detections; therefore, inspection is necessary on both sides. 3-102 Can radiographic inspection and ultrasonic inspection be used as substitutes for each other? Answer: Since Article 86 of the Regulations on Safety Supervision of Pressure Vessels specifies different inspection methods for various wall thicknesses of vessels, they cannot be used as substitutes for one another. 3-103 What are the main structural types of spherical shells? Answer: There are mainly three types of segmentation; low-alloy steel with a thickness of 540 MPa and greater than 38 mm ； Butt weld between the embedded splice and the sphere shell ； Welding tests determined that the welds require degassing treatment ； Welds that meet one of the above conditions must undergo hydrogen removal treatment immediately after welding. 3-111 Under what conditions is 100% inspection required for the butt welds of spherical tanks? Answer: Spherical tanks made of carbon steel or 16MnR steel with a steel thickness δs greater than 30 mm shall undergo 100% radiographic or ultrasonic inspection using the inspection methods specified in the drawings, if either of the following conditions is met: a ； b. Steel spherical tanks made of 15MnVR with a steel thickness δs greater than 25 mm, and 15MnVNR steel spheres of any thickness ； c. Steel spherical tanks with a minimum standard tensile strength value of σb > 540 MPa; d. Spherical tanks that are subjected to pressure testing ； e. The diagram indicates spherical tanks used to hold flammable media and those with extremely or highly hazardous toxicity ； f. Spherical tanks specified in the drawings to be inspected 100%. 3-112 What are the regulations regarding testing of steel materials prone to delayed cracking? Answer: Non-destructive testing of the welds should be carried out at least 36 hours after welding is completed. 3-113 How to conduct a hydraulic test on spherical tanks? Answer: In accordance with the following provisions of GBl2337—98: 8.10.1 After the spherical tank has been installed, it shall undergo pressure testing and airtightness testing as specified in the drawings. 8.10.2 During pressure testing, a pressure gauge with the same range and calibrated properly must be installed at both the top and bottom of the spherical tank. It is advisable to choose a pressure gauge with a range of about 2 times the test pressure, but it should not be less than 1.5 times or more than 4 times the test pressure. It is advisable that the diameter of the pressure gauge be no less than 150 mm. The test pressure is determined based on the reading of the pressure gauge at the top of the spherical tank. 8.10.3 Compressed air at a pressure of 0.4~0.5 MPa should be introduced into the opening reinforcement rings of the spherical tanks prior to the pressure test to check the quality of the welds. 8.10.4 Hydrostatic testing 8.10.4.1 Water is generally used for hydrostatic testing; other liquids that do not pose a risk may be used if necessary. 8.10.4.2 The pressure of the test solution shall be in accordance with the provisions of 3.8.1 of this standard. Test temperature: a. The temperature of the liquid during testing should be below its flash point or boiling point. b. For hydraulic testing of spherical tanks made of carbon steel, 16MnR, and normalized 15MnVB steel, the liquid temperature must not be below 5℃ ； For other low-alloy steel spherical tanks (excluding cryogenic spherical tanks), the liquid temperature during the hydrostatic test must not be lower than 15°C. If factors such as steel grade and plate thickness cause an increase in the material’s ductile transition temperature, the temperature of the testing liquid must be increased accordingly. 8.10.4.4 Test requirements a. An exhaust port should be provided at the top of the spherical tank during testing. The air inside the spherical tank must be exhausted when filling it. During the testing process, the outer surface of the spherical tank should be kept dry ； b. During the test, the pressure should be increased gradually; once it reaches 50% of the test pressure, it should be held at that level for 15 minutes. Afterward, a leak check should be conducted on all welds and connections of the spherical tank, and once no leaks are detected, the pressure can be increased further ； c. When the pressure rises to 90% of the test pressure, hold it for 15 minutes, conduct another leak check; once no leaks are detected, increase the pressure further ； d. When the pressure rises to the test pressure, hold it for 30 minutes, then reduce the pressure to the design pressure and conduct an inspection; no leaks are considered acceptable ； e. After the hydraulic test is completed, the liquid should be drained, and the tank should be dried out using compressed air. During drainage, it is strictly prohibited to discharge it on-site. 8.10-4.5 Requirements for foundation settlement. a. During the process of filling and draining the spherical tank, the settlement of the foundation should be monitored during the following steps ； Before filling with liquid ； When the liquid level reaches 1/3 of the sphere shell diameter ； When the liquid level reaches 2/3 of the sphere shell diameter ； 24 hours after filling. After draining the liquid. b. The settlement of each pillar foundation should be measured, and the settlements of all pillar foundations should be uniform. After fluid discharge, the basic settlement shall not exceed