From time to time, a heat ex-changer is designed carefully yet fails to achieve the desired performance by a wide mar-gin, achieving, say, only half

the duty. With an understanding of some of the more common reasons why this might happen, design-ers can avoid these problems in the first place, and troubleshooters can recognize the root causes quickly.

Exchangers for single-phase opera-tion, condensing and boiling are con-sidered in that order here; but as we shall see, exchangers often handle a combination of these, and it is not always obvious which process is caus-ing the problem. In fact, some of these problems are quite unexpected and can even take experienced designers by surprise.

It must be recognized that the most important cause of problems in ex-changers is excessive fouling. Other articles, books, and conferences have been dedicated to this problem, so fouling will not be addressed here. Instead we consider those exchangers that have failed for some reason other than fouling.

SINGLE PHASEGood flow patterns are keyIn a single-phase exchanger, most problems arise when the unit is de-

signed to meet an unrealistically high thermal effectiveness. See Realistic Expectations, p. 45, for the upper lim-its of common heat exchangers. Oc-casionally, people inadvertently try to exceed these values. While not neces-sarily impossible, such an approach really pushes ones luck.

To achieve realistically high ther-mal effectiveness, countercurrent flow is normally essential. Also, both streams must be distributed evenly across any flow path and among all parallel flow paths. Furthermore, there must be no axial mixing. Flow that achieves these characteristics to-gether is normally referred to as plug flow. Plug flow, while necessary, is not a sufficient condition on its own to en-sure high thermal effectiveness. All parallel flow paths must also undergo identical heat transfer processes.

In general, shell-and-tube heat ex-changers are not good at achieving plug flow and identical heat transfer in parallel paths, which is why they

cannot achieve as high a thermal ef-fectiveness as other types. The main problems arise on the shell side.

In trying to achieve 90% effective-ness with a shell-and-tube exchanger, it is essential to ensure good arrange-ment of nozzles and headers so that equal flow occurs in all tubes. On the shell side, care is needed in the choice of baffle pitch and cut to minimize re-circulation paths behind baffles and thereby minimize the resultant axial mixing. It is very important to avoid bypass flows. It may even be necessary to go to a special design like a twisted-tube exchanger. For a practical example of this scenario, see Example 1, p. 45.

The Z and U flow arrangementsPlate-heat-exchanger manufacturers know well that of the two flow ar-rangements shown in Figure 1, the U arrangement gives a better flow distribution in the plate pack than the Z arrangement. One might think that the Z arrangement would be bet-

Feature Report

Sometimes fouling is not the problem. To get closer

to your design duty, consider these practical

design tips

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44 CHEMICAL ENGINEERING WWW.CHE.COM DECEMBER 2004

Heat Exchanger Duty: Going for Gold

David ButterworthHeat Transfer and Fluid Flow Service

FIGURE 1. In the plate exchanger with a Z arrangement, momentum changes throughout the exchanger give rise to a larger (unwanted) variation between its header pressures (graphs above and below each diagram) than that of a U ar-rangement. Therefore, the U arrangement is preferred for its more identical behav-ior of parallel channels

ter, since it gives equal-length flow paths for each parallel stream. (See Example 2, p. 47, for an illustration of how misleading this concept can be.) However, we have to look at the mo-mentum changes in the header as well as the frictional losses.

The small graphs over and under each figure show the pressure change in the headers when momentum

change effects are taken into account. It can be seen that the Z arrangement gives rise to a larger variation in the pressure difference that drives the flow in the parallel channels.

