Mechanical Design of Shell and Tube Heat Exchanger

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Heat exchanger is a device, which is used for transfer of heat from fluid to another, usually separated by walls. Whenever a temperature gradient exists between two fluids, energy is transferred by heat transfer process.


Based on the applications, heat exchangers are classified as boilers, condensers, heaters, coolers, recuperators etc. Depending on the configuration of fluid flow paths, heat exchangers are classified as parallel flow(co-flow) heat exchangers, counter current(counter flow) heat exchangers and single pass cross flow heat exchangers and multi pass cross flow heat exchangers. The most important difference between the four types lie on the relative amount of heat transfer surface area required to produce a given temperature rise for a given temperature difference between the two fluid streams where they benter the heat exchanger. Heat exchangers are employed in varied installations such as steam power plant, chemical processing plants, building heating, air-conditioning, refrigeration system etc to carry away the heat carried by the gases and it cools the gasses to a sufficiently low temperature, using a suitable fluid.

Regenerative Heat Exchangers

In Regenerative Heat exchangers the hot and cold fluid flow through the one and same passage and heated surface is alternately exposed to the hot and cold fluids. If the periods of heating and cooling are of equal duration, continuous heating requires two apparatus in which the hot fluid is cooled in one apparatus the cold fluid is


heated in the other. After that apparatus is switched over and heat transfer process is reversed. In Regenerative exchangers the process of heat transfer is transient. The temperature varies as it cools or heats variation in wall temperature is accompanied by change in fluid temperature with time and along the heating surface as well. Regenerators are mainly used in the branches of industry where there is waste gas at high temperature and is required to heat air at a high temperature i.e. blast furnaces, open hearth furnaces, coke and glass manufacturing. The performance of regenerators depends on many factors as thickness of packing, its conductivity, accumulating capacity, duration of periods and fluid temperature. While in operating conditions the heat transfer co-efficient may vary due to the burning of gas in the regenerator.

Recuperative heat exchangers

The two fluids performing the exchange of heat in the exchangers can flow (a) with each other in the same direction (parallel flow) or in opposite direction (counter flow) or (b) at right angles to one another (cross flow) with both types of flow a single or multi pass arrangement is possible. The element from which a recuperative matrix is built up is mainly of two kinds, the fluid flows along the tubes on the inside and along across the tubes on the outside. The fluid flows between consecutive plates arranged at a certain distance apart. To reduce the equivalent diameter, the flow channels between the plates may be sub divided in different ways by a folded or corrugated plate arranged between any two parallel plates and thus forming a multiplicity of parallel flow channels, the shapes of which depend on the kind and shape of corrugations. Heat transfer and flow resistance depend on size, shape and


arrangement of above elements. Any reduction in diameter leads to an increase in the number of tubes for given mass flow and requires new methods for fixing in tubes in plates of the headers. Further reduction in the equivalent diameter of flow channels about 6mm to 3mm leads to the use of plate type matrix


Mixed type Heat Exchangers

The direct contact type heat exchanger is one in which the two fluids are not separated from one another. If heat is to be transferred between gas and liquid, the gas is either bubbled through the liquid or is sprayed in form of drops through the other. In this heat transfer takes place with mass transfer. The heat is carried by the evaporation of cooled water carried with air. Its performance not only depends upon the temperature difference but also on relative humidity of air. Common examples of this type are feed heaters, cooling towers and evaporative condensers. The influence of flow path arrangement on heat transfer area is dependent on the temperature rise to be achieved for a given inlet temperature difference the flow path arrangement does not affect heat transfer area. Parallel flow arrangement should be restricted to this area. The counter flow arrangement requires least area throughout range while cross flow arrangement requires slightly larger than counter flow but is much better than parallel flow arrangement. Whenever temperature changes in one or both fluid streams closely approach the temperature difference between the entering fluid systems only counter flow arrangement has to be employed. The cross flow arrangement is classified as single pass and multi pass cross flow. If both fluids traverse the exchanger only once; the arrangement is called single pass heat exchanger. If the fluids are made to shuttle back and forth across the heat transfer


matrix more than once that arrangement is called multi pass heat exchanger. If the fluid path on hot side is so arranged as to make two passes and on the cold side 4-passes, then the heat exchanger is called 2-4 heat exchanger.


Heat exchangers are practical devices used to transfer energy from one fluid to another to get fluid streams to the right temperature for the next process to condense vapours, to evaporate liquids, to recover heat to use elsewhere and to drive a power cycle.

Shell and Tube Heat Exchangers:

Shell and tube heat exchangers come in a wide range of sizes and lengths to many needs. This large unit features tantalum sheet and tubing to provide long life and corrosion resistance in severe environments. They are widely used in process industries, especially petro chemical and petroleum refineries, the use of shell and tube heat exchangers range from chillers heat removers etc to reboilers, process steam coolers etc. Also there are no moving parts. The shell and tube heat exchangers are grouped into three as “R” heat exchangers, “C” heat exchangers and “B” heat exchangers, according to service standards set by TEMA Class “R” heat exchangers are designed for severe service requirements class “B”, for moderate service requirements and class “C” for chemical process service.Although copper tubes and steel shells are the most common materials of construction a wide range of metals are available for handling various fluids and gasses as brass, aluminium, stainless steel, titanium and other alloys. Some application use glass or plastic tubes to resist the attack of extremely corrosive substance or to avoid affecting the


flavour of food.The recent innovations in heat transfer technology had led to greater efficiency of shell and tube type heat exchangers. Multiple pass, multiple module constriction help to achieve a significant amount of heat transfer in a limited amount of space. Even though sufficient space must be left for the cleaning of tubes and the removal of tube bundles for repair, the units can be located just about anywhere. The shell and tube heat exchanger is shown in figure 1.

Figure 1. Shell & Tube Heat Exchanger

Multiple pass, multiple module constriction help to achieve a significant amount of heat transfer in a limited amount of space. Even though sufficient space must be left for the cleaning of tubes and the removal of tube bundles for repair, the units can be


located just about anywhere.

The main components of shell and tube heat exchangers as mentioned above are the shell, tube bundle, tube sheets, baffles, channels, flanges and nozzles.


The shell consists of a cylinder made from seamless pipe rolled and welded with a bolting flange at each end. It is often designed so as to withstand a pressure, one and half times its rated pressure. Shells are often designated by letters E, F, G, H, J, K and X.

