Heat Exchanger Design Manual Rev1

HEAT EXCHANGER DESIGN MANUAL PROCESS DEPARTMENTREVISION DATE :1 : 06.03.2003 Process Dept Heat Exchanger Design Manual 6 th March, 2003 Page 2 Foreword This Heat Exchanger design Manual discusses about the Heat exchanger design related to refinery services. The discussion starts with general information on the Heat exchangers and proceeds into the detailed design aspects focused on Shell and Tube Heat Exchangers. Discussions on Thermal design and some mechanical aspects are also included in the manual. A study report on optimization of exchanger design is included in the final parts of the manual for readers better understanding on the cost implications for exchanger design. Process Dept Heat Exchanger Design Manual 6 th March, 2003 Page 3 Table of Contents 1 INTRODUCTION-----------------------------------------------------------------------------------------------------------------------4 2 CLASSIFICATION---------------------------------------------------------------------------------------------------------------------5 3 BASIC DESIGN THEORY----------------------------------------------------------------------------------------------------------10 4 SELECTION OF HEAT EXCHANGERS----------------------------------------------------------------------------------------13 5 TUBE SIDE DESIGN -----------------------------------------------------------------------------------------------------------------16 6 SHELL SIDE DESIGN ---------------------------------------------------------------------------------------------------------------19 7 NOZZLES-------------------------------------------------------------------------------------------------------------------------------27 8 MISCELLANEOUS-------------------------------------------------------------------------------------------------------------------29 9 VIBRATIONS---------------------------------------------------------------------------------------------------------------------------31 10 HEAT EXCHANGERS OPTIMIZATION-------------------------------------------------------------------------------------32 11 GENERIC GUIDELINES FOR HEAT EXCHANGER DESIGNER-----------------------------------------------------49 14 TYPICAL FOULING RESISTANCES USED FOR DESIGNING HEAT EXCHANGER-------------------------57 15 MECHANICAL ASPECTS--------------------------------------------------------------------------------------------------------58 A direct transfer type of heat exchanger is one in which the cold & hot fluid flows simultaneously through the device and heat is transferred through a wall separating the fluids. This is not generally used in refineries. the hot fluid storing heat in it and the cold fluid extracting heat from it. Each kind of exchanger is suitable for a typical application. The hot & cold fluid flows alternately through the matrix. Thus. Among the above three types direct type heat exchanger is commonly used in the refinery. . Cooling towers and scrubbers are two examples of direct contact heat exchangers. Moreover. which is used for transferring the thermal energy between two or more fluids that have different temperatures with or without phase change. If heat is to be transferred between a gas & liquid. Shell and tube heat exchanger is of this type. This manual will however discuss only the shell & tube heat exchanger design in detail. 2003 Page 4 Heat exchanger is a device. the gas is either bubbled through the liquid or the liquid is sprayed in the form of droplets into the gas. Direct contact type. Other two types of exchangers have there own limitations. they may be classified as Direct transfer type. and environmental control and transport industries has called forth an equally wide variety of heat exchanger configurations to meet the requirements. A storage type of heat exchanger is the one in which the heat transfer from hot fluid to cold fluid occurs through a coupling medium in the form of porous solid matrix. process. it also includes condensers. Storage type. chemical and petrochemical applications. Most heat exchangers employ this mode of heat transfer. reboilers and vapourizers in addition to heaters and coolers where there is no phase change. The wide range of heat transfer services that are required in the energy. Direct contact heat exchanger is the one in which the fluids are not separated. the design of storage type and direct type exchanger are not common and thus designed by special vendors.Process Dept 1 Introduction Heat Exchanger Design Manual 6 th March. Based on the transfer process occurring in heat exchangers. 2003 Page 5 2 Classification Now based on the geometry of the equipment we can classify the exchangers into following types: 2.1 Shell-and-Tube Exchangers Shell-and-tube exchangers are composed of round tubes mounted in a cylindrical shell with the tube axis parallel to that of the shell. Class R designates the unfired shell-and-tube heat exchangers generally used for severe conditions in petroleum and related processing. There is considerable design flexibility because the core geometry can be varied easily by changing the tube diameter. and the third ‘the rear-end head type’. the second ‘the shell type’. rear-end head. stationary or front-end head.Process Dept Heat Exchanger Design Manual 6 th March. shell. to provide for ease of cleaning. The major components of this exchanger are tubes (or tube bundle). fabrication. the other flows across and along the tubes. although elliptical and rectangular tubes have also been used. the first letter indicating ‘the front-end head type’. limited only by the materials of construction. These exchangers may be further classified as shell-and-tube. . and arrangement. double-pipe. etc. TEMA has developed a notation system to designate major types of shell-and-tube exchangers. and material selection for three classes of heat exchangers: R. and B. baffles. to contain operating pressures and temperatures. radioactivity. and tubesheets. 1. toxicity. For instance AEL is fixed tube one shell pass exchanger. depending upon the desired heat-transfer and pressure-drop performance and the methods employed to reduce thermal stresses. A variety of internal constructions are used in shell-and-tube exchangers. C.1 Tubular Heat Exchangers Tubular heat exchangers are generally built of circular tubes. pressures. length. TEMA has also set up mechanical standards for design. to prevent leakage. They can be designed for special operating conditions: heavy fouling. each exchanger is designated by a three letter combination.1. and to control corrosion. spiral and tube – coil exchangers. In this system. Class C designates the unfired shell-and-tube heat exchangers generally used for moderate conditions of commercial and general processing. corrosiveness. and any temperature and pressure differences between the fluids. These exchangers are classified and constructed according to Tubular Exchanger Manufacturers Association (TEMA) standards in the United States or modified TEMA standards in other countries. The details of this are discussed later. One fluid flows inside the tubes. highly viscous flow. Tubular exchangers can be designed for any operating temperatures. Process Dept Heat Exchanger Design Manual 6 th March. these exchangers cannot accommodate high pressure and temperature differences. Double-pipe exchangers are generally used for the small capacity applications where the total heattransfer surface area is 20 m2 (215 ft2) or less. while the other heat exchangers employ ferrous materials. However. with separate shell on the each leg of the hairpin & special cover over the U bend of the hairpin. 2003 Page 6 Class B designates the unfired shell-and-tube heat exchangers used for chemical process service. This is perhaps the simplest heat exchanger. they are at times used for application up to 50 m2 (500 ft2).3 Extended surface heat exchangers . The plate-type exchanger can be used when there are several hot or cold process streams present. Packinox TM is an example of this type of exchanger manufacturer. These types of exchangers are not common in the refinery industry for its limitation on clean fluids and high capital investment. Generally. This configuration is appropriate when one or both of the fluids are at very high pressure because containment in small diameter pipe or tubing is less costly than containment in large diameter shell. Generally. The number of tubes in the bundle is usually much less than conventional shell & tube exchanger. 1. The required thermal duty is achieved with the very large surface area per unit volume. where the temperature approaches are very small (about 20°C). Flow distribution is not a problem.1. 2.Type Heat Exchangers Plate-type heat exchangers are usually built of thin plates. Stacks of double-pipe or multitube-type heat exchangers are also used in some process applications with radial or longitudinal fins. this exchanger can find the suitable place. The tubes may be longitudinally finned or can be plain tubes with baffle to give cross flow. 2. which are either smooth or corrugated and are either flat or wound in an exchanger. and cleaning is done easily by disassembly. Class B exchangers employ non-ferrous materials.2 Double-Pipe Heat Exchangers Double-pipe heat exchangers usually consist of concentric pipes.2 Plate. However. These exchangers are principally used for clean fluids when the temperature difference between process streams is small. Multitube hairpin exchanger consists of bundle of hairpin tubes. One fluid flows in the inner pipe and the other flows counter-currently in the annulus between the pipes. Based on the process function also we can classify the heat exchanger to the following: 2. The spray condenser is the most common direct contact condenser in which sub-cooled liquid is sprayed into the vapour. power generation. The most common types of condensers used in the industry are shell & tube type exchangers. automotive. which are employed to extend the heat transfer surfaces. and the condensate may be a single component or multi-component (miscible and/or im-miscible) pure vapour with or without non-condensable gases. In a direct contact condenser. are known as fins. but should be high enough so that it does not flood with condensate.Process Dept Heat Exchanger Design Manual 6 th March. One of the most important design features of a condenser is to provide a vent for the removal of non-condensable gases (regardless of how small they may be in the vapour). In refinery application. applications are limited due to mixing of the process stream with the coolant. the condensing vapour is injected into the pool of liquid coolant. . they extend the surface and enhance the heat exchanger area. the condensing process stream and coolant come into direct contact with each other and heat is transferred. chemical process. In addition to condensation. and many other industries. 2003 Page 7 When additional metal pieces are attached to ordinary heat-transfer surfaces such as pipes & tubes. have high heat-transfer rates per unit volume. Extended surface exchangers are used for very high heat loads. condensation occurs either inside or outside the tubes. air conditioning. where air is used as cooling media. which will be discussed later. and fouling is not a problem. These condensers are inexpensive. The vent should be located near the coldest part of the condenser to avoid escape of vapour. The fundamentals of the design of Air Fin Cooler are same as of shell and tube heat exchanger.4 Condensers Condensers are widely used in refinery industries. the tubes being either horizontal or vertical. Condensation may be total (as with a pure vapour) or partial (single or multi-component vapours with or without non-condensable). Pieces. there extended surface exchanger is commonly used. de-superheating and/or sub-cooling may also take place in a condenser. In a pool condenser. Broadly speaking. However. In the process industry. they are either direct contact type or indirect contact type (stream separated). The kettle reboiler (TEMA K shell). Generally. which don’t have tubes. H.3 Natural Circulation (Thermosiphon) Reboiler In a kettle reboiler. it is a two-phase (vapour – liquid) mixture. 2. the tube pitch-to-diameter ratio is kept between 1. in other thermosiphon reboilers. and X. the vapour at the exit of the boiling liquid side is dry or almost dry. the kettle reboiler is considered as a pool-boiling device.5 Liquid to vapour phase change exchangers These are mainly evaporators. The tube bundle diameter ranges from 50 to 70 % of the shell diameter. The liquid is fed at a rate greater than the vaporisation rate so that low volatile components do not build up in the shell. or powder.or multi-segmental baffles). placed in an oversized shell. however flow boiling prevails in the tube bundle. the nozzle in this space is used to drain the excess liquid. Vaporisers which supply heat to generate vapours in a distillation column & are located at the bottom of a distillation column are termed as reboilers. The liquid to be vaporised enters from bottom and covers the tube bundle. 2003 Page 8 An impingement plate should be placed under the vapour inlet nozzle to prevent tube erosion and possible flow-induced vibrations.5 and 2. and almost dry vapour exits from the top nozzle. G. usually consists of horizontal bundle of heated U tubes (two tube passes). Typical weir heights exceed bundle heights by 50 – 150 mm (2 – 6 in). Proper design of the inlet and outlet piping to the reboiler is essential to maintain minimum pressure drop in the piping. The final products of evaporators are typically concentrates. the X shell having the lowest pressure drop on the shell side. reboilers. J. one or more vapour nozzles are used. To maintain a high liquid circulation rate. Depending upon the length of the kettle and a need to reduce liquid entrainment. . the vapour occupies the upper space in the shell. circular in cross section. a multiple tube-pass floating head bundle may also be used. A variety of shell types have been used depending upon the applications such as E (single. The large empty space in the shell acts as a vapour disengaging space. 1.1. The small hold-up space beyond the weir is used to control the removal of excess liquid (which includes non-volatiles). crystals. Evaporators are mainly used for concentrating a chemical solution by evaporation.Process Dept Heat Exchanger Design Manual 6 th March. has a TEMA E shell with single-pass tubes in which boiling occurs. The boiling fluid is on the tube side in most cases. 2003 Page 9 The vertical thermosiphon reboiler. or E shell.4 Forced Circulation Reboilers These reboilers are used for heavy fouling. H. and boiling takes place on the shell side. G.6 times the tube length. and the column must be elevated above the ground. . Vapour – liquid mixture leaves the reboiler and separation takes place in the column. It uses refrigerant such as Ammonia or Freon.7 m (8 – 12 ft). For vacuum applications.7 Heaters Its function is to impart sensible heat to a liquid or a gas by means of condensing steam or Thermicfluid.  The horizontal thermosiphon reboiler.3 – 0.4 – 3. very little boiling (< 1 %) takes place within the reboiler due to the high circulation rate. the liquid level may be dropped to 0. The inlet pipe area is kept small (up to 50 % of the tube area) to enable flow to be restricted in case of flow instability. and tube lengths are usually 2. Generally. 