Commit 6886a1
2025-03-13 11:02:00 João Lopes: Deleted Permutadores_de_Carcaca_e_Tubo.md| Utilidades industriais/Equipamentos/Permutadores_de_Carcaca_e_Tubo.md .. /dev/null | |
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| - | > Departamento de Engenharia Química |
| - | > |
| - | > Mestrado Integrado em Engenharia Química |
| - | > |
| - | > Integração e intensificação de processos |
| - | > |
| - | > Shell and Tubes Heat Exchangers |
| - | > |
| - | > **Docente** **Responsável:** |
| - | > |
| - | > Nuno Manuel Clemente de Oliveira |
| - | > |
| - | > **Integrantes** **do** **grupo:** |
| - | |
| - | João Victor Vieira |
| - | |
| - | > Matteo Gecchele |
| - | |
| - | <img src="./xvze01o4.png" |
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| - | > **Introduction** **&** **Structure** |
| - | > |
| - | > The most common type of heat exchanger is the shell-and-tube, usually |
| - | > used in a lot of industrial applications. This type of heat exchanger |
| - | > has large number of tubes, sometimes several hundred, packed in a |
| - | > shell with their axes parallel to that of the shell. The heat transfer |
| - | > takes place between two fluid, one flowing inside the tubes and one |
| - | > flowing outside the tubes through the shell. Baffles are commonly |
| - | > placed in the shell to force the shell-side fluid to flow across the |
| - | > shell to enhance heat transfer, to maintain uniform spacing between |
| - | > the tubes and, also in order to maintain the turbulent flow inside the |
| - | > exchanger. The baffle spacing is usually not greater than a distance |
| - | > equal to the inside diameter or closer than a distance equal to |
| - | > one-fifth the inside diameter of the shell. |
| - | > |
| - | > Usually the shell-and-tube heat exchangers have large size and weight, |
| - | > and for this reason they are not using in automotive and aircraft |
| - | > applications. At both ends of the shell, the tubes open to some large |
| - | > flow areas, called headers, where the tube-side fluid accumulates |
| - | > before entering the tubes and after leaving them. |
| - | > |
| - | > Shell-and-tube heat exchangers are further classified according to the |
| - | > number of shell and tube passes involved. Heat exchangers in which all |
| - | > the tubes make one U-turn in the shell, for example, are called |
| - | > one-shell-pass and two-tube-passes heat exchangers. Likewise, a heat |
| - | > exchanger that involves two passes in the shell and four passes in the |
| - | > tubes is called a two-shell- passes and four-tube-passes heat |
| - | > exchanger. |
| - | |
| - | <img src="./tnci3l31.png" |
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| - | > **Operation** **principle** |
| - | > |
| - | > In order to calculate the temperature difference ∆𝑡 in a 1-2 |
| - | > exchanger, it is necessary to make some assumptions: |
| - | > |
| - | > 1\. The shell fluid temperature is an average isothermal temperature |
| - | > at any cross section |
| - | > |
| - | > 2\. There is an equal amount of heating surface in each pass 3. The |
| - | > overall coefficient of heat transfer is constant |
| - | > |
| - | > 4\. The specific heat of each fluid is constant 5. The flowrate of |
| - | > each fluid is constant |
| - | > |
| - | > 6\. There are not phase change (evaporation or condensation) in a part |
| - | > of the exchanger |
| - | > |
| - | > 7\. Heat losses are negligible |
| - | > |
| - | > The overall heat balance where ∆𝑡 is the true difference of |
| - | > temperatures, is: |
| - | > |
| - | > 𝑄 = 𝑈𝐴∆𝑡 = 𝑊𝐶(𝑇 − 𝑇 ) = 𝑤𝑐(𝑡2 − 𝑡1) where U is the heat transfer |
| - | > coefficient and A is the surface of contact. |
| - | > |
| - | > Shell-and-tube heat exchangers are complicated devices and the |
| - | > simplified approaches should be used with care. In fact, it is assumed |
| - | > that the overall heat transfer coefficient U is constant throughout |
| - | > the heat exchanger and that the convection heat transfer coefficients |
| - | > can be predicted using the convection correlations. However, in some |
| - | > practical application, the predicted value of U can exceed 30 percent. |
| - | > Thus, it is natural to tend to overdesign the heat exchangers in order |
| - | > to avoid unpleasant surprises. |
| - | > |
| - | > Heat transfer enhancement in heat exchangers is usually accompanied by |
| - | > increased |
| - | > |
| - | > pressure drop, and this causes higher pumping power. Therefore, any |
| - | > gain from the enhancement in heat transfer should be balanced against |
| - | > the cost of the accompanying pressure drop. Also, some thought should |
| - | > be given to which fluid should pass through the tube side and which |
| - | > through the shell side. Usually, the more viscous fluid is more |
