The original idea for the plate heat exchangers was patented in the latter half of the nineteenth century, the first commercially successful design being introduced in 1923 by Dr. Richard Seligman. The basic design remains unchanged, but continual refinements have boosted operating pressures from 1 to 25 atmospheres in current machines.
The plate and frame heat exchanger (see Figure 1) consists of a frame in which closely spaced metal plates are clamped between a head and follower. The plates have corner ports and are sealed by gaskets around the ports and along the plate edges. A double seal forms pockets open to atmosphere to prevent mixing of product and service liquids in the rare event of leakage past a gasket.
Recent developments have introduced the double wall plate. The plates are grouped into passes with each fluid being directed evenly between the paralleled passages in each pass.
An important, exclusive feature of the plate heat exchanger is that by the use of special connector plates it is possible to provide connections for alternative fluids so that a number of duties can be done in the same frame [Lane (1966), Hargis et al. (1966), and Marriott (1971)].
Plates are made from a range of materials, for example, the "Paraflow" plates are pressed from stainless steel, titanium, Hastelloy, Avesta 254 SMO, Avesta 254 SLX or any material ductile enough to be formed into a pressing. The special design of the trough pattern strengthens the plates, increases the effective heat transfer area and produces turbulence in the liquid flow between plates. Plates are pressed in materials between 0.5 and 1.2 mm thick and plates are available with effective heat transfer area from 0.03 to 3.5 m2. Up to 700 plates can be contained within the frame of the largest Paraflow exchanger, providing over 2400 m2 of surface area. Flow ports and associated pipework are sized in proportion to the plate area and control the maximum liquid throughput.
As detailed in Table 1, various gasket elastomers are available which have chemical and temperature resistance coupled with good sealing properties. The temperatures shown are maximum, therefore possible simultaneous chemical action must be taken into account when selecting the most suitable material for a particular application.
Plate performance is determined by the plate geometry but it is not possible to estimate the film coefficient from the trough dimensions with some accuracy as can be obtained with a tube. The geometrical parameters involved such as plate gap, height, pitch and angle of the trough are too numerous for this to be possible but some work has been done on evaluating the effect of these variables [Kays and London (1958); Maslov (1965)].
In a plate heat exchanger, the heat transfer can best be described by a Dittus-Boelter type equation:
Typical values of the constant and exponents are
d is the equivalent diameter defined in the case of the plate heat exchangers as approximately 2 × the mean gap.
α=heat transfer coefficient ,
λ=thermal conductivity ,
cp=specific heat ,
subscription w = wail.
Typical velocities in plate heat exchangers for waterlike fluids in turbulent flow are 0.3-0.9 m/s but true velocities in certain regions will be higher by a factor of up to 4 due to the effect of the geometry of the plate design. All heat transfer and pressure drop relationships are based on either a velocity calculated from the average plate gap or on the flow rate per passage.
Figure 2 illustrates the effect of velocity on pressure drop and film coefficient. The film coefficients are very high and can be obtained for a moderate pressure drop.
One particularly important feature of the plate heat exchanger is that the turbulence induced by the troughs reduces the Reynolds number at which the flow becomes laminar. Typical values at which the flow becomes laminar varies from about 100 to 400, according to the type of plate. The friction factor is correlated with:
where y varies from 0.1 to 0.4 according to the plate and B is a constant for the plate.
In many applications, the heat transfer surface of the plate is less susceptible to fouling than a tubular unit. This is due to 4 principle advantages of the plate design:
There is a high degree of turbulence, which increases the rate of foulant removal and results in a lower asymptotic value of fouling resistance.
The velocity profile across a plate is good. There are no zones of low velocity compared with certain areas on the shell side of tubular exchangers.
Corrosion is maintained at an absolute minimum by careful selection of use of corrosive resistant materials.
Materials used for pressing the plates have a very smooth surface.
