|
When to Consider By-pass Strips in S & T Bundle
Use a by-pass strip if tubes are removed under a nozzle. Removing tubes leaves an open area where the shell fluid can flow either over or under the bundle.
Consider by-pass strips if the bundle to shell clearance is more than 3/4 inches and the shell fluid is mostly sensible heat transfer.
Especially consider by-pass strips if the shell liquid is a hydrocarbon with an average viscosity greater than 1 centipoise and the tube fluid has a high heat transfer coefficient (example water). In this case, a 5 to 10% increase in heat duty can be achieved by installing by-pass strips.
Lowest Limit of LMTD Correction Factor
What is the lowest LMTD correction facor to be used ?
Here is what several literature sources say:
Heat Exchanger Design Handbook (HEDH)
"F should be kept above 0.75 to 0.80"
Perry's Chemical Engineers' Handbook
"Values of F less than 0.80 (0.75 at the very lowest) are generally unacceptable because
the exchanger configeration chosen is inefficient ..."
In over 50 years of experience, a correction factor of 0.75 is the lowest we have seen a thermal designer use.
Although there was one case where an operating shell-and-tube heat exchanger reflected a lower LMTD correction factor
than 0.75. Another way of looking at the correction factor is to never use a temperature cross of more than
5 degrees F. in a single multi-tube pass shell.
Better Baffle Window Pressure Drop Equation
A new baffle window pressure drop equation has been published in the June 2004 issue of Hydrocarbon Processing. The name of the
article is " More Accurate Exchanger Shell-Side Pressure Drop Calculations". The article can be found on this page with the
subject " Heat Exchanger Articles Published by Dale Gulley". The equation improves the accuracy of the shell side
pressure drop. Refer to the article for more detail. The equation has the following form.
Kp = Pressure loss coefficient for velocity head equation
fi = Friction factor for ideal tube bundle
C1 = Constant based on the type of tube layout
30 deg. triangular 2.2
90 deg. square 3.64
45 deg. sq. rotated 2.29
60 deg. triangular 1.79 estimated
Ncw = Effective number of tube rows crossed in baffle window
D = Distortion factor for ideal fluid stream. It varies with baffle cut
Refer article elsewhere on this site for equation
Sl = Total of leakage areas (in2)
Sw = Net flow area in baffle window (in2)
EXAMPLE
This is taken from the first experimental case in "A Reappraisal of Shellside Flow in Heat Exchangers HTD-Vol. 36".
Average flow of 990,000 lb/hr with a density of 62.4 lb/ft3 is flowing through a 13.25 ID nozzle. The shell ID is 23.25 in. and the
OTL is 22.375 in. The tube OD is 0.75 in. on a tube pitch of 0.9375 in. with 30 degree layout. There are 7 baffles and 26% baffle cut
The following are taken from a tip in this section named "Improve Shell Side Pressure Drop Calculations"
fi = 0.1025
Ncw = 5.96
Sl = 11.0
Sw = 44.47
C1 for a 30 degree layout is 2.2
D = 1 since the fractional baffle cut is 26%
Kp = 0.1025 ( (2.2 x5.96) -2(11/44.47)2) )
Kp = 1.33
Gw = (990000 x 0.04)/44.47 = 890.5 (#/sqft-sec)
ΔPw = Kp x 0.000108 x Gw2/ρ
ΔPw = 1.33 x 0.000108 x (890.5)2 x 7/62.4
ΔPw = 12.78
Weighted MTD
If there is more than a slight curvature in the heat release curve, things get more complicated. Then a step-wise method using local temperatures and local heat transfer coefficients are used to calculate the heat exchanger area. The question is what do you report as the MTD and the correction factor ? There is a reference in TEMA in the temperature relations section T-3.2 that refers to a weighted MTD. The article mentioned was published by Dale Gulley in the June 1966 issue of Hydrocarbon Processing. The article shows how to calculate a weighted MTD and it's correction factor if one is required.
Types of Steam Condensers
Small steam condensers use shell-and-tube heat exchangers while large steam condensers use surface condensers.
A conventional "E" type shell is used when the steam condensing temperature is above approximately 120 F. For lower temperatures, a "X" type shell can be used. There reaches a point where the size or operating pressure requires a surface condenser.
Zone Those Condensers
The heat transfer and pressure drop of a condenser usually should be zoned. A typical heat exchanger that condenses 100% of the vapor will go through 2 or 3 different flow pattern zones before the flow becomes a liquid. There is better accuracy if the flow patterns are determined and their individualistic equations are used.