| - | > suitable for the shell side (larger passage area and lower pressure |
| - | > drop) and the fluid with the higher pressure for the tube side. |
| - | > |
| - | > Usually, it is convenient to relate the equivalent temperature |
| - | > difference to the log |
| - | > |
| - | > mean temperature difference relation for the counter-flow case as |
| - | > |
| - | > ∆ 𝑙𝑚 = 𝐹∆ 𝑙𝑚,𝐶𝐹 |
| - | |
| - | where *F* is the correction factor**,** which depends on the geometry of |
| - | the heat exchanger and the inlet and outlet temperatures of the hot and |
| - | cold fluid streams. The |
| - | |
| - | > ∆𝑇𝑚,𝐶𝐹 is the log mean temperature difference for the case of a |
| - | > counter-flow heat exchanger with the same inlet and outlet |
| - | > temperatures. |
| - | > |
| - | > The correction factor *F* for a shell-and-tube heat exchanger is shown |
| - | > in the figures below versus two temperature ratios *P* and *R* defined |
| - | > as |
| - | > |
| - | > 𝑡2 − 𝑡1 𝑇 − 𝑡1 |
| - | > |
| - | > 𝑇 − 𝑇 𝑡2 − 𝑡1 |
| - | > |
| - | > where the subscripts 1 and 2 represent the inlet and outlet*,* |
| - | > respectively. Note that for |
| - | > |
| - | > a shell-and-tube heat exchanger, *T* and *t* represent the shell-side |
| - | > and tube-side temperatures, respectively. |
| - | |
| - | <img src="./gklfx0zi.png" |
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| - | > **Factors** **that** **influence** **performances** *Fouling:* |
| - | > |
| - | > The performance of heat exchangers usually deteriorates with time as a |
| - | > result of accumulation of deposits on heat transfer surfaces. The |
| - | > layer of deposits represents additional resistance to heat transfer |
| - | > and this causes a decrease of the rate of heat transfer in a heat |
| - | > exchanger. The net effect of these accumulations on heat transfer is |
| - | > represented by a fouling factor, which is a measure of the thermal |
| - | > resistance introduced by fouling. |
| - | > |
| - | > For a shell-and-tube heat exchanger it possible to write the overall |
| - | > heat transfer relation as |
| - | > |
| - | > 𝑈𝐴𝑠 = 𝑈𝐴𝑖 = 𝑈0𝐴0 = 𝑅 = ℎ𝑖𝐴𝑖 + 𝐴𝑖𝑖 + ln𝑈0𝐴0 𝑖) + 𝐴0 + ℎ0𝐴0 |
| - | > |
| - | > where 𝐴𝑖 = 𝐷𝐿 and 𝐴0 = 𝐷0𝐿 L are the areas of inner and outer |
| - | > surfaces, and 𝑅,𝑖 and 𝑅,0 are the fouling factors at those surfaces. |
| - | > |
| - | > *Heat* *transfer* *rate:* |
| - | > |
| - | > The heat transfer rate is the most important parameter of a heat |
| - | > exchanger. A heat exchanger should be capable of transferring heat at |
| - | > the specified rate in order to achieve the desired temperature change |
| - | > of the fluid at the specified mass flow rate. |
| - | > |
| - | > *Size* *and* *Weight:* |
| - | > |
| - | > The heat exchanger is better if it is smaller and lighter, in |
| - | > particular, in the automotive and aerospace industries, where size and |
| - | > weight requirements are most stringent. For this reason, |
| - | > shell-and-tube heat exchangers cannot be used in this type of |
| - | > application. Also, a larger heat exchanger normally carries a higher |
| - | > price tag. The space available for the heat exchanger in some cases |
| - | > limits the length of the tubes that can be used. |
| - | > |
| - | > *Material:* |
| - | > |
| - | > The thermal and structural stress effects need not be considered at |
| - | > pressures below 15 *atm* or temperatures below 150*°C*. But these |
| - | > effects are major considerations above 70 *atm* or 550*°C* and |
| - | > seriously limit the acceptable materials of the heat exchanger. |
| - | > |
| - | > A temperature difference of 50*°C* or more between the tubes and the |
| - | > shell will probably pose differential thermal expansion problems and |
| - | > needs to be considered. In the case of corrosive fluids, we may have |
| - | > to select expensive corrosion-resistant materials such as stainless |
| - | > steel or even titanium. |
| - | > |
| - | > **Cost** |
| - | > |
| - | > The purchase cost of a shell and tube depends on the rear head type |
| - | > and on the heat transfer |
| - | > |
| - | > area (size factor). The relationship between the purchase cost and the |
| - | > size factor is |
| - | > |
| - | > represented in the graph below |
| - | |
| - | <img src="./s1x5d1ti.png" |
| - | style="width:4.86667in;height:3.36917in" /> |
| - | |
| - | > Both fluids are usually forced to flow by pumps or fans that consume |
| - | > electrical power. The annual cost of electricity associated with the |
| - | > operation of the pumps and fans can be determined from |
| - | > |
| - | > 𝑂𝑝𝑒𝑟𝑎𝑡𝑖𝑛𝑔𝐶𝑜𝑠𝑡 = 𝑃𝑢𝑚𝑝𝑖𝑛𝑔𝑃𝑜𝑤𝑒𝑟\[𝑘𝑊\] × 𝐻𝑜𝑢𝑟𝑠𝑜𝑓𝑂𝑝𝑒𝑟𝑎𝑡𝑖𝑜𝑛\[ℎ\] × |
| - | > 𝑃𝑟𝑖𝑐𝑒𝑜𝑓𝐸𝑙𝑒𝑐𝑟𝑖𝑐𝑖𝑡𝑦\[\$ 𝑘𝑊ℎ\] |
| - | > |
| - | > where the pumping power is the total electrical power consumed by the |
