CONTENTS Section Page SCOPE ...3 REFERENCES...3 NUMBER OF SHELLS ...3 TUBE SELECTION...3 TYPE ...3 LENGTH ...4
DIAMETER AND WALL THICKNESS ...4
FERRULES ...5 TEMA TYPES ...5 CHANNELS ...6 TUBE BUNDLES ...6 TYPES...6 TUBE LAYOUT...7 TUBE PITCH ...7 CROSS BAFFLES...8
Position of Baffle Chord ...8
SEAL STRIPS ...8
IMPINGEMENT PROTECTION...9
INTERCHANGEABILITY OF TUBE BUNDLES...9
SHELL TYPES ...9
SINGLE-PASS SHELL (TEMA E)...9
TWO-PASS SHELL (TEMA F)...9
DIVIDED FLOW SHELL (TEMA J) ...10
KETTLE (TEMA K) ...10
CROSSFLOW SHELL (TEMA X) ...10
SPLIT FLOW (TEMA G) AND DOUBLE SPLIT FLOW (TEMA H)...10
TUBE SIDE FLOW ...10
PRESSURE DROP ...11
CORRECTION FOR REDUCED FLOW AREA ...11
ECONOMICS CONSIDERATION OF PRESSURE DROP...11
SHELL SIDE...12
STACKING OF SHELLS ...13
RECOMMENDED GOOD PRACTICE MAINTENANCE FEATURES FOR THE HEAT EXCHANGER DESIGN ...13
CONTENTS (Cont)
Section Page
FLOW INDUCED VIBRATION ... 13
VIBRATION MECHANISMS... 13
TYPES OF DAMAGE ... 14
AREAS OF CONCERN ... 14
DEFINITIONS ... 15
BASIC EQUATIONS ... 15
ANALYSIS OF HTRI VIBRATION PRINTOUT... 16
CORRECTIVE MEASURES... 17
DESIGN GUIDELINES ... 20
TABLES Table 1 Exchanger Tube Data ... 21
Table 2 Thermal Conductivities ... 22
Table 3 Tema Head Type Selection ... 23
Table 4 Maximum Number of Tube Passes... 23
Table 5 Manual Estimation of Tube Count... 24
Table 6 Heat Exchanger Specification Sheet ... 26
Table 7 Heat Exchanger Maintenance Good Practice Designer's Checklist Recommended Mechanical Design Features ... 27
FIGURES Figure 1 Components of Shell and Tube Exchangers... 28
Figure 2 Tema Heat Exchanger Nomenclature... 29
Figure 3 Common Tube Layouts ... 30
Figure 4 Types of Shell Bafffles ... 31
Figure 5A Single Pass Shell (TEMA Type AES)... 32
Figure 5B Divided Flow Shell (TEMA Type AJS)... 32
Figure 5C Two Pass Shell (TEMA Type AFS) ... 33
Figure 6 Location of Bundle Seal Strips... 34
Figure 7 Outermost Tube Limit (OTL) for the Pull through Floating Head Tube Bundle ... 34
Figure 8 Shell Nozzle Correction Factor (C2)... 35
Figure 9 Antivibration Tube Supports... 36
Figure 10 Detuning Baffle ... 37
Revision Memo 12/01 Abbreviation for Tube Pitch changed from TP to PT.
SCOPE
This section presents general design considerations for shell and tube heat exchangers. For a diagram depicting standard components of shell and tube exchangers see Figure 1.
REFERENCES DESIGN PRACTICES (BESIDES OTHER SECTIONS OF SECTION IX) Section I Design Economics
Section XIV Fluid Flow GLOBAL PRACTICE
GP 6-1-1 TEMA Type Shell and Tube Heat Exchangers OTHER LITERATURE
Refinery Construction Materials Manual, EMRE Manual No. EETD 028
User's Manual for Hextran, General Purpose Program for the Design and Rating of Shell and Tube Heat Exchangers Standards of Tubular Exchanger Manufacturer's Association (TEMA), Eighth Edition, 1999
HTRI Computer Programs Support Volume Section F-3, Flow Induced Vibration Analysis
HTRI Reports STV-1, Tube Vibrations in Shell-and-Tube Heat Exchangers, and STV-5, Design of Shell-and-Tube Heat
Exchangers to Avoid Flow-Induced Vibration Problems
NUMBER OF SHELLS
The total number of shells necessary for an exchanger is frequently dictated by how far the outlet temperature of the hot fluid is cooled below the outlet temperature of the cold fluid. This is known as the “extent of the temperature cross." The cross as well as other variables discussed in Section IX-B, determine the value of Fn, the temperature correction factor; this factor should always be equal to or greater than 0.800. (The value of Fn drops slowly between 1.00 and 0.800, but then quickly approaches zero. A value of Fn below 0.800 cannot be predicted accurately from the usual information used in process designs.) In a one-shell exchanger, Fn is at least 0.800 when there is no temperature cross. Increasing the number of shells in series increases the value of Fn, thus permitting a larger temperature cross.The total number of shells also depends on the total surface required, since the size of the individual exchanger is usually limited by handling considerations. If there are no local restrictions due to capacity of tube bundle handling or cleaning equipment, the shell is usually limited to a shell inside diameter of 60 in. (1524 mm). (Note that this limitation does not necessarily apply to fixed tubesheet exchangers in clean service and for kettle type exchangers it applies to the bundle diameter.)
In special cases, such as reboilers and fixed tubesheet exchangers, very large areas per shell are occasionally used. Areas up to 25,000 ft2 (2300 m2), and shell inside diameters up to 100 in. (2540 mm), have been used in fixed tubesheet exchangers. For these cases the individual refineries should be consulted to see if they are equipped to handle oversized tube bundles and shells.
TUBE SELECTION TYPE
Exchanger tubes are commonly available with either a bare or finned outside surface. Selection of the type of surface is based on the application, availability and economics.
The conventional shell and tube exchanger is supplied with bare tubes. These are readily available in any material used in exchanger manufacture and in a wide range of wall thicknesses.
With integral fin tubes, the fins increase the outside area to approximately 2-1/2 times that of a smooth tube. Low-fin tubes should be considered, after an economic evaluation, for services with shellside fouling resistance of 0.0030 hr-ft2-°F/Btu (0.00053 m2-°C/W) or less, unless local field experience precludes their use. Low fins should never be used where the corrosion rate exceeds 2 mils per year, since the fin life will be 3 years or less. Since finned tubes cost 50 - 70% more than bare tubes for the same length and nominal wall thickness, a ratio of total outside resistance (heat transfer film and fouling) to total inside resistance of three or greater, based on bare tubes, is required to justify finned tubes. This ratio occurs frequently in steam-heated reboilers and preheaters, water coolers and condensers handling organic fluids.
TUBE SELECTION (Cont)
Finned tubes may be the economic choice for ratios lower than three if their use requires fewer shells than would be needed for plain tubes. Finned tubes are also particularly useful for debottlenecking, expansions utilizing existing shells and cases where minimum bundle diameter rather than cost is of primary interest. For discussion on the enhancement of heat transfer of both smooth and fin tubes see Section IX-A; for an estimation of tube count, see Table 5 of this section.
Methods for calculating the heat transfer coefficient and pressure drop for low-fin tubes are given in Section IX-G.
LENGTH
The selection of tube length is affected by availability, site preference, and economics. Tube lengths up to 24 ft (7.3 m) are readily obtainable worldwide. Longer tubes [up to 80 ft (24.4 m) for carbon steel and 70 ft (21.3 m) for copper alloys] are available in the United States. However, either 16 ft (4.9 m) or 20 ft (6.1 m) tube length is a common refinery preference. Refer to the refinery for preferable tube lengths. Keep in mind that quite often exchangers are fabricated in one location for use in another location. Some contractors and owners use standard metric lengths, e.g., 6 m instead of 20 ft (6.1 m).
