Exxon Design Practices Part-1 PDF

Course Description Exxon Design Practices Part-1 Course goal The goal of this course is to give the participants an overview and knowledge of the Exxon Design Practices for safe plant design. This course covers primarily the design of Pumps and heat exchanger and their application, and design basi...

Course Description Exxon Design Practices Part-1 Course goal The goal of this course is to give the participants an overview and knowledge of the Exxon Design Practices for safe plant design. This course covers primarily the design of Pumps and heat exchanger and their application, and design basis and design case studies. In addition to Theoretical knowledge, this course imparts hands-on experience of the design strategies involved. It also involves the basics of fluid flow and impact of temperature and pressure on it. Learning objectives Participant profile Upon completion of this course the This training is targeted for CL 21-23 based participants will be able to understand: on area assignment of Process MPTs Basic terms used in Pumps and heat Prerequisites exchanger Types of Pumps and heat exchanger used The participants should be familiar with basic in process designing concepts of process designing. Application of various Pumps and heat Duration exchanger Steps involved in Sizing Pumps and heat The duration is 4 weeks. exchanger Selection of Pumps and heat exchanger Page 1 of 72 Module 1: Exxon Design Practices Part-1 Page 2 of 72  Table of Contents Section 01.................................................................................................................................................. 7 1. Design Temperature & Pressure.............................................................................................. 7 1.1 Introduction to Design Safety:........................................................................................... 7 1.2 General Principles of Safety in Process Design:........................................................... 7 1.3 Minimizing the Risk of Fire, Explosion, or Accident:.................................................. 8 1.4 Equipment Spacing:............................................................................................................... 8 1.5 Designing for Equipment Spacing:.................................................................................. 9 1.6 Layout Considerations:....................................................................................................... 10 1.7 Preventing Equipment Failures:....................................................................................... 11 1.8 Maximum and Minimum Design Temperatures:...................................................... 11 1.9 Design Temperatures Selection:..................................................................................... 11 1.10 Critical Exposure Temperature.................................................................................... 13 1.11 Design Pressure:............................................................................................................... 13 1.1.1. Vacuum........................................................................................................................ 14 1.12 Piping Flange Design:..................................................................................................... 14 1.13 Piping Flange Design Temperature and Pressure Selection:............................ 15 Section 02................................................................................................................................................ 16 2. Fluid Flow........................................................................................................................................ 16 2.1 Types of Fluid Flow:............................................................................................................. 16 2.2 Flow Regimes:........................................................................................................................ 16 2.3 Two Phase Flow Regimes in Horizontal Pipes............................................................ 17 2.4 Flow Regimes in Horizontal or Slightly Inclined Pipes............................................ 17 2.5 Two Phase Flow Regimes in Vertical Pipes................................................................. 20 2.6 Flow Regimes in Vertical Pipes:....................................................................................... 20 2.7 Effects of Fittings on Two-Phase Flow:......................................................................... 22 2.8 Two-phase warnings:.......................................................................................................... 23 2.9 Pressure Drops Calculations:............................................................................................ 23 Page 3 of 72  2.10 Friction Losses:.................................................................................................................. 23 2.11 For Compressible Flow Applications:........................................................................ 25 2.12 Equivalent Length of Fittings....................................................................................... 25 2.13 Quick Reference to Pressure Loss through Fittings:............................................ 26 2.14 Design Basis for “Average” Carbon Steel Lines:.................................................... 27 2.15 Static Head Losses:.......................................................................................................... 27 2.16 Preliminary Hydraulic Calculations:............................................................................ 28 2.17 Typical Process Lines Equivalent Lengths:............................................................... 29 2.18 Final Hydraulic Calculations:......................................................................................... 30 2.19 Case Study:......................................................................................................................... 32 2.20 Choking or Critical Flow................................................................................................. 32 Section 03................................................................................................................................................ 34 3. Pumps............................................................................................................................................... 34 3.1 Types of Pumps:.................................................................................................................... 34 3.2 Centrifugal Pumps:............................................................................................................... 35 3.3 Special Types of Centrifugal Pumps.............................................................................. 35 3.4 Pump Affinity Laws:............................................................................................................. 39 3.5 Centrifugal Pump Design Factors:.................................................................................. 40 3.6 Fluid Properties & Differential Pressure:...................................................................... 40 3.7 Cavitation & NPSH:.............................................................................................................. 40 3.8 Calculating and Approximating NPSH.......................................................................... 41 3.9 NPSH Cheat Sheet:............................................................................................................... 41 3.10 Driver and Utility Requirements:................................................................................. 42 3.11 Motor Sizing( DP XXX-G)............................................................................................... 43 3.12 Design Pressure for Centrifugal Pumps:.................................................................. 44 3.13 Design Pressure for Other Equipment:..................................................................... 44 3.14 Summary of DP Downstream of Centrifugal Pump:............................................ 45 3.15 Case study:.......................................................................................................................... 47 Page 4 of 72  3.16 Mechanical Seals:............................................................................................................. 47 3.17 Mechanical Seals Details:............................................................................................... 48 3.17.1 Flushing Mechanical Seals:................................................................................... 49 3.18 Seal less Pumps:................................................................................................................ 