Thread Content
【Keywords】Pressure vessels 【Paper abstract】Several issues commonly encountered in the manufacturing of pressure vessels are discussed, and personal opinions are presented based on a comprehensive analysis of various existing views, with the aim of reaching a consensus to ensure the quality of product manufacturing. In pressure vessel manufacturing, various issues often arise, such as the use of spin-formed approximately elliptical heads as substitutes for standard elliptical heads, the determination of the minimum thickness for formed heads, the sequence of operations involved in the expansion welding of pipes to tube sheets, and matters related to hydrostatic testing. These issues may seem simple on the surface, but practical implementation presents certain difficulties, and there are even some fundamentally different opinions on them. To this end, the author has conducted a thorough and detailed analysis of the aforementioned issues. Based on an evaluation of existing viewpoints, the author presents their own conclusions regarding these issues, so as to ensure proper understanding and implementation in construction processes and to guarantee the safety and reliability of product quality. 1 Replacement of spun approximate elliptical heads with standard elliptical heads As heads become larger and spin-forming technology is more widely applied, an increasing number of spun heads are being used in pressure vessels. From the perspective of the mechanism and process of spin forming, the heads produced by this method should be classified as disc-shaped heads; therefore, in practice some people believe that they cannot be used as a substitute for elliptical heads [1,2], and there are quite a number of people who hold this view. They believe that since the head is disc-shaped, it should be inspected based on the wall thickness calculated using the strength formulas applicable to disc-shaped heads; in other words, the wall thickness of the head should be thicker than that of an elliptical head conforming to the corresponding standard specifications, regardless of the degree of similarity in actual shape between the spin-formed elliptical head and the standard elliptical head. Contrary to the above view, some people argue [3–5] that regardless of the method used for forming the head, as long as the actual shape of the resulting head falls within the allowable tolerances specified for standard elliptical heads in GB 150, it should be considered a standard elliptical head; therefore, it can be used as a substitute for such heads without the need to increase the wall thickness. Studies of relevant literature show [5,6] that the shapes of five types of deep dish-shaped end caps, namely those with R=0.8Di, r=0.154Di, and R=0.904Di, r=0.173Di, are very similar to those of standard elliptical end caps; the deviations of their theoretical curves do not exceed the allowable shape deviation limits for standard elliptical end caps as specified in GB 150. Therefore, based on the meaning of these regulations, these dish-shaped end caps should also be classified as standard elliptical end caps, and thus can be used as an equivalent substitute for them. Furthermore, from the perspective of stress analysis, the relevant charts in ASME VIII-2 clearly state [7] that the calculation curves for standard dish heads with R=0.9Di and r=0.173Di are actually the same as those for standard elliptical heads; in other words, the two are identical [8]. Meanwhile, for elliptical heads where Di/2hi is not equal to 2, Reference [7] stipulates that they should also be designed using an equivalent dish head or according to the analysis in the appendix. It can be seen that in Reference [7], the design of disc heads and elliptical heads follows the same method, with both being determined based on the curve diagram of disc heads. Therefore, for disc and elliptical end caps with similar or comparable shapes, their design is always the same [8]. The author’s calculations of the wall thicknesses for the approximately elliptical heads with R/Di=0.833 and r/Di=0.156, as well as for the standard elliptical heads, show that, when the diameter and material are the same, the results are very close to those in Reference [7]. Furthermore, the author analyzed the wall thickness calculation formula for the deep-dish head (R=0.8Di, r=0.154Di) in AD [9] and found that the wall thickness of the deep-dish head as specified in AD is generally always less than or equal to that of the standard elliptical head calculated according to GB 150 (since AD does not include the concept of an elliptical head, only deep-dish and shallow-dish heads). It can also be seen from this that, under normal circumstances, it is feasible to use a deep-dish head in place of a standard elliptical head with the same specifications, material, and plate thickness. The wall thickness cannot be simply determined using the appropriate calculation formulas based on the type of end cap; rather, more attention should be paid to the differences between the actual shape and the standard shape, in order to determine whether its strength can be guaranteed. The reason why Papers [1] and [2] arrived at incorrect conclusions is that they focused only on the forming process without paying attention to the actual results, and relied solely on the results of moment analysis