2. In general when a liquid is sub-cooled after condensation or cooling in an air cooler it is then called as trim cooler. but may be on the shell side in special applications. The pump in the liquid feed line to the reboiler is used to circulate the fluid through the system. or when the vaporisation rate is low. viscous. The tube diameter ranges from 25 to 50 mm (1 – 2 in) for low-pressure operation with wide-boiling mixtures. 2.Process Dept  Heat Exchanger Design Manual 6 th March.8 Coolers This term is usually used for exchanger that cools a liquid or gas with water. smaller than obtainable if water were only used as a coolant. 1. has a TEMA X.6 Chillers Its function is to cool a fluid to a temperature. The liquid level for non-vacuum operations is maintained to the upper (top) tubesheet. The reboiler may be horizontal or vertical. or solid-bearing liquids. This term can be applied to exchanger with a purpose of cooling a liquid within range possible by water. single or two-tube passes. This means the liquid level in the base of the distillation column must be higher. 2. Flow is restricted in the inlet line for stable operation and good control of the reboiler. The outlet pipe is kept short with its flow area equal to the flow area of all tubes because of the high exit velocity.1. 2003 Page 10 Basic heat transfer equation is the relationship of the heat transfer rate. For heat exchange across a typical heat-exchanger tube the relationship between the overall coefficient and the individual coefficients. (Pr)0.33 Where the Nusselt number and Reynolds number is based in the tube outside diameter and the flow velocity is based in the minimum cross sectional area at the shell inside diameter. The prime objective in the design of an exchanger is to determine the surface area required for the specified duty (rate of heat transfer) using the temperature differences available. and geometry specifications are needed. heat transfer surface. kcal / hr m °C di = Tube inside diameter.Process Dept 3 Basic Design Theory Heat Exchanger Design Manual 6 th March.33 .6 . which are the reciprocal of the individual resistances.ln(do/di)}/2. process conditions. The . kcal/hr m2 °C ho = Outside fluid film coefficient.kw + (do/di). Q = U *A * MTD Overall heat transfer coefficient should be calculated to determine how much heat transfer area is needed for the required heat transfer rate at a given MTD. physical properties of the material involved. kcal /hr m 2 °C hi = Inside fluid film coefficient. is given by 1/Uo= 1/ho + 1/hod + {do. To calculate overall U.(1/hid) + (do/di). (Re)0. kcal/ hr m2 °C kw = Thermal conductivity of the tube wall material. which is the sum of several individual resistances. kcal / hr m2 °C hid = Inside dirt coefficient. and mean temperature difference (MTD) is as given below.1 Shell side flow correlation The first heat transfer correlation suggested is due to Colburn in 1933 and is of the form NuD = 0. m 3. m do = Tube outside diameter. kcal / hr m2 °C hod = Outside dirt coefficient.(1/hi) Where Uo = The overall heat transfer coefficient based on the outside area of the tube. The overall coefficient is the reciprocal of the overall resistance to heat transfer. For baffled shells.14 In this equation all properties are evaluated at the local bulk mean temperature except viscosity wall. The essential difference is that it incorporates a viscosity correction factor and is therefore valid up to Pr = 16000.6 . the tube heat transfer & pressure drop can be predicted much more accurately and there are large number of correlation for doing this. NuD = 0.7 < Pr<100. (Re)0.027. where bypass and leakage streams diminish the flow effectiveness. (Pr)0. The Dittus Boelter equation is not applicable to fluids having Prandtl number greater than 100 because of the fact that the viscosity of such fluids changes rapidly with temperature.33 . 2003 Page 11 validity was restricted to Re > 2000 and staggered layout.2. (Pr)0.Process Dept Heat Exchanger Design Manual 6 th March. (pr)n In the above equation. (Re)0. Sieder Tate have proposed an equation which is similar to the above and is valid under the same conditions.6 . The most commonly used relation for calculating the heat transfer coefficient in turbulent fully developed flows in smooth circular pipes is the Dittus Boelter equation.3 if the fluid is being cooled. Eventually the correlation form was slightly modified to include the Sieder-Tate form for nonisothermal effects.14 Where all properties are based on the average bulk temperature except viscosity wall. the thermo-physical properties are evaluated at the local bulk mean temperature. The validity was restricted to the turbulent flow of 2000< Re<40000 3. This equation is valid for 0. .6 was introduced. / visco wall ) 0.(visco.4 if the fluid is being heated and 0.33 .2 Tube Side Correlation As compared to shell side. / visco wall) 0. (Re)0. The exponent n of the Prandtl number has the value 0. which is much easier to handle and within the general data scatter produced equal results.023 . NuD = 0. a safety factor of 0. Many of these correlations have been obtained experimentally for laminar & turbulent flow. which is evaluated at the wall temperature.(visco.8 . NuD = 0. which refers to tube wall. Caloric bulk temperature: For non-incremental calculations in which properties vary greatly along the exchanger’s length. thus. which are required for calculation of heat transfer coefficient. 2003 Page 12 Average Bulk Temperature Since physical properties. This is a more representative method used for overall calculations across the entire exchanger. first aim is to determine average fluid temperature on which the fluid properties will be based. However. which provides the driving force for the heat transfer. change with the fluid temperature.∆Tc ) / ln ( ∆Th / ∆Tc) Pure counter flow permits extreme temperature overlap without incurring any special penalty for the above equation. 3. . For convenience. which gives good results in each increment for incremental calculations if the range of physical property variation is not large. the overall heat transfer coefficient requires either Calculating local coefficients at a series of temperatures between the exit and inlet. ∆Tc & ∆Th ∆ Tlm = ( ∆Th . The bulk temperature changes as the fluid heats or cools. if the fluid temperature variation is plotted against the length of the exchanger it is not a straight line. Determining a representative average bulk temperature for average physical property evaluation. ' F' factor is applied which is always less than one.4 Mean temperature difference The heat transfer rate is directly proportional to the temperature difference.3 Heat Exchanger Design Manual 6 th March. the mathematical expression for LMTD is based on the terminal differences. this method is used. Therefore determining actual mean temperature difference between the two streams engaged in heat transfer becomes very essential. followed by graphical or numerical integration. This bulk temperature is defined as the hypothetical temperature at a given cross section in which all fluids flowing past the cross section are completely mixed and in thermodynamic equilibrium. Two methods used for defining average bulk temperature are: Arithmetic average bulk temperature: it is a simple method. In case of exchangers with more than one pass.Process Dept 3. The maximum value of Ft is 1.Process Dept Heat Exchanger Design Manual 6 th March. F shell is rarely used in practice because there are many problems associated with the design. shell side leakage and bypass streams etc. 4 Selection of Heat Exchangers E Shell: The E shell is the most common as it is inexpensive and simple. and should be considered when temperature cross is likely to occur. The value for Ft is above 0. Although ideally this is a desirable flow arrangement. The use of two or more shells in series. . Boiling range is very important in case of vaporization services. In case of. will give closer approach to true counter-current flow. unbalanced thermal expansion in case of large temperature difference between inlet & outlet are encountered. With the F shell. The tubes may have single or multiple passes and are supported by transverse plate baffles.5 Boiling range (BR) Boiling range is nothing but the difference between dew point & bubble point. based on boiling range. vaporization of multicomponent mixtures. Co-current and Cross-flow. 2003 Page 13 Usually actual exchangers are combination of True Counter-current. As such it is required to estimate the “true temperature difference” from the LMTD. there are also additional problems of fabrication and maintenance. It is difficult to remove or replace the tube bundle. This correction factor is applied due to the effect of the composition gradient on the effective ∆T and can be quite severe if the boiling range is large. resulting in two shell passes and nearly a pure counter flow. by applying a correction factor to account the deviation from true counter current flow. 3. number of shell & tube passes. Also problems like internal leakage. ∆Tm = Ft ∆Tlm Ft = Temperature correction Factor The temperature correction factor is function of shell and tube fluid temperatures. F Shell: The F shell has a longitudinal baffle.0 for true counter current flow. or multiple shell side passes. multicomponent mixture correction factor is applied to nucleate boiling heat transfer coefficient. Hence it is very essential to enter boiling range of vaporizing side in HTRI so that the correction facto applied will be realistic.85 for most of the economical designs. N – Fixed tubesheet with removable cover plate. For low pressure applications B – Single tube side joint.or bonnet. Drawback of this type of shell is that no support plate can be given. The double-split flow H shell is similar to the G shell. condensers. They are used when essentially 100% vaporization is required. For application where tube side is corrosive. X shell: For a given flow rate and surface area. or one inlet and two outlet nozzles. It is less expensive than Type A. and G shells. and other phase-change applications. J shell: The divided flow J shell has two inlets and one outlet.1 TEMA Front Head Selection A – Easy to open for tube side access. but with two inlet and two outlet nozzles (and two horizontal baffles) to accommodate high inlet velocities. and it usually has low shell side ∆P compared to that for the E. the cross flow X shell has the lowest shell side pressure drop compared to all other shell configurations except the K shell.which is a heavy item. Tube side is corrosive. It is used as a kettle reboiler in the process industry and as a flooded chiller in the refrigeration industry. but requires piping dismantling of connected piping. it is used for gas heating and cooling with and without finned tubes and for vacuum condensation. F. C–Channel to tubesheet joint is eliminated. It has approximately one eighth the pressure drop of a comparable E shell. The H shell approaches the cross flow nozzle arrangement of the X shell. They are used as horizontal thermosiphon reboilers. Hence. The G and H shells are seldom used for shell side single-phase applications. D – Very High-pressure applications . For Higher-pressure applications.Process Dept Heat Exchanger Design Manual 6 th March. since there is no advantage over E or X shells. It is also used when shell flows are large 4. K shell: The K shell is used for partially vaporizing the shell fluid. and is therefore used for low pressure drop applications such as a condenser in vacuum. so maximum tube length gets limited. toxic or hazardous and when removable tube bundle is required. removal of gaskets and of integral cover. toxic or hazardous and shell side fluid is clean and any leakage possibility is to be eliminated. preferred with clean tube side fluid. 2003 Page 14 G and H shell: The G and H shells are related to the F shell having variants of the longitudinal baffle. It is less expensive than A & normally used for low-pressure operations. Exchangers with good ∆T & no cleaning required on tube side are provided with ‘ U ‘ type of bundle. T. In these kind . P.2 TEMA Rear Head Selection Fixed Head (L. (Low Fouling) L – Same as A type front end. fixed head exchangers can be used. This is applicable only for shell pressures up to approximately 5kg/cm2g. toxic or hazardous and where leakage of shell to tube side fluid and vice versa. frequent need to takeout the tube bundle. High-pressure requirements.Process Dept Heat Exchanger Design Manual 6 th March. In case of floating head exchangers with single pass on tube side.g. ‘ J ’ type of shell can be used. Not commonly used. in cases where shell side cleaning is not required. M. W – Should be used when shell side fluid or both shell and tube side fluid are Dirty. It is also preferred whenever cleaning on shell side is required. 2003 Page 15 4. Exchangers such as process gas trim cooler. normally floating head exchangers are used. BFW exchangers where normally temperature cross is observed. W – For low-pressure application 4. where ∆T is less than 50°C & also does not require cleaning shell side. For low pressure applications M – Same as B type front end. For thermal differential expansion of the shell and tubes is higher and tube side fluid is clean. Fixed tube sheet with expansion joint can be used for high ∆T. For e. Exchangers where very little pressure drop is available on shell side. relatively lesser maintenance requirements T– Pull through floating head. Normal Pressure requirements. with hazardous/ toxic fluid on shell side. U tube – It is a removable bundle without floating head. S. N) – Should be used when thermal differential expansion of the shell and tubes is low and shell side fluid is clean. not requiring frequent maintenance. is to be eliminated.3 EXCHANGER Type Selection Criteria & Special GUIDELINES Whenever ∆T on any side is more than 50°C. N – For application where tube side is Corrosive. (High fouling) P – Pressure is low and shell side fluid is not toxic or hazardous. S – Floating head with backing ring. For Higher-pressure applications. Where risk of internal flange leakage is to be avoided. expansion bellows are necessary. For high-pressure applications or. 40mm. Aluminum. copper. In case of alloyed steel. tantalum. Vertical floating head exchangers are provided with floating head on bottom side for free expansion of the shell. 5. i. they should be selected for heavily fouling fluids. The frequently used Tube OD is 0. 14.625 in (16 mm) to 2 in (50 mm) are used. 16. carbon steel alloys. monel. 1. and other special materials. 12 and 10 BWG. Plain tubes are most common ones. as they will give more compact. For titanium it can be as low as 1 mm.1 Tubes Shell & tube exchangers have either of the two types of tubes. For mechanical cleaning. and 3.25 in. tubes should be welded to tube sheet to prevent leakage of hydrogen to the medium at the other side. The commonly used thickness corresponds to 20. tube thickness used can be 1. 5 Tube Side Design 5. Larger tubes are easier to clean by mechanical methods.11mm. 1. exchangers. the smallest practical size is 0.65mm.75 in. (19 mm). and are more rugged. nickel. 