B. N. Acharya and N. PAL, Contributing Consultants, Mumbai, India; K. DARU, Air Products and Chemicals Inc., Houston, Texas; and A. PANCHAL, PVA Systems, Mumbai, India
Expansion joints are used throughout industry for various purposes. One main purpose is to compensate for relative movements in heat exchangers and piping caused by temperature differences in shell-and-tube parts of the equipment or piping—without the use of expansion joints, it would be impossible to install such equipment or piping in those service conditions. These expansion joints are also used to isolate vibration or mechanical movements from rotating machines to stationary equipment or machines, and to allow vertical movement of equipment connected to piping placed on load cells.
This article focuses on flanged and flued (thick) expansion joints used in a shell-and-tube heat exchanger constructed to Boiler and Pressure Vessel (BPV) Code ASME Sec. VIII, Div. 1. The scope of Mandatory Appendix 5 in ASME BPV Code states that the minimum thickness of a flanged and flued expansion joint should be 3 mm; however, the higher thickness limit is left to interpretations. This article discusses a case of a shell-and-tube heat exchanger that was constructed with an expansion joint made of a 60-mm thick plate and an alternative expansion joint providing a cost-effective and reliable solution.
The thickness of the expansion joint and the configuration of its convolution (radii of inner and outer torus) provide flexibility (i.e., stiffness), which ultimately determines the number of convolutions for the joint to withstand the load/s occurring due to differential thermal expansion between shell and tubes. FIG. 1 shows expansion joint convolution details and geometric proportions.
According to FIG. 1, it is common to provide expansion joint thickness equal to the adjoining shell thickness. It is not required to keep the shell and expansion joint thicknesses the same, provided the requirements of Para. 5-3(f) of Div. 1 are met.
The following reference is from ASME Sec. VIII-1:
Extended straight flanges between the inner torus and the shell and between both outer tori are permissible. An outer shell element between the outer tori is permissible. Extended straight flanges between the inner torus and the shell, between the outer tori and the outer shell element, and between both outer tori that do not have an intermediate outer shell element with lengths in excess of 0.5 √ R × tf satisfy all the requirements of UG-27, where:
R = inside radius of expansion joint straight flange at the point of consideration
= Ra or Rb
tf = uncorroded thickness of expansion joint straight flange.
From this information, it can be concluded that an expansion joint with a flange length ≤ 0.5 √ R × tf can have a thickness less than the shell, provided all other requirements of Appendix 5 are met. The response below from the ASME Code Committee to the authors’ inquiry further clarifies this point.
Standard designation: BPV Sec. VIII Div. 1
Edition/Addenda: 2015
Para./Fig./Table number: Mandatory Appendix 5
Subject description: Flanged and flued expansion joint thickness
Date issued: 05/31/2017
Record number: 17-577
Interpretation number: BPV VIII-1-17-27-R
Question(s) and Reply(ies): Question: For an expansion joint designed in accordance with Mandatory Appendix 5, may the thickness of the flexible elements be different from either the shell attached to the inner torus or a shell element attached between the outer torii, or both, provided the requirements of 5-3(f) are met? Reply: Yes.
Considering this information, an expansion joint can be designed and constructed as per FIG. 2.
Based on FIG. 2, an expansion joint flange can be thinner from the end of the 1:3 taper provided by weld built-up (or equivalent). A taper transition of 1:3 can be accommodated in the flange length (Lf), and the remaining flange length can be of smaller thickness.
For example:
Provided the length Lf < 51.96 mm, the flange thickness tf does not need to meet UG-27.
The thickness difference between the shell and expansion joint flanges is tsh – tf = 30 – 18 = 12 mm, and the distance required to maintain the 1:3 weld built-up taper is 3 × (tsh – tf) = 12 × 3 = 36 mm, which can be easily accommodated within the flange length without affecting the stiffness of the expansion joint.
However, in many practical cases, where the shell-side design pressure is higher and the shell thickness is > 30 mm—certainly, this is not the maximum limit for the constructability of the expansion joint—the expansion joint size (inner and outer torus radii and, accordingly, the overall dimensions) is expected to be larger. Concurrently, the thickness of the expansion joint elements would also be heavier. This causes the stiffness of the flexible element to decrease (since these properties are inversely proportional) and, therefore, will require a larger number of convolutions to compensate for the same amount of differential thermal expansion between shell and tube.
As the thickness increases, the spring rate decreases; so, to compensate for the same amount of differential thermal expansion, the expansion joint will require an increased number of flexible elements compared to one with a thinner-walled expansion joint.
It is preferred that the expansion joint is provided for operating load cases and should be avoided for upset and maintenance conditions, whenever possible. For example, the tube-side steam-out condition is often specified for evaluating the expansion joint requirements. For the tube-side steam-out condition, the tube mean metal temperature is considered as the steam temperature, while the shell mean metal temperature is considered as the ambient temperature—for most of the combination of shell-and-tube material, an expansion joint will be required. If the shell is thicker, then the situation will deteriorate and require many convolutions (discussed later in this article).