to carry out calculations and comparisons using their respective strength formulas. However, when making comparisons, the fact that different types of end caps have different sets of calculation formulas was overlooked; as a result, for what are actually the same end caps, different formulas were used, leading to significant differences in the calculation results. This is somewhat similar to the situation in opening reinforcement calculations, where there are significant differences between the results obtained using the limit analysis method and the equal area method for the same structure. In summary, the author believes that for approximately elliptical heads manufactured by spinning, as long as inspections are carried out using the templates for standard elliptical heads in accordance with GB 150 and the deviations do not exceed those specified by GB 150, such heads should be regarded as standard elliptical heads and can therefore serve as a complete equivalent substitute for them. Of course, after cold spinning, stress-relief heat treatment should be applied to the head in order to avoid excessive residual stresses within it; when necessary, non-destructive testing should also be carried out on the transition zone [3, 10]. 2 Minimum thickness value for formed heads Since the implementation of GB 150-89, several authors have discussed the provisions in GB 150 regarding the minimum thickness value for formed heads [11–13]. Chinese sources [11]–[12] argue that the minimum thickness δmin of the formed head should be ≥ δ (calculated thickness) + C2 (corrosion margin), rather than δmin ≥ δn (nominal thickness) – C1 (negative deviation of the steel plate); furthermore, the minimum thickness or calculated thickness of the head should be indicated on the diagram, rather than the nominal thickness δn. These analyses indicate that, in accordance with GB 150, in many cases it is not only impossible to fully utilize a company’s processing capabilities and technical advantages; moreover, unnecessary increases in plate thickness are required to meet or ensure the specified minimum thickness, resulting in waste. The analytical approach of these articles is correct and in line with the original intention behind the provisions set out in GB 150; however, they all overlook the fact that some manholes need to be provided on the covers and require reinforcement. It is clearly inappropriate to still use the values from References [11]–[12] at this point. Because when calculating the strength of pressure vessels, the additional amounts or rounded values beyond the designed thickness of the head are usually taken into account in the calculations for reinforcement due to openings. If δmin = δ + C2 is still used in such cases, it is obvious that the additional reinforcement material considered in the calculations cannot be guaranteed. Therefore, it is necessary to determine the minimum thickness of the head through reinforcement calculations; reference [13] discusses this issue precisely. Based on the calculation using the equal area method, this paper derives the formula for calculating the minimum thickness of the head: The meanings of the various symbols in equation (1) are specified in GB 150. As can be seen from Equation (1), for end caps that require reinforcement due to openings, the minimum thickness increases by a value of Δδ compared to end caps without openings: (2) When Δδ < 0, Δδ should be set to 0, meaning that no additional reinforcement area is needed; however, this is not mentioned in Reference [13]. Furthermore, Reference [13] also does not discuss the case of seamless heads, so only one aspect of the problem is taken into account. For these reasons, the author has written a dedicated article to conduct a more in-depth and comprehensive discussion on this issue [14]. By considering the requirements regarding strength and stability, an analysis was carried out on the minimum thickness of seamless heads as well as headed heads with reinforcement openings, resulting in several more accurate and practical recommendations: ① For seamless heads, δmin ≥ δ + C2. ②For perforated heads, δmin shall be calculated according to Equation (1), and δmin ≥ δ + C2. ③The pattern should indicate the minimum thickness δmin, rather than the nominal thickness δn, so that manufacturers can accurately determine the actual margin when selecting the blank thickness, and thus decide on the processing method necessary to ensure this minimum thickness. Determining the minimum thickness of the formed head based on these principles not only ensures that the head after stamping meets the requirements specified in the drawings, but also gives enterprises sufficient flexibility to produce heads that are cost-effective; this is precisely the true intent behind GB 150’s regulations regarding the minimum thickness of heads. 3 Sequence of construction for expansion welding of the tube sheet and tubes There are several ways to connect tubes to the tube sheet: expansion welding, welding, strength expansion + sealing welding, and strength welding + gasketed expansion. 