The exchanger cost depends upon the tube length: the longer the tube, the smaller the bundle diameter for the same area. The savings result from a decrease in the cost of shell flanges and tubesheets with only a nominal increase in the cost of the longer shell. In the common range of tube lengths, there is no cost penalty for the longer tubes since length extras are added for steel only over 24 ft (7.3 m) and for copper alloys over 30 ft (9.1 m).
A disadvantage of longer tubes in units (such as condensers) located in a structure is the increased cost of the platforms and additional structure required. Longer tube bundles also require a greater pulling clearance, thereby possibly increasing the plot area requirements.
DIAMETER AND WALL THICKNESS∗∗∗
Exchanger tubing is supplied on the basis of a nominal outside diameter and either a minimum or average wall thickness. For exchanger tubing, the outside tube diameter is fixed. The inside diameter varies with the nominal wall thickness and wall thickness tolerance. Minimum wall tubing has only a plus tolerance on the wall thickness, resulting in the nominal wall thickness being the minimum thickness. Since average wall tubing has a plus-or-minus tolerance, the actual wall thickness can be greater or less than the nominal thickness. The allowable tolerances vary as a function of the tube material, diameter and fabrication method.
For heat exchangers, ExxonMobil usually specifies minimum wall tubing, i.e., the wall thickness at any point must not be thinner than the thickness specified on the exchanger specification sheet. The effect of wall thickness on pressure drop is discussed under PRESSURE DROP in this section.
The following tube diameters and wall thicknesses are preferred in exchangers fabricated by rolling tubes into the tubesheets for the service indicated:
1. Water Service - Nonferrous tubes 3/4 in. OD, 0.065 in. wall thickness; Ferrous tubes 1 in. OD, 0.083 or 0.109 in. wall
thickness.
TUBE SELECTION (Cont)
2. Other Services - Tube diameters, wall thicknesses, layout, and spacing are provided in the following table for various
services placed on the tubeside or shellside: SEVERITY OF SERVICE (FOULING FACTORS; hr⋅⋅⋅⋅ft2-°°°°F/Btu [m2-°°°°C/W]) (ON TUBESIDE) OD, in. (mm) (ON SHELLSIDE) LAYOUT AND SPACING,
in. (mm) BWG
MINIMUM WALL THICKNESS(1),
in. (mm) A. Ferrous Tubes for Oil Service
Non-fouling or fouling < 0.003 (0.00053), mildly corrosive
Non-fouling or fouling < 0.003 (0.00053), corrosive
Extremely fouling ≥ 0.003 (0.00053), mildly corrosive
Extremely fouling ≥ 0.003 (0.00053), corrosive
3/4 (19.05) 3/4 (19.05) 1 (25.4) 1 (25.4) 15/16 (23.81) ∆∆∆∆ 1.00 (25.4) ∆∆∆∆or □□□□ 1.25 (31.75) □□□□ 1.25 (31.75) □□□□ 14 12 12 10 0.083 (2.11) 0.109 (2.77) 0.109 (2.77) 0.134 (3.4) B. Alloy Tubes for General Service
Water service fouling < 0.003 (0.00053) Non-fouling or fouling < 0.003 (0.00053) Extremely fouling ≥ 0.003 (0.00053) 3/4 (19.05) 3/4 (19.05) 1 (25.4) 15/16 (23.81) ∆∆∆∆ 15/16(23.81) or 1.0(25.4) ∆∆∆∆ or □□□□ 1.25 (31.75) □□□□ 16 16 14 0.065 (1.65) 0.065 (1.65) 0.083 (2.11) Note:
(1) When low-finned tubes are specified, the outside diameter of the unfinned portion and the wall thickness at the finned section are required. Although the thickness of the tube in the finned portion is nominally a standard wall thickness, manufacturing tolerances can result in a somewhat thinner wall. It is this minimum wall (given on Table 1 of Section IX-G) that is specified.
Tubing may be supplied on either a minimum or average wall basis. However, the tabulated wall thickness is the minimum acceptable wall thickness.
For special materials (Titanium, High Nickel, etc.) thinner tube walls may be specified.
See the Refinery Construction Materials Manual for minimum thickness as a function of corrosion rate and service.
Refineries may request exceptions to this list on the basis of duplication of existing units or extended periods between exchanger cleaning.
FERRULES
Ferrules are short sleeves inserted into the inlet end of a tube. They are used to prevent erosion of the tube ends due to the inlet turbulence when erosive fluids, such as streams containing solids, are handled. When there is evidence that the tubes will be subject to erosion by solids in the tube side fluid, ferrules should be specified. Ferrule material, length and wall thickness should be given. Also, ferrules are occasionally used in cooling water service to prevent oxygen attack at the tube ends. The ferrules should be secured in place by roller expanding into the tubes. However, ferrules should not be provided if back-flushing is to be used. Assistance in selecting the proper ferrule can be obtained from the Materials Engineering Section of EMRE.
TEMA TYPES
TEMA stands for the Tubular Exchanger Manufacturer's Association, which develops and publishes the basic industrial standards for shell-and-tube heat exchangers. The standards cover heavy-duty heat exchangers (TEMA R) as well as lighter duty heat exchangers (TEMA C and TEMA B). ExxonMobil refineries typically use only the TEMA “R" heat exchangers due to the generally severe requirements of petroleum applications; however, more moderate process services, such as those encountered in Chemical Plants, may warrant consideration of TEMA B or C construction.
Each TEMA heat exchanger is designated by letters representing three major parts: the front end stationary head (commonly referred to as “channel"), the shell, and the rear end head. Each part can be designed in several modifications, commonly referred to as “types." The latest (8th) edition of the TEMA standards designates five types of channels (A, B, C or N, and D); seven types of shells (E, F, G, H, J, K and X); and eight rear end head types (L, M, N, P, S, T, U and W), refer to Figure 2. The
rear end head type determines whether or not the tube bundle is removable from the shell. In any ExxonMobil specification, therefore, the TEMA type heat exchanger is specified by the three single letters which refer to the type of channel, shell and rear end head, respectively; for example, “TEMA type AET."
TEMA TYPES (Cont) CHANNELS
The front end stationary head of shell and tube exchangers is commonly referred to as the channel. The common TEMA channel types and their applications are as follows:
Type A - Features a removable channel with a removable cover plate. It is used with fixed tubesheet, U-tube and removable
bundle exchanger designs. This is the most common stationary head type.
Type B - Features a removable channel with an integral cover. It is used with fixed tubesheet, U-tube and removable bundle
exchanger designs. This type is normally used only when the tube side fouling factor is less than 0.0020 hr-ft2-°F/Btu (0.00035 m2-°C/W) or when flange leaks represent a greater concern than easy access for cleaning and inspection. Note that with this channel, the mechanical cleaning of the bundle requires disassembly of the piping spool piece or elbow.
Type C and Type N - Two types of channels that are integral with the tubesheet. Type C is attached to the shell by a flanged
joint and is used for U-tube and removable bundles. The N type has a tubesheet that is integral with the shell and is used with fixed tubesheet designs. Note that neither the C nor the N type is recommended in normal application because of extremely high maintenance costs for: a) assembling/testing of the heat exchanger, b) tubesheet repair, and c) retubing of the bundle.
Type D - This is a special high-pressure head used when the tube side design pressure exceeds approximately 1000 psi (6900
kPa gage). The channel and tubesheet are integral forged construction. The channel cover is attached by special high-pressure bolting.
TUBE BUNDLES TYPES
Fixed Tubesheet Design (TEMA Rear End Head Types: L, M or N) - Fixed tubesheet exchangers have both tubesheets
attached directly to the shell and are usually among the most economical exchangers for low design pressures. This type of exchanger construction should be considered when no shell side cleaning or inspection is required, or when in-place shell side chemical cleaning is available and applicable. Differential thermal expansion between tubes and shell limits applicability to moderate temperature differences. [An expansion joint may be required when there is a difference of more than 50°F (28°C) between average tube metal temperature and average shell metal temperature.] All start-up, operating, and upset conditions as well as normal operating conditions must be examined in determining the maximum temperature difference. For services that would require an expansion joint, a U-tube bundle may be more economical.
