Research Article  Open Access
Numerical Investigation on Double ShellPass ShellandTube Heat Exchanger with Continuous Helical Baffles
Abstract
A double shellpass shellandtube heat exchanger with continuous helical baffles (STHXCH) has been invented to improve the shellside performance of STHXCH. At the same flow area, the double shellpass STHXCH is compared with a single shellpass STHXCH and a conventional shellandtube heat exchanger with segmental baffles (STHXSG) by means of numerical method. The numerical results show that the shellside heat transfer coefficients of the novel heat exchanger are 12–17% and 14–25% higher than those of STHXSG and single shellpass STHXCH, respectively; the shellside pressure drop of the novel heat exchanger is slightly lower than that of STHXSG and 29–35% higher than that of single shellpass STHXCH. Analyses of shellside flow field show that, under the same flow rate, double shellpass STHXCH has the largest shellside volume average velocity and the most uniform velocity distribution of the three STHXs. The shellside helical flow pattern of double shellpass STHXCH is more similar to longitudinal flow than that of single shellpass STHXCH. Its distribution of fluid mechanical energy dissipation is also uniform. The double shellpass STHXCH might be used to replace the STHXSG in industrial applications to save energy, reduce cost, and prolong the service life.
1. Introduction
A variety of heat exchangers are used in industries, such as shellandtube heat exchangers (STHX), platefin heat exchangers, and finandtube heat exchangers. More than 35–40% of heat exchangers are of the shellandtube type due to their robust geometry construction, easy maintenance, and possible upgrades [1, 2]. However, the traditional STHX with segmental baffles (STHXSG) has many disadvantages such as large back mixing, fouling, high leakage flow, and large crossflow. Especially, segmental baffles bring on significant pressure drop across the exchanger when changing the direction of flow [3, 4]. Over the past decades, different kinds of baffles have been developed, for example, the conventional segmental baffles with different arrangements, the deflecting baffles, the overlap helical baffles, and the rod baffles [5–10].
For STHX with continuous helical baffles (STHXCH), the shellside flow passes a periodic helical path under the action of baffles. Compared to the conventional STHXSG, STHXCH has some advantages such as reduced shellside fouling, increased heattransferratetopressuredrop ratio, reduced bypass effects, and prevention from flowinduced vibration [11, 12]. Lutcha and Nemcansky [11] give two reasons accounting for the improvement of heat exchangers performance with helical baffles: firstly, a near plug flow is formed in the shell side which can increase temperature difference for heat transfer, and, secondly, the rotational flow induced by helical baffles creates a vortex which interacts with the boundary layer on the tube surface and favorably affects the heat transfer coefficient.
Some researches [13, 14] indicated that the larger the helix angle, the better shellside comprehensive performance of STHXCH when helix angle is less than 45°. However, a large helix angle, or in other words a large helix pitch, has some adverse effects: first, the shellside velocity becomes small under the same mass flow rate, which goes against heat transfer; second, the quantity of helical cycle is small, which means the helix flow is possibly not fully developed until it reaches the shellside outlet; third, the unsupported span on the tube bundles is large, which is not favorable for the prevention of fluidinduced vibration in the shell side [14]. To overcome the aforementioned drawbacks of STHXCH with large helix angles, a new type of double shellpass STHXCH is invented in this paper. At the same helix angle and shell diameter, the helix pitch and flow area of double shellpass STHXCH become half of those of single shellpass STHXCH. In order to validate the advantages of double shellpass STHXCH, its performance has been compared with that of a conventional STHXSG and single shellpass STHXCH by numerical method.
2. Physical Model
Physical models for the computational domain are depicted in Figure 1. Three models have the same heat transfer area and shellside flow area. The geometric parameters are summarized in Table 1.

(a) Double shellpass STHXCH
(b) Single shellpass STHXCH
(c) STHXSG
3. Numerical Model and Simulation Method
3.1. Governing Equations
The renormalization group (RNG) turbulence model is adopted because it can provide improved predictions of nearwall flows and flows with high streamline curvature [15]. The general governing equation is as follows: The (RNG) turbulence model governing equations are as follows: For continuity equation, , generalized diffusion coefficient , and source term ; for momentum equation, , generalized diffusion coefficient , and source term ; for energy equation, , generalized diffusion coefficient , and source term . Other relative parameters are as follows:, , , , , , , , , and , among which is viscosity coefficient, is Prandtl number, is turbulent Prandtl number, is pressure, is temperature, and , , and are velocity components.