Tube bundle:

The tube bundle is made of tubes, tube sheet and cross baffles. Different types of tube configurations are available. One common type is the U-tube configurations which is the most economical. It has the fewest components one head assembly, one tube sheet and a shell with flange opening to accept the tube bundle on one end. Even though the arrangement features a removable, replaceable bundle, it is difficult to clean the tubes mechanically. Efforts to overcome this disadvantage resulted in the introduction of straight tube configuration. This configuration consists of a shell assembly with a flange on each end and the tube is fixed at the both ends. The unit is durable, can handle higher pressures and easy to maintain. The major disadvantages include the inability to tolerate large temperature difference between the shell and tube side fluids, the failure or breaking of tubes from the tube sheets due to the differential expansion and the nonreplicable bundle. Hybrid designs are also available to overcome many temperature and pressure drawbacks. The complete


bundle can be removed in a hybrid heat exchanger. The most recent development in shell and tube designs has the introduction of a double wall construction. The outer tube is rolled into one header and the inside one extends past the outer tube and is rolled into a second header. The failure of the tube can be detected by visual inspection or an electronic monitoring device. The double wall construction offers significant protection and safety.

Tube Sheets:

Tube sheets are used to keep the tubes in position. The tubes can be either square pitched or triangular pitched. Due care must be taken in the design of tube sheets as it is affected by longitudinal stresses in shell and tube, tube compressive stress, tube to tube sheet joint loads etc.


Baffles are used to induce turbulence outside the tubes as turbulence increases the heat transfer coefficients. Baffles cause the liquid to flow through the shell at right angles to axes of tubes. The centre to centre distance between baffles is called baffle pitch or baffle spacing. There are several types of baffles like disks and doughnuts, orifice, strip and segmental of which segmental baffles are most commonly used. Segmental baffles are formed by cutting a segment from a disc. Segmental baffles are drilled plates with heights which are generally 75% of the inside diameter of the shell and these are also called 25% cut baffles. The cut portion of baffle is often called the window section. Baffle is efficient and gives good heat transfer rates for pressure drop and power consumed



Flanges are used on the shell of a vessel to permit disassembly and removal or cleaning of internal parts. They are also used for making connection for piping and nozzle attachments. The standard types of flanges for different pressure ratings are welding neck type, slip on type, screwed type, lap joint blind type etc.


The functions of a gasket are to interpose a semi plastic material between the flange facings, by which the material seals (through deformation under load) the minute surface irregularities to prevent the leakage of the fluid. The amount of force required for this purpose is known as yield or seating force. They are of different types; the most commonly used are fabricated with a metal jacket and a soft filler (usually of asbestos). Such gaskets can be used up to temperatures of about 8500 F and require comparatively less bolt load to seat and keep tight.


Channel is a tube side component. It has also got a cylindrical section. The tube side flows through the channel. Partition plates are made use for multipass flow. The effective thickness of the channel cover will be the thickness measured at the bottom of the pass portion groove minus tubeside corrosion allowance.


In the case of heat exchangers, nozzles are the pass ways for the in and out flow of the hot and cold fluids. The strength of shell will be reduced due to the drilling of holes for the insertion of nozzles. If the thickness of the shell is not sufficient to withstand the differential stress thus developed additional metallic plate must be introduced in order to reinforce the shell structure.



The ASME boiler and pressure vessel code (BPVC) section VIII deals with the“Rules for construction of pressure vessels” (unified). ASME BPVC section VIIIcomprises of the following three divisions.

Section VIII : Rules for construction of pressure vessels Division 1 : Rules for construction of pressure vessels Division 2 : Alternate Rules

Division 3 : Alternate rules for construction of high pressure vessels

The ASME BPVC Section VIII, division is the most widely used code for the design and construction of pressure vessels.


The ASME BPVC Section VIII, division 1 adopts the design by formula (DBF) approach. Division 1 uses approximate formulas, charts, and graphs in simple calculations, applies a higher factor- of- safety (resulting in lower allowable stresses), is more tolerant to fabrication techniques and fabrication defects as compared to other divisions of this section. The design basis of division 1 is the “maximum


principal stress theory”. The ASME BPVC Section VIII, division 1is comprises of an introduction,3 subsections, 34 mandatory appendices and 22 nonmandatory appendices.The sructure of division 1 is as follows.

Subsection Pertaining to

Introduction Defines the scope, establishment of design requirements, responsibilities of manufacturer and authorized inspector, standards referenced by this code and units of measurement

Subsection A General Requirements

Part UG General methods of all methods of construction and all materials

Subsection B Requirements pertaining to methods of fabrication of pressure vessels

Part UW Requirements of pressure vessels fabricated by welding

Part UF Requirements of pressure vessels fabricated by forging

Part UB Requirements of pressure vessels fabricated by brazing

Subsection C Requirements pertaining to classes of materials

Part UCS Requirements of pressure vessels constructed of carbon and low alloy steel


Part UNF Requirements for pressure vessels constructed of non ferrous materials Part UHA Requirements for pressure vessels

constructed of high alloy steel

Part UCI Requirements for pressure vessels constructed of cast iron

Part UCL Requirements for pressure vessels constructed of materials with corrosion resistant integral cladding

Part UCD Requirements for pressure vessels constructed of cast ductile iron

Part UHT Requirements for pressure vessels constructed of ferrite steels with tensile properties enhanced by heat treatment Part UHX Rules for shell- and- tube heat


Mandatory Appendices Mandatory Appendix 1 through 34 (34 nos.)

No mandatory Appendix No mandatory Appendix A, C, through H, K, L, M, P, S, T, W, Y DD, EE, FF, GG, HH and JJ (22 nos.)


The following pressure vessels are included in the scope of division 1. Vessels designed for pressure above 15 psig (1.0546kg/cm2 = 1.0342 bar) and not exceeding

3000psig (210.915 kg/cm2 = 206.84 bar). Vessels having inside diameter above

6 inches (150 to 40mm). Unfired steam boilers, evaporators, heat exchangers. The following pressure vessels are excluded from the scope of division 1. Vessels covered by other sections. Pressure containers, which are integral part of rotating machinery. Piping system and components beyond battery limits. Vessels for human occupancy.

Organisation of the ASME boiler and pressure vessel codes:

The ASME BPVC is divided into many sections, divisions, parts, subparts. Some of these sections relate to a specific kind of equipment and applications; others relate to a specific materials and method of application and control of equipment; and others relate to care and inspection of installed equipments.