Shell and tubesheet material specifications incompatible for welding to each other (i.e., steel to aluminum or high copper alloys) might preclude fixed tubesheet construction, necessitate a change in metallurgy, or require a clad tubesheet.
U-tube Design (TEMA Rear End Head Type U) - U-tube exchangers represent the greatest simplicity of design, requiring only
one tubesheet and no expansion joint or costly floating head, while permitting individual tube differential thermal expansion. U-tube exchangers are the least expensive units for high tube side design pressures. The tube bundle can be removed from the shell, but replacement of individual tubes (except for those on the outside of the bundle) is impractical.
External surfaces of the tube bundle are mechanically cleanable. When the tube side fouling factor exceeds 0.0010 hr-ft
2-°F/Btu (0.00017 m2-°C/W), chemical cleaning is preferable, although mechanical cleaning with flexible end-drill shafts is possible. U-tube construction is not used when the tube side fouling factor exceeds 0.0020 hr-ft2-°F/Btu (0.00035 m2-°C/W). Although the bend portion of the U-tube bundle provides heat transfer surface, it is relatively ineffective for sensible heat transfer compared to the straight tube length surface area. Therefore, when the effective surface area for U-tube bundles is calculated, only the surface area of the straight portions of the tubes is included, except for isothermal boiling or condensing on the shell side.
Pull-through Floating Head Design (TEMA Rear End Head Type T) - A pull-through floating head exchanger has stationary
tubesheet at the channel end and a floating tubesheet with a separate floating head cover at the rear end. The bundle can be easily removed from the shell by disassembling only the channel. The floating head flange and bolt design require a relatively large clearance between the bundle and shell, particularly as the design pressures increase. Because of this clearance, the pull-through bundle has fewer tubes per given shell size than other types of construction. The bundle to shell clearance, which decreases the shell side heat transfer capability, should be blocked by sealing strips or dummy tubes to reduce shell side fluid bypassing. Mechanical cleaning of both the shell and tube sides is possible.
TUBE BUNDLES (Cont)
Split-ring Floating Head Design (TEMA Rear End Head Type S) - A split-ring floating head exchanger has a stationary
tubesheet at the channel end and a floating tubesheet that is sandwiched between a split-ring and a separate floating head cover. The floating head assembly moves inside a shell cover of larger diameter than that of the shell. To remove the bundle from the shell, the front end head, rear shell cover, and floating tubesheet cover/split ring assembly must be dismantled. This requirement is the greatest disadvantage of the split-ring design since it requires considerably more time to pull the bundle. Mechanical cleaning of both the shell and tube side is possible.
Outside Packed Floating Head (TEMA Rear Head Type P) and Externally Sealed Floating Tubesheet (TEMA Rear Head Type W) Designs - There are two variations of outside packed floating head designs: the lantern ring type (W) and the
stuffing box type (P). In the lantern ring design, the floating head slides against a lantern ring packing which is compressed between the shell flange and the shell cover flange. The stuffing box design is similar to the lantern ring type, except that the seal is against an extension of the floating tubesheet, and the tubesheet cover is attached to the tubesheet extension by means of a split-ring. Neither of these designs is normally specified by ExxonMobil, since the packing has a tendency to leak, allowing the streams to leak to the atmosphere. These designs should never be used for hydrocarbons or toxic fluids. Mechanical cleaning of both the shell and tubeside is possible.
See Table 3 for guidance in TEMA head type selection. TUBE LAYOUT
There are four types of tube layouts with respect to the shell side cross flow direction between baffle tips: square (90°), rotated square (45°), triangular (30°) and rotated triangular (60°). See Figure 3 for illustrations. The 60° triangular layout is seldom used because its heat transfer characteristics are poor compared to the pressure drop expenditure.
Use of the triangular layout (30°) is preferred for all exchangers (except reboilers with a heat flux exceeding 6000 Btu/hr ⋅ ft2 [19,000 W/m2]) with the shell side fouling factors of 0.0020 hr-ft2-°F/Btu (0.00035 m2-°C/W) or less, provided the deposit is subject to chemical cleaning. An exchanger with triangular layout costs less per sq ft and transfers more heat per sq ft than one with a square or rotated square layout. For this reason, triangular layout is preferred where applicable.
Square or rotated square tube layout should be specified for all shell and tube exchangers when the shell side fouling factor is over 0.0020 hr- ft2-°F/Btu (0.00035 m2-°C/W), whenever mechanical cleaning of the outside of the tubes is required, or when requested by the refinery. Square or rotated square tube layout is also required for reboilers with a heat flux exceeding 6000 Btu/hr-ft2 (19,000 W/m2). Rotated square layouts are preferable for laminar flow because of a higher heat transfer coefficient caused by induced turbulence. In turbulent flow, especially for pressure-drop-limited cases, square layout is preferred because the heat transfer coefficient is equivalent to that for rotated square layout while the pressure drop is somewhat less.
Tube layout for removable bundles may be either square (90°), rotated square (45°) or triangular (30°). Nonremovable bundles (fixed tubesheet exchangers) are always triangular (30°) layout.
TUBE PITCH
The tube pitch (PT) is the center-to-center distance between adjacent tubes (Figure 3). The common pitches used are:
TUBE SIZE TRIANGULAR SQUARE HEAVIEST WALL
1. 3/4 in. OD tubes (19.05 mm) 15/16 in. (23.81 mm) — 0.095 in. (2.41 mm)
2. 3/4 in. OD tubes (19.05 mm) 1 in. (25.4 mm) 1 in. (25.44 mm) 0.109 in. (2.77 mm)
3. 1 in. OD tubes (25.4 mm) 1-1/4 in. (31.75 mm) 1-1/4 in. (31.75 mm) 0.134 in. (3.40 mm)
4. 1-1/2 in. OD tubes (38.1 mm) 1-7/8 in. (47.63 mm) 1-7/8 in. (47.63 mm) 0.165 in. (4.19 mm)
5. 1-1/4 in. OD tubes (31.75 mm) 1-9/16 in. (39.7 mm) 1-9/16 in. (39.7 mm) 0.165 in. (4.19 mm)
6. For tube sizes larger than 1-1/2 in. (38.1 mm), use 1.25 times the outside diameter.
7. In kettle-type reboilers where the pressure is less than 50 psig (345 kPa - gage) and the heat flux is greater than 10,000 Btu/hr•ft2 (31,500 W/m2), use 3/8 in. (9.5 mm) spacing between tubes.
With 3/4 in. (19.05 mm) OD tubes on triangular layout, 15/16 in. (23.81 mm) pitch should be used unless the pitch is limited by the wall thickness required, since this results in a smaller unit due to the higher coefficient and closer spacing. The column “Heaviest Recommended Wall" is based on the maximum allowable tubesheet distortion resulting from rolling the indicated tube into a tubesheet having the minimum permissible ligament width at the listed pitch. The ligament is that portion of the tubesheet between two adjacent tube holes.
TUBE BUNDLES (Cont) CROSS BAFFLES
Cross baffles support the tubes, restrain tube vibration and direct fluid flow on the shell side. Three types of baffles are generally used: segmental, double segmental, and rod baffles. The first two baffle types are illustrated in Figure 4. The rod baffle concept is discussed and illustrated in Section IX-A. In addition, full circle support baffles are occasionally used for tube support in kettle reboilers.
Baffle cut defines the segment of the baffle “cut" away to provide for fluid flow past the chord of the baffle. For segmental
baffles, this is expressed as the ratio of the segment height to shell diameter in percent. Segmental baffle cuts are usually about 25%, although the maximum practical cut for tube support is approximately 45%. Test work (single phase) by Heat Transfer Research, Inc. (HTRI), on segmental baffles indicates that optimum baffle cuts range from 15 - 30% with a slight peak at 25%.
