Without details (flow rate, density, tube size, number of tubes, number of passes, etc.) it is not possible to calculate velocity; however, there are recommended upper and lower limits for such heat exchangers that can be found in references available free online, for example the TEMA manual (TEMA-2019.pdf).
This is a classical heat exchanger design problem: How to determine the performance without knowing the geometry. In Hewitt (ed): Heat Exchanger Design Handbook, Begell House (2008) there is a fairly good 'recipe' on how to approach this for shell-and-tube heat exchangers. (I believe you'll find it in vol.3)
To determine the speed of the heat carrier in the heat exchanger, it is not necessary to know the shape of the heat exchange surface. It is enough to know the cross-sectional area of the channels and the flow rate of the coolant.
Dear Valery Gorobets I don't have any information about area of the channels, but I have the flow rate
If you choose a specific design of a shell-and-wire heat exchanger, you must have the dimensions of the heat exchanger. Based on these dimensions and general recommendations for line-up of tube bundles, you can find the probable cross-section of the channels in this heat exchanger.
How do you calculate viscosity of fluid?
📷
There are several formulas and equations to calculate viscosity, the most common of which is Viscosity = (2 x (ball density – liquid density) x g x a^2) ÷ (9 x v), where g = acceleration due to gravity = 9.8 m/s^2, a = radius of ball bearing, and v = velocity of ball bearing through liquid.
Hello Dr. Suresh Ahuja, thank you for answering
Actually, I meant Velocity
Shell-and-Tube Heat Exchangers
R. Shankar Subramanian
Department of Chemical and Biomolecular Engineering
Clarkson University
Shell-and-tube heat exchangers are used widely in the chemical process industries, especially in refineries, because of the numerous advantages they offer over other types of heat exchangers. A lot of information is available regarding their design and construction. The present notes are intended only to serve as a brief introduction.
For detailed information about analyzing and designing shell-and-tube heat exchangers, consult “The Chemical Engineers’ Handbook” (http://www.knovel.com/knovel2/Toc.jsp?BookID=48 ) (Chapter 11) or any of a variety of sources on heat exchanger design. Mechanical standards for shell-and-tube heat exchangers are set by TEMA (Tubular Exchangers Manufacturers Association) and these supplement the ASME code for such heat exchangers. API (American Petroleum Institute) Standard 660 supplements both of these standards, and chemical and petroleum companies also have their own internal standards in addition.
Advantages
Here are the main advantages of shell-and-tube heat exchangers (Thanks to Professor Ross Taylor for this list).
1. Condensation or boiling heat transfer can be accommodated in either the tubes or the shell, and the orientation can be horizontal or vertical. You may want to check out the orientation of the heat exchanger in our laboratory. Of course, single phases can be handled as well.
2. The pressures and pressure drops can be varied over a wide range.
3. Thermal stresses can be accommodated inexpensively.
4. There is substantial flexibility regarding materials of construction to accommodate corrosion and other concerns. The shell and the tubes can be made of different materials.
5. Extended heat transfer surfaces (fins) can be used to enhance heat transfer.
6. Cleaning and repair are relatively straightforward, because the equipment can be dismantled for this purpose.
Basic considerations
The tube side is used for the fluid that is more likely to foul the walls, or more corrosive, or for the fluid with the higher pressure (less costly). Cleaning of the inside of the tubes is easier than cleaning the outside. When a gas or vapor is used as a heat exchange fluid, it is typically
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introduced on the shell side. Also, high viscosity liquids, for which the pressure drop for flow through the tubes might be prohibitively large, can be introduced on the shell side.
The most common material of construction is carbon steel. Other materials such as stainless steel or copper are used when needed, and the choice is dictated by corrosion concerns as well as mechanical strength requirements. Expansion joints are used to accommodate differential thermal expansion of dissimilar materials.
Heat transfer aspects
The starting point of any heat transfer calculation is the overall energy balance and the rate equation. Assuming only sensible heat is transferred, we can write the heat duty Q as follows.