CONDENSATIONVentingOne of the main causes of problems with condensers is the failure to vent

noncondensable gases. This results in the depression of the dewpoint in the condenser and hence the loss of tem-perature driving force. High noncon-densable-gas concentrations also lower the heat transfer coefficient. Consider these rules-of-thumb for good venting: Vent from the cold end of the con-

denser, where noncondensable con-centrations are highest

Avoid any direct paths (from inlet to vent) that do not cross any tubes

Keep the pressure drop per unit length as uniform as possible along the flow path to drive the noncon-densable gas toward the vent (thus avoiding formation of noncondens-able gas pockets)

Keep the vent clear of the conden-

EXAMPLE 1. SHELL-AND-TUBE GAS HEATER

CHEMICAL ENGINEERING WWW.CHE.COM DECEMBER 2004 45

This case involves a cross-flow shell-and-tube exchanger with two tube-side passes formed with U tubes (TEMA AXU type). The shell side has nitrogen gas, which is to be heated from 21.5C to 135C by condensing steam at 151C inside the tubes. Under the design operating conditions, almost all the thermal resis-tance is on the shell side around 98%. While this arrangement is apparently a simple design, a quick calculation shows that the thermal effectiveness is 88%, a high enough value that any prob-lems arising will have serious effects.

The original design was as shown in Figure A (right). However, during construction of the plant, problems were found with the pip-ing layout so that the nozzles and bundle layout were rearranged as shown in Figure B. Nobody thought to block the interpass by-pass flow path on the shell side, so the exchanger performed far below the design duty. The outlet gas temperature achieved was only 97C as compared with the design value of 135C. The actual heat load was only 66% of the design value, and the overall heat transfer coefficient appeared to be only 42% of the design value.

The first attempt at a solution was to immediately insert a metal box to block the interpass gap, as shown in Figure C. Unfortu-nately, while some improvement was achieved, the problem was not completely solved. The gas outlet temperature was raised to 116C from 97C but still fell short of the design value of 135C. The overall heat-transfer coefficient had increased to 66%.

At one point, the designers decided to replacing the square tube layout with a triangular configuration to increase the number of tubes. It is good that this option was not actually implemented, because it would not have solved the problem.

The problem was instead with the condensing on the tube side. The way the U tubes were arranged meant that there was really a single pass on the tube side rather than two passes. To achieve the design condition, the first tubes encountered by the gas flow would experience a temperature difference of 129.5C, and the last only

16C. This is a ratio of 8.1 to 1. As the design overall coefficient was constant, the same ratio was being expected for the heat loads to the first and last tubes.

In the final and ultimately successful solution, the tube bundle was rotated as shown in Figure D. Calculations proved that to achieve the design condition, the outer U tubes would have to handle about twice the duty of the inner U tubes. This seemed better than before, but we were not sure if this alone would solve the problem. So we went for both a rotation of the bundle and a change to a triangular pitch. The solution was successful, giving an outlet gas temperature of 140C and thermal effectiveness of 91.5%, an improvement for both values over the design value.

Incidentally, after all of the attempts to troubleshoot this ex-changer, it would seem that the original design (Figure A), would have worked perfectly.

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Exchanger type .BYJNVN Shell-and-tube 90%

Plate-and-frame 95%

Plate-fin 98%

Printed circuit 98%

REALISTIC EXPECTATIONS

One important cause of problems is when the exchanger has been designed to have an unrealistically high thermal effectiveness (). The definition of thermal effectiveness is

(1)

where Q is the heat added or removed from the stream, and Qmax is the theoreti-cal maximum amount of heat that can be added or removed.

The most important stream in this context is the one with the highest value of . For single-phase streams with constant specific heat, Equation (1) becomes, for the hot and cold stream, respectively (of a two-stream exchanger):

(2a)

(2b)

When one is trying to achieve a high ther-mal effectiveness, any flaw in exchanger performance will have serious conse-quences. Some exchangers are capable of higher thermal effectiveness than others. Table 1 indicates the maximum percentage that can be realistically achieved for differ-ent types. To achieve the values in Table 1, great care is required in design. Going be-yond these values is not recommended.

sate layer to prevent flooding Be sure that parallel flow paths

have identical dutiesFigure 2 shows a vertical, shell-side condenser with a sensibly positioned vent that is at the cold end of the condenser and high enough up not to flood with condensate. It has been known for people to put the vent at the top of the shell, using some mistaken reasoning that noncondensable gases rise. The problem with that arrange-ment is that the whole of the con-denser must fill with noncondensable gas before the vent starts to operate. More information on the good loca-tion of vents is given elsewhere [1]. Here are explanations of some slightly unexpected venting errors that have caused serious problems.