49 3.19 Pump Hints:........................................................................................................................ 50 Section 04................................................................................................................................................ 51 4. Heat Transfer & Exchangers......................................................................................................... 51 4.1 Background............................................................................................................................. 51 4.2 Designer’s Role...................................................................................................................... 51 4.3 Engineer’s Toolbox............................................................................................................... 51 4.4 Heat Transfer Analogy........................................................................................................ 51 4.5 Heat Transfer Equation....................................................................................................... 52 4.6 Heat Transfer Coefficient Theory.................................................................................... 52 4.7 More on Total Heat Transfer Resistance...................................................................... 52 4.8 The Limiting Resistance...................................................................................................... 53 4.9 The Controlling Coefficient............................................................................................... 54 4.10 “Shortcut” Shell & Tube Design Technique............................................................ 54 4.11 Shortcut HX Design......................................................................................................... 55 4.12 Temperature Driving Force, ΔT................................................................................... 55 4.13 LMTD..................................................................................................................................... 56 4.14 Fn Correction Factor........................................................................................................ 56 4.15 Common Exchanger Types........................................................................................... 57 4.16 Shell and Tube Exchangers........................................................................................... 57 4.17 Major STE Types................................................................................................................ 57 4.18 TEMA Type.......................................................................................................................... 58 4.19 TEMA Heat Exchanger Nomenclature...................................................................... 59 4.20 Shell & Tube Exchangers............................................................................................... 59 4.21 Problem – Bundle Rear End Head Types................................................................. 60 Page 5 of 72 4.22 Preliminary Shell & Tube Decisions........................................................................... 60 4.23 Tube Side Fluid.................................................................................................................. 61 4.24 Tube Length....................................................................................................................... 61 4.25 Tube Layout........................................................................................................................ 61 4.26 Type of Baffle..................................................................................................................... 62 4.27 Rod Baffle Heat Exchanger (RBE)................................................................................ 63 4.28 Nucleate Boiling Tubes (NBT)...................................................................................... 64 4.29 Helical Baffle Exchanger (HBE)..................................................................................... 64 4.30 Twisted-Tube Exchanger (TTE).................................................................................... 65 4.31 Handy Factors.................................................................................................................... 66 4.32 Air-Cooled Exchangers................................................................................................... 66 4.33 Double Pipe / Multitube Hairpin Exchangers......................................................... 67 4.34 Plate Type Heat Exchangers......................................................................................... 68 4.35 Plate and Frame Exchangers........................................................................................ 69 4.36 Welded Plate Heat Exchangers................................................................................... 69 4.37 PACKINOX Plate Exchanger.......................................................................................... 70 4.38 Spiral Heat Exchangers................................................................................................... 70 4.39 Fouling Primer................................................................................................................... 71 Page 6 of 72 Section 01 1. Design Temperature & Pressure 1.1 Introduction to Design Safety: Designing safety into a process is important to everyone including the process design engineer, the people working in the plant, the community and the environment in the surrounding areas. The risk of potential safety problems is measured by following formula: (The probability of the occurrence) x (The consequences of the occurrence). This module describes actions that can be taken at the design level to reduce the risk of potential safety problems. We can reduce this risk by reducing the: Probability At times we can reduce the probability to such a low level that it is no longer considered a plausible event. Consequences: Sometimes we can reduce the potential consequences of the event until it is an event we can tolerate. 1.2 General Principles of Safety in Process Design: The following general principles of safety apply to process design: The design of safety into a process is the responsibility of the process design engineer. Every design must be safe against the reasonable causes of failure. The design must incorporate facilities that are adequate to prevent fires, explosions, accidents and minimize releases. Page 7 of 72  All process designs (grass-roots as well as revisions) must be reviewed with the Safe Operations Committee to ensure safety standards are being met. 1.3 Minimizing the Risk of Fire, Explosion, or Accident: Designing to minimize the risk of fire, explosion or accident consists of: Preventing uncontrolled releases of flammable or toxic materials, such as those that may result from equipment failure, improper materials or incorrect procedures. Minimizing the number of ignition sources. Preventing hydrocarbon/air mixtures in the flammable range from forming within process equipment. Ensuring the plant is safe and operable for personnel. Preventing runaway exothermic chemical reactions. 1.4 Equipment Spacing: Equipment spacing is important in the design of safe plants: Permits access for firefighting. Permits operators to perform emergency shutdown actions in a fire situation. Minimizes involvement of adjacent facilities in a fire, thus preventing further equipment failures. Ensures that critical emergency facilities are not subject to fire damage. Separates continuous ignition sources from probable sources of release of flammable materials. Segregates high risk facilities (such as LPG storage). Avoids danger or nuisance to persons or facilities beyond adjacent property lines. Permits access for normal operation and maintenance. Ensures site security. Page 8 of 72 1.5 Designing for Equipment Spacing: DP XV-G, Figure 1B provides recommended minimum distances between equipment items in onsite. Figure 2B is similar, but gives distances between equipment items in offsite. Page 9 of 72 1.6 Layout Considerations: In addition to requirements for each type of equipment, there are general spacing considerations for: Maximum height. Avoid stacking more than three levels of equipment to minimize potential fire involvement. Stacking. Do not locate equipment over pumps handling flammable materials, air fin exchangers, or heat exchangers or drums containing flammable materials above 600°F (315ºC) Above mentioned example layout shows some of the equipment spacing layout design considerations e.g.: Locate cooling tower on downwind side of unit. Locate air-cooled exchangers on other side