Section Title

Section I Rules for construction of power boilers

Section II Materials

Part A Ferrous materials

Part B Nonferrous materials

Part C Specification for welding rods, electrodes, and filler metals

Part D Properties

Section III Nuclear power plant components

This section is further divided into subsection NCA, Division I, Division II


and Division III.Division I is divided into subsections NB, NC, ND, NE, NF, NG and NH

Section IV Recommended rules for care and operation of heating boilers

Section VII Recommended guidelines for the care of power boilers

Section VIII Rules for construction of Pressure vessels. Division 1,Division 2:Alternate Rules, Division 3:Alternate Rules for construction of high pressure vessels Section IX Welding and Brazing Qualification

Section X Fibre reinforced plastic pressure vessels Section XI Rules for in-service inspection of nuclear

power plant components

Section Rules for construction and continued service of transport tanks

TEMA Standards

The most widely used consensus standard heat exchanger manufacture is the “Standards of Tubular Manufacturer’s Association”. In short the TEMA standards first published in 1941, this standard had evolved into something of an international document. Many countries have accorded it he status of their international codes.


TEMA standards specify three classes of construction namely TEMA-R, TEMA-C, and TEMA-B. The formulas for determining thickness are the same for all TEMA classes; however empirical guidelines for sizing no pressure part items vary. TEMA-R, which specifies the most rugged construction, is widely used in refinery service and nuclear power plant applications. TEMA-C and TEMA-B are used in other industries. TEMA-B has been promulgated as an American National Standard (ANSI B-78).

TEMA Nomenclature

As per TEMA, the STHE is divided into three parts, the front head (stationary head), the shell and rear head (stationary or floating). Exchangers are described by alphabetic codes for the three sections.

Front (stationary) head type : A, B, C, N & D

Shell type : E, F, G, H, J, K & X

Rear head type : L, M, N, P, S, T, U &

The sequence of designating the shell and tube heat exchanger is: first the front (stationary) end – then the shell – and finally the rear end. Various combinations like AES, AEP, CFU, BEM, AKT, AJW, etc are possible. Each of these types has their relative merits and demerits. The one most suitable for the specific service is selected by considering the pros and cons of various constructional features



In designing the heat exchanger, the following requirements were established. Eliminate or at least eliminate fouling by not allowing the product to stick to the heated or cooled surfaces, be opened easily and cleaned thoroughly, eliminate leaking gasket and withstand high pressure.

The shell and tube heat exchanger meets the non fouling requirements by permitting the product liquid flow to be set at velocity that can avoid or at least substantially minimise any deposit even if it means large pressure drop. To prevent any potential deposits from lodging the pockets, corners, crevices and zones, where the velocity cannot be strictly controlled were eliminated. During the washing periods the washing liquids should penetrate thoroughly.

The opening and closing operations were simplified so that even inexperienced workmen could do them easily. The elements of unit were made sturdy material to avoid damage and to satisfy all requirements and calculated pressure ratings, permanently. The flow channels are completely smooth to avoid changes of cross-section. Change in the velocity of flow to utilise the overall pressure drop to generate


actual velocity and not be lost in return flow or at sharp edges. In this way many irregularities in the flow patterns and changes of cross-sections are avoided.

The pressure rating of each element must be calculated individually, independent of others. The design is flexible enough to establish any velocity to be calculated. It is not desirable to have more than one flow path in parallel. For this reason the parallel pipes used in shell and tube heat exchanger avoided

Thermal Design:

The thermal design is very important in design of shell and tube heat exchanger. The thermal design is accomplished using one of the simple methods as narrated by D Q Kern. The heat transfer and coefficient pressure drop as predicted, particularly on the shell side could vary considerably from the actual values obtained in operation. The search has been instituted to develop more accurate predictive methods for thermal design. This is particularly relevant for optimum use of more expensive materials of construction coupled with the necessity for the increased reliability in operation.

The flow distribution, physical property variation, temperature correction, velocity consideration and fouling factors are some of the criteria to be given due weightage to accurately predict thermal performance.

Mechanical Design:

The mechanical designs of heat exchanger are based on reputed codes and standards. The most common standard used in TEMA. The Tubular Exchanger Manufactures Association (TEMA) was founded in the late 1930’s in an attempt establish standards


for high quality shell and tube heat exchanger. TEMA in turn refers to ASME section VIII wherever necessary. ANSI and ASTM (American Standard for Testing Materials) are also referred.

The code provides only basic frame work at minimum acceptable practices with which compliance is necessary to obtain a vessel that is structurally safe at the design temperature and pressure. Additional requirements are left to the judgement of the user and designer. Codes contain guidelines and recommendation covering design, material, fabrication, inspection and testing. Simplified rules based on theory of elasticity and consolidated experience are outlined for calculating thickness of pressure components along with the permitted configurated and recommended shapes. Stress tables for various materials, weld joints details and testing requirements are stipulated.

Heat Exchanger Design-An Over View:

Heat exchanger is the work home of the chemical industry and nuclear and thermal power plants. As it is the most commonly used equipment, it is imperative that improvements are continuously made in the design, for maximum cost effectiveness. The total design involves the thermal design and mechanical design. In thermal design, attempt is made to obtain a value, as realistic as possible, for the overall heat transfer coefficient and pressure drops on the shell and tube sides. The heat transfer correlations for the tube sides have a valid theoretical foundation, but those on the shell side are primarily empirical in nature because of the difficulties encountered in mathematically analysing the shell side flow.


After the thermal design comes the mechanical design. In small heat exchangers, there is no need for stress analysis. However, with increasingly large chemical plants, nuclear and thermal plants, such analysis also comes important. The mechanical design is covered by ASME codes and TEMA standards


Design Specifications:

The design specifications for TEMA class R 610-2438 BEM type heat exchanger is shown below.


1. Component: Shell Cylinder

[As per ASME SECTION VIII, DIVISION II] Material selected: SA 516 GR 60

P= Shell side design pressure = 0.735 Mpa IR=inside radius of shell = 304.8 mm

S=Maximum allowable stress =118MPa [From ASME SECTION II, Part D] E=Joint efficiency =0.85

Corrosion allowance =3


Circumferential stress

t = (P*IR / (S*E-0.6*P)) +CAI+CAO+tol = 5.27 mm [UG-27(c)(1)] Longitudinal stress

t = (P*IR / (2*S*E+0.4*P)) +CAI+CAO+tol = 4.13 mm [UG-27(c)(2)] Actual wall thickness of cylinder: tnom = 12 mm

2. Component: Front and Rear Head Cylinder

[As per ASME SECTION VIII, DIVISION II] Material selected: SA 516 GR 60

P=1.4715 Mpa

S=Maximum allowable stress= 118MPa [From ASME SECTION II, Part D] E= 1

Required wall thickness of the cylinder, greater of: Circumferential stress

t = (P*IR / (S*E-0.6*P)) +CAI+CAO+tol = 6.87 mm [UG-27(c)(1)] Longitudinal stress

t = (P*IR / (2*S*E+0.4*P)) +CAI+CAO+tol = 4.92 mm [UG-27(c) (2)] Actual wall thickness of cylinder: tnom = 12 mm