The most important of these is turbulence. HTRI (Heat Trasfer Research Incorporated) has shown that for tubular heat exchangers, fouling is a function of low velocities and friction factor. Although flow velocities are low with the plate heat exchanger, friction factors are very high, and this results in lower fouling resistance. The effect of velocity and turbulence is plotted in Figure 3. The lower fouling characteristics of the plate heat exchanger compared to the tubular has been verified by HTRI's work [Suitor (1976)].
Tests have been carried out which tend to confirm that fouling varies for different plates, with the more turbulent type of plate providing lower fouling resistances.
Higher overall heat transfer coefficients are obtained with the plate heat exchanger compared with the tubular for a similar loss of pressure because the shell side of tubular exchanger is basically a poor design from a thermal point of view. Considerable pressure drop is used without much benefit in heat transfer due to the turbulence in the separated region at the rear of the tube. Additionally, large areas of tubes even in a well-designed tubular unit are partially bypassed by liquid and low heat transfer area are thus created. Bypassing in a plate type exchanger is less of a problem and more use is made of the flow separation which occurs over the plate troughs since the reattachment point on the plate gives rise to an area of very high heat transfer.
For most duties, the fluids have to make fewer passes across the piates than would be required through tubes or in passes across the shell. Since in many cases a plate unit can carry out the duty with one pass for both fluids, the reduction in the number of required passes means less pressure lost due to entrance and exit losses and therefore more effective use of the pressure.
For condensing duties where permissible pressure loss is less than 7 kPascals the tubular unit is most efficient. Under such pressure conditions only a portion of the length of a plate heat exchanger plate would be used and a substantial surface area would be wasted. However, when less restrictive pressure drops are available the plate heat exchanger becomes an excellent condenser, since very high heat transfer coefficients are obtained and the condensation can be carried out in a single pass across the plate.
The pressure drop of condensing steam in the passage of plate heat exchangers has been investigated experimentally for a series of different Paraflow plates.
It is interesting to note that for a set of steam flow rates and given duty the steam pressure drop is higher when the liquid and steam are in countercurrent rather than cocurrent flow.
It can be shown that for equal duties and flow the temperature difference for countercurrent flow is lower at the steam inlet than at the outlet, with most of the steam condensation taking place in the lower half of the plate. The reverse holds true for cocurrent flow. In this case, most of the steam condenses in the top half of the plate, the mean vapor velocity is lower and a reduction in pressure drop of between 10-40% occurs. This difference in pressure drop becomes lower for duties where the final approach temperature between the steam and process fluid becomes larger.
The pressure drop of condensing steam is a function of steam flow rate, pressure and temperature difference. Since the steam pressure drop affects the saturation temperature of the steam, the mean temperature difference, in turn, becomes a function of steam pressure drop. This is particularly important when vacuum steam is being used, since small changes in steam pressure can give significant alterations in the temperature at which the steam condenses.
Plate Heat Exchangers also are used for gas cooling. The problems are similar to those of steam heating since the gas velocity changes along the length of the plate due either to condensation or to pressure fluctuations. Designs usually are restricted by pressure drop, therefore machines with low-pressure drop plates are recommended. A typical allowable pressure loss would be 3.5 kPascals with low gas velocities giving overall heat transfer coefficients in the region of 300 W/m2K.
The plate heat exchanger can also be used for evaporation of highly viscous fluids when the evaporation occurs in plate or the liquid flashes after leaving the plate. Applications generally have been restricted to the soap and food industries. The advantage of these units is their ability to concentrate viscous fluids of up to pascal seconds.
The unit is particularly suitable for high concentration especially as a finishing stage to a larger evaporator where the quantity of vapor is low and can be handled by the comparatively small ports of the plate [Jackson and Trouper (1966)].