Parallel condensing paths with different dutiesFigure 3 illustrates a problem that can occur with air-cooled condensers. In this particular case, the condenser has one tube-side pass and two rows, but the problem can occur with other ar-rangements as well. In the illustrated case, the temperature difference of the bottom row is greater than that of the top row. Because of the difference in heat loads, all of the vapor entering the bottom row condenses, while some in the top row does not. In compensa-tion, the bottom row sucks in the vapor from the outlet header that has not been condensed in the top row. As the

vapor enters the bottom row from both ends, the noncondensable gases cannot be vented and will be trapped, causing a deterioration in performance of the bottom pass.

The problem of parallel paths with different duties can occur in many situations (also illustrated in Exam-ple 1). Additionally, Figure 4 shows a TEMA J-type shell with condensation on the shell side. The problem in this arrangement is that there is only one tube-side pass for the coolant flow. Thus, one half of the exchanger has a higher heat load than the other, caus-ing a pocket of noncondensable gas to form away from the vent. Incidentally, the vent is actually in a sensible posi-tion for designs with many tube-side passes, which would cause the two halves to have nearly the same duty.

Zero pressure drop condenser with dead zonesIt is possible, particularly in vacuum condensers, to have zero pressure drop, because the frictional losses bal-ance the momentum recovery that results from vapor deceleration. The problem with this condition, of course, is that there is nothing to drive out the gas. Increasing the venting is not a good solution, because rather than clearing the dead tube, it tends to suck out vapor from the operating tube. The problem may not manifest itself at first, since time is needed for the noncondensable gases to accumulate. In fact, the problem could go unno-ticed for years causing a significant loss of thermal efficiency throughout the plant if, say, only one of two con-densers in parallel is affected.

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46 CHEMICAL ENGINEERING WWW.CHE.COM DECEMBER 2004

FIGURE 2. The placement of vents in vertical, shell-side condensers should be at the cold end of the condenser, just high enough to avoid condensate flooding

FIGURE 3. A lesser heat load in the top row of a parallel path inhibits complete condensation of vapor, which ultimately limits performance of the bottom pass

FIGURE 4. With only a single tube pass for coolant flow, this exchanger has a higher heat load on one half than on the other. This causes uneven condensation and inefficient heat transfer

FIGURE 5. With this configuration, the large vapor inlet flow and rapid condensa-tion achieve a large enough deceleration and pressure recovery to overcome the frictional pressure drop. To be stable, the zero-flow tube has to be filled with non-condensable, inert gas

So, with vacuum operation, never design a condenser to have zero pres-sure drop. Also, follow the good vent-ing rule that the pressure drop per unit length through the condenser is kept as uniform as possible. For con-densers in parallel, always have sepa-rate vents, rather than relying on a manifold system.

In Figure 5, for instance, all flow is running through the bottom tube. With this configuration, the large vapor inlet flow and rapid condensa-tion achieve a large enough decelera-tion and pressure recovery to over-come the frictional pressure drop. To be stable, the zero-flow tube has to be filled with noncondensable, inert gas.

Condensate not drainingCondensation gives a high heat-trans-fer coefficient, whereas a stagnant pool of condensate gives a very low heat-transfer coefficient for the tubes sit-ting in the pool. So, if the condensate is not draining properly, the efficiency of heat transfer will be compromised. The problem is somewhat common during plant startup, because debris that is generated during construction gets lodged in the outlet line. If that is not the cause, the problem may be arising because the condensate outlet line is too small.