3. Component: Front and Rear Head Cover

ASME Section VIII-1 2004 A06 UG-32 Formed Heads, and Sections, Pressure on Concave Side

Ellipsoidal Cover Internal Pressure with t/L >= 0.002 Material: SA 516 GR 60


Design pressure P = 0.15 kg/mm2 Design temperature T = 170 C

Radiography = Full Joint efficiency E = 1 Design stress S = 12.022 kg/mm2

TEMA min. thickness tm = 9.5 mm Inside corrosion allowance CAI = 3 mm Major/minor rat. D/2h = 2.0

Forming tolerance Tol = 0 mm Corroded min. thk t = 3.85 mm radius L = 554 mm Ratio t/L = 0.01263 Outside diameter OD = 629.6 mm Corroded diameter ID = 615.6 mm Proportion factor K = 0.1667*(2+ (D/2h) ^2) = 1.0002 Required wall thickness of the cover:

4. Component: Tubes

[As per ASME SECTION VIII, DIVISION II] Material selected: SA179

P=1.471 Mpa OR=25 mm S = 92.4 Mpa E= 1


Required wall thickness of the cylinder, greater of: Circumferential stress

t = (P*OR / (S*E+0.4*P)) +CAI+CAO+tol = 0.2 mm [APP.1-1(A)] Longitudinal stress

t = (P*IR / (2*S*E+0.4*P)) +cai+cao+tol = - [UG-27(c) (2)] Actual wall thickness of cylinder: tnom = 1.25 mm

ASME Section VIII-1 2004 A06 UG-28 Thickness of Shells under External Pressure Material: SA-179

Design pressure P = 0.075 kg/mm2 Design temperature T = 170 C Inside corr. allow. CAI = 0 mm Corrosion allowance CAO = 0 mm Radiography = Full Material tol. Tol = 0 mm

Cylinder outside diameter Do = 25 mm Cylinder length EP L = 2438 mm

Nominal thickness tnom = 1.25 mm (tnom-CAI-CAO-Tol) t = 1.25 mm L/Do ratio Ldo = 97.52

Do/t Dot = 20.08

(2*S) or (0.9*yield) SE = - Modules of elasticity ME = 19649 kg/mm2 A factor, A = 0.002734 [From ASME SECTION II, Part D, figure G] B factor CS-1, B = 8.52 [From ASME SECTION II, Part D, figure CS-1] Max allowed external pressure:


Actual external design pressure: P = 0.075 kg/mm2

t = (P*ID*K / (2*S*E-0.2*P)) +CAI+CAO+tol = 6.85 mm [App. 1-4(c)] Actual wall thickness of cover: tnom = 10 mm

5. Tube-to-Tubesheet Welds

ASME Section VIII Div.1 2004 A06 UW-20 Tube-To-Tubesheet Welds Fig UW-20.1 Sketch (a) Full Strength G

Tubesheet material: SA-105 K03504 Forgings Tubes material: SA-179 K01200 Smls. Tube Allowable stress Tubes St = 14.06 Kg/mm2 Allowable Stress tubes Sa = 9.42 Kg/mm2 Allowable stress weld Sw = 9.42 Kg/mm2 Tube OD do = 25 mm

Tube thickness t = 1.25 mm

Design temperature Tubesheet = 170 C Design temp. Tubes = 170 C

Fillet weld leg af = 2.05 mm Groove weld leg ag = 0 mm

Minimum length ac acmin = 2.04 mm Total length ac = af+ag = 2.05 mm

Fillet weld strength Ff = 0.55*Pi*af*(do+0.67*af)*Sw Ff = 882 kgf Groove weld strength Fg = 0.85*Pi*ag*(do+0.67*ag)*Sw Fg = 0 kgf Tube strength Ft = Pi * t * (do - t) * Sa Ft = 875 kgf


Design Strength Fd = 875 kgf

Fillet weld strength Ff = min (Ff, Ft) Ff = 875 kgf Groove weld strength Fg = min (Fg, Ft) Fg = 0 kgf Weld strength factor fw = Sa / Sw

fw = 1

Ratio fd = Fd / Ft fd = 1

Ratio ff = 1 - Fg / (fd * Ft) ff = 1

Minimum required length of the weld leg(s), ar

ar = SQRT((0.75*do)^2 + 2.73*t(do-t)*fw*fd) - 0.75*do ar = 2.04 mm

UW-18(d) - Allowable load on fillet/groove welds Weld Leg = 2.05 mm

Allowable Load = PI * do * Weld Leg * Sw * 0.55 = 836 kgf Maximum Allowable Axial Loads, Lmax

Pressure only = LmaxP = Ft LmaxP = 875 kgf Other loads = LmaxO = 2*Ft LmaxO = 1751 kgf Total weld throat dimension = 1.45 mm

6. Component: Front Pass Partition

Pass Partition Plate Max. Allowed Pressure Differential (TEMA 1999 RCB-9.132) Pass plate material: SA 516 GR 60 Plate


Allow. Pressure drop qa = 0 kg/cm2 Design stress S = 12.02 Kg/mm2 Max. Allowable pressure drop:

q = (1.5*S*((t-c)/b) ^2)/B = see table below

Table1: B-Factor

4. Component: Flange Design

ASME Section VIII-1 2004 A06 App. 2 Bolted Flange with Ring Type Gaskets Flange type: Integral tapered hub - code fig.2-4(6)

Flange material: SA-105 K03504 Forgings Int. design pressure PI = 0.15 Kg/mm2 Design temperature T = 170 C

Ext. design pressure PE = 0 Kg/mm2 B1 = B+g1 or B +go

B1 = 624.6 mm

Inside corrosion allowance CAI = 3 mm Outside corrosion allowance CAO = 0 mm Stress (operating) SFO = 14.06 Kg/mm2 Stress (atmos.) SFA = 14.06 Kg/mm2


Outside diameter A = 737 mm Inside diameter B = 615.6 mm Hub thickness g1 = 31.75 mm Bolt circle diameter C = 695 mm Hub thickness at attach. go = 9 mm Mean gasket diameter G = 638 mm Weld leg/hub length h = 70 mm Hub to bolt circle R = 7.95 mm Bolt circle to OD = 21 mm Flange thickness t = 55 mm