Db/1000 (where Db is the diameter of the base circle), and the difference in settlement between adjacent pillar bases shall not be more than 2 mm. When exceeded, measures should be taken to address it. 3-114 What range of pressures does the Code for Pressure Vessels specify as those for which a vessel is considered a high-pressure vessel? According to the Code, pressure vessels with a design pressure (P) within the following range are classified as high-pressure vessels: 10 Mpa ≤ P < [specific value] Mpa. 3-115 What is the formula used to calculate the wall thickness of cylindrical parts of multi-layer high-pressure vessels as specified in GBl50--98? What are the limitations regarding the maximum design pressure for which this formula is applicable? Answer: The formula for calculating the wall thickness of cylindrical parts under internal pressure for multi-layer high-pressure vessels is as follows: δ = [formula]. Where: Pc – design pressure, in MPa ； Di——inner diameter of the cylinder, mm ； φ—weld coefficient ； t—is the allowable stress of the cylinder material at the design temperature, in Mpa. For multi-layer containers, the value of tφ is determined using the following formula: tφ = tiφi + t0φ0, where δi is the thickness of the inner layer of the multi-layer container, in mm ； δ0 — total thickness of the laminate, δn — nominal thickness of the cylinder, t0 — allowable stress of the inner liner material at the design temperature, ti — allowable stress of the laminate material at the design temperature, in MPa ； φi——Welding coefficient of the inner cylinder in multi-layer containers. φ0—is the weld coefficient for the laminate layer. The above formula for calculating the cylinder wall thickness is applicable to the range where the design pressure Pc ≤ 0.4tφ and D0/Di ≤ 1.5 (Do – outer diameter of the cylinder, Di – inner diameter of the cylinder). ) 3-116 What are the various forms of single-layer high-pressure vessel shells? Answer: The single-layer high-pressure vessel shells come in the following forms: fully forged type, single-layer rolled and welded type, and electroslag remelting type. 3—117 What are the disadvantages of single-layer rolled welded high-pressure vessels? Single-layer rolled welded high-pressure vessels have the following disadvantages: a. The cylinder of a single-layer high-pressure vessel is formed by rolling and welding thick plates, whose properties are far inferior to those of thin plates. There are significant differences in properties along the thickness direction, with considerable variations in ductility and toughness both in the rolling direction and perpendicular to it; the properties in the thickness direction are even worse. b. Due to the plate thickness, metal elements tend to segregate; the content, distribution, and shape of impurities are uneven. Additionally, during heat treatment, the quenching effects on the interior and surface differ because of the thick plate, resulting in uneven material properties inside and outside. This makes cracks more likely to occur during welding. c. Thick plates have higher transformation temperatures, increasing the likelihood of brittle failure. 3--118 What are the common defects that occur in the butt deep-groove weld seams of multi-layer high-pressure vessels, especially those with multi-layer wrapping? How to overcome it? Answer: The welding defects that tend to occur more frequently in multi-layer circumferential welds are undercutting or slag inclusions at the interfaces between the layers. To overcome this defect, the method of pre-welding the end face can be employed. 3-119 Multi-layer high-pressure vessels are generally equipped with vent holes at the shell sections. Why? Answer: The purposes and functions of these vent holes are as follows: a. During ring welding, the gas trapped between layers can escape freely, which helps to improve the quality of the welding ； b. During operation and temperature changes, the gas in the interlayer can expand freely, which reduces the adverse effects caused by gaps ； c. It can serve as an alarm system: in the event of a leak in the inner cylinder, the leaked material can be discharged from the device quickly, allowing it to be detected and addressed promptly ； d. In a high-pressure vessel with a hydrogen atmosphere, if hydrogen diffuses throughout the entire vessel, it can be discharged through vent holes to prevent its accumulation. 3-120 What are the failure modes of pressure vessels? Answer: A pressure vessel loses its ability to function properly due to excessive mechanical or thermal loads; this is referred to as failure. There are three forms: 1. Strength failure: The container undergoes excessive plastic deformation or rupture under load. 2. Stiffness failure: The container undergoes excessive elastic deformation, resulting in difficulties in transportation and installation or a loss of its normal operational capacity. 3. Stable failure: The container suddenly changes shape under load, resulting in the loss of its functional capacity. The design of pressure vessels must take into account these three possible failure modes and consider them comprehensively to ensure the proper operation of the equipment. 