Using Turbulators for Tube Side Laminar Flow
If the flow inside the tubes of a heat exchanger is in laminar or viscous flow, take a look at enhancing the
heat transfer. One simple and inexpensive device is the twisted-tape insert. Using twisted-tape inserts for
laminar flow in new heat exchangers results in cost savings and smaller heat exchangers. Twisted-tape inserts
can be used in existing heat exchangers to make a significant increase in capacity. The amount of increase in
heat exchanged depends on whether the increase in pressure drop can be tolerated. If there is not a pressure drop
limitation, there can be as much as a 50% increase in capacity,
Here are the recommended guide lines for using twisted tape inserts:
1. Pressure drop in the tube side without inserts is less than 3 to 4 PSI.
2. Minimum fluid viscosity of 2 centipoise unless there is a very low velocity
3. Use a minimum tube diameter of 5/8” for .001 fouling. Use a minimum of 1” diameter for .0015 fouling. It is
not recommended
to use turbulators in a service
that has a fouling factor greater than .0015.
These guide lines for tube diameter are due to fouling being more of a problem with turbulators in small tubes.
Suggestions to Reduce Fan Drive Noise
The most effective solution is to reduce the fan speed by changing the drive ratio between the fan and the motor. Other suggestions are to reduce the fan blade angle or change to a fan with more blades.
Use Superficial velocities to Calculate Best Heat Transfer Flow Pattern
The best heat transfer occurs when there is an annular flow pattern. Then there is a relatively thin liquid film and little vapor in contact with the heat transfer surface. How do you tell if the flow is annular? It will be when the superficial gas velocity is above the following value:
If the superficial liquid velocity is below 0.30 ft/sec.
VgMax = 72 -148VL +100 VL2.
If the superficial liquid velocity is above 0.30 ft/sec.
VgMax = 28.1 +28VL +1.12VL2.
where VL is the superficial liquid velocity, ft/sec.
Heat Exchanger Articles Published by Dale Gulley
1. "More Accurate Exchanger Shell-and-Tube Pressure Drop Calculations", Hydrocarbon Processing, June 2004
2. "Troubleshooting Shell-and-Tube Heat Exchangers", Hydrocarbon Processing, September 1996
3. "Computers help Design Tubesheets", The Oil & Gas Journal, May 20,1974
4. "Computer Programs aid Design Work", The Oil & Gas Journal, Jan. 13,1969
5. "How to Calculate Weighted MTD's", Petroleum Refiner, July 1966
6. "How to Figure True Temperature Difference in Shell-and-Tube Exchangers", The Oil & Gas Journal, Sept. 14,
1964
7. "Make This Correction Factor Chart to Find Divided Flow Exchanger MTD", Petro/Chem Engineer, July 1962
8. "Use Computers to Select Exchangers", Petroleum Refiner, July 1960
Copies of the articles are available in .pdf format
Maximum Exhaust Gas Temperature for Steel Fin Tubes
Here is an approximation of the maximum exhaust temperature for steel fin tubes when generating steam. Otherwise the fins would need to be the more expensive 409SS material. This is based on the typical 2 inch O.D. tubing with 1 inch fins and 6 to 7 fins/inch.
MaxTg = 1090 -0.23Btemp
where MaxTg = maximum gas temperature in F.
Btemp = water boiling temperature in F.
When do Bare Tubes become More Efficient Than Fin Tubes?
If the inside heat transfer coefficient beomes too low, fin tubes can become inefficient. This can be the case in heavy oil coolers.If it is expected that the heat transfer coefficient is below approximately 20 Btu/hr-ft2-F, investigate both bare and fin tubes.
Maximum Motor HP for a Fan
Adding more HP to a fan will only work up to a point. The fan efficiency reaches a peak. Then increasing the HP will produce no more air. An estimate for this HP is:
Max HP = 17 +8.4(Fan Diam -3.5)
This is for fan diameters greater than 3.5 ft.
LMTD Correction Factor Charts for TEMA G and J Shell Types
There are LMTD correction factor charts in TEMA for a single type G shell and two in series of type J shells. For charts of more shells in series, contact support@gulleyassociates.com.
Shell Side Impingement Protection
There may be tube vibration or errosion if the shell-side fluid velocity is above a maximum value, These values can be found in TEMA section RCB-4.61 & 4.62. In the eighth edition the maximum values can be found on page 35.