()(),,,,,,hotphothotinhotoutcoldpcoldcoldoutcoldinQmCTTmCTT=−=−
lmQUAFT=Δ
The various symbols in these equations have their usual meanings. The new symbol F stands for a correction factor that must be used with the log mean temperature difference for a countercurrent heat exchanger to accommodate the fact that the flow of the two streams here is more complicated than simple countercurrent or cocurrent flow. Consider the simplest possible shell-and-tube heat exchanger, called 1-1, which means that there is a single shell “pass” and a single tube “pass.” The sketch schematically illustrates this concept in plan view. Note that the contact is not really countercurrent, because the shell fluid flows across the bank of tubes, and there are baffles on the shell side to assure that the fluid does not bypass the tube bank. The entire bundle of tubes (typically in the hundreds) is illustrated by a single line in the sketch. The baffle cuts are aligned vertically to permit dirt particles settling out of the shell side fluid to be washed away.
1T2T1t2tBaffle
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The convention in shell-and-tube heat exchangers is as follows:
1:T inlet temperature of the shell-side (or hot) fluid
2:T exit temperature of the shell-side (or hot) fluid
1:t inlet temperature of the tube-side (or cold) fluid
2:t exit temperature of the tube-side (or cold) fluid
Thus,
()()12211221lnlmTtTtTTtTt−−−Δ=−−
The fraction of the circular area that is open in a baffle is identified by a “percentage cut” and we refer to the types of baffles shown as “segmented” baffles. For the shell side, in evaluating the Reynolds number, we must find the cross-flow velocity across a bundle of tubes that occurs between a pair of baffles, and determine the value of this velocity where the space for the flow of the fluid is the smallest (maximum velocity). For the length scale, the tube outside diameter is employed.
Most shell-and-tube heat exchangers have multiple “passes” to enhance the heat transfer. Here is an example of a 1-2 (1 shell pass and 2 tube passes) heat exchanger.
As you can see, in a 1-2 heat exchanger, the tube-side fluid flows the entire length of the shell, turns around and flows all the way back. It is possible to have more than two tube passes. Multiple shell passes also are possible, but involve fabrication that is more complex and is usually avoided, if possible.
1T2T1t2tBaffle
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Correction factors to be used in the rate equation have been worked out by analysis, subject to a set of simplifying assumptions, for a variety of situations. In the olden days, the formulae for them were considered too cumbersome to use. Therefore graphs were prepared plotting (),FPR, where 2111ttPTt−=− and 1221TTRtt−=− are parameters on which F depends. Figures C4.a-d in Appendix C of the textbook by Mills display such graphs. Nowadays, one can compute these factors quickly with a pocket calculator. Given next are the two common factors.
2122211ln111ln1RPRPRFARAR−+−−−=++−+
()2242211ln2111ln1RPRPRFABRABR−+−−−=++++−+
where ()()221,11ARBPPRPP=−−=−−
The first and second subscripts on the factor Fcorrespond to the number of shell and tube passes, respectively. The simplifying assumptions mentioned in the previous paragraph, given in Perry’s Handbook, are as follows.
1. The heat exchanger is at steady state.
2. The specific heat of each stream remains constant throughout the exchanger.
3. The overall heat transfer coefficient U is constant.
4. All elements of a given fluid stream experience the same thermal history as they pass through the heat exchanger (see footnote in Perry for a discussion regarding the violation of this assumption in shell-and-tube heat exchangers).
5. Heat losses are negligible.
The formula given above for 12F− also applies for one shell pass and 2, 4, (or any multiple of 2) tube passes. Likewise, the formula for 24F− also applies for two shell passes and 4, 8, (or any multiple of 4) tube passes.
In designing heat exchangers, one should avoid the steep portion of the curves of Fversus P, because small errors in estimating P can cause large changes in the value of F. A misleading rule of thumb is that 0.8F≥, but the correct idea is that the region of steep fall-off in the curves should be avoided.
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Heat Transfer Coefficients
The evaluation of the overall heat transfer coefficient is an important part of the thermal design and analysis of a heat exchanger. You’ll find several tables of typical overall heat transfer coefficients in shell-and-tube heat exchangers in Chapter 11 of Perry’s Handbook. The following generic result can be written for the overall heat transfer coefficient oU based on the outside surface area of the tubes, which is the heat transfer surface.