3-121 What are the main differences between the conventional design method and the analytical design method for pressure vessels? Answer: The main design methods for pressure vessels at present are the conventional design method and the analytical design method. The conventional design method uses elastic failure as a criterion and film stress as a basis to calculate the thickness of the component. The maximum stress is limited to not exceed a certain allowable value (usually 1 times the allowable stress). Local stresses such as high edge stresses present in the container are represented in the form of stress enhancement factors, and the maximum stress obtained after calculating these local stresses is assigned the same allowable stress value as that for thin films. The thickness of the internal pressure cylinders and shells in GBl50 is calculated based on the membrane stress within those components (the primary overall membrane stress), with it being controlled at around 1 time the allowable stress level. For ellipsoidal heads, as well as dish-shaped heads, the thickness is determined by taking into account the local stresses resulting from the interaction between the head and the cylinder wall; the maximum stress, which is obtained by adding these local stresses to the membrane stress, is then controlled to be no more than 1 time the allowable stress. The conventional design method is simple, but it is not entirely rational and tends to be conservative. The analytical design method is based on plastic and elastoplastic failure criteria; it takes into account various stresses within the container, such as overall membrane stress, edge stress, and peak stress, to carry out accurate calculations. The stresses are classified, and different strength constraints are applied according to the various failure modes caused by each type of stress, thereby enabling the calculation of the component’s thickness. Containers designed according to this method are more scientific, reasonable, safe, and reliable, and can also achieve certain economic benefits. The thickness calculation for various components in the JB4732 standard is based on stress analysis and employs the third strength theory. Among them, although the calculation formulas for internal pressure cylinders and spherical shells are formally identical to the corresponding formulas in GBl50, their mathematical meanings are completely different. Analytical design, by distinguishing between different types of stresses and loads, fully utilizes the load-bearing capacity of materials; as a result, it imposes high technical requirements on materials as well as on manufacturing and inspection processes. 3-122 How are thin-walled containers and thick-walled containers distinguished? What is the theoretical basis for their strength design? What are the differences between them? Answer: The outer diameter of the container (D.). When the ratio K = Do/Di to its inner diameter (Di) is ≤ 1.2, it is called a thin-walled vessel. When K>1.2, it is a thick-walled vessel. The theoretical basis for the strength design of thin-walled containers is the moment-free theory of rotating thin shells, with the assumption of a straight normal line being employed ； The stresses calculated in this way are all film stresses that are uniformly distributed along the wall thickness, and the radial stresses perpendicular to the container wall surface are ignored. It is an approximate calculation method, but it can be kept within the error range acceptable in engineering applications. The theoretical basis for the strength design of thick-walled vessels is the Lamé formula derived from stress analysis in elasticity theory. The stress calculated in this way is a triaxial stress. Among them, the circumferential and radial stresses are non-linearly distributed along the wall thickness; when subjected to internal pressure, the absolute value of the stress is highest on the inner wall and lowest on the outer wall. However, their axial stress is still evenly distributed along the wall thickness. The stress in thick-walled cylinders as presented by the Latin American formula agrees well with the actual situation, reflecting the objective distribution law of stress. It is applicable to both thick-walled and thin-walled containers. For containers under internal pressure, the circumferential film stress calculated using the film theory is lower than the maximum circumferential stress on the inner wall obtained from the Lame formula, and this error increases as the value of K increases. When K=1.5, the circumferential stress calculated based on the film theory using the inner diameter is 23% lower than the circumferential stress of the inner wall calculated using the Lame formula. When based on the mean diameter, the circumferential stress calculated using the film theory is only 3.8% lower than the circumferential stress on the inner wall calculated using the Lame formula. For ordinary pressure vessels, this error is within the allowable range. To this end, GBl50 adopts a membrane theory formula based on the mean diameter for the calculation of internal pressure cylinders. Its applicable condition is K≤1.5, which is equivalent to Pc≤0.4tφ. 3-123 What are the similarities and differences in the membrane stresses of cylinders and spherical shells under internal pressure? Answer: Similarities: Both experience biaxial membrane stresses, and these stresses are uniform throughout. · Difference: The circumferential film stress in the cylinder is twice the axial stress. The biaxial membrane stress in the spherical shell is equal to the axial stress in the equidiameter cylinder. For this reason, under the same diameter and pressure, the wall thickness required for a spherical shell is only half that of a cylinder. 