The authors were involved in a project in which the detailed engineering company provided a standard list of load conditions, and the shell-side and tube-side steam-out conditions were part of those conditions. It was surprising to note that most of the exchangers required expansion joints and, in most of those, three to four sets of flexible elements became necessary. After consultation with the plant operations team, the steam-out sequence was set so that steam will be simultaneously introduced on both the shell and tube sides, ensuring the same mean metal temperature and resulting in essentially close to nil differential thermal expansion. Based on this, the expansion joints were removed from most of the heat exchangers.
The authors also encountered a practical case where a heavy expansion joint was provided on a heat exchanger with a 55-mm shell-side thickness. The expansion joint thickness was kept at 55 mm (like the shell thickness) and designed in accordance with Mandatory Appendix 5, which required five sets of flexible element convolutions consuming more than half of the shell length. FIG. 3 is a 3D representation of a heat exchanger as per as-built dimensions—some of the external accessories are not shown for better clarity. FIG. 4 shows a close-up view of the expansion joint.
FIG. 3 illustrates a 55-mm thick expansion joint with five convolutions. The construction of the convolutions required a starting plate thickness of 60 mm (considering a thinning allowance during flanging) that is also in circular rings, leaving considerable material wastage from a rectangular plate. The joint convolutions were fabricated by a flanging operation of 20 edges and two transition pieces. Subsequently, post-forming normalizing heat treatment was performed on the formed convolutions. After completion of the austenitizing heat treatment, the convolutions were welded together, followed by non-destructive examination of 10 numbers of circumferential weld seams. This directly impacted the fabrication cost of the exchanger. After analyzing similar types of exchangers operating in a plant at a different location with nearly the same amount of differential thermal expansion, it was found that exchangers with two convolutions of toroidal bellow are satisfactory.
A toroidal expansion joint is permitted as per Mandatory Appendix 26. Due to the tubular construction, it can withstand high pressure with thinner walls. FIG. 5 shows the construction of the toroidal expansion joint, illustrating the toroidal flexible element and collar (stub-end) on both sides to be welded to the shell.
None of the engineering publications above state what the shell thickness of a toroidal bellow should be considered in lieu of a flanged and flued-type expansion joint. However, based on the authors’ engineering experience and confirmed by the performance of similar heat exchangers in operating units, when the shell is heavy—it is difficult to quantify the shell thickness limit, but as a general guide, for a shell thickness heavier than 40 mm—the use of a toroidal expansion bellow in place of a flanged and flued expansion joint should be evaluated. Toroidal expansion joints have been used on shell thicknesses up to 90 mm in some cases and have been operating successfully.
Takeaway. When a flanged and flued (thick) expansion joint is used, it does not need to be as thick as the shell, and its thickness can be less than the shell thickness, as explained.
However, when the shell is heavy (or heavier than 40 mm), a toroidal expansion joint will serve the same purpose, providing safe and reliable performance and many other advantages, including cost and good engineering practices. HP
B. N. ACHARYA has work experience in static equipment design with UHDE India Ltd. and Reliance Industries Ltd., India. With more than 40 yr of experience in the design and troubleshooting of heat exchangers and high-pressure reactors, he has been active in executing greenfield, revamp and modernization projects. Acharya holds a BS degree in mechanical engineering and an MS degree in thermal engineering from the Indian Institute of Science, Bangalore.
NITISH PAL has more than 40 yr of experience in the design, detailing and troubleshooting of various process equipment in refining, petrochemical, fertilizer and pharmaceutical plants. His expertise includes executing lump-sum turnkey projects and project management, and he is a specialist for double-wall storage tanks and air-cooled heat exchangers. Pal is a graduate in chemical engineering from Jadavpur University, India, and his work experience includes Richardson and Cruddas, India, Toyo Engineering India Ltd. and Reliance Industries Ltd.
KUNTAK DARU is a Pressure Vessel Engineer with Air Products and Chemicals Inc. He works in the mechanical equipment division of Air Products, supporting engineering and design activities of AP projects worldwide from the company’s Houston, Texas location. Prior to working with Air Products, Daru was a Technical Director in mechanical engineering at Fluor Corp. He has more than 30 yr of experience in the design, engineering and specification of static equipment in the refining, oil and gas, petrochemical and chemical industries. Daru has been published more than 15 times in his field and holds a BS degree in mechanical engineering from SVNIT in Surat, India. Daru is a Registered Professional Engineer in the states of Texas and Louisiana, U.S. The author can be reached at Darukm@airproducts.com or Kuntak.Daru@Gmail.com.
ANILKUMAR PANCHAL has more than 15 yr of experience as a Mechanical Design Engineer, including hands-on work experience in pharmaceutical solid and semi-solid product plants, butyl rubber, gasification, ethane cracker, chlorinated PVC and various revamp projects. He holds a BS degree in mechanical engineering and an MS degree in industrial process equipment design. Panchal is a certified Chartered Engineer and Professional Engineer in India, and has work experience at L&T Heavy Engineering, Bechtel India Ltd., Jacobs Engineering, Reliance Industries Ltd. and KIPIC, Kuwait. The author can be reached at anil0208@gmail.com.