The most common impingement protection is a plate baffle that is slighty above the tube bundle. But this type of protection has some drawbacks. It has a relatively higher pressure drop than most other methods and the tubes on the first several rows tend to vibrate. Other types of impingement protection are:
1. Plate within a nozzle enlarger
2. Solid rods instead of tubes for the first 2 or 3 rows.
3. Snap-on tube protectors on top of the tubes in the firrst 2 or 3 rows
4. Small angle iron types setting on top of the tubes in the firrst 2 or 3 rows
5. Vapor belt
Kettle Reboiler - Problem Shell Nozzle Arrangement
Sometimes you see kettle reboilers where the inlet nozzle is directly under the outlet vapor nozzle. This arrangement creates extra turbulence under the vapor nozzle which effects the amount of liquid entrainment in the outlet vapor. It is safer to use the conventional nozzle arrangement where the inlet is some lateral distance away unless a demister pad is used..
Another problem with the vertical nozzle arrangement is when the kettle bundle is relatively long and there is a single pair of nozzles. Then there is not good flow distribution. The boiling zones near the ends of the bundle will have lower fluid circulation rates and lower heat transfer.
Minimum Recirculation Rate in Thermosyphon Reboilers
When does a recirculalation rate become too low (high % vaporization) ? When this happens the tube wall is no longer wet and the heat transfer diminishes. The guidelines in the literature show the lowest permissable recirculation rates give from 25 to 40% vaporization for hydrocarbons. It has been observed that this threashold is when the outlet two phase density (volume basis) is below 1.0 lb/cu-ft. Nearly all thermosyphons have outlet densities above this value.
Features of a New S & T bundle to Replace Bundle That Vibrated
1. If possible, design for lower cross flow velocity with special baffles.
2. Make sure that impingement plate is very secure.
3. Use a tube/baffle clearance of 1/64.
4. Use thicker baffles.
5. Use closer baffle/shell clearance.
6. Use thicker tubes.
7. If tubes are lowfins, have the tubing bare where it goes through the baffles.
When
When Will Exchangers With Lowfins be More Economical Than Exchangers With Bare Tubes?
1. If the shell size is a least 2 sizes smaller (pipe size).
2. If the shell size is at least 14" O.D.
3. If there are fewer exchangers. when using lowfins
4. When (total shell resistance/total tube resistance) is greater than 0.4
Calculate Shell Nozzle Pressure Drop
Shell nozzle pressure drop calculation methods are difficult to find in the open literature. The nozzle pressure drops
are difficult to predict accurately. There is a complex flow pattern of a tube matrix, bundle bypassing and recirculation.
Because of this, it is possible to have pressure loss coefficients greater than the customary 1.5 velocity heads for
sharp edge expansion/contraction edges.
If the bundle entrance area is equal to or greater than the inlet nozzle flow area, use a pressure loss coefficient of 1.0.
If the bundle exit area is equal to or greater than the exit nozzle area, us a pressure loss coefficient of 0.58. There
are indications that it should be larger.The following procedure is for the situation where the nozzle flow area is
greater than the entrance or exit area and the bundles do not have an impingement plate. If there is an
impingement plate, there will have to be added a turning loss to the calculation below.If the two shell side nozzles
are not the same size, calculate the inlet pressure drop and take 2/3 of it and make a separate calculated pressure drop
for the outlet and take 1/3 of it.
Shell Entrance or Exit Area
1. Calculate the bundle bypass area Sb = π x Dn x h
2. Calculate the slot area Aslot = 0.7854Dn2 (Pt -Dt)/(F2 x Pt)
3. Calculate the shell entrance and exit area.(As)
As = Sb + Aslot
(refer TEMA RGP-RCB-4.621 & 4.622)
4. Calculate ratio of Sb to total area FR = Sb/As
5. Kn = 0.65 +2.14 (FR -0.4)
(minimum Kn = 0.8, maximum = 1.8)
6. ΔPn = Kn x .000108Vs2 x density
(ΔPn = total of both nozzles)
where
ΔPn = Total nozzle pressure drop (lb/ft2)
Dn = Nozzle ID (in)
Ds = Shell ID (in)
Dt = Tube outside diameter (in)
F2 = 0.707 for 45 degree pitch, all others use 1.0
h = 0.5(Ds-OTL)(in)
Kn = Pressure loss coefficient
OTL= Outer tube limit diameter (in)
Pt = Tube center to center pitch (in)
Vs = velocity in the entrance/exit area (ft/sec)
EXAMPLE
990,000 lb/hr with a density of 62.4 lb/ft3 is flowing through a 13.25 in. ID nozzle. The shell ID is 23.25 in. and the
OTL is 22.375 in. The tube OD is 0.75 in. on a tube pitch of 0.9375 in. with 30 degree layout.