Failure to drain condensate is a problem that sometime occurs with vertical thermosiphon reboilers heated with condensing service steam. Usu-ally, the reboiler has been designed with an unrealistically high fouling resistance, so that, initially, while clean, it will over-perform.

To bring the reboiler back to the de-sired performance, the common reac-tion is to reduce the steam pressure, which in turn lowers the condensing temperature. However, the steam trap

may not have been designed to oper-ate at the lower pressure. If not, the condensate will not drain properly, and the shell will fill with condensate until the pressure from the liquid head is high enough to force the steam trap to work. This problem manifests itself periodically, as a slow deterioration in performance and a jump back to good performance. Such unstable behavior in the reboiler can be violent enough to cause instabilities in the distillation column to which it is attached.

BOILINGUnpredictable nucleate-boiling heat-transfer coefficientsNucleate boiling is strongly depen-dant on the microscopic nature of the surface and the wetability of that sur-face with the liquid being boiled. Such properties are variable and difficult to predict. It has been shown [2] that all nucleate boiling correlations can be transformed to the following form:

(3)

where is the heat transfer coefficient and q is the heat flux. The quantities A and n are constants for a given fluid boiling on a given surface at a given pressure. Typically, n is around 2/3.

Various correlations are available to give A, but they are not very accurate. Errors in the prediction of A of a factor of two or more are common.

The heat transfer coefficient is de-fined in terms of the wall superheat, Ts, as follows:

(4)

Taking n as 2/3 and eliminating be-tween these two equations gives the heat flux as follows:

(5)

Hence, for a nucleate-boiling con-trolled design, any error in A will re-sult in a large error in local heat flux, and therefore, in the total heat load.

Admittedly, most boiling equipment is not entirely controlled by nucleate boiling. Yet, the magnitude of unpre-dictability can still lead to serious

CHEMICAL ENGINEERING WWW.CHE.COM DECEMBER 2004 47

EXAMPLE 2. FLOW MALDISTRIBUTION IN AN AIR-COOLED HEAT EXCHANGER

In the initial specification of a two-pass air-cooled heat exchanger, the design engineer was told that even flow distribution on the tube side would be vital to the performance of the ex-changer. The designer therefore came up with the design shown in the top il-lustration to the right. By having one inlet nozzle and two outlet nozzles, and by putting vertical baffles in the return header, the designer ensured that all the flow paths were the same length, believing that this would give a good flow distribution. The problem was that he had set up a Z flowpattern, which gives rise to large variations between the header pressures and ultimately does not maintain identical behavior in the parallel flow channels.

The obvious solution was to convert the Z flow arrangement to a U arrangement (actually two U arrangements in parallel). This could easily be achieved by replacing the one cen-tral inlet nozzle with two one at each end of the header. When other plant personnel weighed in, however, they insisted on a change that did not affect the pipework.

In the end, a rather complicated solution was found, as shown in the bottom illustration. The vertical baffles in the return header were replaced by a horizontal baffle, and piping was integrated inside the unit to take the flow from the center upper compartment and feed it back into the two ends of the lower compartment. The resultant flow arrangement was therefore two U arrangements in series (and two in parallel).

An important lesson to learn from this experience is that achieving good flow distribution is often more complicated than it seems. Momentum changes and their downstream effects on header flow (vis--vis header-pressure variation) must be taken into account.

FIGURE 6. Variation of temperature along a vertical thermosiphon reboiler

errors if corrective measures are not in place. Fortunately, the solution is quite simple. From Equation (5), it is clear that only a small adjustment in temperature difference is needed to overcome any problems due to a bad prediction of A. There is usually some control on the hot-fluid temperature, which is often condensing service steam, and hence on the overall tem-perature difference and wall super-heat.

The lesson to take away here is that if nucleate boiling plays a significant role in an exchanger performance, make sure that the temperature difference can be controlled. This lesson was forgot-ten for a period in the 1980s when pinch technology was becoming popular as a method of saving energy. People were replacing service steam with another process stream to do the heating. This replacement caused a loss of control of the temperature difference and resulted in failure of some boiling equipment.

