Some of the important design tips for heat exchanger design are:
1) Fluids that are corrosive, fouling, scaling, high pressure drop or under high pressure are usually placed in tube side
2) Hot, viscous and condensing fluids are typically placed on the shell side
3) Pressure drops are about 1.5 psi (0.1 bar) for boiling/vaporization and 3-10 psi (0.2-0.7 bar) for other services
4) The minimum approach temperature for shell and tube exchangers is about 20°F (10 °C) for fluids and 10°F (5°C) for refrigerants.
5) Cooling tower water is typically available at a maximum temperature of 90°F (30°C) and should be returned to the tower no higher than 115°F (45°C)
6) Double pipe heat exchangers may be a good choice for areas from 100 to 200 ft2 (9.3-18.6 m2)
7) Spiral heat exchangers are often used to slurry interchangers and other services containing solids
8) Plate heat exchanger with gaskets can be used up to 320°F (160°C) and are often used for interchanging duties due to their high efficiencies and ability to "cross" temperatures.
9) For the heat exchanger equation, Q = UAF (LMTD), use F = 0.9 when charts for the LMTD correction factor are not available
10) Shell and Tube heat transfer coefficient for estimation purposes can be found in many reference books
11) Most commonly used tubes are ¾” (19 mm) outer diameter on a 1” triangular spacing at 16 ft (4.9 m) long
12) A 1 ft (300 mm) shell will contain about 100 ft2 (9.3 m2) A 2 ft (600 mm) shell will contain about 400 ft2 (37.2 m2) A 3 ft (900 mm) shell will contain about 1100 ft2 (102 m2)
13) Typical velocities in the tubes should be 3-10 ft/s (1-3 m/s) for liquids and 30-100 ft/s (9-30 m/s) for gases
12.1 DESIGN MARGINE
The following design margins shall be adopted as applicable:
1) Heat exchangers shall be designed with 10% margin on area while using the normal heat duty.
2) While designing heat exchangers, 10% margin on normal duty shall be considered and the exchanger shall be designed with minimum over design factor (say 2% to 5%) on area.
3) If the design duty (10% to 15% over design factor) is used for sizing the heat exchangers, then the exchanger shall be designed with minimum over design factor (say 2% to 5%) on area.
12.2 EVAPORATORS
1) When the boiling point rise is appreciable, the economic number of effects in series with forward feed is 4-6.
2) When the boiling point rise is small, minimum cost is obtained with 8-10 effect in series.
12.3 HEAT EXCHANGER DETAILS
1) TEMA shell types:
2) Baffle cut orientation; parallel, perpendicular and 45° with reference to shell nozzle
3) Tube length : Length up to the tangent point of the outermost tube for U-tubes. The length in the tube sheets should be included. Effective area is excluding the tube sheet.
4) Shell inlet nozzle location
5) U-tube nozzle location
Parallel
0° Perpendicular
90°
45°
U-tube exchanger
Overall tube length Total tube length
code 0
Vertical Horizontal
code code 2
1
Code 0, normal Code 1, in front of Code 2, behind U ACU
TEMA X TEMA G TEMA H
TEMA J21 TEMA J12 TEMA E
TEMA K
6) Shell side inlet nozzle location: Normally it is assumed to be at the tube channel end.
Distance from tangent point to last baffle: Generally the last baffle is placed at the at the tangent point of the U-tube
7) Tube arrangement
8) Impingement device: An impingement plate will be added if the inlet nozzle pV² value is greater than TEMA standards (1500 lb/ft s² or 2232 kg/m s²)
9) Adding a shell in parallel: If pressure drop limitations are not met at the maximum shell diameter permitted, a unit can be added in parallel.
10) Adding a shell in series: If a pure counter flow exchanger has been selected, a shell can be added in series only if the required heat duty can not be achieved at the maximum permitted shell diameter. For multi-tube pass exchangers shells in series can be added if the F-factor is less than 0.7
11) Temperature cross: When the outlet temp of the cold fluid is higher than the outlet temp of hot fluid, it is called the temperature cross. For small exchangers, temperature cross does not have much effect on the type of shells. But for large exchangers shells in series to be used.
12) Baffle design: Baffle window cut should be between 17% to 35%. Baffle spacing should be 20% to 100% of shell ID. For no-tube-in-window exchangers, the ratio of window velocity to cross-flow velocity should normally be 2 to 3. For double segmental baffles (for low pres drop service), baffle spacing should not be too small to avoid ineffective shell side flow patterns.
13) No-tube-in-windows baffle cut design: The baffle cut shall be limited between 15 - 30% of shell diameter
14) Baffle cut out of window: The portion of the baffle which is continuous into the window acts as a sealing strip in the window to force the fluid into the bundle. Continuous baffles should be considered for pull-through floating heads.