Calculate Sb
h = 0.50(23.25-22.375)= 0.4375
Sb = π x 13.25 x 0.4375 = 18.23
Calculate Aslot
Aslot = 0.7854(13.252) (0.9375-0.75)/(1.00 x .9375)= 27.58
Calculate total area As
As = Sb + Aslot = 18.23 + 27.58 = 45.81
Calculate FR
FR = 18.23/45.81 = 0.4
Calculate Kn
Kn = 0.65 +2.14(0.4 -0.4) = 0.65 (use minimum 0.8)
Calculate nozzle pressure drop
Vs = (990000 x 0.04)/(45.81 x 62.4)= 13.85
ΔPn = 0.8 x 0.000108 x 13.852 x 62.4 = 1.03 psi
Comment - Using 1.5 total pressure loss coefficient and the nozzle flow area gives only 0.21 PSI
Maximum Velocity Inside Tubes
An estimate for maximum tube velocity inside steel tubes
Vmax = 80./sqrt(density)
where Vmax = maximum velocity (ft/sec)
density = lb/cu ft.
Quick Estimate for Reflux Condenser LMTD in Air-cooler
This type of service has steam condensing out from a non-condensible gas which is mostly CO2. The condensing curve has a hump which will give a LMTD higher than one calculated from a straight line condensing plot. An equation which makes a quick estimate for the LMTD is: Standard LMTD x Factor
For outlet process temperatures below 153.5 F.
Then LMTD Factor = 1.4 -0.0092(T -110)
Where T = outlet temperature and air inlet temperature is 100 F.
Avoid These Fluids When Using Lowfin Tubing
When a fluid has a high surface tension, the fluid doesn't readily flow from the gap between the fins. This lowers the heat transfer. The types of fluids that are to be avoided are those whose surface tension is above 30 to 40 dynes/cm. This includes such fluids as condensing steam, aqueous solutions with a high % of water, amines and glycols.
Why Did the Performance Decline in a TEMA F,G or H Type Shell?
Has performance declined after the bundle has been pulled and later installed back in the shell? If the longitudinal long baffle is sealed on the sides with leaf seals, they are probably the problem. These thin flexible strips should be positioned so that they form a concave pattern and flex upward. Then when the shell fluid puts pressure on the leafs, they will press harder against the sides of the shell. If there is too much pressure or if the bundle is installed upside down, the leafs will flex downward and the shell fluid will bypass the bundle. Another possibility is that the leaf seals were damaged when the bundle was out of the shell.
Calculate When to Use Seal Bars on a Bundle to Increase Heat Transfer
One of the fluid by-pass streams that lowers the shell-side heat transfer is the stream that flows around the
bundle. To evaluate, calculate a heat transfer varable named FSBP. It is the ratio of the bundle by-pass area to
the crossflow area. The by-pass ratio for triangular and square tube pitch is normally:
FSBP =
Bs(Ds -OTL)
Bs(Ds-OTL) +Bs(OTL-Do)(P-Do)/P
Where shell ID = inside diameter of shell (in)
Bs = distance between baffles (in)
Do = tube OD (in)
Ds = shell ID (in)
OTL = outer diameter of the tube bundle (in)
P = tube spacing (in)
If FSBP is more that 0.15, then seal bars are needed.
Which Stream Goes Inside Tubes for Gas/Gas Exchangers?
If the heat exchanger is counter-current flow, the steam with the highest factor as calculated below goes inside the tubes:
Factor = (flow)2 / density
You can also use this if the molecular weight and temperature are about the same on both sides:
Factor = (flow)2 / pressure
Where:
Density (#/cu. ft.)
Flow (#/hr)
Pressure (Psia)
Allowable Shell Side Pressure Drop if a Multi-leaf (a.k.a. Lamaflex) Long Baffle is Used
Four thin(.oo8") stainless strips are normally used to seal the sides of the long baffle. Because of their flexibility they are not able to withstand large shell side pressure drops. It is best to limit the pressure drop to 5 psi with 7.5 psi being the maximum.
When to Use Bare Tubes in Waste Heat Boilers
Use bare tubes if the bundle is quite small or the gas temperature is greater than 1350 to 1400 F.