of towers away from furnaces. These may be located over pipe rack if DP XV-G requirements are met. Locate pumps on opposite side of towers from furnaces. Locate pumps on same side of pipe rack as suction vessels. Locate fired heaters on upwind side of unit near battery limit. Provide access for tube removal Page 10 of 72 1.7 Preventing Equipment Failures: Equipment failure can result in fire, explosion or accident. Most stationary equipment failures are due to: Overstressing of equipment (over pressure, under pressure, over temperature and under temperature) The environment (corrosion, erosion and cycling) It is important to be sure that: The design temperatures and pressures cover the full range of operating conditions which might be experienced, including upset conditions and unusual operations. The materials of construction must be able to handle the environment to which the equipment will be exposed 1.8 Maximum and Minimum Design Temperatures: The maximum and minimum design temperatures are the metal temperatures that: Equal the most severe combination of temperatures and pressures. Are the same as the fluid temperatures for equipment without internal insulation. Normally, the design of an equipment item has: A maximum design temperature, which often sets the materials of construction and is used to determine allowable stresses. A minimum design temperature, which can set materials of construction and impact testing requirements. It is usually specified as Critical Exposure Temperature (CET). 1.9 Design Temperatures Selection: Determine range of normal operating temperatures and coincident pressures. Include normal operations, startup, shutdown, regeneration, steam out (300 ºF) and other PLANNED scenarios. Identify most severe combination of coincident pressure Page 11 of 72 and temperature and add suitable safety margin to account for deviations from normal. Determine range of temperatures and coincident pressures arising from abnormal operations, e.g. utility failures, operating failures and other UNPLANNED scenarios (excluding fire). Identify most severe combination of coincident temperature and pressure. No need to add safety margin. The design temperatures are set by the most severe conditions arising from normal or abnormal operations. Additional Guidelines: Guidelines for design temperatures up to 650 °F (345 ºC) Allowable stresses for ASME vessels are constant. The usual practice is to add a 50 °F (28 ºC) margin to the design temperature set by planned operations. However, this margin can be reduced if the temperature can be predicted with confidence, or increased if significant uncertainty exists. Guidelines for design temperatures 650 °F to 800 °F (345 ºC to 425 ºC) Allowable stresses for carbon steel (CS) decrease with temperature. Add a minimum margin to cover uncertainty in temperature prediction for the design temperature set by planned operations. Guidelines for design temperatures above 800 °F (425 ºC) These temperatures usually require alloy construction or internal insulation (refractory lining) for CS vessels. Keep in mind: For alloy vessels, add a minimum margin to correct for uncertainties in planned operating temperature. For internally insulated vessels, design metal temperature is usually set at 650 °F (345 ºC) for fluid temperatures above 650 °F (345 ºC). Page 12 of 72 1.10 Critical Exposure Temperature The lowest temperature at which the equipment may be pressurized up to its full design pressure without risk of brittle fracture is known as critical exposure temperature (CET) Maximum safe pressure below the CET determined by mechanical/ materials engineering specialists The CET is derived from either operating, upset or atmospheric conditions. The CET may initially be determined as the lowest metal temperature at which a component will be subject to either a pressure greater than 35% of design pressure or the lowest one-day mean temperature. Used to determine minimum impact testing requirements to avoid brittle fracture The cost impact of CET is usually not significant (for new construction) above - 55 °F (-48 ºC). o Above 120°F (50 ºC), impact testing not required. o Above -20°F (-30 ºC) impact testing normally not required for most steels o Below -20°F (-30 ºC), normalized (killed) CS is used and impact testing is required as per IP 18-6-1 o Below -55°F (-48 ºC), use of alloy may be required 1.11 Design Pressure: Design pressures are the maximum pressure expected in the top of the vessel and are used to determine minimum wall thickness. The designer must also specify the maximum dynamic pressure drop and the maximum liquid static head, if these values are significant. Minimum design pressure is 15 psig (1 kg/cm2) Minimum recommended design pressure for vessel with pressure relief valve discharging to a closed system (flare) is 50 psig (3.5 kg/cm2), and may consider 100 psig (7 kg/cm2) depending on flare backpressure. Margins are added to the maximum expected pressures in most equipment to account for uncertainties in estimating actual pressures. Page 13 of 72  Maximum Operating Pressure, PMAX ≤ 250 PSIG (17.6 kg/cm2) o Direct Acting PRV: PDES = PMAX + 25 PSIG (1.8 kg/cm2) o Pilot Operating PRV: PDES = PMAX + 5 PSIG (0.6 kg/cm2) Maximum operating pressure, PMAX > 250 PSIG (17.6 kg/cm2) o Direct acting PRV: PDES = PMAX / 0.90 o Pilot operated PRV: PDES = PMAX / 0.95 (In normal practice, larger margins are used since at these values, safety valves will begin to simmer) 1.1.1.Vacuum: If the vessel can be under vacuum, the minimum pressure must also be specified. The important questing arise in this scenario is: – How to make sure if vacuum possible? Specify vacuum or partial vacuum if fluid in vessel will have less than 14.7 psia (1.0 kg/cm2) vapor pressure if cooled to ambient temperature. Assume vessel loses flow and heat input. As vessel cools, overhead pressure control valve closes, and vacuum can be created. 1.12 Piping Flange Design: Process designs do not necessarily specify design temperatures and pressures for piping. It must be done by detailed engineering, but process designs do specify ratings for the flanges on piping and equipment. These ratings depend upon: o The design temperature and pressure of the line o The type of flanges and materials of construction Temperature and pressure conditions for these ratings are presented in DP II, Tables 14.4.1 –14.4.11. One can also access flange rating tables from ASME B16.5. Page 14 of 72 1.13 Piping Flange Design Temperature and Pressure Selection: To set flange design temperature, first determine the long term operating temperature plus 50 °F (28 ºC). Determine design conditions from DP II, Tables 14.4.1 to 14.4.8, depending on conditions and materials. Uninsulated flanges can use 90% of the temperature you determine. Design pressure is generally set at the design pressure of connected equipment. Some flanges may see excursions of temperature and pressure above the design value for a small period. This is permissible if the excursions are less than: 33% for short term events (less than 10 hours per event and less than 100 hours per year). 20% for intermediate events (less than 50 hours per event and less than 500 hours per year). If the temperature, pressure or timing do not meet these requirements, raise the flange rating until the above requirements are met. Page 15 of 72 Section 02 2. Fluid Flow 2.1 Types of Fluid Flow: There are three types of fluid flow encountered in typical hydrocarbon processing plants. Each type of flow has different characteristics, handling requirements, and calculation methods. The flow regime can have an important influence on process design of equipment. Single Phase (all liquid or all vapor) : Single phase flow is the most common and easiest to handle. Two phase (liquid and vapor) : Two phase flow is more difficult to handle because the type of flow depends on velocity, ratio of liquid to gas and physical properties. With two phase flow it is often necessary to stay within certain flow regimes. This type of flow is discussed in detail in this module. Solids mixed with liquid and/or vapor : This type of flow is rarely encountered and will not be covered in this module 2.2 Flow Regimes: In two-phase flow, interactions between liquid and vapor phases cause the fluids to flow in various types of patterns, called flow regimes. Only one type of flow exists at a given point in a line at any given time. However, as flow conditions change, the flow regime may change from one type to another. Flow regimes are influenced by: the physical properties of the liquid and vapor flow rates pipe size, roughness and orientation Page 16 of 72 ExxonMobil Engineering Design Practices, Section XIV-D, Figure 1 and Figure 2A provide graphs, or flow regime maps, for determining the flow regime obtained. 