Gasket material: Spiral-Wound Metal Fiber Stainless Gasket outside dia. ODG = 650 mm

Gasket width Wth = 12 mm Gasket thickness = 2 mm Gasket factor m = 3.000

Gasket seating stress y = 7.03 Kg/mm2 Gasket eff. width b = 6 mm

Gasket rib length Rib = 1878 mm Seating width bo = 6 mm

Gasket rib effective width Br = 4.76 mm (Table 2-5.2 facing 1a/1b Col. II) Bolt material: SA-193 G41400 GR B7 Bolt (<= 2 1/2)

Bolt diameter = 16 mm No. of bolts No. = 42 Bolt root area = 138.32 mm2


Stress (atmos.) SA = 17.58 Kg/mm2 Joint-contact compressive load

HP = 6.2832*b*G*PI*m+2*Br*m*PI*RIB = 18873 kgf Hydrostatic end force H = 0.7854*G*G*PI = 47954 kgf Hydrostatic end force He= 0.7854*G*G*PE = 0 kgf Operating conditions:

Min. calc. bolt load WM1 = HP+H = 66827 kgf

Min. used bolt load WM1 = max of 2 mating flanges = 66827 kgf Bolting up conditions:

Minimum bolt load WM2 = b*3.1416*G*Y+Br*Y*RIB = 147434 kgf Min. used bolt load WM2 = max of 2 mating flanges = 147434 kgf Required bolt area AM = WM2/SA or WM1/SB = 8388 mm2 Available bolt area AB = No.Bolt*Area = 5809.6 mm2 Design bolt load W = 0.5*(AM+AB)*SA = 124774 kgf

Minimum gasket width NMIN = AB*SA/ (6.283*y*G) = 3.62 mm Loads: Integral Flange Calculations (Operating conditions):

Hydrostatic end load HD = 0.785*B*B*PI = 44646 kgf Hydrostatic end load HDe= 0.785*B*B*PE = 0 kgf Gasket load HG = WM1-H = 18873 kgf

Result hydrostatic force HT = H-HD = 3308 kgf Bolting up conditions:

Gasket load HG = W = 124774 kgf Operating conditions:


Gasket load lever arm hg = (C-G)/2 = 28.5 mm

Result hydrostatic lever arm ht = (R+g1+hg)/2.0 = 34.1 mm Bolting up conditions:

Gasket load lever arm hg = (C-G)/2 = 28.5 mm Operating conditions:

Hydrostatic moment MD = HD*hd = 1064 kgf*m Gasket moment MG = HG*hg = 538 kgf*m

Result hydrostatic moment MT = HT*ht = 113 kgf*m

Total operating moment MOP = MD+MG+MT = 1714 kgf*m

Total operating moment MOPe= HDe(hd-hg)+HTe(ht-hg) = 0 kgf*m Bolting up conditions:

Bolt up moment MATM = W*hg = 3556 kgf*m

Effective bolt moment MB = MATM*SFO/SFA = 3556 kgf*m Total moment MO = MOP or MB = 3556 kgf*m

Cf= 1

Bolt spacing correction M = MO*Cf = 3556 kgf*m (TEMA 1999 RCB-11.23) Flange shape constants:

K = A/B = 1.1972 ho = SQ (B*G0) = 74.4339 TF = Fig.2-7.1 = 1.8401 h/ho = h/ho = 0.9404 Z = Fig.2-7.1 = 5.6158 F = Fig.2-7.2 = 0.6901 Y = Fig.2-7.1 = 10.8856 V = Fig.2-7.3 = 0.07 U = Fig.2-7.1 = 11.9621 f = Fig.2-7.6 = 1.3099 G1/G0 = G1/Go = 3.5278 e = F/ho = 0.0093 t = 55 mm


D = U*ho*g0*g0/V = 1029855 Alpha = t*e+1.0 = 1.5099 Beta = 1.333*t*e+1.0 = 1.6797 Gamma = Alpha/TF = 0.8206 Delta = t*t*t/D = 0.1616 Lambda = Gamma +Delta = 0.9821 Stress calculations, Allowable stress:

Longitudinal hub SH = (f*M)/(Lambda*g1**2*B1) = 7.53 Kg/mm2 1.5*SFO=21.09Kg/mm2 Radial SR = Beta*M/(Lambda*t**2*B) = 3.27 Kg/mm2 SFO =14.06 Kg/mm2 Tangential ST1 = M*Y/(t**2*B)-(Z*SR) = 2.45 Kg/mm2 SFO =14.06 Kg/mm2 (Greater) ST2 = (SH+SR)/2 or (SH+ST1)/2 = 5.4 Kg/mm2 SFO =14.06 Kg/mm2

ASME Section VIII Div.1 2004 A06, Appendix 2, 2-14 Flange Rigidity Factor V = 0.07

Factor L = 0.9821

Modulus elastic design T Ed = 19649 Kg/mm2 Modulus elastic atm. T Ea = 20530 Kg/mm2 Thickness g0 = 9 mm Factor h0 = 74.4 mm Factor KI = 0.3 Factor KL = 0.2 Corrosion allowance CA = 3 mm Factor K = 1.1972 Thickness T = 55 mm


Rigidity index, J, integral flange type

Gasket seating J = 52.14 * Ma * V / (L * E * G0 ** 2 * ho * KI) = 0.356 Operating J = 52.14 * Mo * V / (L * E * G0 ** 2 * ho * KI) = 0.1793

5. Component: Tubesheets:

The materials of construction and design conditions for tubesheet design are given below.

Table 2: Materials and Design Conditions


MATERIALS SA-516 GR 60 SA-516 GR 60 SA-105 SA-179


0.075 0.15 0.15 0.15

DESIGN TEMP. (deg C) 65 170 170 170


12.02 12.02 14.06 9.42

MEAN METAL TEMP. (deg C) 50 50 170 165 MOD.OF ELAS/ M.M.T.(Kg/mm2) 20361 20361 19649 19675 COEF.TH.EXP/M.M.T. (mm/mm/C) 0.0000117 0.0000117 0.0000125 0.0000125 CORROSION ALLOWANCE (mm) 3 3 3 3



220 220 205 180

RCB-7.134 Tubesheet Formula - Tubesheet Flange Extension RCB-7.1341 Fixed Tubesheet or Floating Tubesheet Exchangers Design temperature TS = 170 C TS allowable stress S = 14.06 Kg/mm2 Tubesheet OD A = 737 mm Reaction diameter G = 616 mm Ratio A/G r = 1.1972 Equivalent diameter DL = 415.9 mm Flange moment M = 3556 kgf*m

Relative expansion between shell and tubes (TEMA T-4.5) Shell metal temp. = 28.9 C

Tube metal temp. = 143.9 C Tube length L = 2438 mm

DeltaL = (Alphas*Thetas-Alphat*Thetat)*L = -3.56 mm RCB-7.13 Required Effective Tubesheet Thickness

Tubesheet details with effective thicknesses (no corrosion added), mm Effective thickness definition as per TEMA 1999 RCB-7.12

Corroded conditions refer to head and shell dimensions only Bending:


Factor Eta = 0.4195 Shear:

T = 0.31*DL*(P/S)/ (1-do/Pitch)

The tubesheet required effective thickness under corroded and uncorroded conditions are given in table.