Superheat required to initiate boilingA wall superheat is required to initi-ate bubble nucleation and, therefore, to enable boiling. At ambient and higher pressures, this critical super-heat is small and will be exceeded by the available superheat. However, with increasing vacuum, the required superheat for nucleation increases. So, for example, the superheat for water at atmospheric pressure is typically around 4C for a modest heat flux. However, at a tenth of an atmosphere it will be around 10C.

Again, the actual superheat is dif-ficult to predict and depends on fine features of the surface. Care must therefore be taken in vacuum boil-ers that there is enough temperature difference to initiate boiling. Alterna-tively, the need for nucleate boiling can be avoided by, say, using a falling-film evaporator.

Loss of temperature difference in reboilersFigure 6 shows a sketch of the tem-perature variation versus height in a vertical thermosiphon reboiler. The liquid entering the reboiler must be at a higher pressure than at the sur-face of the liquid in order to meet the

hydrostatic head needed to drive the natural circulation. This means that the liquid is subcooled at the inlet, and there must be a liquid-heating region in the bottom portion of the tubes.

Moving up the tube, the liquid tem-perature rises and the pressure falls until saturation is reached and bulk boiling commences. From there (with a relatively pure fluid), the boiling temperature falls with the falling pressure. From the figure, it can be seen that the average temperature difference is lower than the inlet and outlet values. In addition to having a lower temperature difference, the heat transfer coefficients in the liquid heat-ing region are lower than if boiling were to start at the entry to the tube. So, the heat transfer performance is reduced by these two effects.

It is important to note here that these hydrostatic effects are more dramatic under vacuum conditions because the hydrostatic pressure differences are a larger fraction of the total pressure. That state of affairs, combined with the need for sufficient superheat to give nucleation, can make vacuum re-boilers unpredictable. In some cases, falling film evaporators can be a better design for vacuum operation.

While this example focused on ver-tical reboilers, the same problem can occur in kettle and horizontal (shell-side) reboilers. In the case of horizon-tal, tube-side reboilers, care must be taken if the shell diameter is large, be-

cause then, the lower tubes in the bun-dle will behave very differently from the higher tubes. Check the saturation temperature for tube-side liquid at the top and bottom of the bundle, and go to a smaller shell diameter if the dif-ference in saturation temperature is large compared with the driving tem-perature difference.

Recirculation in vertical thermosiphon reboilersThere needs to be enough vapor leaving the tubes to drag the liquid smoothly out. If not, unstable opera-tion will result. It is also important to get the liquid smoothly out of the reboiler through the outlet nozzle and pipework (see Example 3, above).

Problematic operation in this re-gard is characterized by large diam-eter reboilers with poor and fluctuat-ing performance. They are certainly unstable, and may actually shake as well as upset the performance of the distillation column.

That problem is somewhat common. It occurs more often in reboiler appli-cations that caused problems in the past. Upon replacement of the reboiler, the temptation is to make the next re-boiler bigger in an attempt to avoid the problem. The tubes, of course, cannot be made longer without the expense of raising the distillation column, so more tubes are added. But, this tactic merely makes the problem worse.

To avoid this problem, use a flow pat-

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EXAMPLE 3. INSTABILITY IN A VERTICAL THERMOSIPHON REBOILER.

A new reboiler designer discovered instability in one of the plant's vertical thermosiphon reboilers, in that it had problems getting liq-uid smoothly out of the exchanger. The designer concluded that the solution was having high velocities in the two-phase region, which would have to be paid for with increased pressure drop there. Also a better outlet nozzle design was in-vestigated.

So far so good, but then the designer concluded that pressure drop in the liquid region was wasted and decided to install large inlet nozzles and pipework (figure). Consequently, the reboil-ers were very unstable and the problem had to be rectified by putting valves in the inlet lines so that a liquid pressure drop was introduced and could be adjusted until the instability ceased.

tern map or other criterion that will ensure that annular flow is achieved at the end of the tubes.