3.1 Weld first, then expand [15,16] In the process of welding first and then expanding, it is easier to clean the grooves on the tube sheet before welding. During welding, the air present in the gap between the tube and the tube sheet can be removed from both the front and back sides, which is very beneficial for preventing pores in the welds and ensuring the quality of the welded joint. At the same time, post-swelling prevents the residual stresses remaining after swelling from relaxing, thus avoiding relaxation caused by the high temperatures of welding. However, for tube-to-tube sheet joints with poor weldability, microcracks are likely to form in the weld seam during expansion bonding, and in some cases the weld seam may even be torn apart. In such cases, deep expansion should be employed (that is, the pipe opening should not expand by more than 10–15 mm), so as to keep the expansion area away from the weld seam and thereby reduce the impact of expansion on the weld seam; this is also the biggest drawback of the process of welding first and then expanding. Experimental studies in Reference [15] show that when the expansion followed by welding process is used, the leakage rate after welding the tube to the tube sheet is about 10 times higher than that when the welding is carried out first and then expansion. Moreover, inspection results indicate that the weld surfaces are uniform, have a metallic luster, and exhibit good shape; there are very few pores and lack of fusion detected through coloring inspections. Therefore, the process of welding first and then expanding is also commonly used abroad. 3.2 Expansion followed by welding [15] The expansion-then-welding process is used; however, since a large amount of dirt such as oil and rust remains at the pipe ends and at the groove areas during expansion, and although cleaning is carried out before welding, it is difficult to ensure thorough cleaning of the grooves due to the narrow pipe bridges and the fact that the pipes extend beyond the tube sheets. During welding, these residual impurities undergo intense chemical changes; water and air expand locally due to the heat, creating pressure in the gaps between the tubes and their holes. As the back side of these areas becomes blocked, the pressurized gases can only escape from the side of the weld seam. Metal that is in a molten state during welding has no strength at all, so gases can easily pass through the weld seam, especially at the end of the weld. The gas rushing out of the weld bead causes the weld metal to boil, resulting in an uneven weld surface that may even appear honeycombed. At the same time, it also causes oxidation of the weld surface, leading to defects such as lack of fusion. During the cooling process of the weld, some gases fail to escape from the surface of the weld in time, thus forming pores inside the weld. Furthermore, the high temperatures generated during welding can cause the already expanded areas to deform, reducing the residual stresses and elastic deformations resulting from the expansion process; this may lead to a decrease or even disappearance of the clamping force. The experimental results in Reference [15] show that the leakage rate of the expand-then-weld process is about 10 times that of the weld-then-expand process. Our long-term extensive production experience has also shown that the method of expanding first and then welding does have many shortcomings, especially when the welding process properties are poor; this is particularly true in cases such as the combination of 20MnMo, 15CrMo, and austenitic stainless steel pipes. Based on the above analysis, although the process of expanding first and then welding can be used, manufacturing practices both domestically and internationally show that the process of welding first and then expanding offers greater advantages. Therefore, the author believes that during design and manufacturing, the process of welding first followed by expansion should be given priority; in cases where the weldability of the tube material is poor, a region of 10–15 mm without expansion can be left at the tube end. 4 Material substitution Material substitution is inevitable in manufacturing. Although the \"Regulations on Safety Supervision of Pressure Vessels\" lay down some rules regarding material substitution, these are only principle-based; in practice, the problems that arise as a result of such substitution often go unnoticed. 4.1 Better in Place of Worse and Thicker in Place of Thinner The so-called \"better in place of worse\" refers to the use of materials of a higher grade or with better performance to replace those of a lower grade or with poorer performance ； Thick-for-thin substitution involves using material of a thicker grade of the same steel type in place of material of a thinner grade. The aforementioned substitutions are often considered to be feasible at will, thereby ignoring the existence of related issues, which sometimes even turn into serious problems. When thick plates are used in place of thin ones, it often requires changes to the connection structure; for example, in the connection between a thickened head and the cylinder, the head usually needs to be trimmed on the outside. For equipment with a cylinder made of pipes, when the cylinder wall is thickened, the joint between the cylinder and the end cap sometimes also requires internal trimming on the side of the cylinder. These same problems exist in the butt-welded joint structure of the cylinder with the tube sheet and flat cover. When the thickness increases significantly, changes in the welding