,0,111ooffioolmiiAArRRUhkAhAΔ=++++
In the above equation, oh is the heat transfer coefficient for the fluid flowing in the shell, ih is the heat transfer coefficient for the fluid flowing through the tubes, iA and oA are the inside and outside surface areas of a tube, respectively, and lmA is their log mean. The fouling resistances on a unit area basis are,0fR for the shell side, and ,fiR for the tube side. Accumulated information on fouling resistances can be found in the Standards published by TEMA.
The inside heat transfer coefficient ih can be evaluated using the standard approach for predicting heat transfer in flow through tubes, including applying a viscosity correction where possible. Typically, turbulent flow can be expected, and a good design would aim to arrange for turbulent flow, because of the substantial enhancement in heat transfer provided by eddy transport. Predicting the shell-side heat transfer coefficient oh is more involved, because the flow passage is not simple, even in the absence of baffles. The presence of baffles needs to be taken into account in calculating the fluid velocity across the tube bank. Heat transfer correlations for flow through tube banks are used, such as those given in the book by Holman (1). These correlations assume flow normal to the long axes of a set of tubes placed in a geometrical array. The correlation given in Holman’s book is
1/3RePrnoohDNuCk==
The Reynolds number maxReoDVρμ=, where oD is the outside diameter of a tube. maxV is the “maximum” velocity of the fluid through the tube bank. To find it, first, the cross-flow area must be evaluated. This is given as clearanceCrossflowareaShellIDBafflespacingpitch=×× where the clearance l and pitch nS (normal to the flow direction) are illustrated in the sketch on the next page for tubes in a square pitch.
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The clearance nolSD=−. When the volumetric flow rate of the shell-side fluid is divided by the cross-flow area defined here, it yields the “maximum velocity” through the tube bank, maxV. The symbols ,,andkρμ represent the thermal conductivity, density, and viscosity of the shell-side fluid, respectively, and all the properties should be evaluated at the arithmetic average temperature of that fluid between the two end temperatures. The symbol Pr stands for the Prandtl number of the shell-side fluid. The exponent n and the multiplicative constant C depend on the pitch to tube OD ratio, and are given in a table provided in Holman’s book. An excerpt from the table for tubes on a rectangular pitch (in-line tube rows) is given below.
Values of the constant C /noSD
1.25
1.5
2.0
3.0
/poSD
1.25
0.386
0.305
0.111
To estimate fluid velocity in a heat exchanger without knowing the surface area, you can make assumptions and employ simplified methods. Assuming a known heat transfer rate (Q) and overall heat transfer coefficient (U), rearrange the heat exchanger equation to solve for surface area (A). Although the surface area cannot be directly determined without more information, you can estimate fluid velocity by assuming a velocity profile, using the cross-sectional area of the exchanger, or relying on known inlet or outlet velocities. These approaches involve simplifications and assumptions about fluid behavior within the exchanger, emphasizing the need for detailed design specifications or manufacturer information for more accurate calculations.
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Here's a general approach to estimating the fluid velocity without knowing the surface area:
1. Determine the flow rate: If you know the volumetric flow rate (Q) or mass flow rate (m) of the fluid passing through the exchanger, you can use this information in the calculation.
2. Estimate the cross-sectional area: If you can approximate the shape of the flow channel in the exchanger, you can estimate the cross-sectional area (A) based on that assumption. For example, if the flow channel is rectangular, you can estimate the area using the width and height of the channel. If the channel is circular, you can estimate the area using the radius or diameter of the channel.
3. Calculate the fluid velocity: Once you have an estimated cross-sectional area, you can calculate the fluid velocity (V) using the equation V = Q / A, where Q is the flow rate and A is the cross-sectional area.
It's important to note that this method provides an estimation and may not be as accurate as directly measuring the surface area of the exchanger. Additionally, the actual flow path and configuration of the exchanger may impact the fluid velocity distribution, so it's recommended to consult specific design guidelines or seek expert advice if possible.
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