1. Basic Design Procedure
2. Rating of the Heat Exchanger Design Shell Diameter
3. Insufficient Thermal Rating
4. Excessive Pressure Drop Rating Note: There is always a compromise process between the thermal and Pressure drop ratings
5. Fluids Allocation
6. Cooling water is usually placed in the tubes due to its tendency to corrode carbon steel and to form scale, which is difficult to remove from the exterior tube surfaces. It is generally less expensive to confine a high-pressure stream in the tubes rather than in the shell. In summary
7. Tubing selection – The most frequently used tube sizes are 3/4 and 1 in. – For water service, 3/4 in., 16 BWG tubes are recommended. – For oil (liquid hydrocarbon) service, 3/4 in., 14 BWG tubes are recommended if the fluid is non-fouling. – For fouling fluids, I in., 14 BWG tubes should be used. – Tube lengths typically range from 8 to 30 ft, and sometimes longer depending on the type of construction and the tubing material. – A good value to start with is 16 or 20 ft.
8. Number of tubes
9. Tube layout – Triangular and square layouts are the most common with a pitch of 1.0 in. (for 3/4-in. tubes) or 1.25 in. (for 1-in. tubes). Rotated square pitch is also used but rotated triangular pitch is seldom used. – However, the clearance between tubes is typically the larger of 0.25 in. and 0.25 Do, and with triangular pitch this is not sufficient to allow cleaning lanes between the tube rows. – Although chemical cleaning may be possible, triangular pitch is usually restricted to services with clean shell-side fluids. – Rotated square pitch provides some enhancement in the heat- transfer coefficient (along with higher pressure drop) compared with square pitch, while still providing cleaning lanes between the tubes. – This configuration is especially useful when the shell-side Reynolds number is relatively low (less than about 2000).
10. Tube layout (c) triangle pitch
11. Tube Pitch
12. Tube Length
13. Tube & Header Plate Deformation Thermal expansion of tubes needs to be taken into account for heat exchangers operating at elevated temperatures Tube elongation due to thermal expansion causes: Header plate deformation Shell wall deformation near the header plate Fatigue strength of the tube, header plate and shell joint needs to be considered when using Longer tubes High operating tube side temperatures Cyclic thermal loads
14. Tube passes – For typical low-viscosity process streams, it is highly desirable to maintain fully developed turbulent flow in the tubes since turbulent flow provides the most effective heat transfer. – Once the tube size and number of tubes have been determined, the number of tube passes can be chosen to give an appropriate Reynolds number, i.e., – Except for single-pass exchangers, an even number of tube passes is always used so that the tube-side fluid enters and exits at the same header. – Fluid velocity can also be used as a criterion for setting the number of tube-side passes. It is desirable to maintain the liquid velocity in the tubes in the range of about 3-8 ft/s. Too low a velocity can cause excessive fouling, while a very high velocity can cause erosion of the tube wall.
15. Maximum Recommended Velocities for Water in Heat- Exchanger Tubes – Maximum velocities for water are given in the Table shown below. For liquids other than water, multiply the values from the table by the factor: – For gases flowing in plain carbon steel tubing, the following equation can be used to estimate the maximum velocity: – For tubing materials other than plain carbon steel, assume the maximum velocities are in the same ratio as given in Table
16. Fouling Shell and tubes can handle fouling but it can be reduced by keeping velocities sufficiently high to avoid deposits avoiding stagnant regions where dirt will collect avoiding hot spots where scaling might occur avoiding cold spots where liquids might freeze or where corrosive products may condense for gases Flow-induced vibration Two types - RESONANCE and INSTABILITY Resonance occurs when the natural frequency coincides with a resonant frequency Fluid elastic instability Both depend on span length and velocity - Velocity Velocity Resonance Instability Tubedisplacement
17. Avoiding vibration Inlet support baffles - partial baffles in first few tube rows under the nozzles Double segmental baffles - approximately halve cross flow velocity but also reduce heat transfer coefficients Patent tube-support devices No tubes in the window (with intermediate support baffles) J-Shell - velocity is halved for same baffle spacing as an E shell but decreased heat transfer coefficients (Information about J- and E- types will be given in the next lecture about Shell and Tube HE.
18. Inlet support baffles Double-segmental baffles No tubes in the window - with intermediate support baffles Tubes Windows with no tubes Intermediate baffles Avoiding vibration
19. Baffles and tubesheets – Single segmental baffles are standard and by far the most widely used. In order to provide good flow distribution on the shell side, the spacing between baffles should be between 0.2 and 1.0 shell diameters (but not less than 2 in.). – However, the maximum baffle spacing may be limited by tube support and vibration considerations to less than one shell diameter. – Although baffle spacing and baffle cut are apparently independent parameters, in practice they are highly correlated.