1.05mm). carbon. For example coker naphtha on shell side. It is also specified as ‘ strength weld’ on datasheet. 2003 Page 16 of exchangers’ liquid outlet is given from bottom in order to have free draining. 20mm. stainless steel. cupro-nickel. 18. 0.75 in. available in Carbon steel. (19. to 1 in.625 in. Hence normally expansion bellows are given in the inlet side. Typical tube thickness used are 2 mm or 2. Whenever partial pressure of hydrogen is more than 5 bar on any side. and therefore cheaper. (16mm). For chemical cleaning.77mm. 2. 0. smaller sizes can be used provided that the tubes never plug completely.6 mm. Exchanger who requires frequent cleaning on shell side are given with ‘S’ type or ‘T’ of rear head.89mm. titanium.e. 2.3 Tube Thickness The tube thickness is selected to withstand the internal pressure and give adequate corrosion allowance. brass and alloys. (16 to 25 mm) are preferred for most duties. 5. (32mm).2 Diameter Tube diameters in the range 0. Aluminum alloys.0 in.25mm. glass. namely (a) plain or bare and (b) finned – external or internal.625 in. The smaller diameters 0.Process Dept Heat Exchanger Design Manual 6 th March. (25mm) or 1. .5 mm for 20 mm or 25 mm tube OD for CS. H – band The number of passes in an exchanger is the number of times one fluid passes though the other fluid compartment.tube heat exchangers commonly are 6 ft. 10 ft.1. 8 ft. The standard design has one.2 to 24. (12. 5. 14 ft. (3. In multi-pass designs.829 m). Ribbon Quadrant H . but longer exchangers are made with tube up to 40 to 80 ft.1. The fluid in the tube is usually directed to flow back and forth in a number of “passes” through groups of tubes arranged in parallel to increase the length of the flow path. Higher velocities in the tube result in higher heat transfer coefficients.048 m). The tube lengths for both straight and U.877m). Quadrant. Ribbon.6 Tube pass geometry There are three main types of tube passes. (2.Process Dept Heat Exchanger Design Manual 6 th March.096 mm). two or four tube passes.5 1. at the expense of increased pressure drop compared to that for low velocity. It is also useful in utilising maximum pressure drop available. (4.4 m). odd passes are uncommon. (4. Tube length also depends on clients' preference such as 6m or 9 m. Exchangers are built with from one to up to about sixteen tube passes. (3.Band . 1. Not normally used. an even number of passes are generally used. 12 ft.658 m).4 Tube Length Tube length used should be maximum as possible so that minimum number of tubes are required which ultimately lead to a smaller size of exchanger.267). 2003 Page 17 5. 20 ft (6. and may result in mechanical and thermal problems in fabrication and operation. The number of passes is selected to give the required tube-side design velocity and taking maximum advantage of available pressure drop. 16 ft. (1.5 Tube Arrangement Tube-Side passes Number of passes can be defined as number of times material flows along the exchanger.438 m). For identical tube pitch and flow rates. Thus the 90° layout will have the lowest heat-transfer coefficient and the lowest pressure drop. the 90° layout. square pattern must be used. and 90°. As described earlier. at the typical tube pitch of 1. rotated triangular (60°). However. since access lanes are not available.35mm (1/4 in) is provided. is preferred. a square layout is typically employed. a minimum-cleaning lane of 6. a triangular pattern produces high turbulence and therefore a high heat transfer coefficients. Chemical cleaning does not require access lanes.25 times the tube OD.7 Heat Exchanger Design Manual 6 th March. Where boiling is necessary. or rotated square (45°) pattern. The triangular pitch is generally used in the fixed tubesheet design because no cleaning is needed. the 45° layout is preferred for laminar or turbulent flow of a single-phase fluid. 45°. For services that require mechanical cleaning on the shell side. the 90° layout is used for turbulent flow. U-tube construction for clean services on the tube side. which provides vapour escape lanes. fixed-tubesheet construction is usually employed for clean services on the shell side. so a triangular layout may be used for dirty shell side services provided chemical cleaning is suitable and effective. For dirty shell side services. A rotated triangular pattern seldom offers any advantage over a triangular pattern.Process Dept 1. Furthermore. Here. Consequently. square (90°).1. The triangular pitch (or rotated triangular) pattern will accommodate more tubes than a square (or rotated square) pattern and provide a more compact arrangement. usually resulting in a smaller shell. For clean services on both . and floating-head construction for dirty services on both the shell side and tube side. it does not permit mechanical cleaning of tubes. a triangular layout is limited to clean shell side service. For dirty fluids with low shell side Reynolds number (<2000) its usually advantageous to use this type as it produces higher turbulence and higher efficiency. Where mechanical cleaning is required. and for condensing fluid on the shell side. If the pressure drop is constrained on the shell side. the tube layouts in decreasing order of shell side heat-transfer coefficient and pressure drop are 30°. and its use is not very popular. 60°. 2003 Page 18 Tube layout patterns The tubes in an exchanger are usually arranged in a triangular (30°). 25 times the tube OD. Additionally adopting other methods to reduce pressure drop will result in a cheaper design. For example. 2003 Page 19 shell side and tube side. They provide support to tubes and minimize the tube vibrations. Support baffles are also provided for extreme vibration condition. .25 times the tube OD. enable a desirable velocity to be maintained for the shell-side fluid.1 Baffles Baffles are used to support tubes. TEMA additionally recommends a minimum-cleaning lane of 0. 1. They protect the tubes in the top row from erosion. F shell). Triangular pitch typically holds 15% more tubes than square pitch for same OTL. They are used to control overall flow direction of Shell fluid (Longitudinal baffle.25 mm (25 X 1. For U-tubes exchangers.4 for laminar flow. a 25 mm tube pitch is usually employed for 20mm OD tubes and 31. either fixed tubesheet or U-tube construction may be used. a triangular pattern may be used provided the shell side stream is clean and a square (or rotated square) pattern if it is dirty. whichever is larger. however. For square pattern. although Utube is preferable since it permits differential expansion between the shell and tube. TEMA specifies a minimum tube pitch of 1. (6mm) between adjacent tubes. Hence a triangular tube pattern may be used for fixed-tubesheet exchanger and square (or rotated square) pattern for floating head exchangers. These are provided in the shell to direct the fluid stream across the tubes. the optimum tube-pitch to tube OD ratio for conversion of pressure drop to heat transfer is typically 1.Process Dept Heat Exchanger Design Manual 6 th March. to increase the fluid velocity and so improve the rate of heat transfer by increase in turbulence. One can decrease shell side pressure drop by increasing the tube pitch. and prevent failure of tubes due to flow-induced vibration. Thus. For a triangular pattern.25 is for 25 mm tube. 20mm tube tubes should be laid on a 26mm (20mm + 6 mm) square pitch. Thus the minimum tube pitch for square patterns is 1. or the tube OD plus 6 mm.25 to 1. 6 Shell Side Design 6.8 Tube Pitch Tube pitch is defined as the shortest distance between two adjacent tubes.25 in. and vibration due to the impact of high velocity fluid jet from the nozzle to the tubes (Impingement Baffle).25) square pitch. the same is not recommended as it increases the shell diameter. As far as thermal hydraulics is concerned.1. but 25mm tubes should be laid on a 31.35 for turbulent flow and around 1. tube diameter & pitch size. cavitation. 6. Various clearances are required for manufacturing. and H’ shells.2 Transverse Baffle The transverse baffles may be classified as plate baffles and rod (or strip) baffles. 1. expressed as the percentage of the shell inside diameter. The segmental baffle is a circular disk (with tube holes) with a segment removed. For example. F. The flow in a rod-baffled heat exchanger is parallel to the tubes. It is also difficult to remove or replace the tube bundle with such baffles. and H shells have longitudinal baffles. spacing.1 Longitudinal Baffle These are plates inserted in the shell longitudinally for the purpose of controlling the overall flow direction of the shell fluid as in ‘F. The rod baffles are used to support the tubes and to increase the turbulence of the shell fluid. and flow-induced vibration. and cut are largely determined by flow rate. Three clearances associated with a plate baffled . which is responsible to direct the shell side fluid across the tubes and thus mainly responsible for bringing effective heat transfer. G. 6. The limiting factor for baffle selection is usually the pressure drop and the flow induced vibrations.1. The choice of baffle type.Process Dept Heat Exchanger Design Manual 6 th March. G. vary from 15 % to 45 %. 2003 Page 20 Here we will discuss baffles primarily with respect to its role in Heat Transfer.1. The single.3 Plate Baffles The two types of plate baffles are either segmental or disk-and-doughnut. The transverse baffle is the one. The purpose of longitudinal baffles is to control the overall flow direction of the shell fluid. 6. Fluid can leak through these clearance passages and reduce the heat-transfer effectiveness.1. Even though one of the major functions of the plate baffle is to induce cross flow (flow normal to the tubes) to improve heat-transfer performance. and the problem of flow-induced tube vibration is virtually eliminated by the baffle support of the tubes. The baffle cuts. Although it helps bringing nearly counter current flow but its use is restricted to low pressure drop and low temperature difference applications. Their major disadvantage is because of increased leakage that may be caused if it is not sealed properly.9 Types Baffles may be classified into primarily two types transverse and longitudinal.1.1.1.and double-segmental baffles are most frequently used. allowable pressure drop. the most common being 20 – 35 %. The plate baffles are used to support the tubes. and to increase the turbulence of the shell fluid. The single-segmental baffle is generally simply referred to as a segmental baffle. this objective is only approximated. tube support.1. to direct the fluid in the tube bundle at approximately 90° to the tubes. Since the tubes in the window zone are supported at a distance of two or more times the baffle spacing. The double-segmental baffle. The resultant design is referred to as the segmental baffle with notubes-in-window (NTIW). Generally. The low velocity regions in the baffle corners do not exist for NTIW. If the shell side operating pressure is high. also referred to as a strip baffle. this no-tubes-in-window design is very expensive. and the diameter of the hole of the "doughnut" is smaller than the half shell diameter. This baffle design provides a lower pressure drop compared to a single-segmental baffle for the same unsupported tube span.Process Dept Heat Exchanger Design Manual 6 th March.4 Segmental Baffles with No tubes in Windows. the disk diameter is larger than the half shell diameter. Thus the loss of heat-transfer surface in the window region is partially compensated. If the primary function of baffles is to support the tubes.1. The lower pressure drop asks for large baffle spacing. they are referred to as support plates. and eliminates the tube bundle-to-shell bypass stream C. resulting in good flow characteristics and less fouling.1. They may be thicker than the baffle. 2003 Page 21 shell-and-tube exchanger are tube-to-baffle clearance. This can only be envisaged for twophase or gas as shell side fluid. and provide greater stiffness to the bundle.and doughnut-shaped baffles. Hence an exchanger with this type of baffle can handle larger fluid flows on the shell side. provide a lower pressure drop. and permit closer tube support to prevent tube vibrations. To eliminate the susceptibility of tube vibrations and to reduce the shell side pressure drop. 6. the tubes in the window zone are removed and support plates are used to reduce the unsupported span of the remaining tubes. This kind of arrangement mainly comes in to picture because of vibration problem. Multisegmental baffles have a strong parallel flow component. these are most susceptible to vibration. have less tube-to-baffle clearance. baffle-to-shell clearance.5 The disk-and-doughnut baffle The disk and doughnut baffle is made up of alternate disk. 6. . and bundle-toshell clearance.1. However. the shell size must be increased to compensate for the loss in surface area in the window zone.1. provides lower shell side pressure drop than that for the single-segmental baffle for the same unsupported tube span. two vertical baffles constrain all tubes from two sides. cavitation. in a square layout. adequate areas should be provided both between the nozzle and plate and between the plate . so that the baffle touches and constrains all tubes on four sides. This is not common in the refinery application since the fluid is highly fouling and minimum diameter is required ¾”. The rod baffle exchanger has several advantages over the plate-baffled exchanger.8 mm (3/16 in) or 6.152 m (6 in) for single-phase applications and 0. Generally the tube diameters and tube pitches are selected so that the rod diameter is 4. The baffle spacing is generally kept at 0. resulting in less fouling and corrosion. However. (3) There are no stagnant flow areas with the rod baffles.7 Impingement Baffle (Impingement Plate) Impingement baffles or plates are generally used in the shell side just below the inlet nozzle. The following two horizontal baffles constrain all tubes from the remaining two sides (90° to the first side). the rod baffle exchanger will require longer tubes of smaller diameter. Alternatively. The location of this baffle within the shell is critical in order to minimize the associated pressure drop and high escape velocity.1. A set of four-rod baffles (two vertical and two horizontal) is repeated in the exchanger.305 m (12 in) for condensers and reboilers. The most common cause of tube failure is the improper location and size of the impingement plate. 6. 