Double segmental baffle cut is expressed as the ratio of total window area to exchanger cross-sectional area in percent. Normally the window areas for the single central baffle and the area of the central hole in the double baffle are equal and are 40% of the exchanger sectional area. This allows a baffle overlap of approximately 10% of the exchanger cross-sectional area on each side of the exchanger. However, there must be enough overlap so that at least two rows of tubes are supported by adjacent segments.
Baffle pitch is the longitudinal spacing between baffles. The maximum baffle pitch is a function of tube size and metallurgy
and, for no change of phase flow, of shell diameter. See Table 4 in Section IX-D which gives maximum allowable values as a function of tube diameter. If there is no charge of phase in the shell side fluid, the baffle pitch should not exceed the shell inside diameter. Otherwise, the fluid would tend to flow parallel with the tubes, rather than perpendicular to them, resulting in a poorer heat transfer coefficient. With shell side condensation or vaporization, the maximum baffle spacing is a function of tube diameter only, as indicated in Table 4 of Section IX-D. Also, refer to TEMA Table RCB-4.52 for maximum unsupported straight tube spans.
The minimum allowable baffle pitch is 20% of the shell inside diameter or 2 in. (50 mm), whichever is greater. A very tight baffle pitch tends to force the shell side fluid into leakage and bypass streams. This decreases the effective cross flow stream, resulting in a decreased heat transfer coefficient.
Position of Baffle Chord
1. Square and Rotated Square Tube Layout
a. The chord should be vertical (usually with a large baffle cut) in condensers, vaporizers and units handling fluids containing suspended solids. This arrangement minimizes “pockets" which would trap vapors or sediment, reducing the effective surface area.
b. The baffle chord should be horizontal when sediment-free, single phase shell side fluid is being cooled or heated through a wide temperature range [200 to 300°F ( 110 to 165°C)] in one shell. This avoids fluid stratification.
c. The baffle chord can also be placed at a 45° angle to grade. This position is termed “on the bias". When sedimentation and stratification possibilities co-exist, a 45° chord at a suitable baffle cut provides a reasonable compromise solution to both problems. A 45°chord is also useful to prevent the accummulation of heavy components in an exchanger when a wide boiling range fluid is being partially vaporized. With horizontal pass partitions, the 45° chord mitigates the effect on heat transfer of fluid flow through the pass lanes.
2. 30°°°° Triangular Tube Layout
a. When a sediment-free shell fluid is being cooled through a wide temperature range [200 - 300°F ( 110- 165°C)] in one shell, the chord of the baffle should be horizontal, to avoid fluid stratification (see Par. 1. b. above).
b. The chord of the baffle should be vertical for all other services (see Par. 1. a. above). SEAL STRIPS
Seal strips are flat strips of metal (or tie rods) that extend the length of the shell in order to prevent the shell side fluid from flowing through the clearance between the tube bundle and the inner wall of the shell. The seal strips are located between the baffle chords of adjacent baffles, in the clearance between the tube bundles and the shell. Figure 6 illustrates a typical seal strip installation. Note that seal strips are normally installed in pairs, a strip on each side of the bundle.
Seal strips are normally installed on pull-through and split ring bundles because the clearance between the tube bundle and the shell is quite large [2 to 5 in. (50 to 125 mm)]. They are also installed on other types of tube bundles when the clearance between the outer tubes and the shell exceeds one half the tube pitch. Seal strips are not used in kettle reboilers or in units with isothermal condensation occurring on the shell side. Refer to GP 6-1-1 for specific requirements concerning seal strip
TUBE BUNDLES (Cont) IMPINGEMENT PROTECTION
Impingement plate or other protection devices are normally installed on the top row of tubes under the shell inlet nozzle or in shell side inlet nozzles to protect the bundle against impingement by the incoming fluid when the fluid: (a) is condensing, (b) is a liquid vapor mixture, (c) contains abrasive material, or (d) is entering at high velocity. In addition, TEMA requires bundle impingement protection when nozzle values of ρV2 exceed:
1. 1500 lb/ft-sec2 (2230 kg/m-s2) for noncorrosive, nonabrasive, single phase fluids. 2. 500 lb/ft-sec2 (744 kg/m-s2) for all other liquids, including liquid at its boiling point.
The minimum bundle entrance area should equal or exceed the inlet nozzle area and should not produce a value of ρV2 greater than 4000 lb/ft-sec2 (5950 kg/m-s2) per TEMA. An impingement baffle can be a flat plate, a curved plate or an angle bar; angle bars must be installed at each tube individually. The use of perforated impingement baffles is prohibited by GP
6-1-1. In order to maintain a maximum tube count, the impingement plate is sometimes located in a conical nozzle opening or
in a dome cap above the shell. The impingement protection device material should be of at least the quality of the tube material.
INTERCHANGEABILITY OF TUBE BUNDLES
There is an incentive to standardize tube bundle design within a project, to reduce the refinery's expensive inventory of spare bundles and other exchanger parts. Although desirable, standardization from project to project is less subject to control since the exchanger fabrication may not be by the same manufacturer. It is not considered desirable to overdesign a heat exchanger solely for bundle interchangeability, unless the required surfaces are nearly equal (within approximately 5%) and the design pressures and temperatures are similar.
Tube bundles must be of the same general design and must have identical tubesheets to be physically interchangeable, including the same grooving for pass partitions even though the two units are designed for a different number of passes.
SHELL TYPES SINGLE-PASS SHELL (TEMA E)
The single-pass shell is the most common shell construction for shell and tube heat exchangers. The shell side inlet and outlet nozzles are located at opposite ends of the shell. The nozzles can be placed on opposite or same sides of the shell, depending on the number and type of baffles used. A typical one shell-pass exchanger with horizontal cut segmental baffles is illustrated in Figure 5A.
TWO-PASS SHELL (TEMA F)
A two-pass shell requires the use of a longitudinal baffle to direct the shell side flow. An exchanger with two shell passes is illustrated in Figure 5C. Note that both the shell inlet and outlet nozzles are adjacent to the stationary tubesheet. To avoid excessively thick longitudinal baffles, two-pass shells should not be used with a shell side pressure drop greater than 10 psi (70 kPa) [5 psi (35 kPa) is preferred]. Also, a shell side temperature range exceeding 350°F (195°C) should be avoided since greater temperature ranges result in excessive heat leakage through the baffle, as well as thermal stresses in the baffle, shell and tubesheet.
The longitudinal baffle can be of either welded or removable design. Since there are severe design difficulties and cost penalties associated with the use of welded baffles in floating head exchangers, this type of design should be used only with fixed tubesheet units that do not require expansion joints. A welded baffle may be used in a four-pass U-tube design while retaining the ability to remove the bundle from the exchanger. Any welded baffle is limited to a minimum shell ID of 16 in. (406.4 mm) due to welding feasibility. If an ID of less than 16 in. is desired, a double pipe design could be considered. If a longitudinal baffle is to be used in a floating head exchanger, it should be a removable design. Removable longitudinal baffles require the use of flexible, light gage sealing strips or a packing device between the baffle and the shell, to reduce fluid leakage from one side to the other.
Care must be taken during maintenance operations in removal or insertion of bundles with removable longitudinal baffles to avoid damaging the sealing device. Damaged seals will result in poor heat transfer performance and it is recommended that they be replaced every time the bundle is removed. The shell ID must be ground smooth during fabrication to assure a good seal between the sealing device and the shell.
SHELL TYPES (Cont)
Recent developments in large diameter multi-tube hairpin exchangers have turned hairpins into viable alternatives to F-shells and should be considered where applicable.