Effect of 1st Tube Rows on Shell Nozzle Pressure Drop
Usually when shell-and-tube heat exchangers are designed, the tube layout is made so that the shell entrance area is
approximately equal to the shell nozzle flow area. The average distance to the 1st tube row is Dn/4 where Dn is the
inside diameter of the shell nozzle. In this case the pressure loss coefficient is 1.0 for the pressure drop calculation
for the shell nozzle entrance.
If the shell nozzles are greater than 2" and tubes are not omitted from the tube layout, the nozzle entrance pressure drop
can be significantly higher than the normal calculation based on the nozzle flow area. In a case of a 6" shell nozzle and
no tubes were omitted in a BEM type heat exchanger, the pressure drop was 3 times higher than that calculated with just
the nozzle flow area. For more information you can refer to the tip "Calculate Shell Nozzle Pressure Drop" on this page.
Estimate - Latent Heat of Hydrocarbons
An equation from the Bureau of Standards Miscellaneous Publication No. 97 can be used when the Specific Gravity is greater than 0.67 or less than 0.93. It is:
Lat heat = (111 -.09T)/SG60
Where:
Lat heat = Latent Heat in Btu/lb
T = temperature in F.
SG60 = Specific gravity @ 60 F
For hdrocarbons below a Specific Gravity of 0.67 and pressures below 50 psia. Use
Lat heat = 172 -0.195T
Kettle - Solutions to Liquid Carryover
1. If there is excess surface and the liquid is clean - lower the liquid level.
2. Add a mist eliminator
3. Add more vapor outlet nozzles.
L/D Equation For Heat Transfer Coefficient Inside Tubing
For Reynolds numbers below 10,000 there is a L/D effect on the heat transfer coefficient inside tubing. If you use the full tube length for L, you may be too conservative. There will be turbulation at the tube entrance before laminar flow is fully developed. The turbulent length needs to be subtracted from the full tube length.
Use the following for tube sizes 1.0 inch or less.
L = Tube Length -0.0027DiRe
where L = variable to use in L/D expression, ft
Tube Length, ft
Di = tube I.D., in
Re = Reynolds number
Estimate - Pool Boiling Heat Transfer Coefficient for Hydrocarbons
Boil h = 22(Δt)1.25
Δt is tube wall temperature - liquid temperature
Where Boil h = heat transfer coefficient, Btu/(hr)(ft2)(OF)
t = temperature, OF
Estimate - Condensing Heat Transfer Coefficient for Hydrocarbons Inside Tubing
Cond h = 4.15W0.8
Where Cond h = condensing heat transfer coefficient, Btu/(hr)(ft2)(OF)
W = lbs/hr/tube
Sulfur Condenser - Tube Velocity Limits
For good operation of a sulfur condenser the design velocities inside the tubes should be within certain limits. The velocity range is between 1.5 and 6.0 lb/sq ft-sec.
Below this range there will be slugging. Above this range there is sulfur fogging.
Undersurfaced S&T Quote
What to ask the vendor if his quote is undersurfaced.
1. Are there seal strips? If so, how many?
2. What tube hole clearance was used in the baffles
Minimum Number Shells in Series
The conventional way of determining the minimum shells in series when it is non-counterflow is to calculate 2 variables and use the TEMA curves. Then chose the chart with an LMTD correction factor that exceeds a value greater than 0.80.
Another method is to use the temperature cross in the proposed operating temperatures. The amount of temperature cross determines the minimum number of shells in series. Temperature cross is the amount the cold side outlet temperature
exceeds the hot side outlet temperature. The maximum temperature cross for shells in series are as follows:
1 Shell Tcross = 0 - 3 F.
2 Shells in series Tcross = Range
1.8+0.11Range
3 Shells in series Tcross = Range
1.49+0.0032Range
Where:
Tcross = temperature cross (t2 - T2) = F.
Range = temperature drop of hot fluid =F.
Example
The temperature range on the hot side is from 180 to 100F.
The cooling water enters at 85F. If the cooling water outlet temperature is 115F.
Can this be done with 2 shells in series?
Tcross = (180 -100)
1.8+0.11(80)
Tcross = 29.9 F.
Since the actual temperature cross is 15 F. and the maximum is 29.9 F., the minimum number of shells in series is 2.
Improve Shell Side Pressure Drop Calculations
The shell side pressure drop calculation can be improved by better equations for the baffle window and
the nozzle pressure drops. Both of these methods can be found elsewhere on this web page.