2003 Page 22 The disadvantages of this design are: (1) all tie rods to hold the baffles are within the tube bundle. They protect the tubes in the top row from erosion. Thus. the rod baffle exchanger will result in a smaller unit for the same heat transfer and pressure drop. The rod diameter is equal to the tube spacing (zero clearance).1. and vibration due to the impact of high velocity fluid jet from the nozzle to the tubes.1. as follows: (1) it eliminates flow-induced tube vibrations since the tubes are rigidly supported at four successive points.6 Rod Baffles. If the tube side fluid is controlling and has a pressure drop limitation. the rod baffle exchanger may not be applicable.4 mm (1/4 in). and (2) the central tubes are supported by the disk baffles. (2) The pressure drop on the shell side is about one-half of that with a double-segmental baffle at the same flow rate and heat transfer rate. 6.Process Dept Heat Exchanger Design Manual 6 th March. For this purpose. which in turn are supported only by tubes in the overlap region of the larger diameter disk over the doughnut hole.1. A rod baffle is made up of parallel rods (a rod matrix) mounted on a baffle ring. and improved heat transfer compared to that for a plate baffle exchanger. One of the most common forms of this baffle is a solid circular plate located under the inlet nozzle just in front of the first tube row. which is less efficient than cross flow.) whichever is greater. ρV2 the shell or bundle entrance or exit area should be such that they do not produce a value of in excess of 4000 (5953). as it effects cross flow velocity (Velocity across the tubes). The TEMA standards specify for Segmental baffles minimum baffle spacing as one-fifth of the shell inside diameter or 50.Process Dept Heat Exchanger Design Manual 6 th March. For all other gases and vapors. no-tubes-in-window or pure cross-flow design should be tried. Higher baffle spacing will lead to predominantly longitudinal flow. The maximum baffle spacing is the shell inside diameter.8 mm (2 in) is required for cleaning the bundle. Spacing closer than one fifth of the shell diameter provide added leakage. This is the most vital parameter in STHE design. . ρ = Density in Pounds per Cubic Feet (Kilograms per Cubic Meter) V = Linear Velocity Feet per second (Meters per Second) 6.1. including those at its boiling point ρV2 exceeding 500 (744). and large unsupported tube spans which will make the exchanger prone to tube failure due to flow induced vibration.8 mm (2 in. Closer spacing results in poor bundle penetration by the shell side fluid and difficulty in mechanical cleaning of outside of tubes. If the foregoing limits on the baffle spacing do not satisfy other design constraints such as Dpmax or tube vibration. Single phase Fluids with entrance line values of ρV2 exceeding 1500 (2232).2 Baffle Spacing (Number of Baffles) Baffle Spacing is Centerline to centerline distance between adjacent baffles. TEMA Standards specify an impingement protection is required for following cases: For Non abrasive. TEMA gives the maximum permissible unsupported tube spans for tube support plates (or Baffles acting as support plates) specified in TEMA standards. Where. A spacing of 50. For all other Liquids. 2003 Page 23 and tube bundle. which nullifies the heat-transfer advantage of closer spacing. This also effects the unsupported tube length and thereby vibration. and for liquid –vapor mixtures. including nominally saturated vapors. therefore heat transfer and pressure drop. One needs to take extra care for specifying baffle cut for cases with no tubes in Windows. It is strongly recommended that baffle cuts between 20% to 35% be employed for single segmental baffle. a baffle cut of 20 to 25 percent will be the optimum. Flow across the Tubes) and Window Flow (i.Process Dept Heat Exchanger Design Manual 6 th March. It can vary between 15 to 45% of shell ID in case of single segmental baffle. Thus the optimum ratio lies when there is highest efficiency of conversion of pressure drop to heat transfer. 6. .e. through the Baffle cut area). At large baffle spacing. If fouling is of prime concern.4 Baffle cut orientation Baffle cut orientation can be Vertical. This optimum ratio is normally between 0. giving good heat-transfer rates. When this is not possible.1. It is preferable to have the same within 20%of each other. Large or small spacing coupled with large baffle cuts is undesirable because of the increased potential of fouling associated with stagnant flow areas. When the baffle spacing is reduced the pressure drop increases at a much faster rate than does the heat transfer coefficient. without excessive drop. spanning the effective tube length. The baffle cut and spacing should be designed such that the flow velocity is approximately the same for the Cross-Flow (i. For Double and Triple segmental baffle one should try to select the cut such that equal flow area exists at each baffle position. the percent cut may approach 45 – 49 % in order to avoid excessive pressure drop across the windows compared to the bundle. 6. 2003 Page 24 Baffle normally shall be spaced uniformly. Horizontal or Inclined (45°C). Generally.e. shall be located as close as practically possible to the shell nozzle. Both very small and very large baffle cuts result in inefficient heat transfer on the shell side. the baffle cut should be kept < 25 % to maintain high velocity through the window zone.3 to 0. baffles nearest the ends of the shell. In case parameters such as Pressure drop or heat transfer do not meet other methods should be adopted rather that changing baffle cut beyond these values.6 of shell ID. and/ or tubesheets. and the remaining baffle are spaced normally. such as using double segmental baffles or divided flow or even cross flow.3 Baffle Cut Baffle Cut is the height of the segment that is cut in each baffle to permit the shell side fluid to flow across the baffle.1. This is expressed as a percent of the Shell ID. The number of baffles required gets fixed based on selected tube length and baffle spacing. Depending on the same one can decide the number of ‘F’ stream seal rods that are required. orientation. spacing and clearances can decide how effective is the actual heat transfer across the exchanger. 2003 Page 25 The direction of the baffle cut is selected as either horizontal or vertical for a single-phase fluid (liquid or gas) on the shell side. Depending upon baffle cut direction the horizontal pass partitions will have different effect. because of ease of fabrication). The direction of the baffle cut is selected as vertical for the following shell side applications: Condensation (for better drainage) Evaporation/boiling (to promote more uniform flow) Entrained solids in liquid (to provide the least interference for solids to fall out) Multishell pass exchangers (includes F shell. however this is not the governing criterion. The selection of baffle cut. For vertical baffle cuts Horizontal partitions are parallel to flow direction Vertical partitions are perpendicular to flow direction For Horizontal baffle cuts Horizontal partitions are perpendicular to flow direction Vertical partitions are parallel to flow direction Thus while selecting tube layout one should take care of baffle cut orientation and try to minimize the ‘F’ Stream.Process Dept Heat Exchanger Design Manual 6 th March. One should be cautious while placing the first baffle with respect to inlet-outlet nozzle position. .1.5 Baffle Clearances Baffles have maximum impact on the various shell side flow streams. Inclined baffle orientation may as well be selected. With inclined (45°C) Baffle cuts one can assume that F stream by pass as negligible. It should be placed such that minimum bypassing occurs. the direction of the baffle cut should be horizontal for better mixing. This type should however be used with single shell pass and preferable single-phase fluid on shell side. 6. For example if the inlet nozzle is in vertical position it is better to have horizontal single-segmental baffles. The inlet outlet nozzle orientation should be selected in order to ensure shell side fluid properly passes through the tube bundle. For a very viscous liquid. Effect of baffle cut orientation is also dependent of inlet outlet nozzle position. TEMA indicates the minimum Baffle and Support plate thickness depending on shell ID and plate spacing. .91 mm (36 in) or for the tube OD > 31. However. This results in the largest bundle-to-shell clearance and the shell must be made greater than in the fixed and U-tube designs to accommodate the floating-head flange. It permits double the value. The tube-to-baffle hole clearance should be kept at a minimum to reduce tube vibration and resultant damage as well as to minimize the 'A' leakage stream.91 m (36 in) or for the tube OD 31. hence. Tube to Transverse Baffle Hole Clearance This is the factor. and provide greater stiffness to the bundle. Its flow is not only dependent on the clearance but also on the baffle thickness. 2003 Page 26 Shell to Transverse Baffle Clearance This is the factor which brings about the Leakage ‘E’ stream (discussed lr). The clearance in terms of the difference between the baffle hole ID and tube OD is 0.8 mm (1. because of floating head bonnet flange and bolt circle. many tubes are omitted from the tube bundle near the shell. If the baffles are acting as support plates they may be thicker than the baffle.2 Sealing Trips Sealing trips are trips placed along the shell to reduce the bundle to shell leakage and improves the shell side heat transfer coefficient. localised high velocities near the sealing strips could cause flow-induced tube vibration. for cases where its impact on Mean Temperature Difference and shell side coefficient is not significant.8 mm (1. have less tube-tobaffle-hole clearance. All support plates should also be accounted for.80 mm (1/32 in) for the unsupported tube length L < 0.25 in). which brings about the Leakage ‘A’ stream.Process Dept Heat Exchanger Design Manual 6 th March.40 mm (1/64 in) for unsupported tube length L > 0.25 in) or smaller. The baffle-toshell clearance should be kept to a minimum to minimize the E leakage stream. This allows the shell side fluid to bypass the tubes. 6. In some cases like floating head type. In order not to reduce the exchanger performance. This stream is fairly effective in heat transfer as it is contact with the tubes. 0. proper care must be exercised for the design. But usually these values are assumed for all practical purposes. sealing strips (or dummy tubes or tie rods) in the bypass area are essential. 1. Nozzle sizes are related to the piping size. Sealing strips are normally given longitudinally between first baffle to last baffle. ‘F’ stream is reduced by introducing dummy tubes or seal rods. cross flow (B stream). Out of these fluid flow streams. Not effective in heat transfer.10 Size Standard pipe sizes will be used for the inlet and outlet nozzles. B stream: Between tubes across the bundle. which is also called as bypass. This fraction is also effective for heat transfer since it is in contact with tube surface.7 in order to minimise shell side flow leakage. the exchanger shell design. ‘C’ stream.3 to 0.Process Dept 6. 2003 Page 27 Shell Flow distribution Fluid flows through the shell in 5 paths: A stream: Through gap between tube & baffle. mainly leads to good heat transfer. ‘A’ stream can be minimised by giving tight tolerance to tube OD & baffle hole OD. This .3 Heat Exchanger Design Manual 6 th March. No tube side fluid passe s through them. ‘E’ stream can be minimised by reducing the tolerance between bundle & shell ID.B leakage Turbulent 40 – 70 % 15 – 20 % 6 – 20 % 9 – 20 % Laminar 25 – 50 % 20 – 30 % 6 – 40 % 4 – 10 % 7 Nozzles 7.B Leakage T . is reduced by introducing sealing strips. In good design this stream is maintained between 0. This stream is partially effective in heat transfer since it contacts the outer surface of the bundle E stream: Through gap between baffle & shell. Flows ( percent ) for industrial exchangers ( HTRI data) Stream Cross flow (B) Bypass (C+F) S . F stream: Through pass partition lane. Most effective stream for heat transfer. Partially effective in heat transfer. and the escape flow area into the tube bundle. They physically block the flow but have no holes in the tubesheet.1 Shell Side 1. C Stream: Through gap between bundle & shell. If the baffle cut is parallel to the centerline of the nozzle. In case of Horizontal baffle cut and top/bottom inlet nozzle arrangement depending upon whether selected baffles are odd or even in number the outlet nozzle position has to be determined. For other cases ease of flow should be given preference. It should be placed such that minimum bypassing occurs. Sometimes tubes may be removed to obtain sufficient area. It is the total free area between a nozzle and the projected area on the tube bundle. In case if required nozzle size is very large it may be helpful and economical to use TEMA ‘J’ or ‘H’ type of shell. One should also take care that the percentage pressure drop across the exchanger nozzles is not very high as this drop is not utilised in heat transfer.11 Orientation Another important parameter is inlet nozzle orientation. thus giving loss in its efficiency. performance can be enhanced or reduced due to natural convection and other effects. One should be cautious while placing the first baffle with respect to inlet-outlet nozzle position. . 2003 Page 28 escape area should have ρ v2< 6000kgm–1s–2 (4000 lbmft–1s–2). Another important feature of nozzle size is its impact on inlet-outlet baffle spacing. For some cases. 1. bypassing of the inlet and outlet baffle spacing can take place. The direction of shell side flow with respect to gravity is also dependent on inlet nozzle location. One may select inlet-outlet nozzle position such that the hot fluid travels from top to bottom (such as condensers) and cold fluid from bottom to top (such as reboilers). Large inlet or outlet baffle spacing may cause vibrations related problems and a well-designed exchanger may turn out to be unsuitable because of this reason. Nozzles have an impact on the exchanger thermal design as they may influence its geometry. It is usually dependent on the baffle cut orientation. especially for 90°-tube pattern with laminar flow on shell side. If its necessary to select this arrangement one should take appropriate penalty factor as the performance of unit gets reduced.1. This however should be selected only if other parameters do not get effected. outlet nozzle position should accordingly be decided.Process Dept Heat Exchanger Design Manual 6 th March. Therefore one should take this factor into account before getting to finer optimization details. With vertical cut baffles and side inlet nozzle arrangement. Process Dept 1.2 Tube side Standard pipe sizes will be used for the inlet and outlet nozzles. 7. orientation depending on tube passes and partition plate locations. When in need of stacked exchangers. In case of odd number of passes the inlet-outlet nozzles are located on opposite sides. opposite sides or in the center. the shell side inlet nozzle location is on same side or opposite side of the tube side inlet nozzle. When ‘U’ tube or Floating head full support plate is located care should be taken to locate the nozzle (if at all present) before (i.12 Location Heat Exchanger Design Manual 6 th March. it is desired that the exchangers are designed such that the exchangers can be mounted directly nozzle to nozzle. However. 8. For 1-1 pass exchangers shell side tube side inlet nozzle should be placed on opposite sides. Their number is important in thermal design. 2003 Page 29 Nozzle location is very important in case of stacked exchangers. The spacing should be specified accordingly. Tube side nozzles are generally located on the circumference. One should take care that percentage drop across tube side nozzles is not very high as compared to that on exchanger as this pressure drop is not utilized for heat transfer. most of the times the number of passes are even and they are located on the same sides. For values refer TEMA. Depending on shell type selected the inlet-outlet nozzle location are on same sides. 