A two-pass unit may be used when the LMTD correction factor Fn is less than 0.8 for a one-shell pass unit. A two-shell pass unit with a welded longitudinal baffle is satisfactory if Fn for two-shell passes is equal to or greater than 0.85, and a unit with a removable baffle is satisfactory if Fn for two-shell passes is 0.90 or greater. These limitations on Fn are the result of heat loss through the longitudinal baffle and, in the case of a removable baffle, fluid leakage around it.
DIVIDED FLOW SHELL (TEMA J)
A divided flow shell has a central inlet nozzle and two outlet nozzles, or vice versa. A divided flow exchanger is illustrated in Figure 5B. Typically, this type of shell is used to reduce pressure drop and to avoid vibration damage in a condensing service. The following list shows exchanger types ranked in order of decreasing pressure drop:
1. E shell with segmental baffles. 2. E shell with double segmental baffles. 3. J shell with segmental baffles. 4. J shell with double segmental baffles. 5. E shells in parallel with segmental baffles. 6. E shells in parallel with double segmental baffles. 7. J shells in parallel with segmental baffles. 8. J shells in parallel with double segmental baffles.
For most designs, double segmental baffles are used with J shells.
Double segmental baffles in a divided flow exchanger normally have a vertical cut since J shells are most often in condensing or vaporizing service. A full circle support baffle is usually provided at the center nozzle. The baffle arrangement requires that there be an odd total number of baffles, as well as an odd number of baffles in each end of the shell. The baffles on each side of the central baffle and the last baffle at the ends of the shell have solid centers with cut-away edges to optimize flow through regions of the tube bundle at the shell nozzles.
KETTLE (TEMA K)
This type is used for vaporizing services (reboilers, steam generators and refrigeration services). A dome area is provided to separate the vapor from the entrained liquid. Process reboilers are provided with a weir plate to maintain a constant liquid level 1 in. (25mm) to 2 in. (50mm) above the top row of tubes.
CROSSFLOW SHELL (TEMA X)
This type is used when low pressure drop is required or when the shell nozzles are unusually large and is a possible solution to vibration problems since it can be designed for low cross flow velocity.
SPLIT FLOW (TEMA G) AND DOUBLE SPLIT FLOW (TEMA H)
These types are normally used for thermosyphon reboilers. G type shells are typically used for tube lengths up to 10 ft (3.05 m) and H type shells are typically used for tube lengths between 10 ft (3.05 m) and 20 ft (6.1m).
TUBE SIDE FLOW
Whichever fluid appears higher on the following list will ordinarily be passed through the tubes: 1. Cooling water.
2. Corrosive fluid (especially those requiring higher alloy materials) or a fluid likely to deposit coke, sediment or other solids. 3. A fluid which is fouling.
4. The less viscous of the two fluids.
5. The fluid under higher pressure [for extremely high pressures, 1000 psi (6900 kPa gage), it may be economical to treat this item as Number 1 on this list].
6. The hotter fluid.
TUBE SIDE FLOW (Cont) Several exceptions to this list are:
1. Condensing vapors are normally passed through the shell. 2. Condensing steam is normally passed through the tubes.
3. If the temperature change of one fluid is very large [greater than approximately 300 to 350°F ( 165 to 195°C)], that fluid is usually passed through the shell, rather than the tubes, if more than one tube pass is to be used. This minimizes the construction problems caused by thermal expansion. Also, to avoid thermal stress problems, fluids with greater than 350°F (195°C) temperature change cannot be passed through the shell side of a two-pass shell (GP 6-1-1).
4. If one of the fluids is clean [fouling factor 0.0010 hr-ft2-°F/Btu (0.00018 m2-°C/W) or less] and is somewhat corrosive (see the Refinery Construction Materials Manual), this fluid is passed through the tubes and U-tube construction is used since it is typically more economical to alloy up the tubeside versus the shellside to easily handle the corrosive fluid.
PRESSURE DROP
The calculation of pressure drop is covered in greater detail under Calculation of Surface, Sections IX-D, IX-E and IX-F. General considerations applicable to shell and tube exchangers are given in the following paragraphs.
CORRECTION FOR REDUCED FLOW AREA
Tube Side - GP 6-1-1 requires that exchanger vendors provide tubes having no point of the wall thinner than the specified
thickness. For this reason, actual inside diameters are always smaller than standard diameters. This deviation from the standard diameter is greater for carbon steel than for copper alloy tubes. The tube side pressure drop equations used in this manual include a factor that corrects for the increased wall thickness and for fouling on the inside of the tubes. In other words, tube side pressure drop calculations are based on minimum wall tubes with a 10 to 20% increase (depending on tube inside diameter) for fouling deposits. Table 4 of Section IX-D shows the values for this factor. An additional method to correct for reduced flow area using improved correlations can be found on Figure 4 and 4A in Section IX-B.
Low-fin tubes require special consideration. The inside diameters listed in Table 1 of Section IX-G must be used for pressure drop and heat transfer calculations.
Shell Side - The equation for shell pressure drop (Section IX-D) includes a factor that accounts for fouling of the outside of the tubes. This fouling reduces the free flow area, thereby increasing the pressure drop. Typical values of the factor are given in Table 4 of Section IX-D.
Note that fouling affects both pressure drop and heat transfer. The corrections discussed here address the effect on pressure drop only. The effect of fouling on heat transfer is discussed in Section IX-B.
ECONOMICS CONSIDERATION OF PRESSURE DROP
The design pressure drop is usually determined by an economic balance between: (1) the increased cost of pumping and of certain exchanger components, and (2) the decreased surface area. As the number of tube passes is increased or the baffle pitch is decreased, the fluid velocity becomes greater, thereby increasing the pressure drop and the convective heat transfer coefficient. Increased velocity also decreases the rate of fouling.
For pressured streams, the pressure drop should be maximized. Where there are no material or process restrictions on maximum velocity, a reasonable limitation for liquids is about 10 to 15 ft/sec (3 to 5 m/sec). For gases and vapors, velocities up to 100 ft/sec (30 m/sec) are common. Typical pressure drops are shown in Table 4 in Section IX-B. Designs are typically
based on pressure drop allowances set by experience, and many systems exist where pressure drop is dictated by process concerns.
For pumped streams, unless otherwise limited, the preliminary design pressure drop is that required to fully load the pump driver. The pressure drop should be finalized by checking the return on incremental investment by comparing the first case with that using the next larger size driver, loaded either fully or to the extent permitted by velocity limitations. Items to be considered in the comparison are a possible change in line rating or pump type.
Tube Side (Number of Passes) - For coolers and condensers using water, specify enough tube passes to maximize the
utilization of the available pressure drop as limited by materials. The permissible maximum and minimum water velocities vary with the tube material and water type as shown in Table 4 of Section IX-D. Note that for cooling water service care must be taken to avoid using excessive pressure drop in elevated exchangers due to the need to ensure that vacuum conditions do not occur.
For hydrocarbon services, specify sufficient passes to make the tube velocity high enough to prevent laminar or transitional flow (or as shown on Figure 1.8 of Section IX-D, (Re) > 3700).
PRESSURE DROP (Cont)
In some cases the exchanger terminal conditions will be such that the type of flow could change from laminar to turbulent (or vice versa) within the unit. This must be avoided, since prediction of heat transfer coefficients in the transition region is unreliable. For some services, such as heavy fuel oil, this may be overcome by fluxing with a suitable fluid of lower viscosity. Ordinarily no fewer than two nor more than eight tube passes are used. With more than eight or fewer than two passes, the construction becomes complicated and fabrication costs tend to become higher. Note that two, four, or eight tube-pass arrangements are sometimes easily interchanged as long as the exchangers remain within the allowable pressure drops. However, in special services, one pass or more than eight passes can sometimes be justified.
See Table 4 for the maximum number of tube passes, normally used for various shell diameters. Restrictions on tube pass arrangements for particular exchanger designs are as follows:
1. Fixed Tubesheet Exchangers - Any number of tube passes, odd or even, is possible. The most common arrangements
are single pass or an even number of multiple passes.