2.3 Two Phase Flow Regimes in Horizontal Pipes 2.4 Flow Regimes in Horizontal or Slightly Inclined Pipes Seven flow regimes have been defined to describe flow found in horizontal or slightly inclined pipes. They are shown here in order of increasing vapor velocity. Bubble Flow : Liquid occupies the bulk of the cross-section and vapor flows in the form of bubbles along the top of the pipe. Vapor and liquid velocities are approximately equal. If the bubbles become dispersed through the liquid, this is sometimes called froth flow. In uphill flow bubbles retain their identity over a wider Page 17 of 72 range of conditions. In downhill flow the behavior is displaced in the direction of plug flow Plug Flow : As the vapor rate increases, the bubbles coalesce, and alternating plugs of vapor and liquid flow along the top of the pipe with liquid remaining the continuous phase along the bottom. In an uphill orientation the behavior is displaced in the direction of bubble flow; downhill stratified flow is favored. Stratified Flow As the vapor rate continues to increase, the plugs become a continuous phase. Vapor flows along the top of the pipe and liquid flows along the bottom. The interface between phases is relatively smooth and the fraction occupied by each phase remains constant. In uphill flow, stratified flow rarely occurs, with wavy flow being favored. Downhill, stratified flow is somewhat enhanced, as long as the inclination is not too steep. Wavy Flow As the vapor rate increases still further, the vapor moves appreciably faster than the liquid, and the resulting friction at the interface forms liquid waves. The wave amplitude increases with increasing vapor rate. Wavy flow can occur uphill, but over a narrower range of conditions than in horizontal pipe. Downhill, the waves are milder for a given vapor rate and the transition to slug flow, if it occurs at all, takes place at higher vapor rates than in horizontal pipe. Slug Flow Page 18 of 72 When the vapor rate reaches a certain critical value, the crests of the liquid waves touch the top of the pipe and form frothy slugs. The velocity of these slugs, and that of the alternating vapor slugs, is greater than the average liquid velocity. In the body of a vapor slug the liquid level is depressed so that vapor occupies a large part of the flow area at that point. Since slug flow may lead to pulsation and vibration in bends, valves and other flow restrictions, it should be avoided where possible. Uphill, slug flow is initiated at lower vapor rates than in horizontal pipe. Downhill, it takes higher vapor rates to establish slug flow than in horizontal pipe, and the behavior is displaced in the direction of annular flow. Annular Flow The liquid flows as an annular film of varying thickness along the wall, while the vapor flows as a high-speed core down the middle. There is a great deal of slip between phases. Part of the liquid is sheared off from the film by the vapor and is carried along in the core as entrained droplets. The annular film on the wall is thicker at the bottom of the pipe than at the top, the difference decreasing with distance from slug flow conditions. Downstream of bends, most of the liquid will be at the outer wall. In annular flow, the effects of friction pressure drop and momentum outweigh the effect of gravity, so that pipe orientation and direction of flow have less influence than in the previous flow regimes. Annular flow is a very stable flow regime. For this reason and because vapor-liquid mass transfer is favored, this flow regime is advantageous for some chemical reactions. Spray Flow This type of flow is also known as mist flow or dispersed flow. When the vapor velocity in annular flow becomes high enough, all the liquid film is torn away from the wall and is carried by the vapor as entrained droplets. This flow regime is almost completely independent of pipe orientation or direction of flow. Page 19 of 72 2.5 Two Phase Flow Regimes in Vertical Pipes 2.6 Flow Regimes in Vertical Pipes: Flow behavior in vertical pipes, where gravity plays an important role, has been less extensively investigated than has flow in horizontal pipes. Most of the available information on vertical flow pertains to up flow. Conditions under which certain flow regimes exist depend largely on the orientation of the pipe and the direction of flow. Five flow regimes have been defined to describe vertical flow. They are listed here in order of increasing vapor velocity. Note: Direction of flow is shown upwards. Bubble Flow: Upward flowing liquid is the continuous phase, with dispersed bubbles of vapor rising through it. The velocity of the bubbles exceeds that of the liquid, because of buoyancy. As vapor flow rate is increased, the sizes, number and velocity of the bubbles exceeds that of the liquid, because of buoyancy. As vapor flow rate is increased, the sizes, number and velocity of the bubbles increase. The bubbles Page 20 of 72 retain their identity, without coalescing into slugs, at higher vapor rates than in horizontal flow. Slug Flow: As the vapor rate increases, bubbles coalesce into slugs which occupy the bulk of the cross-sectional area. Alternating slugs of vapor and liquid move up the pipe with some bubbles of vapor entrained in the liquid slugs. Surrounding each vapor slug is a laminar film of liquid which flows toward the bottom of the slug. As the vapor rate is increased, the lengths and velocity of the vapor slugs increase. Slug flow can occur in the downward direction, but is usually not initiated in that position. However, if slug flow is well established in an upward leg of a coil, it will persist in a following downward leg, if other conditions remain the same. In designing for two-phase flow, it is normal practice to try to avoid slug flow, since this regime can lead to serious pressure fluctuations and vibration, especially at vessel inlets and in bends, valves and other flow restrictions. This could lead to serious equipment deterioration or operation problems. When slug flow cannot be avoided (for instance, in thermosyphon reboilers), one should avoid flow restrictions and use long-radius bends to make turns as smooth as possible. Froth Flow: As the vapor rate increases further, the laminar liquid film is destroyed by vapor turbulence and the vapor slugs become more irregular. Mixing of vapor bubbles with the liquid increases and a turbulent, disordered pattern is formed with ever shortening liquid slugs separating successive vapor slugs. The transition to annular flow is the point at which liquid separation between vapor slugs disappears and the vapor slugs coalesce into a continuous, central core of vapor. Since froth flow has much in common with slug flow, the two regimes are often lumped together and called slug flow. In the downward direction, froth flow behaves much the same as slug flow does, except that the former is more easily initiated in this position, particularly if conditions are bordering on those for annular flow. Page 21 of 72  Annular Flow: This flow regime is similar to annular flow in horizontal pipe, except that the slip between phases is affected by gravity. In up flow, the annular liquid film is slowed down by gravity, which increases the difference in velocities between vapor and liquid. In downflow, the reverse is true, with gravity speeding up the liquid and reducing the difference in velocities between vapor and liquid. On the other hand, the liquid film thickness is more uniform around the circumference of the pipe than in horizontal flow. Mist Flow ; This flow regime is essentially the same as spray flow in horizontal pipe. The very high vapor rates required to completely disperse the liquid essentially eliminate the effects of orientation and direction of flow. In identification of vertical two-phase flow regimes, annular and mist flow are often considered together and called annular-mist. 2.7 Effects of Fittings on Two-Phase Flow: Fittings: Fittings may strongly affect the flow of vapor-liquid mixtures. Bends: Bends tend to cause phase segregation. Because of differences in momentum, liquid tends to move to the outer wall while vapor tends to stay close to the inner wall. The segregation can be minimized by the use of blanked-off tees instead of elbows. The fluid should enter through one straight-through end and exit through the branch. Flow Restrictions: Orifices, valves, and other flow restrictions tend to disperse the two phases into each other Page 22 of 72 2.8 Two-phase warnings: Avoid slug flow Pulsation and vibration from bends, valves, etc. leads to damage Splitting flow is difficult due to phase segregation Liquid will keep going straight Lack of pressure recovery Vertical static head is lost in downward flow pipes No pressure recovery factor for orifices, nozzles and venturis Higher pressure drop than single phase for same stream PEGASYS note: Use Homogeneous when in “spray or bubble regimes” Use Duklerswhen two phases would be at different velocities 2.9 Pressure Drops Calculations: The pressure difference between two points in a flowing system can be determined by Bernoulli’s Theorem: Simplified it says: Pressure Difference = ∆Elevation Head + ∆Velocity Head + Frictional Losses For incompressible flow applications, velocity is constant (in a fixed pipe diameter) and there is no change in Velocity Head. For now, we will concentrate on incompressible flow and come back to compressible flow later. 