3.2. Basic Assumptions and Boundary Conditions
To simplify the numerical simulation while still keeping the basic characteristics of the process, the following assumptions are made: (1) both the fluid flow and heat transfer processes are turbulent and in steady state; (2) the leakage flow between the tube and the baffle and that between the baffle and the shell are neglected; (3) effects of gravity and buoyancy forces are neglected; (4) the tube wall temperatures are kept constant in the whole shell side; (5) the heat exchanger is well insulated hence, the heat loss to the environment is totally neglected; (6) and the working fluid is heattransfer oil. Its viscosity changes in a large extent with the variation of temperature, and hence a quadratic function is fitted between viscosity and temperature, and the other physical properties are regarded as constant.
The shellside inlet is set as velocity inlet with a prescribed flow rate and temperature (°C). The outlet port is set as a pressure boundary condition, which means that a static pressure and a proper backflow are specified. The temperature of the tube walls is set as a constant of 40°C and the other surfaces are set as nonslip, adiabatic, and impermeable.
3.3. Mesh Generation and Numerical Method
Due to the complicated structure of STHXCH, the computational domain is meshed with unstructured tetrahedral and pyramidal elements which are generated by ICEM 12.0. Mesh adoption is also used for refining and coarsening local mesh according to gradient of variables. In order to ensure the accuracy of numerical results, a careful test for the mesh independence of the numerical solutions was conducted. In the test, three different mesh systems with 9.8 million, 13.6 million, and 17.8 million elements are adopted for calculation of the whole heat exchanger, and the difference in the overall pressure drop and the shellside heat transfer coefficient between the last two mesh systems is less than 2%. The local meshes and velocity distributions are shown in Figure 2.
The commercial code ANSYS CFX 12.0 is adopted to simulate the flow and heat transfer in the computational model. The governing equations are discredited by the finite volume method. The convergence criterion is that the flow field and mass residual should be less than 10^{−6} for the energy residual less than 10^{−7} for the energy equation. A parallel computation is performed on four DELL workstations with two QuadCore CPUs and 16 GB memory each by using CFX, and every simulation case takes approximately 36 h to get converged solutions.
3.4. Validation of Numerical Method
Validation of the numerical method was made using experimental results from the literature [16]. Figure 3 provides the comparisons between experimental date and simulation results using the numerical method in this paper. It can be observed that for both fluid pressure drop and heat transfer their variation trends with mass flow rate in good agreement. Quantitatively, the difference in pressure drop is 6.5%~17.8% and the difference in the heat transfer coefficients ranges 2.7%~5.1%. Obviously the numerical method in this paper is reliable and applicable.
(a) Overall pressure drop
(b) Heat transfer coefficient
4. Results and Discussion
4.1. Heat Transfer and Pressure Drop
Heat exchanger rate of shellside fluid is as follows: The shellside heat transfer coefficient is equal to where is the heat exchange area based on the outer diameter of tube , is the number of tubes, is the effective length of tubes, is the temperature of tube walls and the subscripts and refer to shell side and tube side, respectively.
The variation trends of shellside heat transfer and pressure drop with mass flow rate are shown in Figures 4 and 5. It can be seen that, at the same mass flow rate and flow area, the shellside heat transfer coefficients and heat transfer rate of double shellpass STHXCH are 12–17% and 14–25% higher than those of STHXSG and single shellpass STHXCH, respectively; the shellside pressure drop of double shellpass STHXCH is slightly lower than that of STHXSG and 29–35% higher than that of single shellpass STHXCH.
(a) Heat transfer rate versus flow rate
(b) Heat transfer coefficient versus flow rate
4.2. Flow Field Analysis
In Figure 6 the shellside volume average velocity data are presented. It can be clearly observed that the shellside volume average velocity of double shellpass STHXCH is much higher than those of the other two types of STHX in spite of the same flow rate and flow area. That is one of the reasons why double shellpass STHXCH has the best heat transfer performance.
The velocity distributions of the three STHXs are shown in Figure 7 with same mass flow rate. It can be found that there is nearly no back flow regions and dead zones existed in the shell pass of STHXCHs. The velocity distribution of double shellpass STHXCH is the most uniform of three STHXs.