8 Miscellaneous 8. as spaces for them need to be kept in the . one should take care of this factor while selecting exchanger type.1.1 Outer tube limit (OTL): The outer tube limit contains all tubes of the tube layout. This factor can often influence the Mean temperature difference. Distance between OTL & shell diameter is very important & is different for floating head. They may also be located axially if required. fixed head or U tube bundle.e.2 Tie Rods and Spacers The baffles and support plate are held securely in position by tying them with tie rods and spacers. on the side of shell) the support plate. Depending upon whether first tube pass is required to be co-current or counter. Here again one should try to select inlet-outlet nozzle position such that the hot fluid travels from top to bottom (such as condensers) and cold fluid from bottom to top (such as reboilers). The number of rods required will depend on the shell diameter. The recommended numbers for a particular diameter are tabulated in accordance to that of TEMA. thus effecting the overall geometry. .Process Dept Heat Exchanger Design Manual 6 th March. 2003 Page 30 layout. Some measures such as changing span lengths can be taken to avoid vibrations. After trying all these options. in which tube span at the inlet is equal to inlet baffle spacing. Different types of vibration mechanisms are as follows. Acoustic resonance is very important in case of gases. it is only for checking bundle entrance/exit velocities for perpendicular baffle cut exchangers. It occurs when the frequency of an acoustic wave in the heat exchanger coincides with tube natural frequency. Sometimes reducing clearance between tube & baffle or increasing tube pitch also helps in minimizing tube vibrations. There are different methods to avoid fluid elastic instability such as increase tube natural frequency by decreasing the span lengths or increasing the tube diameter. In this case increasing the flow velocity makes vibrations much worse. The tube vibration problem is complicated because it concerns fluid dynamics. This is not a resonance effect & occurs above a critical flow velocity. in order to get more realistic bundle entrance velocities & amplitudes. If fluid elastic instability is found at the shell entrance. structural dynamics. tubes in the window region can be removed so that all tubes are supported & no tube has double length span. HTRI vibration analysis uses longest tube span at the inlet space (inlet baffle spacing + central baffle spacing) for the bundle entrance analysis. Vortex shedding Acoustic resonance Turbulent buffeting Flow pulsation Fluid elastic instability is important for both gases & liquids. exchanger can be assumed to be NTIW design with same number of tube spans. and the mechanical properties of the metals involved. In such cases. However. Fluid elastic instability.Process Dept 9 Vibrations Heat Exchanger Design Manual 6 th March. This little conservative with perpendicular cut baffle. 2003 Page 31 Fluid flowing through a heat exchanger can cause the heat exchanger tubes to vibrate. if still vibrations exist. Even if acoustic wave does not cause any vibrations. Vortex shedding is caused by the periodic shedding of the vortices from the tubes and can lead to damage of tubes if vibrations coincide with the tube natural frequency. But here care should be taken that. then height under inlet nozzle can be increased. If still vibration is not going then rod type of impingement baffle can be used. . it can lead to intolerable noise. The turbulence in the flowing fluid contains a broad range of frequencies and can coincide with the tube natural frequency to cause tube vibrations.   • • • Baffle tip cross velocity ratio < 0. cross flow amplitude is checked for its ratio with tube gap to be less than 0.1 Rating Rating is nothing but performance estimation for a given exchanger geometry where process conditions & heat balance is known & check is made to see whether heat exchanger will perform to the desired conditions. Vibration output is generally checked for following specific points in HTRI output: • • Unsupported lengths span to TEMA maximum span should not exceed 0. 10 Heat Exchangers Optimization Heat exchanger thermal checking can be of two different kinds.Process Dept Heat Exchanger Design Manual 6 th March. This can become very important in case of two-phase flow or reciprocating machine discharge exchangers.5. rating and design.8 Acoustic vibrations : Vortex shedding ratio & turbulent buffeting ratio should not be more than 0.8 & Chen number to be less than 1300 For tube vibration check: Vortex shedding ratio should be less than 0. 10. cross flow RHO-V-SQR value in any pf the span should no exceed 3000 kpa.8 at the designing stage at any point of time. In all the two cases basic guidelines are essentially same. it should be checked for following ratios. Turbulent buffeting mechanism is very important in case of two-phase flow. . If not then. A general criterion for placing deresonating baffle is parallel to direction of cross flow & generally it is placed 5 % off center. Flow pulsation is because of periodic variations in the flow. Generally deresonating baffles are used for avoiding acoustic vibration problem. Fluid elastic instability is very important vibration problem.1.8 Average cross flow velocity ratio < 0. In any case. 2003 Page 32 It can be avoided by changing acoustic wave by changing span lengths. number of tube passes. The parameters of type (a) and (b) as above form a part of optimization process where “Pinch Technology” is now days employed. bundles. tube pitch. meaning “best” in case of a shell & tube heat exchanger. operating cost (pressure drop). maintenance cost (velocity. This is a new area dealing with the optimization of the entire plant and not just one heat exchanger. baffle cut. It basically starts with allocation of fluid. 2003 Page 33 Thus we can say that rating is nothing but identification of resistance distribution. pressure drop In both the case of thermal design. Performance related parameters such as pumping energy cost of one or both fluids. They are considered only in those contexts. length.2 Sizing Sizing is a series of sequential ratings. type and number of baffles. Overdesign can be given on the surface. checking shell side flow distribution & carrying out vibration analysis. Geometric parameters such as shell diameter. 10. amount of heat to be exchanged with reference to the overall process. are relevant. 10. minimise initial cost. This is the additional performance of the exchanger over desired performance. type of tube layout (square. or duty or flow as per requirement. flow rate/s of one or both fluids etc. baffle design. number of shell passes (with longitudinal baffle/s). It involves optimization of exchanger to get desired results in best possible way. shell type. The topic of this paper is optimization of single equipment and hence these parameters are not discussed here. there are some overlapping areas where the parameters of (a) and (b) type . Only the geometric parameters are considered. Thus optimization involves minimising the size of the unit. tube length. tube diameter. optimum process design would mean that combination of various parameters yielding the most suitable piece of equipment to perform the given duty. etc. However. The parameters under designer’s control may be of different types: Purely process parameters such as inlet and/or outlet temperatures of one or both fluids. tube diameter. layout etc.Process Dept Heat Exchanger Design Manual 6 th March. triangular or their rotated versions). investigation of pressure drop utilisation & its distribution. over design is required to consider to absorb any upset to the process condition.3 Optimization (based on a technical paper) The word “Optimum” is originally Latin. It is therefore necessary to first pinpoint what is most suitable or desirable. this cost is usually low. Usually these annual costs are expressed as percentage of fixed cost of the heat exchanger. this is called as the fixed cost of the exchanger. In case of heat exchangers. Only fixed cost to be minimized. The fixed cost of the exchanger is strongly related with its heat transfer surface. minimum or lowest cost is taken as the most desirable attribute of the design. The diameter of a heat . it is no wonder that the fixed cost of a heat exchanger is directly proportional to its area. In some cases. This payment is to be made every year till the loan is repaid. Therefore. All these annual costs are related with fixed cost. The next costliest components in a heat exchanger are tube sheets and main flanges. Then again. depending upon which cost is to be minimized. Annualized total cost to be minimized. especially in corrosive service. This fixed cost. at least part of the capital is to be borrowed. implies certain annual cost. repairs and maintenance cost could be significant. A) Minimization Of Annualized Total Cost Purchase and installation of a heat exchanger requires some capital expenditure. 2003 Page 34 Coming back to the definition of optimum once again. Interest is to be paid on the borrowings. its value depreciates. depreciation also becomes annual cost. Cost of these components depends on the diameter of the heat exchanger. On usage of the equipment. For the purchase of the heat exchanger.Process Dept Heat Exchanger Design Manual 6 th March. As all the heat transfer surface is provided in the form of tubes. the “most suitable” selection becomes a vague reference. This is needed only once. It is general accepted that tubes form the single largest cost component in a shell & tube heat exchanger. This becomes the annual cost. there is repairs and maintenance to be provided. Here again. there are two ways a design can be optimized. Hence. in-turn. Generally. Process Dept Heat Exchanger Design Manual 6 th March. This pressure drop plus the pressure drop in the pipe leading to and after the heat exchanger is to be made up by a pump or a blower. If this relationship is plotted in the form of a graph. it is known that the pressure drop for a fluid to flow through a pipe is inversely proportional to the fifth power of its (pipe’s) internal diameter. its cost also increases. Combining these two. Thus. Heat exchanger is static equipment. As the area of the heat exchanger increases. 2003 Page 35 exchanger is in turn a function of its area. Operating a heat exchanger necessarily involves flow of two fluids through it.e. of course. at . This. this pushing through the heat exchanger involves spending energy. the cost is naturally nil. The curve of operating cost as a function of heat transfer area is shown in figure-2. Therefore. The concepts of the foregoing paragraphs indicate that annualized fixed cost of a heat exchanger is a direct function of its area.5 th power of its heat transfer area. is over simplification. At zero area. Gravity flow for liquids and pressure flow for vapours and gases are not uncommon but in majority of cases. At the other extreme. This fact introduces some amount of non-linearity in the operating cost curve. The parameters of the type (b) mentioned in the previous topic thus become relevant in considering the variable cost. Cost of this energy becomes the operating cost of the heat exchanger. (the area for the flow of fluids is zero and hence) this cost becomes infinite. major cost components in a heat exchanger are all functions of its heat transfer area. the curve appears as shown in figure-1. Thus. the operating cost) comes not from operating the equipment itself but from operating other machinery to make use of the static equipment! From flow of fluids concept. This establishes the shape of the curve. The passage of the fluid through the heat exchanger means a drop in its pressure. This is the reason they were called performance-related parameters. Common sense indicates that the area of a heat exchanger is proportional to the square of its shell diameter. it stands to reason that the operating cost (i. the only variable cost (i. Tube side operating cost is also a function of the number of tube passes.e the pumping energy cost) of a heat exchanger is inversely proportional to approximately the 2. The flow is by pumping (for liquids) or by blowers or compressors (for gases). At zero area. at least one of the fluids is pushed through the heat exchanger. optimum value of pressure drop for the second (tube) side is found out. Thus. A new value of pressure drop is now assumed for one of the sides. 2003 Page 36 infinitely large area. For ready reference. The pressure drop value (for shell side here) corresponding with this minimum is the optimum pressure drop for this (shell) side. Among the recorded annual cost values. By several trials. . With the know percentage of the fixed cost as the annual cost. Adding the annualised fixed cost to the annual operating cost gives the total annual cost. This procedure is repeated a number of times with changed values of the shell side pressure drop. this cost becomes zero. This assumption is reasonable because the optimization of the process takes place before optimization of the individual equipment. There is a minimum on the curve 3. shell side) keeping the other side pressure drop (tube side) as constant. This (optimum) value of (shell side) pressure drop found as above is now kept constant and the other (tube) side pressure drop is now varied for optimization. This total cost is to be minimized for optimum process design of the heat exchanger. The total annual cost is then worked out and recorded. we are assuming that the type (a) parameters of the previous topic have been frozen. The way to reach this is given in the following paragraphs. Object of the process designing activity is to locate this minimum. minimum is located. The shape resembles what is known as a “rectangular hyperbola”. This is so because the parameters do not vary continuously but have discrete values. Attempts are then made to reach as near the minimum as possible.Process Dept Heat Exchanger Design Manual 6 th March. Combining figure 1 and 2 cost into figure-3. Each assumed value of pressure drop corresponds with a particular annual operating cost. the annual cost is also worked out. It is assumed that the flow rates and inlet-outlet temperatures are fixed by process requirements. (say. minimum heat transfer area solution is worked out. Entire set of calculations is repeated and the total annual cost is again recorded. Practically. keeping the tube side pressure drop constant. Fixed cost for this solution is calculated. A particular value of pressure drop is then assumed for each fluid. the total cost function would appear to be the curve 3. This establishes the asymptotic shape of the variable cost curve. it is not possible to reach the exact minima. curves 1 and 2 are also given in the same figure. With the assumed pressure drops. These values then become the real (or absolute) optimum values.e. The type (a) and (b) parameters being already frozen. geometrical parameters are to be selected in such a way as to result in a minimum cost heat exchanger for the given duty. Minimization of the fixed cost is very common. when the first (shell) side pressure drop was optimized. i. Sometimes. flow rates. The process is repeated till for both sides. square or rotated versions Tube diameter and thickness As the cost of energy is proportional to the product of pressure drop and flow rate. cost. Therefore. That is. Fir fixed cost minimization. For example. for some process reason. each heat exchanger must be optimized for minimum fixed cost. fixed cost minimization is required even during the total cost minimization.4 Minimization Of Only Fixed Cost The total annual cost minimization of previous section is somewhat rare. 10. 