2.10 Friction Losses: Frictional pressure drop can be read from charts available in the Design Practices or can be calculated with the following equation: Where: Page 23 of 72 (∆P)f= Frictional Pressure Drop, psi (kPa) (kg/cm2) K=13.4 (customary), 3.24x1012 (metric), 3.3x1010(kg/cm2) f=Friction factor, dimensionless L=Pipe Length, ft (m) W=Mass Flow Rate, klb/hr (kg/s) ρ=Density, lb. /ft3 (kg/m3) d=Inside diameter of pipe or equivalent hydraulic diameter, in (mm) The experimental factor in the pressure drop equation is the Friction Factor, f, which is a function of the Reynolds number and relative pipe wall roughness. For practical purposes; note the relation between the frictional pressure drop, the line length, the mass flow and the pipe diameter: Page 24 of 72  ∆Pf α L ∆Pf α W2 ∆Pf α 1/d5 2.11 For Compressible Flow Applications: The pressure drop equation is still valid but we must use an accurate density which changes with pressure. Good Rule of Thumb: Equation is accurate if ∆Pf< 10 % of upstream pressure and average ρis used. If this condition cannot be satisfied, break line length into smaller segments and calculate ∆P for each, starting at a point of known pressure. 2.12 Equivalent Length of Fittings For piping systems, we must include the resistances of pipe fittings (i.e. valves, elbows, tees, etc.). Pipe fittings are accounted for in the pressure drop equation by defining an equivalent pipe length representing the fitting. The total line length including fittings becomes the “Equivalent Length”(LEQ). Where: LEQ= Equivalent line length including fittings, ft. L=Line Length excluding fittings, ft. D=Inside diameter of pipe or equivalent hydraulic diameter, ft. f=Fanning friction factor, dimensionless. ΣK=Sum of resistance coefficients of all fittings, dimensionless Note: The Crane friction factor f = 4 x Fanning friction factor Page 25 of 72 2.13 Quick Reference to Pressure Loss through Fittings: 1. All values are based on ff= 0.005. 2. Refer to Crane, pages A-26 thru A-31, for additional resistance coefficient (K) values. Page 26 of 72 2.14 Design Basis for “Average” Carbon Steel Lines: 2.15 Static Head Losses: Pressure balances in refinery and petrochemical plant applications must account for elevation head, i.e., Static Head losses. Static Head of a fluid is the pressure resulting from the weight of a column of the fluid. As the elevation of a fluid relative to some datum point(typically grade) changes, the fluid’s Static Head changes. In engineering calculations, Static Head difference is calculated as follows: ∆PS.H.= ρ(Z2-Z1)/144 Psi ∆PS.H.= s.g. (Z2-Z1)/2.31, liquid only Psi ∆PS.H.= 9.81 x 10-3ρ(Z2-Z1) kPa**important** ∆PS.H.= 9.90 x 10-5ρ(Z2-Z1) kg/cm2**important** Where: Page 27 of 72 Z1Z2, = Fluid elevations above datum point, ft (m). ρ= Density, lb/ft3(kg/m3) s.g. =specific gravity 2.16 Preliminary Hydraulic Calculations: There are four steps to follow when designing a small section or loop of a plant. 1. Consider all of the equipment contained in the section. 2. Attribute an approximate pressure drop to each piece of equipment and all of the piping and instrumentation. DP Section XIV-B, Table 1B and Table 2 propose approximate lengths and pressure drops which can be used for preliminary calculations. These tables are especially useful in early work when plant layout and pipe routings have not been established. For very preliminary calculations when system details are not available, use multipliers (below) to convert line lengths to equivalent lengths. These multipliers are an estimated correction for the effect of piping fittings: Offsite 1.2 to 1.8 **important** Onsite 3.0 to 6.0**important** 3. Calculate the effect of changes in static head. Page 28 of 72 4. Estimate the pressure drops in the circuit. 2.17 Typical Process Lines Equivalent Lengths: Use table 2 attached below when do not have piping isometrics Notice that “D”I s in inches, but “equivalent length” is in feet (There is no metric table) Page 29 of 72 2.18 Final Hydraulic Calculations: For revamps & final designs, it is often necessary to calculate the pressure drop based on the actual piping layout as opposed to making an allowance based on typical equivalent lengths. Two approached can be used: By hand using the formulas in Design Practices Section XIV C. Pegasys software provides a more convenient method of making these calculations. Page 30 of 72 Attachment-1 Case Study Page 31 of 72 2.19 Case Study: Using Restriction Orifices to Absorb Excess Pressure Drop Restriction Orifices (R.O.) provide a convenient way to absorb excess pressure drop in a system or limit the flow rate of some medium such as purge air or steam. Two R.O. examples are given below. This application controls the flow of steam to the low-pressure drum. This is much less expensive than a flow rate controller on the steam. This application limits the flow rate of the high-pressure gas to the second drum. This results in a much smaller safety valve on the second drum and associated equipment. This application limits flow into lower pressure system potentially reducing mechanical design requirements and downstream safety issues. 2.20 Choking or Critical Flow When a gas accelerates through a restriction, its density decreases and its velocity increases. Since the mass flow per unit area (mass velocity) is a function of both the density and velocity, a critical cross-sectional area exists at which the mass velocity is the maximum. In this area the velocity can be sonic, and further decreasing the downstream pressure may not increase the mass flow. This is referred to as choked or critical flow. Critical flow can cause vibration problems downstream. To calculate sonic velocity, use the simplified equation below or Equation 9 in ExxonMobil Engineering Design Practices Section XIV-C (see next page) Where: Vc=Critical (sonic) velocity, ft/s k=C p/Cv z=Compressibility Factor T=Temperature, ºR (ºF+460) MW=Molecular Weight Page 32 of 72 Rule of Thumb: Sonic velocity when pressure drops by more than 50% across restriction Page 33 of 72 Section 03 3. Pumps 3.1 Types of Pumps: The types of pumps most often encountered in a refinery or chemical plant can be classified as either kinetic or positive displacement pumps Kinetic Pumps In a kinetic pump energy is added continuously to increase the fluid's velocity within the pump to values in excess of those in the discharge pipe. Passageways within the pump then reduce the velocity until it matches that in the discharge pipe. Bernoulli's law states that as the velocity head of the fluid is reduced, the pressure head must increase. Therefore, in a kinetic pump the kinetic energy or velocity energy of the fluid is increased and then converted to potential or pressure energy. Most of the commonly used pumps chemical plants are centrifugal pumps, in which the kinetic energy is imparted to the fluid by a rotating impeller, which generates centrifugal force. Positive Displacement Pumps In a positive displacement pump the volume containing the liquid is decreased until the resulting liquid pressure is equal to the pressure in the discharge system. That is, the liquid is compressed mechanically, causing a direct increase in potential energy. Most positive displacement pumps are reciprocating pumps, where the displacement is accomplished by linear motion of a piston in a cylinder. Rotary pumps are another common type of positive displacement pump. In these pumps displacement is caused by circular motion. Page 34 of 72 3.2 Centrifugal Pumps: This type of pump is used almost exclusively in refineries and chemical plants. Centrifugal pumps have few moving parts and therefore tend to have greater reliability and lower maintenance costs than positive displacement pumps. A centrifugal pump generally has a spare unless the plant can run without it. Design Considerations: Pressure rise is dependent on density, because velocity is converted to static head in feet of fluid. Head developed is determined by speed (RPM), impeller diameter, and number of impellers (Stages). Multistage pumps can develop head up to 6000 ft (1800m), about 1300 - 2500 psi (90-175 kg/cm2) for hydrocarbons. Minimum capacity without recirculation is usually limited to 15 - 20 gpm (4 m3/h) Centrifugal pumps lose efficiency as viscosity increases. 