(a) Double shellpass STHXCH
(b) Single shellpass STHXCH
(c) STHXSG
The local velocity vector distributions on the axial sections of shell are shown in Figure 8. For both STHXCH, the shellside fluids pass through the tube bundles basically in a helical pattern and rush the heat exchange tubes with an inclination angle. On the one hand, helical flow avoids abrupt turns of flow. On the other hand, it changes the crosssection shape of tube in flow direction into ellipse. Therefore, it can reduce the pressure drop in shell side and the vibration of tube bundle. It also can be found that, in each shellpass of double shellpass STHXCH, the flow directions are opposite totally. In addition, the angle between flow direction and axis of tube of double shellpass STHXCH is much smaller than that of single shellpass STHXCH. It means that the double shellpass STHXCH is more similar to the longitudinal flow heat exchanger than the single shellpass STHXCH. For STHXSG, the shellside fluid passes through the tube bundles in a dramatic zigzag pattern, which causes large flow resistance and high risk of vibration failure on tube bundle. Furthermore, obvious dead zones are formed at the corners between baffles and shell wall. Flow stagnation in dead zones goes against heat transfer and increases fouling resistance. Therefore, the shellside pressure drop of double shellpass STHXCH angles is slightly lower than that of STHXSG.
(a) Double shellpass STHXCH
(b) Single shellpass STHXCH
(c) STHXSG
4.3. Mechanical Energy Dissipation
Viscousness of the fluid leads to internal friction when the fluid flows. It is the fundamental reason of the generation of flow resistance. Fixed walls or another type of solid surfaces provide the conditions for the production of flow resistance. The final effect of flow resistance is the dissipation of fluid mechanical energy. Within a (RNG) approach for turbulent and steady flow, the local mechanical energy dissipation rate of fluid elements is formed by two parts. They are The first term equal to viscous dissipation rate , which is calculated by local timeaveraging velocity and fluid dynamic viscosity. The second term can be calculated by as shown in [17]. Here is the local turbulent dissipation rate, calculated with a turbulence model. The distributions of the three STHXs are shown in Figure 9. It can be seen that, in flow fulldeveloped region, the distribution of double shellpass STHXCH is the most uniform of the three STHXs. For single shellpass STHXCH, the fluid mechanical energy dissipation concentrated in the central region of the shell. For STHXSG, dissipation of fluid mechanical energy concentrated in crossflow tube banks region of the shell.
(a) Double shellpass STHXCH
(b) Single shellpass STHXCH
(c) STHXSG
4.4. Comprehensive Performance
Figure 10 provides the comparison of shellside heat transfer coefficient and heat transfer rate among the three types of STHX within the range of pressure drop tested.
(a) Shellside heat transfer coefficient versus pressure drop
(b) Shellside heat transfer rate versus pressure drop
It can be found from Figure 10 that under the same pressure drop, the shellside heat transfer coefficient of double shellpass STHXCH is 11–18% and 5–13% higher than that of single shellpass STHXCH and STHXSG, respectively. Then, the double shellpass STHXCH has the best shellside comprehensive performance and the STHXSG has the worst shellside comprehensive performance. The double shellpass STHXCH might be used to replace the STHXSG in industrial applications to save energy, reduce cost, and prolong the service life.
5. Conclusion
In this paper, a novel double shellpass STHXCH is investigated with numerical method and compared with a single shellpass STHXCH and a STHXSG. Three models have the same shellside flow area. The conclusions are summarized as follows: (1) under the same flow rate, the shell side heat transfer coefficient of double shellpass STHXCH is 14–25% and 12–17% higher than that of STHXSG and single shellpass STHXCH, respectively; the shellside pressure drop of double shellpass STHXCH is 29–35% higher than that of single shellpass STHXCH and slightly lower than that of STHXSG. (2) Under the same shell side pressure drop, the shellside volume average velocity of double shellpass STHXCH is the highest and the velocity distribution is the most uniform. (3) In flow fulldeveloped region of double shellpass STHXCH, the distribution of fluid mechanical energy dissipation rate is uniform in the three STHXs. (4) Under the same pressure drop, double shellpass STHXCH has the best heat transfer performance.
Acknowledgment
The authors would like to acknowledge the financial support from the National Basic Research Program of China (973 Program) (no. 2007CB206900).