2003 Page 37 Earlier.e. effect of variation in the flow rate (with or without change in its outlet temperature) can be studied and the calculations for the total annual cost are made just as before. inlet and outlet temperatures. the constant value for the other (tube) side pressure drop may not have been at its optimum. for a given fixed pressure drop value. we are talking about type © i. Furthermore. It can be safely said that as a designing standard. the optimization procedure becomes relatively simpler. this type of optimization is very common. even with the help of a computer. One more cycle of trials is conducted to locate (perhaps) a new optimum value of pressure drop for the first side. both side fouling factors and both side pressure drops are finalized. type (a) parameters are involved. the optimum values found as above do not change. Till then. It can be readily perceived that a lot of trials may be needed to reach the optimum.Process Dept Heat Exchanger Design Manual 6 th March. which one on shell side and which one on tube side Tube length Type of layout – triangular. even flow rate variation introduces variations in the pumping energy . The parameters under the designer’s control are: Placement of fluids. one of the sides pressure drops is fixed. all earlier values can only be called as temporary or local optimum values. instead of type (b) parameters being involved in the optimisation. The entire procedure is quite tedious. geometrical parameters. Now. Only if. 2003 Page 38 Number of baffles. one aim to increase the heat transfer co-efficient for unit pressure drop is higher on tube side compared with the same on shell side. Other considerations may also dictate placement. Shell side has no such access except in U-tubes or full floating head (removable tube bundles). This facilitates provision of adequate turbulence by increasing number of baffles. There are some particular constraints also which have to be considered for each individual parameter. The restrictions re called as constraints. cooling water (which is likely to deposit scales) is generally placed on tube side. Generally smaller flow rate fluid is placed on the shell side.5 Placement of Fluids Some guidelines are given in the textbooks. This facilitates mechanical cleaning of tubes from inside.Process Dept Pitch for the layout Heat Exchanger Design Manual 6 th March. not all the parameters are under designer’s control. 10. For instance. For process requirements or plant space requirements. Space considerations dictate that the diameter of the heat exchanger be less than a certain value Tube side fluid velocity should not be below a certain value to avoid erosion or abrasion of tube wall. Some of the very common constraints areSpace considerations dictate that the length of the heat exchanger be less than a certain value. The optimization is then called constrained optimization. some of the parameters are not to be varied. Pressure drop on tube side should be below a certain given value Mass flow velocity on shell side should be below the given value Mass flow velocity through the baffle window should be below the given value These constraints are general constraints. i. But increased turbulence. These will be discussed along with discussion on each parameter. . Even then. It no long remains true optimization.e. baffles spacing Baffle cut Number of shell passes Number of tube passes Truly speaking. Let us now see the effect of each parameter on the optimization. mechanical cleaning from outside is quite difficult. Process Dept Heat Exchanger Design Manual 6 th March, 2003 Page 39 In general, highly befouling fluids that need frequent mechanical cleaning of the heat exchanger are usually placed on tube side. Highly corrosive fluids are preferably placed on tube side. Whichever side a corrosive fluid is placed, the tubes, where the heat transfer takes place, have to be of corrosion resistant material. Even the tubesheet has to be in corrosion resistant material. But, compared with the cost of entire shell, the cost of channel and heads in corrosion resistant material is substantially less. Fluids with very high operating pressure (when the other fluid has relatively much less pressure) are preferably placed on tube side. There are two reasons for this. First, only channel and heads are to be in higher thickness as against the entire shell. Second the tube thickness to withstand internal pressure. Of course, the diameter of tubes is generally small and the thickness required to withstand higher pressure is not much more than that for low pressure but when all the tubes are thicker, it may add considerably to the cost of the heat exchanger. 10.6 Tube Length For the same heat transfer area, less number of longer tubes are required. This makes it possible to fit the tubes in a smaller shell. As the cost of tube sheet, flanges and the shell are dependent on the inside diameter of the shell, this results in a less expensive heat exchanger. It is however not possible to increase the tube length indefinitely. cumbersome to handle. Too long an exchanger is Restriction on maximum tube length may be there from plant layout considerations (item (I) above). For a longer heat exchanger (with smaller shell diameter), the pressure drop, especially on the shell side, is much higher. If the allowable pressure drop limit is relatively low, long tube lengths do not yield an economical solution. Except for expensive and made-to-order tubes, there are certain standard lengths only in which the tubes are available. Therefore, variation of tube length is generally in steps. Infinitesimal variation is not possible. 10.7 Tube Diameter, Pitch And Layout Larger the tube diameter, it is easier to clean the tube mechanically from inside. On fouling service therefore, it is desirable to go for larger tube diameters. On cooling water service, it is desirable to go for tubes with minimum outside diameter of 25.4 mm Process Dept Heat Exchanger Design Manual 6 th March, 2003 Page 40 Larger the tube diameter, in general, more thickness is required to withstand the pressure, whether it acts internally or externally. For a moderate pressure and limited range of tube diameters under consideration, there may not be any difference in the tube thickness required. Larger the tube diameter, less is the total heat transfer area accommodated in a given shell. Therefore, a solution satisfying the process duty is reached with larger diameter shell. Also, wherever minimum velocity constraint is specified (e.g. item (iii) as above), use of large diameter tubes usually fails to yield an optimum solution. Pitch is usually 1.2, 1.25 or 1.33 times the tube Outside Diameter (O.D) as per the standards of TEMA (Tubular Exchangers Manufacturers’ Association). Larger the pitch, less number of tubes can be fitted in a shell of given diameter. Hence, smaller pitch yields a more economical solution. This, however, is subject to availability of adequate pressure drop on shell side. When unusually small pressure drop limit is specified on shell side fluid, larger pitch may be desirable. Triangular pitch is more compact. More number of tubes can be accommodated in the given shell, as compared with square pitch layouts (for same tube OD and pitch). Correspondingly, the shell side becomes more crowded and the pressure drop on shell side is more. Heat exchangers requiring shell side cleaning use removable tube bundles and also use square pitch layouts/ These bundles are relatively easier to clean than with triangular layout, or, use of only particular tube OD and pitch). Correspondingly, the shell side becomes more crowded and the pressure drop on shell side is more. There are certain combinations that are industry standards¾” (19.05mm) OD tubes on 15/16” (23.81mm) triangular pitch 10.05mm OD tubes on 1” (25.4mm) triangular pitch 19.05mm OD tubes on 25.4mm square pitch 25.4mm OD tubes on 11/4” (31.75mm) triangular pitch 25.4mm OD tubes on 31.75mm square pitch Sometimes, additional constraints are placed on the design, such as, use of only particular type (square or triangular) of layout, or, use of only particular tube OD, or, use of only one specific combination of OD, pitch and type. Process Dept Heat Exchanger Design Manual 6 th March, 2003 Page 41 10.8 Baffle Spacing Baffles usually mean a small percentage of total cost. This does not mean that their effect on the cost can be ignored. Difference in the cost of 6 baffles and 9 baffles may not be significant but difference between the cost of 6 and 30 befalls may be very significant. Significant or not, every baffle added is an extra cost not only of the baffle itself but of the spacer tubes, tie rods and the labour cost for assembly into the shell also. Furthermore, considering that the baffles and tube sheets are drilled together, every baffle adds to the thickness of the bunch for drilling. More the number of baffles, higher is the heat transfer co-efficient and so(higher) is the pressure drop too. This leads to the philosophical observation that nothing is free in nature. One has to pay the price for better heat transfer co-efficient. Increasing the number of baffles does not always result in less heat transfer area. For condensers without desuperheating and subcooling (on shell side) there is no change in the heat transfer area. For boiling in a kettle type reboilers also, there is no change in the heat transfer area. In both these cases, the baffles act more as tube bundle supports than heat transfer promoters. In case of condensation in presence of non-condensable and all applications of sensible heat transfer (including desuperheating and/or subcooling in total condensers), the additional turbulence helps to increase the heat transfer co-efficient. In all the cases when the number of baffles has been increased, the price of added pressure drop has to be paid, whether there has been an increase in the heat transfer co-efficient or not! Good engineering practice dictates that there are limits to the range over which the number of baffles can be varied. Standards of TEMA suggest that baffles should not be placed closer than D/5 or 2” (50mm), whichever is more. Here, D is the internal diameter of the shell. Similarly, baffles should not be placed farther than D or 24” (600 mm whichever is lower). This 600mm is not an absolute value. It depends upon the tube diameter and the unsupported span the tube can sustain without deflecting too much. Also, these are suggestions only. There can be exceptions. Then again, the mass velocity constraints [item(vii)above] impose further limits on the range over which the baffle spacing can be varied. g. it is possible to get lower pressure drop even with more number of baffles. the lowest cut/s may not be admissible. considering all the constraints. starting at minimum number and increasing this number by one each time. it is more prudent to work out the maximum and minimum number of baffles that can be placed. beyond 40% (some people take this value at 45%) cut value. 10.3% is also is also included as a distinct value within this range.9 Baffle Cut Only bearing this factor has on optimization is in cases where pressure drop on shell side becomes critical. Therefore. 33. 2003 Page 42 The baffles spacing cannot be varied infinitesimally. Generally. Again. the spacing becomes a step value type variable. In this case. I wish to emphasise that without adequate allowable pressure drop on shell side fluid. variations in cut can apply. All this discussion assumes that design algorithms are available in which effect of baffle cut on heat transfer co-efficient and pressure drop can be calculated. The range of baffle cuts over which the variation is 15% to 40% in steps of 5%. disc and doughnut) also. the extent to which this can be done is limited because the contribution of baffle window pressure drop in the overall pressure drop on shell side is less than the cross flow pressure drop. the procedure for calculation of heat transfer co-efficient and pressure drop is different. where constraint item (viii) as above applies. The number of baffles must be an integer. Many manufacturers and designers prefer to maintain the baffle cut at only 25% which happens to be the most common value and for this. vary the number of baffles within this range. Of course. It is an uncommon procedure and hence not easily available in the standard textbooks. Then. heat transfer co-efficient and pressure drop calculations are readily available. For other types (e. For this type of flow.Process Dept Heat Exchanger Design Manual 6 th March. The percentage baffle cut need not imply only segmental baffles. it is meaningless to try and optimize the number of baffles. For segmental baffles. . In some cases. by increasing the baffle cut. the shell side flow becomes more axial than cross. It is desirable to place more number of baffles to get smaller heat transfer area but the pressure drop limit is somewhat unsuitable. To place these baffles in the shell in (as far as possible) an equidistant manner. the range becomes narrower. In such a case. ) is small enough. generally above 1. For these two ranges. The new pressure drop is nearly eight times the original. the partition can be welded to the shell. configuration before checking suitability for process duty. the shell partition plate can be put using sealing strips. If the shell I. the number of times the fluid has to change direction at baffle becomes double. on account of large temperaturte overlap. For intermediate diameters 400<I. 2003 Page 43 As a matter of detailing. Another requirement before considering a second shell pass is the availability of liberal pressure drop on shell side fluid. 10. 