structure often occur as well; for example, the welds between the nozzles and the shell, as well as the butt welds, may change from a single V-groove to an X-groove. Another issue with using thicker plates instead of thinner ones is that it may lead to insufficient strength [17]. This is because as the plate thickness increases, the allowable stress of the material tends to decrease. For example, in the case of 16MnR, when the thickness increases from 16 mm to 18 mm, the allowable stress drops from 170 MPa to 163 MPa ； When 20R increases from 16 mm to 18 mm at 100 ℃, the allowable stress decreases from 132 MPa to 126 MPa. This requires special attention during the manufacturing of heads, as a certain amount of extra thickness is often added to the blank when cutting the heads in order to ensure the minimum thickness of the head after stamping; this can result in insufficient strength of the stamped head [17]. Therefore, when thick layers are over thin layers under these critical conditions, the strength must also be verified. For components such as expansion joints, bellows, flexible thin sheet plates, and thin sheet plates, it is generally not advisable to use a thicker material instead of a thinner one; as the thickness of the component increases, its rigidity also increases, thereby reducing the effectiveness of deformation compensation. The opening-reinforcing plate should not be made excessively thick either, as a too-thick reinforcing plate will cause high stress concentration at the periphery where it connects to the cylinder, leading to cracking at the weld toe. It can be seen from this that using thicker layers instead of thinner ones is not always beneficial; therefore, special attention should be paid to the aforementioned issues when making such substitutions. In general, selecting the better option over the worse is always beneficial, as it can increase the strength reserve of equipment and enhance its safety and reliability; however, this is not the case in some special processing environments. In wet H2S environments, it is clear that carbon steel offers better resistance to H2S-induced SCC compared to materials such as 15MnVR, 16MnR, and 20MnMo [18]. The same is true for liquid ammonia environments; since SCC also occurs in 16MnR under such conditions, using low-alloy steels such as 16MnR instead of steels from the 20R, 20g, and Q235 series in these environments makes it more likely to encounter problems. Secondly, for cases where the design requires a low yield-to-tensile strength ratio of σs/σb, attention should also be paid to the principle of replacing inferior options with superior ones, such as large-opening reinforcement structures, reinforcement plate structures, and reinforcement structures designed using the limit method. Another issue to consider when using materials with a higher strength grade as substitutes is weldability, as generally the higher the strength grade, the worse the weldability. Using an even higher-grade material as a substitute in such cases will make welding even more difficult. Furthermore, for components such as expansion joints, rupture discs, and flexible tube sheets, it is not permissible in principle to use inferior materials in place of superior ones; otherwise, recalculation must be carried out using alternative materials, with their thicknesses appropriately reduced, as failing to do so may lead to the failure of these components and the adjacent areas. In addition to the issues mentioned above, those who choose alternatives should also consider the cost-effectiveness of the equipment, as such choices will undoubtedly increase the cost of the equipment; especially when it comes to selecting materials for the alternative equipment, it is necessary to weigh the pros and cons carefully. 4.2 Other issues ① The welding process should be modified accordingly based on the actual material used. ②When higher-grade materials are used in place of lower-grade ones, the testing and acceptance procedures shall still be carried out in accordance with those for lower-grade materials, without the need to raise the acceptance standards. ③Different materials have varying reserves of low-temperature toughness, and accordingly the minimum hydrostatic test temperature may change; in such cases, it is necessary to strictly follow the provisions of GB 150. Furthermore, if the plate thickness exceeds the thickness of cold-rolled sheets specified in GB 150, stress-relief heat treatment must be applied to the cylinder. Ultrasonic testing is also required when the steel plate reaches a certain thickness. If necessary, the hydrostatic test pressure should also be increased appropriately; in some cases, this even leads to significant changes in the equipment’s structure. 