Table 3: Required Effective Tubesheet Thickness





Stresses, Kg/mm2 mean stress exceeds allowable).

Table 4: Shell and Tube Stresses



SHELL LONGITUDINAL STRESS Kg/mm2 4.54 5.13 0 0 SHELL COMPRESSIVE STRESS Kg/mm2 0 0 0 0 TUBE LONGITUDINAL STRESS Kg/mm2 1.36 1.73 0 0 TUBE COMPRESSIVE STRESS Kg/mm2 -22.1 -21.46 0 0 TUBE-TO-TUBESHEET LOAD Wj kgf 126.8 160.9 0 0 EFF.PRES.SHELL SIDE (BEND.) Kg/mm2 0.31 0.254 0 0 EFF.PRES.TUBE SIDE (BEND.) Kg/mm2 0.239 0.185 0 0 EFF.PRES.SHELL SIDE (SHEAR) Kg/ mm2 0.31 0.254 0 0 EFF.PRES.TUBE SIDE (SHEAR) Kg/mm2 0.262 0.208 0 0


RCB-7.161 Equivalent Differential Expansion Pressure, Pd, Kg/mm2 Tube OD do = 25 mm Tube thickness t = 1.25 mm Tube Number N = 264 Tube pitch = 31.25 mm Tube Length Lt = 2438 mm Modulus of Elasticity E = 19649 Kg/mm2 Modulus of Elasticity Es = 20361 Kg/mm2 Modulus of Elasticity Et = 19675 Kg/mm2 Pd = 4*J*Es*ts*(DeltaL/Lt)/ (Do-3*ts)*(1+J*K*Fq) J = Sj*L/Sj*L+Pi*(Do-ts)*ts*Es K = Es*ts*(Do-ts)/ET*tt*N*(do-tt) Fq = 0.25+ (F-0.6)*((300*ts*Es/K*L*E)*(G/T)^3)^0.25

Table 5: Equivalent Differential Expansion Pressure



FACTOR F SHELL SIDE F FS mm 1.0 1.0 1.0 1.0 FACTOR F TUBE SIDE F = Ft mm 1.0 1.0 1.0 1.0


609.6 615.6 609.6 615.6

DIA. G TUBE SIDE G = Gt mm 609.6 615.6 609.6 615.6 SHELL OD DO mm 633.601 633.601 633.601 633.601 SHELL THICKNESS Ts mm 12.0 9.0 12.0 9.0


mm Sj EFFECTIVE TUBE LENGTH L mm 2332 2346 0 0 J=1;W/O EXP.JOINT J 1.0 1.0 0 0 J=0;Sj< (DO-TS) *TS*ES/10*L 0 0 0 0 RIGIDITY FACTOR K 0.9887 0.7451 0.9887 0.7451 Fq kg/mm2 3.0676 3.5613 0 0 Pd kg/mm2 -0.593 -0.483 0 0

RCB-7.162 Equivalent Bolting Pressure, Pb, Kg/mm2

Table 6: Equivalent Bolting Pressure



EQUIV. BOLTING PRESSURE Pbt 0.047 0.046 0.047 0.046 EQUIV. BOLTING PRESSURE Pbs 0.097 0.095 0.097 0.095 Operating moment M1 = 1714 kgf*m Bolting-up moment M2 = 3556 kgf*m Operating - Pbt = Bolting up - Pbs =

RCB-7.163 Effective Shell Side Design Pressure, P, Kg/mm2

Table 7: Effective Shell Side Design Pressure


CONDITION UNCORRODED CORRODED UNCORRODED CORRODED P = (Ps'-Pd)/2 0.31 0.254 0 0 P = Ps' 0.026 0.025 0 0 P = PBs 0.097 0.095 0 0 P = (Ps'-Pd-PBs)/2 0.261 0.207 0 0 P = (PBs+Pd)/2 -0.248 -0.194 0 0 P = Ps'-PBs -0.071 -0.07 0 0 G = Gs = Shell I.D., mm 609.6 615.6 609.6 615.6 fs = 1-N*(do/G)**2 0.556 0.5646 0.556 0.5646 Dj = expansion joint ID, mm 609.6 615.6 0 0 Ps' 0.026 0.025 0 0

RCB-7.164 Effective Tube Side Design Pressure, P, Kg/mm2

Table 8: Effective Tube Side Design Pressure



P = (Pt'+PBt+Pd)/2, If Ps' is positive -0.239 -0.185 0 0 P = Pt'+PBt, If Ps' is positive 0.116 0.113 0 0 P = (Pt'-Ps'+PBt+Pd)/ 2, If Ps' is negative -0.252 -0.198 0 0 P = Pt'-Ps'+PBt, If Ps' is negative 0.089 0.088 0 0 P = Pt+(Ps/2)*((Dj/ 0 0 0 0


G)**2-1)+PBt, When J=0 and Ps and Pt are

both positive ft = 1-N*((do-2*tt)/G) **2 0.64 0.647 0.64 0.647 G = Gs = Shell I.D., mm 609.6 615.6 609.6 615.6 Pt' 609.6 615.6 0 0

RCB-7.22 Shell Longitudinal Stress, Ss, Kg/mm2 Ss = (Cs * (Do - ts) * (Ps*)) / 4 * ts

Table 9: Shell Longitudinal Stress



Tensile stress (shell), Kg/mm2 Allowable stress 12.02 12.02 12.02 12.02 Tensile stress Ss 4.54 5.13 0 0 Compressive stress (shell), Kg/mm2 Allowable stress 11.64 12.02 11.64 12.02 Compressive stress Ss 0 0 0 0 Ps* = Pt-Pt' 0.081 0.083 0 0 Ps* = Ps' 0.026 0.025 0 0 Ps* = -Pd 0.593 0.483 0.64 0.647