3.3 Special Types of Centrifugal Pumps Integral Gear Driven “Sundyne”: Speed increasing gear box integral with pump. Single stage design up to 2100 psig (148 kg/cm2g) discharge pressure, 1000 psig suction pressure (70 kg/cm2g), and 650°F (343ºC). Multistage design up to 3750 psig (265 kg/cm2g) discharge pressure, 1000 psig suction pressure (70 kg/cm2g), and 500°F (260ºC). Page 35 of 72  High heads achieved by operating impellers at speeds to 25000 rpm. Such operating speeds result in high NPSHR. Low cost pump for high pressure applications. Caution: Pump must be operated very stable to avoid damage. Pressure surges can damage pump. Consider providing warehouse spare in addition to installed spare. Vertical “In-Line”: This type of pump bolts directly into piping like a valve. Small footprint. Single sealing chamber. Choice of horizontal or vertical pump usually based on economic and space considerations. Design up to 500 psig (35 kg/cm2g) discharge pressure, 300 psig (21 kg /cm2g) suction pressures, and 600°F (315ºC). May not be suitable for large motors. Canned: Casing buried, with first impeller below grade. Used for tight NPSHA services and in high head services where the large number of stages required does not allow horizontal orientation of the pump. Generally avoid since installation requires excavation. Design up to 1300 psig (91 kg/cm2g) discharge pressure, 300 psig (21 kg /cm2g) suction pressures, and 500°F (260ºC). Page 36 of 72 2.1. Reciprocating Pumps: Reciprocating pumps are often used in to handle viscous fluids and in sludge and slurry service. Compared to centrifugal pumps, reciprocating pumps: Are more efficient Are poorer at handling liquids containing solids that erode valves and seals Have higher maintenance costs and lower availability because of pulsating flows and the large number of moving parts Design Considerations: Its performance curve is constant capacity, variable head. It produces a pulsating flow. Pulsations are minimized by use of multiple plungers (duplex, triplex, etc.). Pressure rise is independent of fluid density, so these pumps are used where a wide range of fluids must be handled. Reciprocating pumps are used for low capacities and high heads. They need safety valve protection. 2.2. Centrifugal Pumps Design: When designing a system utilizing a centrifugal pump, keep in mind that centrifugal pump performance changes with operating conditions, as illustrated by the pump curves on the next screen. This is a typical centrifugal pump curve at a constant speed (rotational velocity). As the flow rate through the pump increases, the generated head decreases. The efficiency curve reaches a maximum, hopefully near the design point. The brake horsepower curve increases continuously with flow rate. Page 37 of 72 Centrifugal Pumps Impeller Change: If the pump speed is changed (i.e. variable speed drive or larger impeller), the head and flow rate will also change. Page 38 of 72 Centrifugal Pumps Design: This curve shows how a control valve affects flow through a pumped system. The control valve pressure drop decreases the flow rate through the system. When the control valve opens fully, the flow rate increases from the "operating flow rate" to the "maximum flow rate" the point where the system pressure drop equals the pump discharge pressure. 3.4 Pump Affinity Laws: Relationships called "Pump affinity laws” show the effect of changing speed or impeller diameter on pump performance. Pump affinity laws are useful when looking at a new service for an existing pump. FLOW Q ∼ (Peripheral Speed) {Q2=Q1 (N2/N1) or Q2=Q1 (D2/D1)} HEAD H ~ (Peripheral Speed) 2 Page 39 of 72  {H2=H1 (N2/N1) 2 or H2=H1 (D2/D1) 2} POWER HP ~ (Peripheral Speed) 3 {HP2=HP1 (N2/N1)3 or HP2=HP1 (D2/D1) 3 Where D=impeller diameter and N=impeller RPM 3.5 Centrifugal Pump Design Factors: There are five major factors that affect the design of centrifugal pumps: – Fluid properties – Pump differential pressure – Net positive suction head – Pump casing design pressure – Driver and utility requirements These factors are described in the following pages. 3.6 Fluid Properties & Differential Pressure: The fluid properties most important to pump design are fluid density and viscosity: For a fixed speed pump and a fixed capacity, the pump will develop the same head regardless of the density. However, pressure rise is directly related to the fluid density. The performance of centrifugal pumps deteriorates with increased viscosity. Differential pressure is the difference between the discharge and suction pressures and is a measure of the amount of energy that the pump must supply. 3.7 Cavitation & NPSH: Cavitation refers to the formation and subsequent implosion of vapor bubbles. Cavitation occurs when the static pressure of the liquid falls to or below the vapor pressure in a moving liquid system. The vapor bubbles formed in the cavitation are Page 40 of 72 subsequently imploded by increasing static pressure. Cavitation commonly occurs in and around the impeller of a centrifugal pump. For satisfactory pump operation the liquid must flow without vaporizing from the pump inlet to a point within the impeller eye where the vanes begin to impart energy to the liquid. The force tending to suppress cavitation is the margin by which the local static pressure of the liquid exceeds the liquid vapor pressure at that temperature. When converted to terms of head o liquid, this pressure margin is termed the net positive suction head, or NPSH. To avoid cavitation, the available net positive suction head. (NPSH) must be greater than that required by the pump. 3.8 Calculating and Approximating NPSH NPSH is calculated by the following equation: Where: hp = Pressure in suction vessel vapor phase. hvpa = Vapor pressure of the liquid. hst = Height of liquid above pump suction. his = All suction line pressure losses. NPSH is a characteristic individual to each pump and is determined by vendor testing. The approximate NPSH required for a given flow and head is presented in Figure 4B of DP Section X-A. Pump specification must include the available NPSH. Apply a 10% safety factor on the calculated NPSH. If the available NPSH is less than the required NPSH, consider increasing the elevation of the suction drum. If the available NPSH is greater than 25 ft (7.6m), specify NPSH = 25 ft (7.6m). 3.9 NPSH Cheat Sheet: Assume pump is 2’ (0.6m) above grade, unless large pump, which will be more. 1 psi (0.07 kg/cm2) (min) for permanent suction strainer. Don’t use permanent strainer if solids are not expected. 10% safety margin normally, 25% for sulfolone, 5 psi (0.35 kg/cm2) for BFW (boiler feed water) Page 41 of 72  Use 32D + 200’ suction line equivalent length if no isometrics available. Fluid usually at bubble point. Can be lower for Stripper bottoms. Account for acceleration in reciprocating pumps. See specialist if cost to meet NPSH too high. Do not specify greater than 25’ (7.6m). Design suction line straight into pump (~5Ø) (during detailed design) Never allow suction line elevation to go above (or even close) to vessel tangent line (no vapor pockets) (during model review) Use vessel tangent line if possible instead of LLL. Late Breaking News: DP X-D recommends extra 3’ (1m) NPSH margin when selecting pump 3.10 Driver and Utility Requirements: Normally, electric motor drives are used except for pumps that must operate during a power failure. The following design information is helpful when developing specifications for electric motor pumps. Pump brake kW is defined by: Minimum driver size: To determine operating load (kWH): Page 42 of 72 To determine pump connected load (kWH): 3.11 Motor Sizing( DP XXX-G) Page 43 of 72 3.12 Design Pressure for Centrifugal Pumps: Casing design pressure for centrifugal pumps is the maximum suction pressure plus the maximum differential pressure. The maximum differential for new pumps is limited by GP 10-1-1 to 1.2 times the head at rated capacity. In addition, centrifugal pumps should be capable of at least a 5% head increase by replacing the impeller (API 610). The result is a recommended maximum differential of 1.26 (1.2 x 1.05 = 1.26) times the rated differential pressure of the pump. The maximum suction pressure is usually set by having the PR valve blowing and having a high liquid level in the suction vessel. Variable speed pumps, which can operate continuously at 105% of rated speed should be designed for at least 1.39 (1.2 x 1.05 x (1.05)2=1.39) times the rated differential pressure of the pump. If the pump can operate with more than one fluid, consider the fluids with the highest and lowest specific gravity, since these will result in the highest and lowest differential pressures. 