10. If it is not possible to provide two shell passes but the process requirement is very much there. This becomes difficult when the cut is large.D.D. 35% or 40%. For some applications. This is as recommended by TEMA. Common values of number of tube passes used in the industry are 1. If the shell inside diameter (I. the passes in these heat exchangers are very high. costly and perhaps unnecessary. the sealing between the passes is not satisfactory and there are leakage’s of fluid from one pass to the other between the partition and the shell wall. Alternatively.000mm.D. some prefer to omit 40% and more cut baffles from optimisation options. The alternative of two shell passes becomes relevant only when there is a large temperature cross for the fluids. it is possible to think of a shell with two passes made with longitudinal baffle. which the standard 1-2 pass configuration heat exchanger is unable to handle effectively.10 Number Of Shell Passes Usually.000 mm.11 Number Of Tube Passes This parameter plays an important role in the optimisation of the design and is at the disposal of the designer.2.4 & 6 and 8. say. often about 16 or 20. For this reason also.g. the exact working out of the tube counts is a This gives a definitive picture of the . is large enough for a person to go inside. tube count procedures are also available. For these values of number of passes and the five standard tube layouts mentioned earlier. generally upto 400mm. two or more shells in series are thought of instead of two shell passes in one shell.Process Dept Heat Exchanger Design Manual 6 th March. This aspect becomes critical as the shell side cross flow area becomes half what it would be without the shell baffle and at the same time. juice heaters in sugar industry. tube count tables are easily available in the textbooks and handbooks. In these cases. This is because providing a pass partition on shell side is difficult. e. single shell pass heat exchangers are used in the industry. it should be ensured that each baffle is supported at minimum three points. < 1. it is not easy to get the tube counts nor is it easy to calculate the value.Process Dept Heat Exchanger Design Manual 6 th March. where the shell side heat transfer co-efficient is very good. This places some restrictions on the degree of control over the optimization process. pass partition plates have to be provided on the tube side (inside the channels). For these passes also. This benefit is not without its usual price of higher-pressure drop. Use of odd number of passes. the tube side pressure drop that has been increased to achieve this co-efficient continues to use more energy and consequent more operating cost.. For processes in which the shell side heat transfer co-efficient is very poor. 3 or 5 is very rare in the chemical industry/ it is certainly not totally absent. This is because the turbulence on tube side increases. this cost benefit ratio is better than the same available on shell side. the pressure drop increases by about eight fold while the heat transfer co-efficient increases by only about 1. For creating tube side passes. As with many parameters in designing of heat exchangers. Thus. the effect is very nominal for large diameter shells. More the number of passes. it is always beneficial to increase the heat transfer co-efficient on tube side by increasing the number of tube passes. It is to be worked out by plotting and counting process. No standard table or simple procedure is available for tube count. More the number of passes. condensers for vapour. e. This makes it necessary to remove certain number of tubes to accommodate the partition plates. Conversely. this parameter also can be varied in certain steps. Even then. worked out on case-to-case basis. The case presented as figure-4 will illustrate this point numerically.g. there is a reduction in the available heat transfer area within the given shell of particular diameter. By doubling the number of passes.g.74 times. On the other hand. Even for cooling of non-viscous liquids (on shell side) with cooling water. Increasing the number of passes does not always help. more tubes are to be removed. Infinitesimal variation is not possible. more is the heat transfer co-efficient and less is the heat transfer area required to perform the duty. substantial increase in the tube side co-efficient also fails to make enough impact. The area reduction is considerable for shells of smaller diameter. it is always possible to derive some benefit by increasing the tube passes. e. 2003 Page 44 tedious job. This feasibility comes from . Process Dept Heat Exchanger Design Manual 6 th March. in vertical thermosiphon type (tube side) reboilers. only single pass design is acceptable. The very construction makes it impossible to have single pass. Different manufacturers also set different standards. In all these cases. For example. Second half of figure-4 illustrates this point numerically. There may be some constraints over the number of passes. There are no known standards. it is generally possible to accommodate more number of tube passes than when larger diameter tubes are used. it becomes difficult and cumbersome (and sometimes even impossible) to accommodate the partition plates for large number of passes. Usually the designer’s opinion of what constitutes good engineering practice becomes the guiding principle. for small shells. It may be theoretically possible to have 6 or 8 passes but practically it is very difficult to provide these number of tube passes in a U-tubes heat exchanger. full variation of tube passes during optimization is not possible. Therefore. large pressure drop is available on cooling water. it is a standard practice to place an upper limit on the number of tube passes. only 2 or 4 pass design is required. Another factor that comes into play is the tube diameter. if the tube diameter is small. In U-tube bundle type heat exchangers. In a shell of given inside diameter. . 2003 Page 45 the fact that heat transfer co-efficient on both sides become comparable and at the same time. For small diameter shells. for low-pressure drop designs. . Alternatively. This point has been discussed adequately in the previous section. progressively. adequacy of heat transfer area and pressure drops within allowable limits. 2003 Page 46 First attempt for finding the optimum solution must start with the minimum number of shells.Process Dept 10. or. Having started with minimum number of shells of a certain I. The last is a very rare occurrence. which satisfies the given duty. one has to take recourse to using more number of shells in series. The reason is not easily apparent. well and good. Otherwise. is then selected.. the designer now has to select all the other parameters within the shell in such a way that minimum area can be fitted in. By taking two shells. For problems involving large temperature overlap for fluids. sometimes this solution comes out to be less expensive than trying to fit the design in one shell. one can ensure that always an optimum (or nearly an optimum) design is reached. When the flow rate of one (or both) of the fluids is very large and when the corresponding allowable pressure drop is less. Keeping only one shell is not a difficult task to fulfil for most of the case an engineer encounters normally. After fixing all the parameters (no doubt within the given constraints). If this combination of parameters is suitable from process performance point of view. The first combination. The pressure drop is one fourth! This makes it economical when overall heat transfer area requirement and its cost are considered.. other combinations with increasing heat transfer area are tried out. the flow in each is halved.D. This way. No doubt. i. or both. one may think of split flow or divided flow patterns in a single shell. it is sometimes better to think of two shells in parallel.12 General Strategy Heat Exchanger Design Manual 6 th March. the geometry is now checked for satisfying the performance.e. generally it is an expensive alternative but surprisingly. This strategy is consistent with the knowledge that tube sheets and flanges cost double for two shells if the required heat transfer area is provided in two shells instead of in one shell. using shells with two shell passes. 3 TOTAL Cm ANNUAL COST FIXED OPRATING Ao HEAT TRANSFER AREA .Process Dept Heat Exchanger Design Manual 6 th March.2 ANNUAL COST HEAT TRANSFER AREA TOTAL COST FIGURE .1 FIXED COST ANNUAL COST HEAT TRANSFER AREA OPERATING COST FIGURE . 2003 Page 47 FIGURE . S.2 97 97. Cu S.S.6 2255.S.6 2669. Cu S. Cu 2000 1000 5000 666.9 1482.13 Cost benefit analysis of heat transfer improvement and effect of tube material Shell side film Tube side film Overall clean Overall clean Remarks coefficient ho coefficient coefficient coefficient c kcal/hr m2°c hio kcal/hr w/o tube Uc tube Ucr m2°c kcal/hr m2°c kcal/hr m2°c 94.S.4 805.S.S.2 94. Cu S. Cu S.9 95.6 98 620 648.9 3000 Tube conductivity in Kcal/ hr (m2/m) °C SS 15 MS 40 Cu 330 Tube thickness 16 BWG (1.Process Dept Heat Exchanger Design Manual 6 th March.6 829. Cu S.7 833.5 763. M.S.2 2000 100 5000 98 95.8 664.4 1566. 2003 Page 48 10.S. M. M. M. M.65 mm) .3 2000 100 5000 1578. M.S.9 1339.S.S.S.7 S.6 2995. hydrocarbon conductivity is lower than water conductivity making design more conservative. Corrosive fluids to be taken at tube side. without water. Maximum number of tube passes can be used keeping watch on the tubes side pressure drop & design velocity. Reason behind is normally water forms slugs while flowing with hydrocarbon therefore heat transfer coefficient is mainly governed by hydrocarbons & not by water. Moreover. However. shells in parallel can be used.Process Dept Heat Exchanger Design Manual 6 th March. Floating head heat exchangers are typically limited to the shell ID of 1. F type of shell can also be used to reduce more number of shells. High-pressure fluids preferably to be kept on tube side to minimize the shell Fluids with high viscosity to be kept on shell side so that turbulence can be induced Two-phase flow to be kept on shell side. Whenever heating cooling curve property points given are more than 10. In case of exchangers where very low-pressure drop is allowable. But design can be started by using 20mm tube OD. & points are decided. But we can still go for higher shell ID in case of fixed head heat exchanger. • • • • Longer shell is always cheaper than big shell diameter.e. • • • • • For pressure above 15bar. thickness.  by introduction of baffles. streams being heated flow from bottom to top or streams being cooled normally flow from top to bottom. It can also be used to improve velocity & hence heat transfer coefficients. The fluid of smaller volumetric flow rate to be taken on shell side. F shell with 2 passes on tube side acts as a pure counter current exchanger. 2003 Page 49 11 Generic guidelines for heat exchanger designer • • As a normal practice.  • Property input to HTRI: Whenever there is water present along with liquid hydrocarbon. Selection of tube or shell side:     The fluid with high fouling tendency preferably to be taken at tube side. always conductivity of the dry stream needs to be inputted i. . plot of enthalpy versus temperature is plotted. Typical tube sizes used are 20mm or 25mm. More than one number of shell can be used as unit in order to avoid local temperature cross.4-1. property curve is inputted at only one pressure level. Tube length should be maximized as far as possible to utilize maximum available pressure drop. but below 15bar it is advisable to enter at least two-pressure level property curve with same inlet temperature.5m. 2003 Page 50 Tubes can be arranged in square. • • Any change in the baffle geometry should be done by watching pressure drop on shell side. • Sealing strips or seal rods can be inserted in order to minimize bypass fraction (stream B) & inline bypass fractions (stream F). or H2 is present in the process.35 mm.     Square pitch with 20mm tube OD: 26 mm tube pitch Square pitch with 25mm tube OD: 32 mm tube pitch Triangular pitch with 20mm tube OD: 25 mm tube pitch Triangular pitch with 25mm tube OD: 31 mm tube pitch • In case H2S. Also some cases such as. . Baffle cut can vary between 15 % to 45% of the shell inside diameter. • Normally used tube pitches. both very small & very large baffle cuts are detrimental to efficient heat transfer. • • Baffle cut can be adjusted to maximize cross flow. • • Single segmental baffle can be used for normal exchanger designs. care should be taken that gravity separation zone is avoided on the shell side. Horizontal baffle cut can be used for vaporizing services. Similarly horizontal baffle cut is also called as perpendicular cut or transverse baffle. Number of baffles & baffle pitch can be decided based on allowable shell side pressure drop. Minimum baffle spacing should not be less than one-fifth of shell diameter or 50. Baffle cut orientation should be based on process conditions. as a tube pitch is the exchanger is under H2 service. double segmental baffle can be used. welded tubes are required. Triangular pitch provides more compact design & good heat transfer coefficient. For welded tubes minimum tube gap between adjacent tubes should be at least 6. 3 rd point covered above will also require at least 26 mm. Normally phase change operation on shell side is provided with vertical baffle cut. low allowable pressure drop on shell side or shell side vaporisation. triangular rotated square or rotated triangular patterns. But. But whenever shell side cleaning is required square pitch is preferred which enables better cleaning of the shell side. In cases where allowable pressure drop on shell side is low. square pitch can be preferred.8 mm (2 inch) whichever is greater. maintain allowable pressure drop and avoid bypass & leakage fractions. It is also called as parallel cut or inline baffle. Generally. In that case. is recommended that only baffle cuts between 20% and 35% be employed. If horizontal cut is given for condensing service.Process Dept • Heat Exchanger Design Manual 6 th March. 1 What exchanger designer should analyze in HTFS / HTRI output  Output produced by HTFS / HTRI reflects all the data fed by designer.2. material such as aluminum brass 90:10 or 70:30 cupronickel or titanium is used. admiralty metal is normally specified.  Check whether fouling coefficients considered are appropriate & are as per Technip standard or client requirements. The minimum recommended liquid velocity inside tubes is 1.2 In general very high velocities lead to erosion. Fluid Cooling water Sea water Sea water Brackish water • Tube material Carbon steel / Admiralty 90:10 Cupronickel Aluminum brass Carbon steel Preferred velocity (m/sec) 2 . Based on clients requirement area ratio need to be decided (for example 1. 11.5 2 – 2.15).0m/sec &. carbon steel is normally specified.10 .0m/sec. However some clients prefer admiralty brass. generally nozzle size should be such that pressure drop contributed by nozzle should not be more than 15% of exchanger pressure drop. 2003 Page 51 Pressure drop across nozzles is considered as a part of heat exchanger pressure drop hence. 10% or 15%. as it is their hardware. 1. • • Exchanger which are subject to steam out or where there is a possibility of being subject to vacuum conditions during operation or shutdown should be designed for full vacuum.5 1.  Area ratio of actual to required should be at least 1in case of HTFS. dirty heat transfer coefficient gives the real picture of the design. Clean coefficient is without considering any fouling & service coefficient is a back-calculated value from the heat transfer equations. For other fresh water services. while maximum is 2. For HTRI % overdesign should be be checked i. hence it should be checked whether the data considered by HTFS is appropriate. For sea or brackish water. .e. For cooling water exchangers with treated or inhibited cooling water service.  Calculated pressure drops on both the sides (shell / tube) should be less than allowable pressure drops for their respective side.5 – 2 1.  