5 Hydrostatic test The hydrostatic test, as the final step in equipment manufacturing, not only checks the strength of the equipment but also evaluates the density of the welds or the tightness of the sealing structures. It can also reduce or eliminate residual stresses and blunt the tips of defects (cracks), thereby preventing the propagation of cracks at lower operating pressures or slowing down their propagation rate, thus increasing their lifespan. It can significantly enhance its fatigue life at a reasonable overload ratio, as well as improve the load-bearing capacity of pressure vessels; the burst pressure will increase markedly [19–21]. It is evident that pressure testing is of great significance, playing an important role in the safe use of pressure vessels. However, due to the insufficient rigor of the regulations specified in GB 150-89, some misunderstandings have arisen in its application, resulting in pressure testing becoming meaningless in certain cases. Furthermore, regarding the pressure testing of jacketed equipment, the provisions in the standard specifications appear to be too arbitrary, causing unnecessary difficulties in manufacturing. Furthermore, it is also a question worth exploring whether the test pressure should be adjusted when the shell material is replaced. 5.1 Values of test pressure and limits on test stress [22] GB 150-89 specifies that the hydrostatic test stress should be: p_test = 1.25 × p_design [σ]/[σ]t (3) When a vertical container is tested while placed horizontally, the static pressure of the liquid column during the test must also be taken into account; thus: p_test = 1.25 × p_design [σ]/[σ]t + γH (4) GB 150-89 requires that the test stress be increased to 1.25 times the design stress in order to assess the strength of the equipment, the integrity of the sealing structures, and the quality of the welds. The analysis in Ref [22] shows that the value calculated using Equation (3) represents only the minimum requirement for hydrostatic testing. For vertical containers containing both liquid and gas or being filled with liquid, Equation (3) is insufficient to meet the requirements; in such cases, the calculated pressure given by the following equation should be used instead of the design pressure: p_calculated = p_design × γ′h (5), where γ′ is the density of the liquid material ； h is the filling height of the liquid material. If the container is pressurized while lying horizontally at this time, its test pressure should be [22]: p_test = 1.25 × p_calculated [σ] / [σ]t – γH, where γ is the density of the medium used for the hydraulic test ； H is the total height of the empty container [22]. Clearly, equation (6) is more appropriate than equation (4), because using equation (4) to calculate the test pressure may result in an increase in the thickness of the equipment required to meet the hydrostatic testing requirements. It is therefore reasonable to use p instead of p_set for calculating the test pressure. ASME [23], BS 5500 [24], AD [7] etc. also use p for such calculations. Although ASME uses p_set as well, it is only applied to vertical containers that are not filled with liquid, and it is not relevant to cases involving the static pressure of a liquid column. Looking at the standards abroad, the test pressures specified there are often very high, and the actual test stress multiples also approach the values prescribed by the standards. As specified in BS 5500, its p-value is given by equation (7), where S is the nominal thickness and C is the additional amount. To truly put the container to the test, the margin built into the shell design is used to withstand increased test stresses. It can be seen that the regulations in various countries set the test pressure specified in the standards as a lower limit; basically, p is used in place of p_set, and the actual test pressures are quite high. The purpose is to ensure that the test stress is several times higher than the design stress in order to test the container, meaning that the actual test pressures are generally greater than or equal to p_test. However, GB 150-89 does not seem to be very clear, to the point that many people in practice believe that the calculated values specified in GB 150-89 should not be exceeded. Moreover, GB 150-89 specifies that pressure testing should be carried out using the p value calculated accordingly; in low-pressure high-tower equipment, the thickness of the equipment is often increased to meet the requirements for pressure testing, which is clearly unscientific. Some of the aforementioned issues have been revised in the published GB 150-1998［25］, but there still seem to be some problems regarding the regulations for pressure testing vertical vessels when placed horizontally. There are many articles analyzing and discussing the issue of stress control during pressure testing; it is generally believed that calculating the test pressure according to GB 150-89 eliminates the need for stress verification, as the test stress cannot exceed 0.9σs or 0.8σs. However, in their analyses, these articles merely discuss the test pressure by using the calculated values specified in GB 150-89. With the implementation of GB 150-1998, there is a trend toward higher actual test pressures; as a result, under these actual test pressures, the test stress may exceed 0.9σs or 0.8σs, thereby affecting the safety of the equipment. Looking at the regulations in other countries, they have also imposed restrictions. Although ASME VIII-1 does not impose direct restrictions [23], it states that if the test pressure calculated according to its formulas is exceeded, intentionally or unintentionally, resulting in significant permanent deformation (plastic deformation) of the vessel, the inspector has the right to refuse acceptance; thus, restrictions are still imposed on the test stress in practice. Therefore, the author believes that the test stress should still be limited in accordance with GB 150-1998. To increase the test pressure, the hydraulic test pressure can be calculated using the formula derived by the author when necessary [19], in order to ensure that the test stress σT ≤ 0.9σs; see equation (8) for the meaning of the symbols, as described in reference [19]. 