2. U-tube Exchangers - Any even number of tube passes is possible, but normally the maximum recommended is six,
because of construction considerations.
3. Pull-through and Split-ring Floating Head Exchangers - Any even number of tube passes is possible. Single-tube-pass
designs require special expansion joints , and are not generally used.
4. Lantern Ring (TEMA Type W) Externally Sealed Floating Tubesheet Exchangers - Only single or double tube-pass
arrangements are possible for this type of construction.
5. Stuffing Box (TEMA Type P) Outside Packed Floating Head Exchangers - Any even or odd number of tube passes is
possible. SHELL SIDE
1. Number of Shell Passes - The vast majority of exchangers are of one-pass shell (Type E) design, or variants thereof
(Types J, K or X). Two-pass shells (Type F), or variants (Types G and H), are less frequently utilized. Multiple pass shells require a longitudinal baffle of either welded or removable design (see previous discussion of two-pass shells).
Where there are very large surface requirements, especially in services requiring multiple exchangers in series to meet Fn requirements, multiple shell-pass units employing longitudinal baffles may be the most economical choice. However, it is difficult to estimate accurately the cost of maintaining exchangers with longitudinal baffles. In addition, if the longitudinal baffle seal strips are damaged, the shellside bypass streams could result in large performance debits. Therefore, it is recommended that the exchanger be specified as a one-pass shell unit if possible. If specific project conditions require the features of a two-pass shell, multi-tube hairpin designs should also be considered.
2. Cross Baffles - Segmental baffles are normally specified, unless shell side pressure drop is excessive, even at maximum
baffle spacing. In such cases, double segmental baffles should be used because they result in lower pressure drop with only a slight decrease in the shell side heat transfer coefficient.
3. “No-Tubes-in-Window" Design - This design is a bundle with segmental baffles that have no tubes in the “windows" left
by the baffle cut. Each tube is supported by every baffle, whereas, in conventional bundles, the tubes in the window region are supported only by every other baffle. The “no-tube-in-window" bundle is used either to minimize pressure drop or to minimize vibration damage to tubes. Intermediate tube support plates may be inserted between NTIW baffles to further increase tube natural frequency, without impacting exchanger performance.
4. Nozzles - Use two shell outlet nozzles (one at each end of the shell) and a center inlet nozzle, or vice versa, if other
means of reducing pressure drop are ineffective. This is termed a “divided flow" (TEMA J) shell.
For condensers or other exchangers handling vapors, an enlarged shell section at the inlet nozzle (“vapor belt") is sometimes used. Vapor belts are recommended if the inlet nozzle is large compared to the shell diameter. This design will prevent deleting tubes to provide the required flow area between the shell and the vapor impingement baffle and will provide inlet vapor distribution to the bundle.
STACKING OF SHELLS
The choice of whether to stack shells depends on maintenance considerations and the amount of plot area available. Stacking of shells requires less area and frequently less piping. Normally, shells are not stacked more than two high. However, stacked heat exchangers are more costly to maintain because they are harder to access.
If sufficient plot area is available, the following should govern:
1. When the fluids are known to be clean and noncorrosive (refer to Construction Materials Manual), the shells should usually be stacked.
2. If the fluids are moderately clean or slightly corrosive, the shells may be stacked.
3. If the fluids are very dirty or corrosive, stacked shells may be used as long as maintenance requirements are taken into consideration, but the maximum recommended number of stacked shells is two.
When multiple shells are specified, the stacked arrangement should be noted on the exchanger specification sheet. The following guidelines should be considered when determining the maximum number of shells to be stacked: 1. When shell ID is less than 24 in. (610 mm), the maximum number of stacked shells is four.
2. When shell ID is between 24 and 40 in. (610 to 1010 mm), the maximum number of stacked shells is three. 3. When the shell ID is more than 40 in. (1010 mm), the maximum number of stacked shells is two.
RECOMMENDED GOOD PRACTICE
MAINTENANCE FEATURES FOR THE HEAT EXCHANGER DESIGN
Considerable lifetime maintenance savings can be realized if the designer utilizes recommendations of the Heat Exchanger
Maintenance Good Practice Designer's Checklist (Table 7). The table presents typical maintenance and heat exchanger mechanical design features that will either minimize or eliminate costly problems.
It is recommended that designers review their proposed heat exchanger configuration against Table 7 in order to develop a maintenance effective heat exchanger. However, it is also important to remember that the configurations suggested by this table are “wants" and not “musts." Specific local requirements in any design may overrule the suggestions given in the table, and the table is not all inclusive of possible solutions.
FLOW INDUCED VIBRATION
In a shell and tube heat exchanger, higher velocities will produce higher heat transfer coefficients. However, higher velocities may also produce undesirable vibration. The vibration can be severe enough to cause mechanical failure, unacceptable flow fluctuations or noise levels exceeding OSHA guidelines. A brief discussion of the various vibration mechanisms along with possible corrective measures is provided below. Also basic equations showing, directional only, relationship between vibration parameters, exchanger geometry and fluid characteristics are presented.
VIBRATION MECHANISMS
There are various vibration mechanisms that excite, or impart energy, to a tube, thereby inducing and sustaining tube vibration. The mechanisms described below act alone or in combination to cause damage.
1. Acoustic Wave - An acoustic wave is considered to be an oscillating column of gas standing at a right angle to the flow
direction. If the acoustic wave is in resonance with tube natural frequency, damage to the tube bundle may occur, as well as damage to associated piping and exchanger foundation and supports.
2. Vortex Shedding - The flow of fluid across a tube produces vortices in the wake downstream of the tube, causing tube
vibration. The vortex-induced vibration is not strong enough to produce large stresses in the tube unless the vortex shedding frequency is in resonance with the natural frequency of the tube. Resonance at full load or steady state flows must be avoided.
3. Turbulent Buffeting - The constant change in velocity of the fluid flowing across a tube causes high turbulences and
eddies around the tube. The energy received by the tube depends on the frequency of the turbulence patterns vs. the tube natural frequency.
4. Fluidelastic Whirling - This results when fluctuations in the balance of fluid flow on either side of a tube induce a
self-corrective action that depends on fluid flow velocity and the inertia (damping) of the system. (If the self-corrective action is faster than the cause, the system becomes unstable.) The velocity at which fluidelastic whirling starts is called critical velocity, or Connors critical velocity. Fluidelastic whirling causes tubes to move in an orbital motion.
5. Longitudinal Flow - Longitudinal flow, or flow parallel to the axis of a tube, can also produce tube vibration, but the
velocities required to produce damage are extremely high compared to velocities in crossflow. Therefore, this mechanism is not discussed further.
FLOW INDUCED VIBRATION (Cont) TYPES OF DAMAGE
Various types of damage are caused by vibration, almost all of which can be attributed to fatigue. It should also be noted that stresses associated with vibration may increase the corrosion rates of component materials. Some common types of damage that have occurred are described below:
1. Baffle Damage - This describes the thinning and eventual failure of a tube where is passes through baffles. As a tube
vibrates, it hits the baffle resulting in wear.
2. Collision Damage - This type of damage results when tubes hit each other due to large amplitudes. The main reason for
this type of damage is high crossflow velocities, accompanied by large unsupported spans.
3. Clamping Damage At Tubesheets - Clamping of tubes at the tubesheet increases the natural frequency of the tubes in
the spans closer to the tubesheet. At the same time, mechanical considerations require larger inlet and outlet spacings, thus reducing the tube natural frequency. Tubes are expanded into tubesheet grooves by plastic deformation caused by the tube expander. Improper tube expansion leaves large residual stresses at the tube-to-tubesheet joint. Also, due to discontinuity in shape or material at the tube-to-tubesheet joints, any thermal stresses caused by differential thermal growth between tubes are magnified by the stress concentration factors. All of the above factors can cause the stress values to be higher at the tubesheets and may cause the tube to snap where it emerges from the tubesheet. In addition, stresses in the tube due to lateral deflection are maximum at the tubesheet (i.e., cantilever verses simple support) contributing to possible fatigue failure.