Ps* = Pt-Pt'-Pd 0.674 0.566 0 0

Ps* = Ps'-Pd 0.619 0.508 0 0

Ps* = Pt-Pt'+Ps'-Pd 0.7 0.591 0 0

RCB-7.23 Tube Longitudinal Stress-Periphery of Bundle, St, Kg/mm2 St = (Ct * Fq * (Pt*) * G^2) / 4 * N * tt * (do-tt)

Table 10:Tube Longitudinal Stress-Periphery of Bundle



G = Gs = Shell I.D 609.6 615.6 609.6 615.6

Fs = 3.25-0.5*Fq 1.72 1.47

Tensile stress (tubes), Kg/mm2 Allowable stress 9.42 9.42 9.42 9.42 Tensile stress St 1.36 1.73 0 0 Compressive stress (tubes), Kg/mm2 Allowable stress Sc 9.42 9.42 0 0 Compressive stress St -22.1 -21.4 0 0 P2 = Pt'-(ft*Pt/Fq) 0.037 0.04 0 0 P3 = Ps'-(fs*Ps/Fq) 0.013 0.013 0 0 Pt* = P2 0.037 0.04 0 0 Pt* = -P3 -0.013 -0.013 0 0 Pt* = Pd -0.593 -0.483 0 0 Pt* = P2-P3 0.025 0.027 0 0 Pt* = P2+Pd -0.555 -0.443 0 0 Pt* = -P3+Pd -0.605 -0.497 0 0 Pt* = P2-P3+Pd -0.568 -0.456 0 0


RCB-7.24 Allowable Tube Compressive Stress-Periphery of Bundle, Sc, Kg/mm2 Sc = Pi^2 * Et/(Fs*(kl/r)^2 when Cc <= kl/r k = 0.80

Sc = (Sy/Fs)*(1-(kl/r)/ (2*Cc)) when Cc > kl/r l = 425.5 mm Cc = Sqrt(2*Pi^2*Et/Sy)

Cc = 156 kl/r = 40.47

r = 0.25*Sqrt(do^2+(do-2*tt)^2) = 8.41 mm

RCB-7.25 Tube-to-Tubesheet Joint Loads-Periphery of Bundle, Wj, kgf Wj = Pi * Fq * (Pt*) G^2 / (4 * N)

Table 11: Tube -to-Tubesheet Joint Loads-Periphery of Bundle



G = Gs = Shell I.D., mm 609.6 615.6 609.6 615.6 Tube-to-tubesheet load, Wj 126.8 160.9 0 0 Pt* = P2 0.037 0.04 0 0 Pt* = -P3 -0.013 -0.013 0 0 Pt* = P2-P3 0.025 0.027 0 0

RCB-7.25 Tube-to-Tubesheet Joint Loads-Periphery of Bundle, Wj, kgf Allowable Loads per ASME Section VIII Div. 1 2004 A06 Appendix A

Table 12: Joint Types

Type Joint


No Test Test

fr Lmax fr Lmax


welded only b Seal welded only 0.55 481 0.7 613 e strength welded and expanded 0.8 700 1 875 f Seal welded and exp.with 2 grooves 0.75 657 0.95 832 g Seal welded and exp.with 1 groove 0.65 569 0.85 744 h Seal welded and exp.with no grooves 0.5 438 0.7 613 I Expanded with 2 grooves 0.7 613 0.9 788 j Expanded with 1 groove 0.65 569 0.8 700 k Expanded with no grooves 0.5 438 0.6 525 .

For joints types a,b,b-1,c,d,e : Lmax = At*Sa*fr For joints types f,g,h, : Lmax = At*Sa*fe*fr*fy For joints types i,j,k : Lmax = At*Sa*fe*fr*fy,ft Cross-sectional area At = 92.91 mm2


Factor fe (l/do or 1) fe = 1 Ratio fy fy = 1.38

ft = (Po+Pt)/Po ft = 1

Min Yield Str SigmaM = 18.28 kg/mm2 (ft = 1 if max exceeded)

Tube OD do = 25 mm Tube thickness t = 1.25 mm

Tubes yield str (min) st = 18.28 kg/mm2 Joint operating Temp T = 1.2 C

Tubes Mod.Elasticity EtT = 19649 kg/mm2 TubSh Mod.Elast. EsT = 19649 kg/mm2 Tubes Coef.Th.Exp. at = 0.0000125 TubSh Coef.Th.Exp. as = 0.0000125

Po = (4*(do*t-t^2)*st)/do^2 Po = 3.46 kg/mm2 Pt = ((T-Tamb)*(at-as)*(EtT*EsT)/(EtT+EsT) For joint types i, j, k:

Po + Pt <= 0.58*SigmaM 3.46 kg/mm2 <= 10.6 kg/mm2

6. Component: Nozzle

ASME VIII-1 2004 A06 UG-27 Thickness of Cylinders under Internal Pressure Material: SA-106 K03006 GR B Smls. Pipe

Design pressure P = 0.075 kg/mm2 Design temperature T = 65 C


Radiography = Spot Joint efficiency E = 1 Design stress S = 12.022 kg/mm2

Inside corrosion allowance CAI = 3 mm Outside corrosion allowance CAO = 0 mm Material tolerance = 1.09 mm

Minimum thickness tmin = 6.02 mm Outside diameter OD = 60.33 mm Corroded radius OR = 30.16 mm

Minimum thickness not less than UG-45(a), UG-16(b), UG-45(b): UG-45(a) internal pressure:

t = (P*OR / (S*E+0.4*P)) +CAI+CAO+tol = 4.28 mm APP.1-1(A) UG-45(a) external pressure+CAI+CAO+tol = 0 mm

UG-16(b) minimum thickness+CAI+CAO+tol = 5.68 mm UG-45(b) Smaller of: t = 6.02 mm

UG-45(b) (4) standard pipe*0.875+CAI+CAO+tol = 7.51 mm UG-45(b) Greater of: t = 6.02 mm

UG-45(b) (1) +CAI+CAO+tol = 6.02 mm UG-45(b) (2) +CAI+CAO+tol = 0 mm Minimum thickness: tmin = 6.02 mm Nominal thickness: tnom = 8.74 mm

7. Component: Reinforcement Nozzle

ASME Section VIII-1 2004 A06 UG-37 Reinforcement Required for Openings in Shells and Formed Heads


Int. design pressure PI = 0.075 Kg/mm2 Ext. design pressure PE = 0 Kg/mm2

Design temperature T = 65 C Fig.UW-16.1 Sketch (c) Vessel material: SA-516 K02100 GR 60 Plate

Inside corrosion allowance CAI = 3 mm Outside corrosion allowance CAO = 0 mm Vessel design stress Sv = 12.02 Kg/mm2 Joint efficiency E = 1