The baffle window pressure drop in the open literature is a function only of the number of tubes crossed and the velocity
in the window. It does not take into account a friction factor, type of tube pattern or fluid eddies.
When there are no tubes removed under the shell nozzles and the nozzles are large, using the nozzle flow area can result in wrong
pressure drop calculations.
This is taken from the first experimental case in "A Reappraisal of Shellside Flow in Heat Exchangers HTD-Vol. 36".
Average flow of 990,000 lb/hr with a density of 62.4 lb/ft3 is flowing through a 13.25 ID nozzle. The shell ID is 23.25 in. and the
OTL is 22.375 in. The effective tube length is 11.729 ft. The tube OD is 0.75 in. on a tube pitch of 0.9375 in. with 30 degree layout. There are 7 baffles and 26% baffle cut
From the following the cross flow pressure drop is calculated:
Bs = 17.6 in
fi = 0.1025 Ideal tube bank correlation ( J. Taborek)
Nc = 13.75
Rb = 0.536
Re = 40,249
Rl = 0.615
ΔPc = 6.41 psi
ΔPshell = ΔPc + ΔPw + ΔPn
From other tips: ΔPw = 12.78
ΔPn = 1.03
ΔPshell = 6.41 +12.78 +1.03 = 20.2 psi
Experimental = 20.3 psi
How to Calculate Excess Surface and Overdesign Surface
Excess surface = 100. x Aactual -Acalculated
Acalculated
Where
Aactual = actual heat transfer surface
Acalulated = surface calculated from design overall heat transfer coefficient
To calculate overdesign surface use the clean overall heat transfer coefficient for Acalculated
Minimum Flow Area For Shell Side Inlet Nozzle
For single phase liquids and no impingement plate
Minimum area = Flow(#/hr) x .04
38.73Sq.Root(ρ)
For boilng liquids and no impingement plate
Minimum area = Flow(#/hr) x .04
22.36Sq.Root(ρ)
Where:
Minimum area = minimum nozzle area at shell entry = sq. inches
ρ = density (lb/cu.ft.)
Minimum Velocity Inside Tubing
For Slurries
The minimum velocity for slurries inside tubes for shell-and-tube is 4 ft/sec. This is for a fine material like a catalyst.
For slurries there is a special Reynolds number used for calculating the settling velocity. For more
information on slurries refer to chapter C11 in the piping handbook.
How to Calculate the Performance of Heat Exchangers With Plugged Tubes
After a heat exchanger goes into operation it may develope leaks in the tube walls. The following procedure calculates the new heat load and new overall heat transfer coefficient.
1. Using the actual overall heat transfer coefficient (U). calculate the heat transfer resistances that exclude the tubeside resistance
Rother = 1/U -1/hio
2. Calculate new hio and new surface using usable number of tubes
3. Calculate new U
Unew = 1/(1/hio + Rother)
4. Calculate new heat load from new surface and new U
Estimate - Optimum Flow Velocity for Gas Inside Tubes
Since the design of heat exchangers is a trial and error solution, a good starting point is desired.
Usually the design starts with an estimated overall heat transfer coefficient. If you don't know a good starting
value for this coefficient the equations
presented here give this starting point with simple equations.
In the design of heat exchangers using up the maximum allowable pressure drops gives the highest heat transfer
for single phase fluids.
The equations below estimates the tube velocity(W)for a gas that will meet the maximum allowable pressure drop.
From W you can calculate the tube count or heat transfer coefficient. For a given tube length the following equation
gives the optimum tube velocity for turbulent flow. Gases will be in turbulent flow more than 99% of the time.
If your calculated tubeside velocity is below what the following equation calculates, you need more tube travel
where tube travel is in the form of number of tube passes or total tube length(s) for countercurrent flow.
These equations can be used for two phase flow as long as the two phase viscosity is less than 0.015 cp,
For 3/4 inch tubes with 0.06 tube wall
W = 1600(ΔPρ/L)0.555
For 1.0 inch tubes with 0.06 tube wall
W = 3500(ΔPρ/L)0.555
Where:
L = total tube lengths in ft.
(Add 8 x tube ID for turning losses for each tube pass)
W = lb/hr/tube
ΔP = allowable pressure drop inside tubes in psi (deduct 15% for nozzle pressure drops)
ρ = density in lb/cu.ft.