4. Acoustic Wave - As discussed, this is an oscillating gas column wave being excited by phased vortex shedding. The
generated sound will not affect the tube bundle unless the acoustic resonant frequency approaches the tube natural frequency, in which case the probability of tube failure will be high.
5. Stress Risers - Nonhomogeneous material properties may contribute to failure due to reduced allowable stress, or by
creating a high stress concentration factor. Also, poor quality control at the fabrication stage may lead to high residual stresses which, coupled with other stresses during operation, can lead to fatigue failure that may resemble a vibration failure.
AREAS OF CONCERN
Most tube failures can be attributed to high turbulence combined with large unsupported tube spans, and accompanying lower tube natural frequencies.
1. Velocity - From the discussion so far, it is apparent that the shell side crossflow velocity is of prime importance in
determining various vibration parameters. Shell side vibration analysis includes the affect of leakage stream on cross flow velocities . (Refer to Section IX-D for a description of various leakage streams.)
2. Inlet and Outlet Spaces - Due to fabrication limitations, the inlet and outlet spaces of most shell-and-tube exchangers are
larger than the central baffle spaces. This results in lowering of the tube natural frequency. Also in this region are the inlet and outlet nozzles and bundle entrance and exit zones. Normally an impingement plate is present in the inlet zone. To maximize tube count, tie rods are often placed in the entrance and exit zones. Therefore, the inlet and the outlet spaces not only have lower tube natural frequencies, but may also have localized high velocities.
3. Baffle Tips Zone - This zone consists of several tube rows on either side of the baffle tips. In this area, all the
baffle-to-tube leakage streams meet with the crossflow and pass partition leakage streams. Therefore, this region experiences the highest crossflow velocities. Also, in this region flow direction and velocity change from crossflow to window or longitudinal flow, producing additional turbulence.
4. U-Bend Region - Normally velocities in the U-bend region are much lower, but the unsupported spans are almost always
very large. Also, it is practically impossible to maintain distance between tubes to the close tolerance required in the straight portion of the tubes. Local low velocities also may facilitate sedimentation of solids, if present, which may lead to localized high velocities in other parts of the U-bend region. Special tube supports can be used to reduce span length and stiffen the bends.
5. Bundle Periphery - At the bundle periphery, the “E" stream, or shell-to-bundle stream, mixes with the other streams and
may produce added turbulence. This region also has seal strips and tie rods, which may lead to large localized velocities. In this region, the effect of vibration is not dampened by the presence of tubes in close proximity on all sides.
FLOW INDUCED VIBRATION (Cont) DEFINITIONS
1. Tube Natural Frequency - This is the frequency at which a tube will vibrate if excited and then continue to vibrate without
further excitation.
2. Vortex Shedding Frequency - This is the frequency of alternate shedding of vortices produced in the downstream wake
as flow across a tube separates alternately from the opposite sides of the tube.
3. Turbulent Buffeting Frequency - This is the dominant central frequency of turbulent buffeting.
4. Critical Velocity - This is the threshold crossflow velocity, above which enough energy can be imparted to the tube to
cause sustained tube vibration.
5. Baffle Damage Number - This number indicates the probability of tube damage at the baffle due to vibration.
6. Collision Damage Number - This number indicates the probability of tube vibration due to collision with other tubes.
HTRI has recently discontinued use of damage numbers, however, they may still appear within the industry.
7. Hydrodynamic Mass - This describes an increase in the apparent weight of the vibrating body (tube) due to the
displacement of the medium (shell side fluid) in which the body (tube) is vibrating.
8. Strouhal Number - This number is used to correlate vortex shedding frequency to crossflow velocity and tube diameter.
This number depends on the type (square, triangular, etc.) of pitch and tube pitch to tube diameter ratio.
9. Unsupported Tube Span - This describes the maximum length of a tube without support from a tubesheet, baffle or any
other means of support. ➧ BASIC EQUATIONS
There is no complete agreement between established methods (TEMA, HTRI, etc.) of calculating the vibration parameters. Most equations used are directionally in agreement (i.e., most equations would agree as to whether one parameter would increase or decrease as a given variable increases or decreases). The equations presented here are directional only. These equations contain constants that are dependent on some variables, and the exact relationship is not presented. The reason for this type of presentation is to avoid discrepancy with any individual methods, and to have simple directional relationships.
5 . 0 o 2 L n n W EM S C C f þ ý ü î í ì − = Wo = Wtm + Wt + Cm⋅ Ws o c T S s d V S C f = ⋅ ⋅ o c Tb Tb d V C f = ⋅ s s a a D V C f = ⋅ o n n m xc B CR C C f d V = ⋅ ⋅δ ⋅ ⋅
(
)
2 c BF BB BT BD C C C V S N = ⋅ ⋅ ⋅ρ⋅ ⋅ E S V C C N 4 2 c CB CT CD = ⋅ ⋅ ρ ⋅ ⋅ M E S V d C X 4 2 c o x c = ⋅ ρ ⋅ ⋅ ⋅ ⋅FLOW INDUCED VIBRATION (Cont)
where: Ca = Factor depending on the fraction of shell volume occupied by tubes CB = Instability coefficient depending on the tube layout geometry
CBB = Factor depending on the baffle thickness and the baffle-to-tube clearance CBF = Factor depending on the fatigue stress
CBT = Factor depending on the tube geometry
CCB= Factor depending on the baffle-to-tube and the tube-to-tube clearances CCT = Factor depending on the tube geometry
CL = Factor depending on the tube axial loading Cm = Factor depending on the tube layout geometry
Cn = Factor depending on the baffle layout and the position of the span Cs = Constant based on the units used (Customary or Metric)
CTb = Factor depending on tube layout geometry
Cx = Factor depending on the ratio of forcing frequency to the tube natural frequency and system damping
Cxc = Factor depending on Wo, δ, ρ and the tube layout geometry do = Tube outer diameter
Ds = Distance between walls for the acoustic wave (normally, the shell diameter) E = Modulus of elasticity of the tube material
fa = Acoustic frequency or shell space fn = Fundamental tube natural frequency fs = Vortex shedding frequency
fTb = Dominant turbulent buffeting frequency
m = Exponent depending on Cx, δ and the tube layout angle M = Sectional moment of inertia of the tube
n = Exponent depending on Cx, δ and the tube layout angle NBD= Baffle damage number
NCD= Collision damage number
S = Unsupported tube span (for U-bends depends on bend radius) ST = Strouhal number
Vc = Velocity of fluid flowing across a tube or tube bundle VCR = Critical velocity
Vs = Velocity of sound in the shell side fluid (gas or vapor)
Wo = Overall effective weight per unit length of the tube for vibration
Ws = Weight per unit length of the shell side fluid being displaced by the tube vibration Wt = Weight per unit length of the fluid inside the tube
Wtm = Weight per unit length of the empty tube Xc = Crossflow amplitude
δ = Logarithmic damping factor depending on the properties of the shell side fluid (viscosity, density and phase), do, Wo and fn
ρ = Density of the shell side fluid ANALYSIS OF HTRI VIBRATION PRINTOUT
Various HTRI (Heat Transfer Research, Inc.) programs, used for thermal rating and/or design of the shell and tube exchangers, can perform calculations to determine various vibration parameters. The exact equations used are documented in Section IX-F of the HTRI computer programs support volume. This section does not deal with the accuracy of these equations or detailed calculation procedures, but deals mainly with what those parameters mean and how to interpret the program's vibration analysis output.