Dynamic instability in a vertical thermosiphon reboilerAs a consequence of the compressibil-ity of the two-phase flow in a reboiler, there is a time delay between a change in liquid inlet flow and the increased pressure drop that would otherwise stop the flow increase. For flow oscil-lations of a particular frequency, the time delay will be just enough to re-inforce the flow variation rather than damp it out. Such cases create dy-namic instability.

The problem is normally avoided by having some restriction to in-crease pressure drop in the single phase region. This removes the time lag between the flow change and the pressure-drop change. Mathematical models have been devised to predict the conditions which cause such insta-bilities.

Temperature difference of wide boiling-range mixtures Normally with a wide-boiling-range mixture, the liquid starts to boil at the bubble point of the mixture, and the boiling temperature rises from inlet to outlet of the exchanger. Temperature differences can be calculated from the boiling curve, which gives the rise

in temperature versus the amount of heat added and is typically based on the reboilers inlet flow.

However, it is easy to forget that there is internal circulation in the re-boiler, so the fluid mixture entering the bottom of the bundle is much richer in low-volatility (high boiling point) com-ponents than the inlet stream. Hence, the temperature differences driving the heat transfer are lower than what a boiling curve that is based on the inlet stream would predict. Failure to recognize this effect has been known to cause some kettles to operate below the design performance. Edited by Rebekkah MarshallReferences1. Butterworth, D., Condensers: Thermohy-

draulic Design, pp. 647 678. In: Kaka, S., Bergles, A.E., and Mayinger, F., eds., Heat Exchanger, Thermal-hydraulic Fundamen-tals and Design, Hemisphere Publishing Corp., Washington, New York, London, 1981.

2. Cooper, M.G., Correlations for Nucleate Boil-ing Formulation Using Reduced Proper-ties, Physicochemical Hydrodynamics, Vol. 3, No. 2, pp. 89111, 1982.

NOMENCLATURE

A Variable in nucleate-boil-ing prediction methods, W/(mK)/(W/m2)n

n Index in nucleate boiling pre-diction methods

Q Heat transfer rate: W

T Temperature: oC or K

Heat transfer coefficient: W/m2 K

Thermal effectivenessT Temperature differenceSubscriptscold Cold stream

hot Hot stream

in At inlet

out At outletmax Theoretical maximums Superheat

AuthorDavid Butterworth is a se-nior consultant to the Heat Transfer and Fluid Flow Service (HTFS; 29 Cleve-lands Abingdon Oxfordshire, OX14 2EQ, UK; Phone: +44 (0)1235 525 955; Fax: +44 (0)1235 200 906; Email: davebutterworth@aol.com) After obtaining his B.S. ChE from University College Lon-don (U.K.) in 1966, he joined

the heat transfer team at the Harwell Labora-tory of the United Kingdom Atomic Energy Au-thority where he helped in the foundation of the HTFS, which he went on to manage for 14 years. The objective of HTFS is to undertake research in process heat transfer and use the results in the development of design software for heat ex-changers and fired heaters. For his contribution to process heat transfer, Butterworth received the American Institute of Chemical Engineers D.Q. Kern Award in 1986 and was elected a Fel-low of the UK Royal Academy of Engineering in 1991. He is currently the Chairman of the Aluminum Plate-Fin Heat Exchanger Manufac-turers Association (ALPEMA), a visiting indus-trial professor at Bristol University and Senior Consultant at HTFS, which is now part of Aspen Technology. His other engineering qualifications include Chartered Engineer, European Engineer, Chartered Scientist and Fellow of the UK Insti-tution of Chemical Engineers. When not engaged in engineering activities, he enjoys landscape painting and cooking, is a council member of the UK Herb Society and is a Fellow of the Royal Society of Arts.

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