| - | > motors of the pumps and fans. |
| - | > |
| - | > Minimizing the pressure drop and the mass flow rate of the fluids will |
| - | > minimize the operating cost of the heat exchanger, but it will |
| - | > maximize the size of the heat exchanger and thus the initial cost. As |
| - | > a rule of thumb, doubling the mass flow rate will reduce the initial |
| - | > cost by half but will increase the pumping power requirements by a |
| - | > factor of roughly eight. Typically, fluid velocities encountered in |
| - | > heat exchangers range between 0.7 and 7 m/s for liquids and between 3 |
| - | > and 30 m/s for gases. Low velocities are helpful in avoiding erosion, |
| - | > tube vibrations, and noise as well as pressure drop. |
| - | > |
| - | > **Advantages**: |
| - | > |
| - | > *Size:* |
| - | > |
| - | > Shell-and-tube heat exchangers are capable of providing a larger |
| - | > surface area for heat transfer to take place while having a shorter |
| - | > length overall due to presence of multiple tubes. |
| - | > |
| - | > *Heat* *duty:* |
| - | > |
| - | > Shell-and-tube heat exchangers can handle higher temperatures and |
| - | > pressures and hence higher heat duty. This is because besides |
| - | > providing a higher overall heat transfer coefficient, additions can |
| - | > also be made to negate thermal expansion effects and the thickness can |
| - | > also be varied (more in the next point). |
| - | > |
| - | > *Versatility:* |
| - | > |
| - | > From the design point of view, shell-and-tube heat exchangers are the |
| - | > most versatile of all heat exchangers. Being tubular in shape, heads / |
| - | > closures of required shape and thickness can be used. The number of |
| - | > tubes and tube pitch can be selected according to operating |
| - | > conditions. Expansion bellows can be used to negate thermal expansion |
| - | > effects, baffles if different cuts and spacings can be used to |
| - | > influence the overall heat transfer coefficients and there's even |
| - | > something called a floating head which can be added to negate thermal |
| - | > expansion of the tubes. The number of passes on shell side and tube |
| - | > side can be altered as well. |
| - | > |
| - | > **Disadvantages**: |
| - | > |
| - | > *Size:* |
| - | > |
| - | > This can also be a disadvantage as at lower heat duty, there are more |
| - | > compact heat exchangers such as plate type exchanger. Also, the |
| - | > absence of hairpin bends causes shell-and-tube heat exchangers to take |
| - | > up more space than double pipe heat exchangers in some cases. |
| - | > |
| - | > *Maintenance:* |
| - | > |
| - | > Cleaning of tubes is difficult and fouling is always an issue when |
| - | > overall heat transfer coefficient is addressed. This requires periodic |
| - | > cleaning of the shell as well as the tubes. Cleaning tubes may be more |
| - | > difficult if the pitch is triangular. |
| - | > |
| - | > **Utilities** |
| - | > |
| - | > The selection of utilities to be used in the shell and tubes tube |
| - | > exchanger takes into |
| - | > |
| - | > account the type of industry in which it is being operated and the |
| - | > desired parameters, such as the required power, thermal stability and |
| - | > thermal capacity. |
| - | > |
| - | > *Cooling* *Water*: |
| - | > |
| - | > Cooling water is used to cool and/or condense currents. The cooling |
| - | > water circulates inside heat exchangers. About 80% of the temperature |
| - | > reduction is due to the evaporation of the cooling water and the |
| - | > transfer of heat to the surrounding air. |
| - | > |
| - | > *Steam:* |
| - | > |
| - | > Steam is the most common heat utility used in the chemical industry |
| - | > and can be used to power pumps, compressors and heat exchangers. Using |
| - | > steam allows a more efficient heat source since the heat of |
| - | > condensation of the steam is quite high, which translates into a high |
| - | > yield per utility mass, at a constant temperature. Another reason is |
| - | > that steam is non-flammable, non-toxic and inert to various process |
| - | > fluids (more safe than other utilities like oil). |
| - | > |
| - | > **Conclusion** |
| - | > |
| - | > The simple design of a shell and tube heat exchanger makes it an ideal |
| - | > cooling solution for a wide variety of applications and as a |
| - | > consequence shell-and-tube heat exchangers are very popular and |
| - | > commonly found in industrial use. |
| - | > |
| - | > **References** |
| - | > |
| - | > \[1\] Notes on Transfer Phenomena II, Professor Maria Graça Carvalho, |
| - | > 2018/2019; |
| - | > |
| - | > \[2\] Warren D. Seider, University of Pennsylvania |
| - | > |
| - | > \[3\] Heat Transfer by Changel 2nd Edition |
| - | > |
| - | > \[4\] Heat Transfer by Holman 6th Edition |