Code Y, shell inlet at tube channel end
Code N, shell inlet at tube end
Flow
30°
Pitch Triangular Pitch
45°
Pitch Rotated Square
Pitch
60°
Pitch Rotated Triangular Pitch
90°
Pitch Square Pitch
12.4 HTRI SHELL SIDE FLOW FRACTIONS
This is for shells with single phase fluid flow. The shell flow is broken down into 5 major streams.
1) B-stream : Main cross flow stream through the bundle. B should be at least 60% of the total flow for turbulent flow and 40% for laminar flow. If the baffle spacing is too narrow, more flow will be forced into the A, C and E streams, thereby decreasing the heat transfer.
2) C-stream : Bundle to shell cross flow bypass stream. C should not normally exceed 10%.
Additional sealing strips can be incorporated to decrease this flow fraction. Although this stream is partially effective for heat transfer, a high C-stream flow fraction, especially for pure cross flow shells ("X"), can lead to a severe delta correction to the mean temp difference
3) F-stream : Tube pass partition bypass stream. F should not normally exceed 10%.
Additional seal rods can be incorporated to decrease this flow fraction. Although this stream is partially effective for heat transfer, a high C-stream flow fraction, is not recommended. To block the F-stream flow fraction, program assumes one seal rod of a diameter equal to the tube diameter for each 6 tube rows of cross flow in the exchanger
4) A-stream : Tube-to-baffle hole leakage stream. A-stream is large in narrow baffle spacing where large TEMA clearances apply. However, the A-stream is fairly effective thermally. It will decrease for multi-segmental baffles. Fouling layers might seal this A-stream. The design should be examined by giving a zero tube-to-baffle clearance and the built-up fouling layer thickness for a safe design from a pres drop stand point.
5) E-stream : Baffle-to-shell leakage stream. E-stream is highly ineffective thermally because it does not contact the heat transfer surface; but, since it mixes poorly with the other streams, it can cause distortions of the temp profile. If E-stream is more than 15%, double segmental baffle or other modifications should be tested. If E-stream causes (ST program only) low delta correction factor (< 0.8), corrective action is required.
• F-stream seal rods: Allocate one seal rod of the tube diameter for each 6 rows in cross flow in the exchanger
• Sealing strips: These are metal strip or rod placed between the shell and the bundle which has the effect of forcing the bundle bypass C-stream back into the bundle.
Window
Baffle A
B E
A C
E
A A
F C
C
12.5 SHELL SIDE ANNULAR DISTRIBUTOR
Thermal correction factor "F" = (TUBE) x (BAFFLES) x (F/G) x (HOT/COLD); where TUBE is the uncorrected F-factor based on the no of tube passes, shell style and tempBAFFLES is the correction when there are few baffles. (F/G) is the correction for thermal leakage through longitudinal baffle for TEMA "F","G","H" shells.
(HOT/COLD) is the correction for nonconstant overall h.t.coeff due to diff in the h.t.coeff at the hot and cold ends.
Effective MTD = (LMTD) x F x (DELTA): where DELTA is the profile distortion due to the E- and C-stream leakage
12.6 SHELL SIDE HEAT TRANSFER LIMIT
If there is spare shell side pres drop available, the shell side coefficient can be increased by various methods:
1) Changing the shell type to "F" or "G" can increase the shell side velocity and h.t.coefficient.
But the mean temperature difference may increase.
2) Reducing the tube pitch
3) Decreasing the tube size to accommodate more tubes in a smaller shell 4) Considering finned tubes
5) If DELTA is lower than 0.85, adding a shell in series gives best result, or using sealing strips might improve the performance
12.7 TUBE SIDE HEAT TRANSFER LIMIT
If there is spare tube side pressure drop available, the tube side coefficient can be increased by various methods:
1) Changing the tube length 2) Decreasing tube diameter
3) When in laminar flow switching the tube side fluid to shell side usually results in a more efficient design
4) Increasing the tube pitch, gives less tubes in the given shell ID
Length
Clearance Slot area
12.8 LIQUID DRIVING HEAD FOR THERMOSIPHON EXCHANGERS
1) For horizontal thermosiphon reboilers, most recirculating type feed systems can be designed with kettle type since the height of the outlet piping entering the column is above the liquid level in the column as shown in figure.
2) For thermosiphon reboiler systems for which the reboiler outlet piping enters the distillation column at a height below the liquid level in the trap-out tray as shown in figurethe piping should be checked to ensure that the liquid level does not cover the exit nozzle.
3) For horizontal thermosiphon reboiler designs, the reboiler exit weight fraction should be limited to 0.5 to avoid tube wall dry out.
4) For an effective design, most of the available loop pressure drop is used across the reboiler. As a rule of thumb, this should be around 60 - 70%.However, the inlet and outlet piping design may change this requirement.
Liquid Driving Head (Static Head)
H G
D F
B E
A
C C
Nozzle pipe
Main pipe Main pipe
Header pipe
Vapor+Liquid
Recirculating Feed System
Vapor+
Liquid
Once through Feed System
13. COLUMNS & TOWERS