Example
Use 3/4 inch tubes and 16 foot tubes. The maximum allowable pressure drop inside the tubes is 7 psi (after nozzle deduction) and
the gas density is 2.66 lb/cu.ft. The tube side flow is 195,000 lb/hr. What should be the starting tube count?
Solution
W = 1600(7 x 2.66/(16+5))0.555
W = 1500 lb/hr/tube
Tube count = 195,000/1500 = 130
For a shell-and-tube heat exchanger there is a tip on this site that calculates the shell diameter when given the tube count. It has
the description "Calculate S & T diameter from number of tubes".
Estimate - Critical Heat Flux For Propane Chillers
A simple equation is presented for a kettle reboiler. It is conservative for very small bundles. The crital heat flux depends
on the geometry of the bundle. The following estimate is based on 3/4 inch tubes on
15/16 inch pitch. It is actually good
for any tube diameter with a tube pitch/tube diameter ratio of 1.25 and triangular tube pitch. A boiling temperature of -30 F. is assumed
for the propane.
CHF = 32500
Ds0.25
where
CHF = crital heat flux in Btu/(hr)(ft)2
Ds = shell bundle diameter in inches
example
What is the critical heat flux for a 41 inch diameter bundle?
CHF = 32500
(41)
0.25
CHF = 12,850
Longitudinal Baffle Heat Conduction Cures
With a longitudinal baffle and a long temperature range there can be a problem with heat conduction through
the longitudinal baffle. There will be a loss of thermal efficiency due to the heat conduction. The longitudinal
baffle can be fabricated in one of two ways.
1. Leaving an small enclosed air gap between two longitudinal baffles.
2. Spray an insulating material like Ryton on the longitudinal baffle.
Equations for how baffle cuts are expressed
To convert from diameter cut to area cut
% area cut = -4.3 +0.816Dcut + 0.00563Dcut2
To convert from area cut to diameter cut
% diameter cut = 5.6 +1.06Acut -0.00367Acut2
Where baffle cuts are expressed as a percent
Fouling factors for water(hr-ft2-F/Btu)
0.0005 steam,steam condensate,engine jacket water
0.0010 boiler feed water
0.0015 clean water,moutain water,etc.
0.0020 normal cooling tower water
For cooling water when velocity is 3 -8 ft/sec
Fouling = 0.025/V1.67
Where V =ft/sec
Fouling Factors for Liquid Hydrocarbons(hr-ft2-F/Btu)
0.0010 If sp. gravity At 60F less than 0.80, lube oil and heating oils
0.0020 If sp. gravity At 60F 0.80 -0.87
0.0030 If sp. gravity At 60F 0.87 -1.00
0.0050 Heavy fuel oils
Kettle Reboiler - Supports or Baffles?
For kettle reboilers use segmental baffles instead of full supports if shell fouling factor is greater Than 0.002(hr-ft2-F/Btu)
Design Temperatures of Carbon Steel and Low Alloy Tubes and Tubesheets
Use the higher of the shell-side and tube-side design temperatures up to 650 F. At higher design temperatures
use the arithmetic average of the 2 design temperatures.
Viscous Flow - Use More Pressure Drop Than Usual
High viscosity fluids can have a problem achieving the design heat transfer. The
fluids are usually petroleum based and have an API of 20 or less.
Low pressure drops can cause maldistribution of the tubeside flow which in turn
reduces the heat transfer. That is why you can see allowable pressure drops 2 or 3
times higher than usual. There is a method by A.C. Mueller for calculating this
minimum allowable pressure drop. Another thing that can help is to use more tube
passes and shorter tubes than normal. Also the fluid could be placed in the shell
side if cleanig isn't a problem.
Design Temperatures of Nonferrous Tubes and Tubesheets
Water in the shell-side
Use the arithmetic average of the shell-side and tube-side design temperatures.
Water in the tube-side
Use the higher of the tube-side design temperature or tube-side outlet temperature + 1/3 of the LMTD.
Vertical Thermosyphon-Design for a Smaller Liquid Preheat Zone
At low operating pressures there will be a sensible heat liquid zone with relatively low heat transfer. This is
caused by the fact that a small pressure change will cause a large increase in the boiling point. There has been
a case where 90% of the tube length was in the sub-cooled phase. What can you change that will decrease the size
of the liquid preheat zone and increase the overall heat transfer?