Among three figures of overall heat transfer coefficient (clean / dirty / service).Process Dept • Heat Exchanger Design Manual 6 th March.5-3. In case single-phase exchanger. Hence it is advisable to give maximum number of points between these two to get best possible result. two property points also gives proper result. Best way is to give immediate first preceding property point of bubble/dew point. attempt should be made to avoid it by changing the span length as discussed earlier. side. You may wish to consider putting the high-pressure stream in the tube .45 in good designs.  In general.  Clearances such as tube/baffle.  Minimum baffle spacing should not be less than D/5 or 50mm. bundle/shell or baffle/shell used by HTFS / HTRI are as per TEMA. based on type such as fixed or floating head. if the fatal error exists. Generally it should be more than 0. it is always better to use at least 150mm-baffle spacing.Process Dept Heat Exchanger Design Manual 6 th March. the program will not run successfully.  Heating / cooling curves provided should have as many property points as possible while designing 2-phase exchanger. You The shell side design pressure is greater than may wish to put it on the tube side. Some warnings can be ignored after due consideration such as 1. clearances etc. Based on this number of baffle should be decided. that of the tube side. which is easier to clean.  Shell side flow fractions should be checked for cross flow (stream B). But in practice. The most fouling fluid is on the shell side. These clearances could be changed only after discussion with mechanical department. Thus end lengths can be different for front head & rare head. But warning messages should be read carefully & attempt should be made to rectify them.  Requirement of impingement should be identified based on TEMA guidelines & it should be checked for the given design. nozzle size should be checked for pressure drop not more than 15% of the total. 2.  In case of vaporizers it should checked that bubble point & dew point considered by HTFS / HTRI is same as that in property package.  All other fractions are leakage. shell flange etc. 2003 Page 52  Output should be read for warnings messages / fatal errors.  Inlet / outlet end lengths should be sufficient enough to accommodate nozzle. which should be reduced by changing geometry such as baffle cut. In general.  Output should be checked for vibration analysis & if present any.  Normally in vaporizers it is observed that approximately 90% area is required for between bubble point & dew point. sealing strips etc. Can not force inlet / outlet dome height for the exchanger.Process Dept Heat Exchanger Design Manual 6 th March. shell flange etc. RPM (resistance proration method) & CPM (Composition profile method) are two important methods available for condensation. HTFS Design mode of HTFS is more user friendly & gives better options. sliding strip. RPM is a default method can be used widely & is applicable for most of the cases. Physical property based method which uses theoretical boiling range mixture correction factor. 11. In certain cases. . It is possible to put very minute data in HTRI. Also tube layout should be checked for incorporation of tie-rods.2 Important Points specific to software HTRI In case of vaporizers in HTRI. tube layout produced by XIST / XTLO is identical. While designing exchanger minimum certain end length is necessary for accommodating nozzle. impingement plate. Different constraints can be forced to get desired results. It is not possible to incorporate seal rods in HTFS other than sealing strips. In some of the cases where there is a presence of inert gases. seal rods needs to be incorporated to minimise inline pass partition stream. is used in case of vaporisation & gives satisfactory results. hence it is useful for running existing exchanger such as revamp. In general it can be said that HTRI gives better result in case of phase change operations. HTFS does not give any warning & hence it can lead to a fabrication problem. 2003 Page 53  Tube layout should be checked on OPTU (HTFS) for checking whether tubes are fitting in the given shell ID. whereas CPM should be used in case of presence of inert gases in the condensing side. mole fraction of inert gas should be specified. If in case given end length is not sufficient enough. Unlike TASC / OPTU. boiling range should be necessarily specified since the same affects heat transfer coefficient. If this is omitted HTRI calculates the same but BR calculated can be over or under predicted which will in turn affect heat transfer coefficient. but it asks for accurate VLE data. Hence output given by TASC should be checked in OPTU. Variable baffle spacing can not be used in HTFS. Maximum numbers of tubes. differ in TASC & OPTU. if it is giving vibration simple remedy such as incorporation of additional support plate per baffle spacing solves the problem. 2003 Page 54 In case of NTIW design. . But this kind of option is not available in HTFS. which can be fitted in given shell ID.Process Dept Heat Exchanger Design Manual 6 th March. lower boiling point elevation is there which saves some heat transfer area. sometimes ‘E or J‘ can also be used. Vertical thermosiphn reboilers are cheapest of all since they are generally fixed tube sheet exchangers. & high boiling heat transfer coefficient.9. It has a lower fouling potential as compared with kettle and are lesser sensitive than vertical thermosiphon reboilers.1 Reboiler Selection: Generally kettle type reboilers. vertical & horizontal thermosiphon reboilers are main types of reboilers. Horizontal thermosiphon reboiler: Generally ‘G or H’ shell is used for minimizing the pressure drop. Vertical thermosiphon reboiler: Vertical thermosiphon reboilers are generally single tube pass since boiling is inside the tube. heating medium needs to be clean. . they are very costly & they have very high fouling tendency since there is little turbulence.3 and 2000 kcal/hr m2 °C for Pr = 0. Kettle type reboiler: These can vaporise fully column bottoms & acts as a additional stage for distillation. Disadvantages of this type of reboiler are it give lower heat flux than shellside boiling especially near critical pressure. Maximum allowable heat flux is generally 30000 kcal/hr m2 °C for Pr = 0. They are very sensitive to changes in operating conditions. Horizontal reboilers can achieve high circulation rates. 2003 Page 55 12. Reboiler Heat Exchanger Design Manual 6 th March. Since fixed tube sheet is generally used. Kettle type reboilers can achieve high heat flux. It has highest MTD due to pure counter current operation. high convective heat transfer component. Disadvantages of horizontal thermosiphon reboiler are more fouling on shell side compared with vertical type. Generally for clean heating mediums this can be preferred with U tubes. 12. Since these reboilers require lower static head. But disadvantages of the kettle type are.Process Dept . Fouling is less pronounced since high circulation & shear stress. Since kettle type reboilers principle is based on pool boiling it is less sensitive to hydrodynamics. Kettle type is preferred near critical pressure region over others. 4 for horizontal thermosiphon reboiler. But this method is limited to HTFS software only.4 bar/km Velocity < 1 m / sec Outlet piping: RHO-V-SQ < 6000 kpa RHO-V-SQ (only for vpour) > 100 kpa …required for separation. heat transfer area can be altered. . Once typical inlet / outlet circuit geometry is fixed. This circulation flow is checked for its adequacy.e. Thus thermosiphon reboiler needs to be checked for both recirculation quantity & thermal adequacy. Once exchanger design is ready it is checked for its thermosiphon reboiler duty.35 times the available static head. if two numbers of shells are required then reboiler should be run for the half of the duty & circulation.2 Design of reboilers To start with. recirculaton quantity gets fixed.3 for vertical & 0. 2003 Page 56 12. In such cases. Frictional plus acceleration pressure drop should be less than 0. it is not possible to fit it in one shell. Thermal adequacy of this exchanger is checked by seeing whether reboiler is capable of carrying out given heat duty i. both the pressure levels should have same inlet temperature point.3 to 0. reboiler is designed as a normal exchanger to carry out specified heat duty. Driving force for this circulation is level in column & resistance are pressure loss in exchanger and inlet / outlet pipeline of exchanger.2 to 0. In these cases. General criteria for checking thermosiphon reboiler output: Inlet piping: Pressure drop < 0.Process Dept Heat Exchanger Design Manual 6 th March. Designing thermosiphon reboiler includes fixing of inlet / outlet geometry. if it can generate required amount of vapor. This is because HTRI does not allow common inlet & outlet piping for parallel exchangers. it can be directly runned for thermosiphon reboiler operation. If required. inlet / outlet piping can be changed to achieve stability. If required. In case of HTRI software. Recommended weight fraction of vapor at outlet is 0. In some of the cases where heat transfer area required is very large. In case of reboilers it is very essential to provide heating curves at second pressure level also. 0 Typical values of heat transfer coefficient Shell Side Boiler feed water Boiler feed water Blow-down Demin water Fuel oil Gasoline Hydrogen Ethylene HT-shift Kerosene / gas oil Naphtha Natural gas Nitrogen Process condensate Process steam (superheating) Process steam (superheating) Steam Water Process water Tube side Blow-down HT-shift product Cooling water Water Water Water Cooling water Cooling water Boiler feed water Water Water Brackish water Cooling water Water Ht-shift product Process gas Naphtha Water MP condensate U (kcal/ hr m2 °C) 600 400 500 2000 100 400 500 325 300 250 250 225 175 500 175 250 200 1000 800 14 Typical fouling resistances used for designing heat exchanger Following are the typical values of fouling resistances. Process side Sour Naphtha Fouling factor (Kcal/m2 hr °C) 2000 .Process Dept Heat Exchanger Design Manual 6 th March. 2003 Page 57 13. which should be used unless otherwise dictated by client. Because to calculate the spacing distance the information on tubesheet thickness and flange thickness are necessary which can be calculated during mechanical design. Inlet/outlet baffle spacing 2. Impingement plate placement 4. 1. Dome height top/bottom 3. 2003 Page 58 Sweet naphtha Cooling water Sea water Boiler feed water Blow down water Steam Demin water Fuel oil Gasoline Kerosene LPG MEA solution Methanol Process gas (shift feed / product) Hydrogen / Nitrogen Refinery gas Regenerating gas 15 Mechanical aspects The mechanical aspects to be considered during thermal design of the exchangers and after the Mechanical design of the exchangers are discussed below.Process Dept Heat Exchanger Design Manual 2500 2500 1650 5000 2500 5000 5000 1000 5000 5000 5000 2500 5000 5000 5000 1650 1650 6 th March. . Tubesheet thickness Inlet/outlet baffle spacing Once the central baffle spacing (baffle pitch) has been frozen during the thermal design of the exchangers the inlet and outlet baffle spacing can be calculated approximately. End lengths . 2003 Page 59 During thermal design of shell and tube exchangers inlet baffle spacing can be calculated roughly P t u b e ts h h k e e t N o zd zi a l e I m p in g e m e n t p B a f f l e s p a c i n g a d w i d t h l a t e f l a n t hg ke t u b e p r o j e c t i o n o u t ts u i d b e e s h e e t 3 m m and used in the design of the exchanger.Process Dept Heat Exchanger Design Manual 6 th March. Nozzle diameter. once the inlet baffle spacing and central baffle pitch have given to HTRI/HTFS as input the program calculates the outlet end spacing. will have an effect based on the clearance. Clearance between reinforcement pad and first baffle (30-50mm). In case of floating head exchangers if the shell fluid enters at the floating head side and support plate is present the tube area between the support plate and the tube sheet is considered as lost area and the inlet spacing will be considered only upto support plate instead of tubesheet. Outlet end spacing This will be calculated as same as that of the inlet end spacing. To calculate roughly the inlet spacing sum the following. This can be done by the spreadsheet available in the process console. This distance normally depends on the impingement plate presence and its placement. Flange thickness (take from previous projects for similar exchanger) Clearance between shell flange and nozzle reinforcement pad (50-100mm) Twice the Nozzle dia for Reinforcement pad width on both sides. The check for the ρv² value in this region to be carried out in later stage during mechanical design to ensure that they are with in TEMA range. Because the distribution of the shell side fluid towards the impingement plate and towards the tube bundle. Impingement plate The need for the provision of impingement plate has been discussed before. now the dimensions and placement of the same will be discussed. The impingement plate can be either a circular plate or a rectangular plate and have the dimension of 1” (25 mm) more than the nozzle diameter on all the sides. These dimensions are available exactly only after the mechanical design. reinforcement pad width (twice the nozzle dia). The dome height as calculated by the HTRI/HTFS program must be analyzed and checked for the ρv² values that they are within TEMA range (5953 kg/ms²). shell exit and bundle exit will lead to vibration of tube bundle and unwanted pressure drop at the shell entrance and exit. clearance between the shell flange weld to the reinforcement pad. Insufficient area for the flow of shell side fluid at shell entrance. In case of a floating head exchanger the additional backing ring thickness will be added to the above dimensions. flange thickness. bundle entrance. clearance between the pad and the first baffle.Process Dept Heat Exchanger Design Manual 6 th March. The impingement plate is generally placed on the tube bundle supported by the tie rods (welded). The shell entrance and exit ρv² values are also available in the HTRI output. 2003 Page 60 The end length describes the tube length between the baffle and the tube end. For example the inlet end length for the diagram shown above will be the summation of tube projection (3mm). B S a c h k i n g r i n g g e e l l f la n Dome height Top/Bottom: Dome height describes the flow area for the shell side fluid between the nozzle and the impingement plate. tubesheet thickness. If tie rod is not available on that position then local tie rods will be provided . The minimum tubesheet thickness can be as follows. Thickness – corrosion allowance ≥ Thickness including C. Finally the ρv² check must be done on all sides of the plate by the available spreadsheet in the process console. Tubesheet thickness The tubesheet thickness value will be calculated by the HTRI/HTFS programs and they must be analyzed during the thermal design of the exchanger. . 2003 Page 61 to support the impingement plate. The values in the output to be checked whether they are reasonable by comparing with previous project data. If not then the tubesheet thickness to be manually fed in the program for design.A ≥ Tube O.D ¾” Will be developed further in future.Process Dept Heat Exchanger Design Manual 6 th March. 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