Vessel outside diameter Do = 633.6 mm Corroded radius IR = 307.8 mm

Nominal thickness tnom = 12 mm Reinforcement limit lp = 48.8 mm

Required thickness internal pressure tr = 1.93 mm Required thickness external pressure tre = 0 mm Corroded thickness t = 9 mm

Reinforcement efficiency E1 = 1.0

Attachment Material: SA-106 K03006 GR B Smls. Pipe Inside corrosion allowance CAI = 3 mm

Outside corrosion allowance CAO = 0 mm Nozzle design stress Sn = 12.02 Kg/mm2 Joint efficiency E = 1

Nozzle outside dia. Don = 60.33 mm Corroded radius OR = 30.16 mm Nominal thickness tnom = 8.74 mm


Reinforcement limit ln = 14.34 mm

Required thickness internal pressure trn = 0.19 mm Required thickness external pressure trne = 0 mm Corroded thickness tn = 5.74 mm

Nozzle Projection h = 0 mm Reinforcement element material: Limit of reinforcement Dp = 0 mm Nominal thickness te = 0 mm Outside diameter = 0 mm Design stress Se = 0 Kg/mm2

Minimum weld size tmin = 5.74 mm Leg size (1/2*tmin) (Act) = 0 mm 1/2 * tmin (minimum) = 0 mm 1/2 * tmin (actual) = 0 mm Weld tw (minimum) = 4.02 mm Weld tw (actual) = 0 mm Weld tc (minimum) = 4.02 mm Weld tc (actual) = 4.02 mm

Smaller |6.35 mm | Leg size tw (actual) = 0 mm tc of |0.7 * tmin| Leg size tc (actual) = 5.74 mm Outward nozzle weld L1 = 5.74 mm

fr1 = Sn/Sv = 1.0

Outer element weld L2 = 0 mm fr2 = Sn/Sv = 1.0


Inward nozzle weld L3 = 0 mm fr3 = Sn/Sv or Se/Sv = 1.0 Inward nozzel weld new = 0 mm fr4 = Se/Sv = 1.0

Corroded int.proj.thk ti = 0 mm

Corroded inside diameter d = 48.8 mm

Vessel wall length available for reinforcement 2*Lp-d = 48.8 mm Plane correction factor (Fig.UG-37) F = 1

Offset distance from centerline doff = 0 mm

Reinforcement areas (internal pressure condition) ASME 2004 UG-37 A1 = Vessel wall. Larger of:

|2*(t+tn)*(E1*t-F*tr)-2*tn*(E1*t-F*tr)*(1-fr1)| = 208.5 mm2 A1 = 345.5 mm2

A2 = Nozzle wall outward | 5*(tn-trn)*fr2*t | = 249.7 mm2 Smaller of: | 5*(tn-trn)*fr2*tn | = 159.2 mm2

A2 = 159.2 mm2

A3 = Nozzle wall inward | 5*t*ti*fr2 | = 0 mm2 Smallest of: | 5*ti*ti*fr2 | = 0 mm2

| 2*h*ti*fr2 | = 0 mm2 A3 = 0 mm2

A41 = Outward nozzle weld = (L1^2)*fr3 = 32.9 mm2 A42 = Outer element weld = (L2^2)*fr4 = 0 mm2 A43 = Inward nozzle weld = (L3^2)*fr2 = 0 mm2 A4 = 32.9 mm2


A5 = Reinforcement pad Area = (Dp-d-2*tn)*te*fr4 A5 = 0 mm2 Aa = Area Available = A1+A2+A3+A4+A5 Aa = 537.6 mm2 A = Area required = (d*tr*F) +2*tn*tr*F*(1-f1) A = 94.2 mm2

Per UG-36(c) (3) (a), this opening does not required additional reinforcement other than the inherent in the construction.

Nozzle attachment weld loads - UG-41 - Strength of reinforcement ASME - Weld strength calculations not required per UW-15(b). Weld load for strength path 1-1 (UG-41(b)(1)

W (1-1) = (A2+A5+A41+A42)*Sv W (1-1) = 2310 kgf Weld load for strength path 2-2 (UG-41(b) (1)

W (2-2) = (A2+A3+A41+A43+2*tn*t*fr1)*Sv W (2-2) = 3552 kgf

Nozzle attachment weld loads - ASME 2004 UG-41 - Strength of reinforcement Unit stresses - UW15(c) and UG-45(c)

Inner fillet weld shear = 5.89 Kg/mm2 Groove weld tension = 8.9 Kg/mm2 Nozzle wall shear = 8.42 Kg/mm2 Strength of connection elements Inner fillet weld shear = 3202 kgf Nozzle wall shear = 4139 kgf Groove weld tension = 4835 kgf Possible paths of failure

1-1 4139 + 3202 = 7341 kgf 2-2 3202 + 4835 = 8037 kgf


Welds strong enough if path greater than the smaller of W or W (path) Path 1-1 > W or W11

7341 kgf > 2310 kgf OK Path 2-2 > W or W22 8037 kgf > 3552 kgf OK



Heat exchangers find wide application in various fields of engineering like space heating, refrigeration, air conditioning, power plants, chemical plants, petrochemical plants, petroleum refineries, and natural gas processing. Two fluids, of different starting temperatures, flow through the heat exchanger. One flows through the tubes (the tube side) and the other flows outside the tubes but inside the shell (the shell side). Heat is transferred from one fluid to the other through the tube walls, either from tube side to shell side or vice versa. The fluids can be either liquids or gases on either the shell or the tube side. In order to transfer heat efficiently, a large heat transfer area should be used, leading to the use of many tubes. In this way, waste heat can be put to use. This is an efficient way to conserve energy. The heat exchanger selected for the design is shell and tube heat exchanger of fixed tube sheet type. The various parameters were considered and mechanical design was completed using ASME and TEMA standards. The mechanical design is analysed using B-JAC software and finds out that the design is safe.



1. Mechanical design of heat exchanger & process vessel components – K.P.

Singh & A.I. Soler

2. Heat & Mass Transfer- S.C. Arora & S. Domkundwar 3. ASME Boiler and Pressure Vessel Code

4. TEMA Standards and Codes

5. Heat & Mass Transfer – R.C. Sachdeva 6. Process Heat Transfer- D.Q. Kern



1. Reinforcing Element UG-37.1


3. Chart for Determining Shell Thickness of Components Under External Pressure CS1


9.Figure RCB-7.132


11. Principle Dimensions of Typical Heads Figure 1-4