20. Recommended baffle cut, Bc , as a function of baffle spacing. SBC, for single-phase flow; CV, for condensing vapors Baffle Cut, Bc
21. Nozzles – Nozzles can be sized to meet pressure drop limitations and/or to match process piping. – The fluid entering the shell through the inlet nozzle impinges directly on the tube bundle. If the inlet velocity is too high, excessive tube vibration and/or erosion may result. – TEMA specifications to prevent tube erosion are given in terms of the product of density (lbm/ft3) and nozzle velocity (ft/s) squared
22. Sealing Strip • The function of the sealing strips is to reduce the effect of the bundle bypass stream that flows around the outside of the tube bundle. • They are usually thin strips of metal that fit into slots in the baffles and extend outward toward the shell wall to block the bypass flow and force it back into the tube bundle.
23. Sealing Strip • Typically, one pair is used for every four to ten rows of tubes between the baffle tips. Increasing the number of sealing strips tends to increase the shell side heat- transfer coefficient at the expense of a somewhat larger pressure drop. • They are placed in pairs on opposite sides of the baffles running lengthwise along the bundle. Sealing strips are mainly used in floating- head exchangers, where the clearance between the shell and tube bundle is relatively large.
24. Estimated Sizes of the HE • The initial size (surface area) of a heat exchanger can be estimated from • Tm Mean temperature • Tlm,cf Log Mean Temperature difference based on cross flow HE The overall heat transfer coefficient can be estimated from:
25. Typical Film Heat Transfer Coefficients for Shell-and- Tube Heat Exchangers
26. Total number of tubes
27. Total number of tubes Shell Diameter
Coefficient of Friction consideration for Horizontal equipment Saddle Analysis
As per Saudi Aramco Specification SABP-Q-004 page 14 of 55 (Vessel foundation design Guide) Typical Coefficient of friction as follows Cl 4.6.3
1) No Sliding Plate (Steel Support on Concrete):- 0.6
2) Steel Slide Plate:- 0.4
3) Low-friction slide plate Assemblies:- 0.05 to 0.2
Commercial software Saddle analysis options:-
1) PVElites allows to input user defined coefficient of friction for right and left Saddle( Fixed / Slidding Saddle )
2) Compress has only one saddle pop up menu for Saddle data input.( Either fixed or Slidding )
My Query: - An horizontal vessel with which has fixed saddle at its right and slidding saddle at its left ( Teflon-Teflon ). What coefficient of friction will chose for each saddle analysis by using PVElite and Compress?
My opinion:- 1) For PVElite :-For Fixed saddle -0.6 and for Slidding Saddle :- 0.2 2) For Compress:-
0.6 :-
My Justification:- As the end result , both the saddle geometry will be symmetrical. Hence, Lets focus on the fixed saddle design with considering the governing frictional load for fixed saddle ( if we allow it to move but that never be ) In such case , as per Cl 4.6.3 1) will be the very conservative design.
Correct me if am not in the right track and looking forward to hear your valuable advice and suggestions
-*
The vessel on saddles may experience growth (expansion) in its length due to the thermal expansion of the metal shell material. If the vessel's expansion was constrained by fixed saddles then longitudinal compressive stress could be produced in the shell, which is not desirable.
To accommodate the thermal growth, usually one saddle is provided with the ability to "slide". I have my doubts about how effective most such detailing is, but certainly a saddle mounted on Teflon slide bearings, or similar, will be effective. There is no need to put slide bearings under both saddles; in fact, this could be counter-productive as then any longitudinal forces acting on the vessel, as from wind or seismic, would have to be resisted by any attached piping.
The sliding saddle detail will have some amount of residual friction inherent in the construction. What this means is that there will be some minimum "breakaway" force required in order to move (slide) the saddle. The breakaway force is the product: (coefficient of static friction) x (weight on the saddle base)
As one design condition, the anchor bolts must be designed for forces acting parallel to the vessel axis. The loads would arise from wind, seismic (earthquake), and from constrained thermal expansion. The constrained thermal expansion results in the breakaway force from static friction. The design force for the bolts is the largest of these loads.
Once the longitudinal force from wind, seismic, etc exceeds the friction breakaway force then the sliding saddle will indeed "slide" (assuming all is working as intended) and the anchor bolts at the fixed saddle must resist the full longitudinal load.
Up to the breakaway force both saddles will experience the same longitudinal force due to constrained thermal expansion. The fixed saddle will resist this force with the anchor bolts (and shear bars, if any, and friction on the base plate...but usually only the bolts are considered). The sliding saddle will resist this force through the internal friction of the connection.
In the numerical example cited above the coefficient of friction at the sliding saddle changes from 0.2 to 0.6. The result is that the anchor bolts must be designed for 3 times the shear force. Of course the bolt size may be expected to increase.