5.2 Adjustment and improvement of the pressure testing process for jacketed equipment [26] In accordance with GB 150 and the \"Regulations on Safety Supervision of Pressure Vessels\", for jacketed equipment, the inner cylinder should be assembled, welded, and tested for pressure; only after this is successful should the jacket be assembled and tested as well. If necessary, pressure retention on the inner cylinder is also required during the jacket’s pressure testing (as determined by the design). However, due to the limitations of the actual structure, the inner cylinder often still needs to be pressure-tested again after the jacket is welded, so jacketed equipment usually requires three pressure tests. However, due to the wide variety of jacketed equipment and their different structural designs, for some of these designs, it is entirely unnecessary to perform 3 pressure tests. As shown in Figure 1a and Figure 1b, with this jacket structure, it is possible to inspect the inner cylinder during pressure testing; as long as the jacket pressure is higher than that of the inner cylinder and no pressure needs to be maintained in the inner cylinder during jacket testing, it is entirely feasible to weld the inner cylinder and the jacket together first and then conduct the pressure test. At this point, the jacket should first be pressurized and the inner cylinder enclosed by the jacket inspected; thereafter, the inner cylinder should be pressurized and the parts outside the jacket inspected. Of course, in cases where the pressure inside the inner cylinder is lower than that in the jacket and the inner cylinder still needs to maintain its pressure during jacket pressure testing, it is still necessary to follow GB 150 and the \"Regulations on Safety Inspection of Pressure Vessels\" in principle (except in cases where support rings are used internally for pressure testing to meet stability requirements). When the structure shown in Figure 1 is assembled according to the procedure described by the author in the text and a pressure testing process is carried out, it not only shortens the manufacturing cycle but also reduces the number of pressure tests required. This leads to lower manufacturing costs while still achieving the purpose of pressure testing; therefore, this approach is feasible and acceptable, and it is recommended that the pressure vessel inspection authorities approve it. Figure 1 Jacket structure 5.3 Changes in test pressure values when the main material is replaced In manufacturing, it is common to replace the main material; when a higher-grade material is used in place of a lower-grade one, or thick plates are used in place of thin ones, I believe that the test pressure should be adjusted accordingly: ① By increasing the test pressure, it is possible to truly achieve the purpose of testing and evaluating the container. ②The degree of residual stress elimination during equipment manufacturing can reach 70%–80% [21]. ③It can improve the blasting limit and fatigue life, etc. [19–21]. On the contrary, some devices have very low actual stress levels under test pressure due to reasons such as the choice of materials or the use of substitute materials; these levels are often far below the allowable stress. In such cases, if the test pressure is not increased, the purpose of pressure testing cannot be achieved, and thus the meaning of pressure testing is lost. As can be seen from certain provisions of ASME and BS 5500 mentioned earlier, a thickness allowance is even included when calculating the test pressure, that is, the test pressure is increased by a factor of S/(S-C). For low-pressure, small-diameter equipment, the K value can sometimes be as high as 2–3; therefore, the actual testing pressure may be 3–4 times the design pressure to ensure that the true purpose of the pressure test is achieved. The results of the numerical examples in Author’s paper [19] also confirm this. Therefore, it is recommended that for cases where the main material is of lower quality, where the thickness is reduced, or in the case of some small-diameter low-pressure equipment, factors such as allowable stress and plate thickness variations should be taken into account to appropriately increase the test pressure (ensuring that the test stress does not exceed 0.9σs or 0.8σs), which is beneficial for the safe use of such equipment. 6 Conclusion The author has conducted a thorough analysis and discussion of several common issues in pressure vessel manufacturing, from a theoretical to a practical perspective. Based on an analysis of both the positive and negative aspects of these issues, and adhering to the principles of safety, economy, and rationality, the author has summarized these problems and presented his own views and recommendations. The aim is to enable relevant engineering and technical personnel to view, understand, and address these issues from a broader perspective, thereby achieving a unified understanding. There may be biases in the text; I earnestly ask readers for their understanding and corrections.