3.13 Design Pressure for Other Equipment: Equipment between the pump discharge and a control valve should have amaximum operating pressure equal to the maximum suction pressure plus maximum pump shutoff differential pressure. (“max-max”) Equipment between the pump discharge and a block valve should also have a maximum operating pressure of maximum suction pressure plus maximum pump shutoff differential pressure if closing the valve has the direct result of raising the suction pressure to maximum. If however, closing the block valve does not raise suction pressure to maximum, the maximum operating pressure should be the higher of – Normal pump suction pressure plus maximum pump shutoff differential pressure (“normal-max”) – Maximum pump suction pressure plus normal pump differential pressure (“max-normal”) Page 44 of 72  Design pump suction piping design pressure for parallel pumps at least 75% of discharge design pressure. (0.75=1/1.33, so is good for short term overpressure) 3.14 Summary of DP Downstream of Centrifugal Pump: Pump: “Max-Max” (maximum suction P, maximum pump ΔP) All equipment up to control valve: “Max-Max” Downstream of control valve: o If closing block valve causes maximum suction pressure: “Max-Max” o If not: “Normal – Max” (Normal suction P, maximum pump ΔP) “Max – Normal” (Maximum suction P, normal pump ΔP). When calculating suction pressure during a block in case, assume 30 minutes of liquid accumulation. Late Breaking News: Use electronic governor to limit turbine speed variation in revamps: – Calculate pressure at maximum speed set point – Use 1.005 instead of 1.05 for turbine speed variation (1.2 x 1.05 x (1.005)2=1.273) – Also use actual pump curve, so 1.2 factor (head rise at shutoff) may change and 1.05 (factor for maximum impeller) may be eliminated. Page 45 of 72 Attachment-2 Case Study Page 46 of 72 3.15 Case study: Design Pressure & Temperatures – Downstream of Pumps: Pump Discharge Press = 14 + 1.26(400) = 518 psig *or 69 + 400 = 469 psig * Pump Casing Design Press = 69 + 1.26(400) = 573 psig Pump Suction Design Press = ________ or ___ psig if parallel pumps, __% of discharge, why? Design Pressure of Exch. A = 518 psig Design Pressure of Exch. B = ___ psig Design Temperature of Exch. A = 232 + 50 = 282oF Design Temperature of Exch. B = ___ + __ = ___oF (Exch. A bypassed) why not 232? * This assumes that blocking the outlet of the pump does not result in the suction SV blowing 3.16 Mechanical Seals: A seal is required to prevent process fluid from leaking past the pump shaft. Mechanical seals have many advantages over packed stuffing boxes. Typical stuffing box and mechanical seal configurations are shown below. A single mechanical seal is sufficient for many pump shaft sealing services. Sometimes two mechanical seals are required, with some sealant injected between the seals. When two seals are constructed in a single assembly with some common parts, the combination is termed a double mechanical seal. When the two seals are separated and oriented in the same direction, they are termed tandem mechanical seals. Page 47 of 72 3.17 Mechanical Seals Details: Items to consider – Can the pumped fluid be used as flushing fluid? Boiling point, solids, pressure, pour point, and lubricating value all must be considered – If not, are there restrictions on the fluid? Will it dilute the pumped fluid? Will it contaminate the pumped fluid? Page 48 of 72 If seal leaks, what are the consequences? – This determines type of seal required 3.17.1 Flushing Mechanical Seals: Mechanical seals must be flushed to maintain “sealage”. Mechanical seals are usually flushed with some of the process fluid; i.e., they are "self-flushed". When the process fluid is too dirty for self-flushing, however, an external liquid is employed. Recently a new technology which flushes the seals with gas has been used for difficult services. See DP X-G for additional information on flushing systems 3.18 Seal less Pumps: Most pump leaks and fires are caused by seal failures. Seal less pumps are becoming popular because of their environmental and safety advantages, although they are somewhat less efficient and more expensive than other types of pumps. Page 49 of 72 Their use is expected to grow, especially for handling toxic material and for sites from environmentally sensitive areas. Three basic types of seal less pumps are available: Magnetically driven pumps The drive shaft is connected to the pump impeller through magnetic couplings located at the inner wall of a metal container. Canned pumps An electric motor sheathed in metal serves as the impeller shaft, with the stator located outside the metal can. seal less diaphragm pumps These are reciprocating machines with a flex diaphragm to cause pumping action due to check valves on the inlet or outlet. 3.19 Pump Hints: Remember: pumps produce head, not pressure – Lower fluid density will reduce pressure – Size pump differential pressure for minimum density fluid – Size motor for maximum density fluid – Check downstream equipment design pressure based on maximum density fluid Always use ACTUAL fluid rate, not standard Account for extra flows – Constant low flow recycles – Warm-up lines – Seal flush using pump discharge fluid Page 50 of 72 Section 04 4. Heat Transfer & Exchangers 4.1 Background Heat exchangers are used widely throughout the plant for: Recovery of heat generated by a furnace to minimize utility requirements Heat Integration to optimize energy consumption and cost Cooling / Condensing with cooling water, air and refrigerant Heating / Reboiling with steam or hot oil 4.2 Designer’s Role Design New Heat Exchangers Select optimum exchanger type, configuration and size [surface area] Rate Existing Heat Exchangers Evaluate based on new heat transfer requirements Optimize heat integration for given service 4.3 Engineer’s Toolbox XOM DP Section IX Heat Exchanger Simulation Program Process Simulation Program (HYSIS or PRO/II) Heat Transfer Specialists 4.4 Heat Transfer Analogy Heat Transfer is a function of the ΔT driving force across a resistance. Page 51 of 72 4.5 Heat Transfer Equation Q = Uo × A × Δ T Where: Q = Heat Transferred, [Btu/hr or W] A = Heat Transfer Area, [ft2 or m2] ΔT = Average Temperature Difference [°F or ºC] Uo = Overall Heat Transfer Coefficient, [BTU / hr x ft2 x °F] or [W/m2 ºC] Uo = 1/RT RT = Total Resistance to Heat Transfer 4.6 Heat Transfer Coefficient Theory What can we change to improve the film coefficient? They are a function of: Reynolds Number = DVρ/μ Prandtl Number = (Cp) (μ)/k 4.7 More on Total Heat Transfer Resistance There are (5) resistances; frequently, one of the film resistances, Ro or Rio, is limiting Page 52 of 72 RT = Rio + rio + rw + ro + Ro = 1/ Uo 4.8 The Limiting Resistance Resistances values are given as follows: To improve exchanger performance, concentrate on the limiting resistance(s) (i.e. the outside film resistance in this case). Page 53 of 72 4.9 The Controlling Coefficient Frequently one of the two film coefficients determines the value of the overall coefficient: Hence ho is the “Controlling Coefficient”, and efforts to improve exchanger performance should concentrate on this side of the exchanger. 4.10 “Shortcut” Shell & Tube Design Technique For studies or planning purposes only; Given the required heat duty, Q, calculate the required exchanger area, A: Estimate the overall coefficient (UO) by looking in DP IX-B, Table 1 for coefficients in similar services. Estimate the number of shells required to ensure that the Fn is greater than 0.8 using the Fn charts in DP IX-D, Table 2. Calculate the log mean temperature difference (LMTD). 1. A = Q/ (U × Fn × LMTD) For design, use HEXTRAN or the detailed hand method given in the DPs. Page 54 of 72 4.11 Shortcut HX Design 4.12 Temperature Driving Force, ΔT Depends on the exchanger flow pattern; Countercurrent flow provides a more constant ΔT driving force across the exchanger. Page 55 of 72 4.13 LMTD For true cocurrent and countercurrent flow, the red delta shown previously equals the log mean temperature difference, LMTD. Only valid when heat release curves are linear (i.e. no phase change). 4.14 Fn Correction Factor For cross flow (i.e. not true co- or countercurrent) like in a shell and tube exchanger, a correction factor is applied to the LMTD MTDe = Fn × LMTD Fn depends on the number of shells in series (i.e. shell passes) The minimum allowable Fn is 0.8 to avoid a temperature cross (i.e. hot stream outlet temp. is below the cold stream outlet temp.) The more shells in series, the closer Fn approaches 1.0 Refer to DP IX-D for a graph of Fn values Page 56 of 72 4.15 Common Exchanger Types Shell and Tube Exchangers (STEs) are the most widely used exchanger type in industry Other common types include: Air-cooled Exchangers Double Pipe and Multitube Hairpin Exchangers Tubular Exchanger Enhanced Heat Transfer Technologies (EHT) Spiral Heat Exchangers Plate Fin Exchangers Plate and Frame Exchangers 4.16 Shell and Tube Exchangers Can be designed for moderate to high pressure without excessive cost Generally the first type considered when the required surface area is more than 300 sq. ft. (28 m2) Wide range of sizes available, including very large Inexpensive Easy to clean Shell and Tube Exchangers consists of a bundle of parallel tubes enclosed in a cylindrical shell. 