FLOW INDUCED VIBRATION (Cont)
Vibration parameters are determined at five locations: inlet, central, outlet, shell entrance, and shell outlet regions. To accommodate shell side nozzles, the exchangers will usually have larger inlet and outlet baffle spacings. Vibration analysis considers baffles as supports. Therefore the unsupported length will be different for the inlet zone, the center zone and the outlet zone. In some exchangers, in addition to different unsupported spans, shell side fluid condition may change from one zone to another (i.e., vapor to two-phase to liquid). HTRI is mainly a thermal and hydraulic analysis program and not a detailed mechanical design program. The program logic does not address the need for nozzle reinforcement, which is required for most nozzles. The inlet and outlet spacings calculated by various HTRI programs may be too small for a number of exchangers. It may be necessary to reevaluate various vibration parameters once the actual inlet and outlet zone unsupported tube lengths have been established. Due to the large number of cross-passes or large baffle pitch, the ratio of the actual central spans to program calculated spans is normally very close to 1, and reevaluation of vibration parameters for the center zone is normally not required. Also for the U-bend region, HTRI unsupported spans should be corrected, if required, as specified in Section IX-K.
Each location should be analyzed separately. Tube vibration can be and is, in most instances, a localized phenomenon. All of the zones should be free of vibration problems. Section IX-F of the HTRI Computer Support Volume gives analysis of the program output. Some important parameters that should be checked are as follows:
1. Length / TEMA Maximum Span - If this number is less than 0.8, the possibility of tube vibration is very low. This is based on the fact that very few, if any, exchangers that exhibited vibration problems had unsupported spans shorter than 80% of the TEMA spans.
2. Shell Acoustic Frequency - This frequency is calculated based on the shell diameter, and, in most instances, shell diameter is larger than the baffle spacing. Check unsupported lengths in the inlet, outlet and central zones to ensure that none exceeds the shell diameter. If any of the unsupported lengths exceed the shell diameter, then recalculate acoustic frequency based on this dimension, as an acoustic wave between baffles is possible. This frequency will be the shell acoustic frequency calculated by the program, multiplied by the ratio of shell diameter to unsupported span.
3. Flow Velocities - Check all velocities, including the Window Parallel Velocity, with the critical velocity. If all the velocities are lower than 80% of the critical velocity, the possibility of vibration is low. If only the window velocity is lower than 80% of the critical velocity, and the ratio of window velocity to crossflow velocity is less than 1.5, the possibility of vibration damage is low.
4. Acoustic Vibration Check - Program output lists the frequency ratios as vortex shedding to acoustic frequency, and turbulent buffeting to acoustic frequency. For both these ratios the acoustic frequencies are based on the shell diameter. If unsupported span exceeds shell diameter in any of the regions, as discussed in Item 2 above, these frequency ratios should be recalculated using the acoustic frequency, based on the unsupported length. If all the ratios are less than 0.8, the possibility of acoustic vibration damage is low.
5. Tube Vibration Check - Program output lists the frequency ratios as vortex shedding to tube natural frequency, and turbulent buffeting frequency to tube natural frequency. If these ratios do not show an asterisk, the possibility of vibration damage is very low. If the frequency ratios show an asterisk, check the amplitudes. If the amplitudes are less than 25% of the tube gap, the possibility of vibration damage is low. Crossflow RHO-V-SQ is printed to indicate energy level in the crossflow stream.
6. Bundle Parameters at Nozzles - As discussed earlier, HTRI programs are mainly thermal rating and hydraulic calculation programs. The calculation of heights under nozzles is approximate. It is recommended that these values be input, based on actual tube count, after the design is finalized. This zone must be analyzed separately, using either inlet or outlet spacing as the unsupported span. As a guideline, if the velocity is less than the critical velocity and the values of RHO-V-SQ are less than TEMA recommended maximum values, the possibility of vibration damage is low.
CORRECTIVE MEASURES
In most cases it is possible to analyze an exchanger under various operating conditions for possible vibration problems, and make changes before an exchanger is built. Making changes during design phase is, obviously, the preferred solution. This section lists various corrective measures. Each measure is classified as either a design change or a field change, and also as either a remedy for an acoustic or nonacoustic vibration problem. Sometimes a new bundle must be designed for an existing shell. These cases are considered to be design changes.
1. Reduce Baffle Pitch
This is a design change that will increase the shell side velocities. This increases vortex shedding and turbulent buffeting frequencies, but the tube natural frequency increase will be greater than the increase in vortex shedding and turbulent buffeting frequencies, as seen from the equations presented earlier. This remedy will increase the shell side pressure drop and may not be a practical remedy in many cases. This change is used for nonacoustic vibration problems.
FLOW INDUCED VIBRATION (Cont) 2. Use Double Segmental Baffles
This is a design change that can reduce the shell side velocities and at the same time decrease the unsupported span. This remedy can be used to solve acoustic as well as nonacoustic vibration problems and will, in most cases, result in a lower shell side heat transfer coefficient. When this remedy is used to solve acoustic vibration problems by decreasing shell side velocity, a decrease in shell side heat transfer coefficient will be unavoidable.
3. Increase Tube Pitch
This is a design change. By increasing the tube pitch along with reducing baffle pitch, unsupported span can be reduced. This option is not recommended for solving acoustic vibration problems and, because it will reduce heat transfer area, it is not a good remedy for nonacoustic vibration problems and should be considered only after other remedies have been investigated.
4. Change Shell Type
This is a design change recommended mainly for nonacoustic vibration problems. Use of “J" shell or “H" shell instead of an “E" shell can reduce unsupported span and/or reduce shell velocities.. Use of a “J" shell is a popular option as a “J" shell allows the unsupported span to be reduced to half that of the “E" shell without substantially altering heat transfer and pressure drop. An “H" shell works in almost the same way as a “J" shell for vibration, but has different thermal characteristics; it is not very popular due to the presence of a longitudinal baffle, except for use near isothermal boiling or condensing. Use of an “X" shell allows changing unsupported tube span without any significant effect on heat transfer or pressure drop from other “X" shell designs. Use of “X" shells is normally restricted to very high volumetric flow on the shell side, or for condensers with a small condensing rangeThere are some other shell types used by various vendors, such as “Double J" (two inlets and three outlets, or vice versa). Most of these designs require special considerations and therefore are not discussed further.
5. Increase Tube Thickness
This is mainly a field change and is used for a few localized tubes that have exhibited vibration damage. Increasing tube thickness will increase the natural frequency of the tubes and the tube side pressure drop; therefore it has limited use. 6. Selective Removal of Tubes
This is mainly a field change which will help reduce velocity in restricted flow areas. This is also one of the least desired changes. It not only reduces surface, but also increases shell size bundle bypass and tube side pressure drop. This option can be used for both acoustic and nonacoustic vibration problems, but its main use is for nonacoustic vibration. 7. Change Tube Field Layout Angle
This is a design change. Changing tube field layout angle will change shell side crossflow velocity and Strouhal number. Changing the tube field angle from 45° to 90° is applicable to acoustic vibration problems since the Strouhal number is reduced, resulting in lower vortex shedding frequencies; however, the baffle pitch must be increased to maintain the same crossflow velocity. Changing from 90° to 45° is applicable to non-acoustic vibration problems either due to a lower crossflow velocity which reduces vortex shedding and turbulent buffeting frequencies, or due to a reduced baffle pitch which increases tube natural frequency.
8. Use NTIW (No Tubes in Window) Design
This is a design change used for nonacoustic vibration problems. With this option it is possible to change the unsupported span for the same baffle pitch by the addition of intermediate support plates. Another advantage of this design is that every tube is supported at every baffle or support. The biggest disadvantage is that this design requires a larger shell diameter (normally about 20%). Therefore, this option is not very cost effective.
9. Use Rod Baffle Design
This is a design change. Rod baffle design is a proprietary design. The design procedure is available for users of HTRI programs, but the fabricator must be an authorized licensee. This design is useful for producing low pressure drop, vibration-free designs to handle large volumetric flows on the shell side of an exchanger. Because of the licensing fees, this option will add 10 to 20% to the exchanger cost. An “X" shell or NTIW design should be checked as an alternative before selecting rod baffle design. This design is used mainly to solve nonacoustic vibration problems.