One answer is to evaluate the piping system above the top tubesheet. In order to make an evaluation check
the pressure drop at the outlet. There is on this section of the website equations to calculate the pressure drop
of a nozzle that is at right angle to the top channel. Most vertical thermosyphons have the outlet nozzle
at right angles to the top channel. There may be a simple change of enlarging the outlet nozzle that would be the
cure. But there needs to be a check to make sure the nozzle and connecting piping are not so large that there is
liquid slip. If enlarging the right angle nozzle and piping is not the answer then there are other configerations
that will use less outlet pressure drop. Next the pressure drop of using a B type channel with a long radius ell
could be tried. If this doesn't do it, try a mitered channel design.
Another solution to the problem is to investigate inserts such as swisted tape, wire matrix , or helically
coiled.
Vertical Thermosyphon-Calculate Pressure Drop at The Outlet Nozzle
A rule of thumb is that the pressure drop at the outlet nozzle should not be greater than 30% of the total
static head. There is another tip in this boiling
section about choking the flow with a small outlet nozzle. The inside
flow area of the outlet nozzle should be the same or greater than the total flow area insde the tubing. For
a channel with a side outlet the pressure drop is composed of a turning loss and a contraction loss
The following equations calculate the pressure drop at the outlet. The pressure drop for expansion into
the channel is not included here but is with the tube pressure drop.
Ktr = ___1______
Ds0.3
(If Ktr less than 0.40, use 0.40)
Kc = 0.5(1 - (No/Ds)2)
KT = Ktr + Kc
ΔPn = KT = 0.000108 x Vn2 x ρtp
Where:
Ds = Top channel ID (inches)
Ktr = pressure loss coefficient for turning loss
Kc = pressure loss coefficient for contraction into nozzle
KT = total pressure loss coefficient
No = Outlet nozzle ID (inches)
Vn = velocity thru nozzle (ft/sec)
ρtp = two-phase density (lb/ft3)
ΔPn = pressure drop thru channel and outlet nozzle (Psi)
Estimate - Hydrocarbon Gas Heat Transfer Coefficient in Shell Side
Its difficult to estimate a gas heat transfer coefficient in the shell side because of the many variables.
The following will give you a value within 25%.
Ho = 430.Cp(ΔP/L x ρ)1/3
where
Cp = specific heat (Btu/lb-F)
L = tube length (ft)
ΔP = shell side pressure drop (Psi)
(subtract nozzle losses)
Best Design Feature to Prevent Bundle Vibration
In designing a shell-and-tube heat exchanger, use a 30o triangular tube pitch
if possible. This will lower the vortex shedding frequency which is a direct function of something called a
Strouhal number. The Strouhal number is a constant composed of the vortex shedding frequency, shell side velocity and
tube OD.
The 30o triangular tube pitch has a significantly lower Strouhal number than
other tube pitch types. Using Barrington as a source, for 3/4 inch tubes on 30o triangular tube pitch the Strouhal number
is 0.21. But for 60o rotated triangular tube pitch the Strouhal number is 0.81.
Maldistribution of Tube-side Flow
Too small of a tubeside inlet nozzle can cause maldistribution of the fluid into the tubes and cause lower heat transfer. This will cause the tubeside fluid
to jet into a relatively small amount of tubes. This lowers the flow to a majority of the tubes. To improve the flow distribution you
could install a larger nozzle, an enlarger or a distribution plate.
U-tube Bend Area Equation
To calculate the amount of surface in the u-bends would be difficult if you had to calculate the surface of each tube row and
then add up all the rows for a total. The following is an equation that gives the u-bend surface. It is based on the typical bend radius
of 1.5 x tube OD
U-bend area = (Count1.44 Do P0.78) /73.5
where:
Count = total number of tubes
Do = Tube OD (in.)
P = Tube pitch (in)
Minimum Boiling Temperature Difference
If the boiling temperature is too low there is not full boiling. The following give the boiling temperature difference
where full boiling decreases to partial boiling. This temperature difference depends upon the nature of the fluid being
nucleated.
where:
Water Δ T = 8.0F.
Lt. HC Δ T = 5.0F.
(lowfins Δ T = 1.8F.)
Gas Turbine Exhaust Boiler - Calculation Start
Gas turbine exhaust provides an economic solution to waste heat recovery. One system consists of vertical rows
of fin tubes with a steam drum on top.
The following is the typical number of tube rows to start the calculation of a gas turbine heat recovery system:
Economizer 3 - 4 rows
Boiler
8 - 10 rows
Superheater 2 - 3 rows
Copyright © 2002-2012 Gulley & Associates
|