4.17 Major STE Types Page 57 of 72 Removable Bundle (Used most frequently by ExxonMobil) Bundle is removable for maintenance or mechanical cleaning of the shell Has only one restrained tube sheet at the channel end Thermal expansion problems avoided by use of floating tube sheet or U-tubes Tube inside can be cleaned in place Fixed Tube Sheet Both tube sheets are fastened to shell Tube bundle is not removable Tubes and shell may require an expansion joint Can be used for very large bundles Cheaper than a removable bundle Counter-current design 4.18 TEMA Type “TEMA Type” followed by three letters refers to the type of Front end (channel) arrangement Shell nozzle/baffle arrangement Rear end (floating head end) arrangement These three characteristics are each identified by a single letter of the alphabet The result, for example, would be the entry “TEMA Type AES” in the specification for the heat exchangers. The type MUST be specified. Page 58 of 72 4.19 TEMA Heat Exchanger Nomenclature 4.20 Shell & Tube Exchangers Fixed Tube Sheet (L, M, or N) Cleanest. Consider only when shell side fouling factor ≤ 0.002 and shell side can be chemically cleaned. Normally used for vertical thermo syphon reboilers or very large exchangers. Because of thermal stresses, an expansion joint may be required if the average shell temperature and average tube temperature differ by more than 50°F. U-Tube (U) Least expensive for high tube side design pressure (eliminates floating tube sheet). Normally used when tube side fouling ≤ 0.002 hr-ft2-ºF/Btu (0.0004 m2-ºC/W), except for water. Page 59 of 72 Split Ring Floating Head (S) This type is normally specified unless very frequent mechanical cleaning is required. Pull-Through Floating Head (T) Most expensive type of S & T unit; thermally inefficient because of shell bypassing. Use when both sides must be mechanically cleaned. Shell and Tube Bundle Types 2. 3. 4. 5. 6. 7. 8. 9. 10. 11. 12. 13. 14. 4.21 Problem – Bundle Rear End Head Types What bundle rear end head type will you pick for? Crude preheat: crude on tube side, kerosene on shell side Polymer tower: frequent cleaning both sides High pressure hydrofiner: fouling on one side Vertical thermosyphon reboiler 4.22 Preliminary Shell & Tube Decisions Which fluid to put in the tubes Tube nominal diameter, wall thickness and material Tube length Page 60 of 72  Tube layout Baffle orientation Baffle pitch (spacing) Maximum bundle diameter (bundle weight) 4.23 Tube Side Fluid Between the two streams, the stream with the higher: Corrosion Rate Pressure Fouling Rate Is usually placed on the TUBE SIDE. When these characteristics apply to both streams, the designer uses his judgement. In a service where one stream is changing phase, that stream is assigned to SHELL SIDE. In steam-heated vaporizers/reboilers, the condensing steam is placed in THE TUBES. Streams with very HIGH VISCOSITY are placed on the SHELL SIDE (better coefficient). 4.24 Tube Length Site decision (local preference) Most common length is 20 feet (6 m) Occasionally, 16’ length is used (4.9 m) For special situations, 8’ and 10’ can be considered Longer tube bundles require more plot area for bundle removal. Longer bundles are also more difficult to extract from the shell and to handle. 4.25 Tube Layout There are 3 Main Layouts; as mentioned below Square: Use when shell side fouling factor ro > 0.002 hr-ft2 -ºF/Btu (0.0004 m2 -ºC/W) and shell side must be mechanically cleaned. Page 61 of 72  Reboilers/Vaporizers Rotated Square: Use as square, but preferred when flow is laminar or for vibration problems Triangular 30°: Use when ro ≤ 0.002 Cheapest, so use when applicable 4.26 Type of Baffle Segmental - Most common Double Segmental (modified disc and donut) is used to obtain very low shell-side pressure drop. Tube Supports Only - No real baffles. Occasionally used in certain reboiling or condensing services. Baffle Orientation and Cut Vertical Chord - Most Common Page 62 of 72  Condensers, vaporizers and fluids containing suspended solids Flow is side-to-side Horizontal Chord Sediment-free fluids being cooled through high temperature range (200 to 300°F or 111 to 167ºC) in one shell. Flow is over-under Baffle Cut This is the percent of the baffle which is cut away to permit flow Typical cut is 25% (40% for double segmental baffles). Baffle Pitch Minimum allowable spacing (pitch) is 20% of the shell ID or two inches whichever is greater. Target same velocity between baffles as in baffle window Maximum allowable pitch: For no change of phase, equals shell ID For change of phase 4.27 Rod Baffle Heat Exchanger (RBE) It eliminates tube vibration in shell-and-tube heat exchangers and allows debottlenecking of pressure drop limited exchangers. Typical applications To correct known vibration problems Compressor inter/after coolers (high velocity gas) Reboilers (high velocity vapor or two-phase) Page 63 of 72 Rod baffles replace conventional baffles on S&T tube bundle. 4.28 Nucleate Boiling Tubes (NBT) Increases shell side heat transfer coefficient for boiling services Typical Applications Horizontal reboilers - shell side boiling Vertical reboilers - tube side boiling Excellent in refrigeration systems (C3 reboilers) 4.29 Helical Baffle Exchanger (HBE) Baffles installed to move the flow spirally through the exchanger Reduces stagnant areas which reduces fouling Reduces pressure drop Increases heat transfer rate / pressure drop ratio Reduces vibration by decreasing unsupported tube length in half Mainly used in high viscosity service Page 64 of 72 4.30 Twisted-Tube Exchanger (TTE) Increase number of tubes No baffles, no vibration Shell side must have very low fouling Not sure of tube side cleaning ability Page 65 of 72 4.31 Handy Factors Pressure Drop Safety Factor Shell side Pressure-Drop Multipliers (DP XI-D Table 4) Tube side Pressure-Drop Multiplier (DP XI-D Table 4) 4.32 Air-Cooled Exchangers Used for cooling high to medium temperature streams where heat recovery is not practical. Consist of one or more fans and one or more finned-tube bundles mounted on a frame. Hot fluid passing through the tubes is cooled by air from fans Page 66 of 72 More on Air Cooled Exchangers Less expensive than water-cooled exchangers if new cooling tower or expansion is required Can be countercurrent or co-current to air flow Tubes have circumferential fins to increase the surface area Outlet temperature limited by ambient air temperature; Can be affected by heat losses from surrounding equipment Can be mounted over pipe racks, etc., to conserve plot space Designer provides duty specification and reviews the vendor’s detailed design 4.33 Double Pipe / Multitube Hairpin Exchangers One or more pipes within a larger pipe True countercurrent (or co-current) flow possible Available in standard off-the-shelf sizes Page 67 of 72  Internal pipes can have longitudinal fins, as long as the corrosion rate is less than 7 mils/year (0.2 mm/year) Can connect several standard units in series or parallel Usually not economical if surface area is beyond 400 sq. ft (37 m2 ) Well suited for high-pressure applications Easy to dismantle and clean by removing cover at return bend 4.34 Plate Type Heat Exchangers Formed by adjacent metal plates that are sealed with gaskets or welded Can be true countercurrent flow; Allow closer temp. Approaches than STEs Very compact; can be cheaper than STEs when alloys required Page 68 of 72 4.35 Plate and Frame Exchangers Sealed by gaskets Usually limited to 300°F (150ºC) and 250 psia (17.6 kg/cm 2 a) Limit on particulates due to plate channel spacing (i.e. 0.05 - 0.5 in. or 1.3 – 13 mm) Require a filter upstream Assembly/disassembly takes very long time Best suited for aqueous streams (i.e. some hydrocarbons attack gaskets) Gaskets must meet special fire standards for flammable fluids 4.36 Welded Plate Heat Exchangers Gaskets on one or both sides of the plate are replaced by a weld Easier to assemble, but cannot be mechanically cleaned May be used at higher temp. and pressure Page 69 of 72 4.37 PACKINOX Plate Exchanger Consists of a plate exchanger inside a pressure vessel May be used in higher pressure services Surface areas up to 110,000 sq. ft. (10,220 m2) possible 4.38 Spiral Heat Exchangers Compact, high surface area Can handle high viscosity fluids or solid particles Typical Applications include product cooling, overhead condensers, tar cooling, slurry exchangers. Page 70 of 72 4.39 Fouling Primer Typical factors in DP IX-B, Table 5 Keep velocity high, especially cooling water (normally 4 ft. /sec or 1.2 m/s min) and crude (6.6 ft. /sec or 2.0 m/s min.) Keep cooling water outlet temperature below 130°F (54ºC), and film temperature below 150°F (66ºF). Consider alloy tubes to reduce corrosion which causes fouling Avoid “N” end, “U” tubes, and triangular layout for high fouling services Page 71 of 72  Highest fouling service on tube side Page 72 of 72