3-124 What are the similarities and differences in the membrane stresses of cylinders and conical shells under internal pressure? Why should the semi-apex angle of a conical shell not exceed 60°? Answer: Similarities: Their circumferential stress is equal to twice the radial (axial) stress, and it is distributed uniformly throughout the wall thickness. Difference: In a cylinder, the stresses are uniformly distributed along the axial (longitudinal) direction, whereas in a conical shell, the stresses are linearly distributed along the longitudinal direction. The stress is highest at the larger end and lowest at the smaller end. The stress at the larger end of the conical shell is 1/cosa times the corresponding stress of a cylinder with the same diameter as that end. Here, a is the half-apex angle, which is less than 60°. For this reason, l/cosa > 1; thus, the stress at the larger end of the conical shell is greater than that of a cylinder with the same diameter, and it increases as a increases. When the semi-apex angle of the conical shell is less than 60°, the stresses within the shell are primarily membrane stresses, and the conical shell behaves as a shell structure; therefore, membrane theory can be applied for calculations. When a > 60°, the stress in the shell becomes predominantly bending stress, which makes the thin-shell theory inadequate; therefore, a should not be greater than 60°, otherwise calculations should be performed using the theory of circular plates. 3-125 What are the characteristics of the membrane stresses in spherical shells, dish-shaped shells, and ellipsoidal shells under internal pressure? Answer: In a spherical shell, the membrane stress is equal in both the radial and circumferential directions; it remains constant throughout, being the same at every point, and it is always a tensile stress. Disc-shaped shell: It consists of a spherical part with a larger radius at the center and an annular shell with a smaller radius around the periphery. The stress condition in its central spherical portion is the same as that of a spherical shell. That is, there is the same bidirectional film tensile stress, which remains constant along the spherical portion. However, in the ring shell with a smaller radius (the transition zone), the radial stress is tensile, while the circumferential stress is compressive. Ellipsoidal shell: Only at the vertices of the shell are the normal membrane stresses equal, and both are tensile stresses. Away from the vertex, both longitudinal tensile stress and circumferential (transverse) tensile stress decrease, but the longitudinal stress remains tensile; at the equatorial region, it equals the axial membrane stress in a cylinder of constant diameter. The circumferential (latitudinal) stress in an ellipsoidal shell is tensile near the center of the shell, but it decreases as one moves away from the center, and may change from tensile to compressive stress. This variation depends on the ratio of the long and short axes of the ellipsoidal shell, a/b: when a/b = √2, compressive membrane stress occurs in the circumferential direction of the ellipsoidal shell, and this compressive stress increases as a/b grows; the maximum circumferential compressive stress occurs at the equatorial region. For a standard elliptical head with a/b=2, the maximum tensile membrane stress occurring at the top of the head is exactly equal to the maximum circumferential compressive membrane stress occurring at the bottom edge (the equator) of the head. Its value is exactly equal to the absolute value of the circumferential film stress in a solid circular cylinder. 3-126 What are the characteristics of edge stress? Answer: (1) Self-limiting nature: Edge stress arises in order to ensure deformation compatibility between adjacent elements. When this stress reaches the material’s yield point, plastic flow occurs in the material, thereby achieving deformation compatibility. Once the deformation requirement is met, the plastic flow of the material automatically ceases. Therefore, its stress and deformation energy are automatically restricted. (2) Locality: In general edge stresses, longitudinal bending stress is predominant, but its range of influence is limited, decreasing rapidly as one moves away from the edge. For a cylinder, at a distance of 2.5 from the edge (where R is the cylinder radius and t is the cylinder thickness), the edge bending stress has already dropped to 5% of the maximum stress value. 3-127 Why do both elliptical and disc-shaped end caps have straight edge sections? Answer: This is to prevent the ring weld connecting the end cap to the cylinder from overlapping with the stress concentration area at the edges. Not only can welding defects exist in the circumferential welds, but welding residual stresses are also inevitable. When these stresses coincide with edge stresses, it has a very adverse effect on the structural integrity; for this reason, straight edges are provided on the end caps to improve their stress conditions. 3-128 What is a thin circular plate? What is the theoretical basis for the stress analysis of thin plates? Answer: A thin circular plate refers to a circular plate whose ratio of thickness δ to diameter D lies within the following range: 0.01