One other field suitable for the plate heat exchanger is that of laminar flow heat transfer. It has been previously pointed out that the exchanger can save surface by handling fairly viscous fluids in turbulent flow because the critical Reynolds number is low. Once the viscosity exceeds 20-50 cP, however, most plate heat exchanger designs fall into the viscous flow range. Considering only Newtonian fluids since most chemical duties fall into this category, laminar flow can be said to be one of three types:
Fully developed velocity and temperature profiles (i.e., the limiting Nusselt case);
Fully developed velocity profile with developing temperature profile (i.e., the thermal entrance region); or
The simultaneous development of velocity and temperature profiles.
The first type is of interest only when considering fluids of low Prandtl number, and this does not usually exist with normal plate heat exchanger applications. The third is relevant only for fluids such as gases, which have a Prandtl number of about one.
For type 2 correlations for heat transfer and pressure drop in laminar flow are in the form
Nu = Nusselt number (αd/λ.),
Re = Reynolds number (Vdρ/η),
Pr = Prandtl number (cpη/λ) ,
η/ηp = Sieder Tate correction factor
(see Convective Heat Transfer), and
where f is the friction factor and a is a characteristic of the plate.
From this correlation it is possible to calculate the film heat transfer coefficient, for laminar flow. This coefficient, combined with that of the metal and the calculated coefficient for the service fluid together with the fouling resistance, are then used to produce the overall coefficient. As with turbulent flow, an allowance has to be made to the Log Mean Temperature Difference to allow for either end-effect correction for small plate packs and/or concurrency caused by having concurrent flow in some passes. This is particularly important for laminar flow since these exchangers usually have more than one pass. (See also Heat Exchangers and Mean Temperature Difference; Overall Heat Transfer Coefficient.)
For liquid/liquid duties, the plate heat exchanger will usually give a higher overall heat transfer coefficient and in many cases the required pressure loss will be no higher.
The effective mean temperature difference will usually be higher with the plate heat exchanger.
Although the tube is the best shape of flow conduit for withstanding pressure it is entirely the wrong shap for optimum heat transfer performance since it has the smallest surface area per unit of cross-sectional flow area.
Because of the restrictions in the flow area of the ports on plate units it is usually difficult (unless a moderate pressure loss is available) to produce economic designs when it is necessary to handle large quantities of low- density fluids such as vapors and gases.
A plate heat exchanger is more compact than a tubular and in many instances will occupy less floor space.
From a mechanical viewpoint, the plate passage is not the optimum and gasketed plate units are usually not made to withstand operating pressures much in excess of 25 kgf/cm.
For most materials of construction, sheet metal for plates is less expensive per unit area than tube of the same thickness.
When materials other than carbon steel are required, the plate will usually be more economical than the tube for the application.
When carbon steel construction is acceptable and when a closer temperature approach is not required, the tubular heat exchanger will often be the most economic solution since the plate heat exchanger is rarely made in carbon steel.
Until recently, applications for plate heat exchangers were restricted by the need for the gaskets to be elastomeric. Recent advances is design have introduced brazed plates and welded plates, thereby widening the range of applications.
Hargis, A. M., Beckman, A. T., and Loiacono, J. (1966) The Heat Exchanger, ASME Publication. PET, 21.
Jackson, B. W. and Troupe, R. A. (1966) Plate heat exchanger design by ENTU method chemical, Engineer Prog, Symp. Serv. 62 (No 64), 185 Kays, W. M. and London, A. L. (1958) Compact Heat Exchangers, McGraw-Hill, New York.
Lane, D. E. (1966) Design trends in plate heat exchangers, Chemical Process Engineering Heat Transfer Summary, 127, August.
Marriott, J. (1971) Where and how to use the plate heal exchangers, Chemical Engineering, 78 (no 8), 127.
Maslov, A. (1965) Calculation of the Heat Exchange of Plate Apparatus on the Basis of Diagram (Graphs), Kholod ekh. 6, 25.
Suitor, J. W. (1976) Plate Heat Exchanger Fouling Study, HTRI Report No. F-EX-18.