References
 K. J. Bell, “Heat exchanger design for the process industries,” Journal of Heat Transfer, vol. 126, no. 6, pp. 877–885, 2004. View at: Publisher Site  Google Scholar
 B. I. Master, K. S. Chunangad, A. J. Boxma, D. Kral, and P. Stehlík, “Most frequently used heat exchangers from pioneering research to worldwide applications,” Heat Transfer Engineering, vol. 27, no. 6, pp. 4–11, 2006. View at: Publisher Site  Google Scholar
 H. Li and V. Kottke, “Effect of the leakage on pressure drop and local heat transfer in shellandtube heat exchangers for staggered tube arrangement,” International Journal of Heat and Mass Transfer, vol. 41, no. 2, pp. 425–433, 1998. View at: Google Scholar
 H. Li and V. Kottke, “Visualization and determination of local heat transfer coefficients in shellandtube heat exchangers for staggered tube arrangement by mass transfer measurements,” Experimental Thermal and Fluid Science, vol. 17, no. 3, pp. 210–216, 1998. View at: Publisher Site  Google Scholar
 P. Stehlík and V. V. Wadekar, “Different strategies to improve industrial heat exchange,” Heat Transfer Engineering, vol. 23, no. 6, pp. 36–48, 2002. View at: Publisher Site  Google Scholar
 H. Li and V. Kottke, “Effect of baffle spacing on pressure drop and local heat transfer in shellandtube heat exchangers for staggered tube arrangement,” International Journal of Heat and Mass Transfer, vol. 41, no. 10, pp. 1303–1311, 1998. View at: Google Scholar
 R. L. Webb and N. H. Kim, Principles of Enhanced Heat Transfer, Taylor & Francis, Boca Raton, Fla, USA, 2005.
 E. A. Vasil'Tsov, “Turbulence of flow in mixers with deflecting baffles,” Chemical and Petroleum Engineering, vol. 24, no. 34, pp. 111–115, 1988. View at: Google Scholar
 C. C. Gentry, “RODbaffle heat exchanger technology,” Chemical Engineering Progress, vol. 86, no. 7, pp. 48–57, 1990. View at: Google Scholar
 Q. W. Dong, Y. Q. Wang, and M. S. Liu, “Numerical and experimental investigation of shellside characteristics for RODbaffle heat exchanger,” Applied Thermal Engineering, vol. 28, no. 7, pp. 651–660, 2008. View at: Publisher Site  Google Scholar
 J. Lutcha and J. Nemcansky, “Performance improvement of tubular heat exchangers by helical baffles,” Chemical Engineering Research and Design, vol. 68, no. 3, pp. 263–270, 1990. View at: Google Scholar
 Q. Wang, G. Chen, Q. Chen, and M. Zeng, “Review of Improvements on shellandtube heat exchangers with helical baffles,” Heat Transfer Engineering, vol. 31, no. 10, pp. 836–853, 2010. View at: Publisher Site  Google Scholar
 Y. G. Lei, Y. L. He, R. Li, and Y. F. Gao, “Effects of baffle inclination angle on flow and heat transfer of a heat exchanger with helical baffles,” Chemical Engineering and Processing: Process Intensification, vol. 47, no. 12, pp. 2336–2345, 2008. View at: Publisher Site  Google Scholar
 S. Ji, W. J. Du, and L. Cheng, “Numerical investigation on heat transfer and flow properties in shellside of heat exchanger with continuous helical baffles,” Proceedings of the Chinese Society of Electrical Engineering, vol. 29, no. 32, pp. 66–70, 2009. View at: Google Scholar
 V. Yakhot and L. M. Smith, “The renormalization group, the $\epsilon $expansion and derivation of turbulence models,” Journal of Scientific Computing, vol. 7, no. 1, pp. 35–61, 1992. View at: Publisher Site  Google Scholar
 H. U. Yan, Numerical Simulation of ShellandTube Heat Exchanger, Harbin Institute of Technology, Harbin, China, 2007.
 F. Kock and H. Herwig, “Local entropy production in turbulent shear flows: a highReynolds number model with wall functions,” International Journal of Heat and Mass Transfer, vol. 47, no. 1011, pp. 2205–2215, 2004. View at: Publisher Site  Google Scholar
Copyright
Copyright © 2011 Shui Ji et al. This is an open access article distributed under the Creative Commons Attribution License, which permits unrestricted use, distribution, and reproduction in any medium, provided the original work is properly cited.