IMPROVING THE PREDICTION OF CONVECTION BANK HEAT TRANSFER. A.P. MANN, F. PLAZA and T.F. DIXON. Sugar Research Institute, Mackay, Qld

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1 IMPROVING THE PREDICTION OF CONVECTION BANK HEAT TRANSFER By A.P. MANN, F. PLAZA and T.F. DIXON Sugar Research Institute, Mackay, Qld KEYWORDS: Heat transfer, CFD Modelling, Convection Bank, Boiler. Abstract Many sugar mill boilers with baffled convection banks are susceptible to tube erosion failure. SRI has successfully reduced tube erosion at many sugar factories by redesigning convection banks using the CFD (Computational Fluid Dynamics) code FURNACE. In some cases where relatively small changes have been made to extend the life of a tube bank by a few years, a significant reduction in convection bank heat transfer performance has occurred. This reduces boiler efficiency and increases bagasse consumption. The FURNACE code has been able to predict whether or not modifications to convection bank design will change heat transfer performance but has not been able to accurately predict the magnitude of these changes. This paper describes modifications that have been made to the FURNACE code to improve the prediction of heat transfer performance. Comparisons with boiler operating data are used to assess the upgraded FURNACE code. In this work the ability of the FURNACE code to predict the changes in convection bank heat transfer performance caused by baffle modifications has been improved significantly. The upgraded FURNACE code successfully predicted most of the increase in convection bank exit gas temperature caused by convection bank baffle modifications in two boilers. For one boiler the model predicted a convection bank gas exit temperature rise of 33 o C versus an actual temperature rise of 36 o C. The predicted convection bank gas exit temperature rise for the other boiler was 76 o C and the actual temperature rise was 92 o C. Further testing of the model will be performed using data from other boilers with a wider range of convection bank baffle configurations. The model will be applied in the future to achieve the combined aims of improved heat transfer and reduced wear in convection banks and other boiler components. Introduction Many boilers in the Australian sugar industry (especially the older units) have baffled convection banks. Baffled banks are generally lower cost than baffleless convection banks because they are usually bottom supported, but they are also far more susceptible to tube erosion; a significant maintenance cost in many sugar factories. Tube erosion has been identified as a problem in sugar mill boilers with baffled convection banks for many years (Moir and Mason, 1982). More recently, SRI has had considerable success using the CFD code FURNACE to reduce tube wear rates by modifying the design of existing convection banks (Dixon and Plaza, 1995; Plaza et al., 1999). With the completion of a recent ARC

2 project with the University of Sydney, the FURNACE code now has sub-grid modelling capability and can predict the tube wear rates in specified parts of a convection bank directly (Mann et al., 2001). In some cases, where the factory has requested minimal changes to the convection bank design to extend tube life for a few years, significant reductions in convection bank heat transfer and boiler efficiency have occurred. It should be pointed out that, if the boiler is fitted with an air heater and/or economiser downstream of the convection bank, then any reduction in convection bank heat transfer will be largely offset by increased heat transfer in these heat exchangers, particularly the economiser. The reduction in boiler efficiency is therefore significantly less than the change in flue gas temperature at the convection bank exit would indicate. For factories with bagasse shortages or those wanting to maximise income from cogeneration, even small reductions in boiler efficiency can affect factory operations. Being able to accurately predict the effect, if any, of different convection bank baffle arrangements on heat transfer performance would greatly assist mill engineers to decide whether or not to go ahead with convection bank modifications to prolong tube life. Improvements in heat transfer may also be achieved. Background Many flow and heat transfer processes occur when flue gas passes over a bank of tubes. Heat transfer rates are affected by the thermal properties of the flue gas, the gas velocity, turbulence level and the arrangement of tubes in the bank. Current practice among boiler design engineers is to use empirical correlations for heat transfer across tube banks using assumed gas velocity distributions (Anon., 1992). Scaling factors are used to match the predicted heat transfer rates with boiler heat transfer data. These scaling factors compensate for the uncertainties associated with the heat transfer correlations used and the assumed gas velocity distributions. Such an approach is usually adequate for the purposes of sizing the main boiler heat transfer elements. However, problems arise with this method when it is not obvious how changing the length, location and orientation of baffles in a convection bank will affect flow patterns across the tubes. Alternatively, a CFD model that represents each individual tube in a bank can be used to predict large and small scale flow patterns and therefore heat transfer rates. No assumptions about gas flow patterns or empirical heat transfer correlations would be required. Unfortunately, flow cell sizes of 1 mm or smaller would be required to adequately represent the important flow processes. For a typically sized convection bank, this would increase computer RAM requirements to over 1 terabyte (10 12 bytes) and dramatically increase computer run times. This approach is therefore not currently feasible. The approach followed by SRI in the past for tube erosion modelling work was to represent the convection bank by a set of vertical plates evenly spaced across the width of the boiler. These plates would approximate the flow resistance and increase in gas velocity associated with the presence of the convection bank tubes. The FURNACE code uses wall functions, which approximate the velocity distribution in turbulent boundary layers, to predict heat transfer rates. As the vertical plates only roughly approximate the geometry of the convection bank tubes, scaling factors are required to match the predicted and measured convection bank exit temperatures. This technique of representing the convection bank by a series of vertical plates has been shown to be adequate for modelling tube erosion. For heat transfer modelling, however, it suffers from being unable to take into account the effects of changing flow angle and the variation of turbulence throughout the bank on heat transfer. Furthermore, the predicted heat transfer rates and therefore the scaling factors required to match predicted and measured convection bank outlet temperatures depend on the number and size of the vertical plates used in the model.

3 Clearly, all the techniques discussed here for predicting convection bank heat transfer have their deficiencies. The approach developed in this work and described in the next section addresses most of these deficiencies and gives improved prediction of the effect of changed baffle arrangements on convection bank heat transfer performance. Modifications to the FURNACE model FURNACE is a three dimensional CFD code that can predict gas and particle flow patterns, combustion and radiation in an industrial boiler. It was developed at the University of Sydney (Boyd and Kent, 1986) primarily for coal fired boilers and has been modified over a number of years to model the flow and combustion processes in bagasse fired boilers (Luo, 1993; Woodfield et al., 1998; Mann et al., 2001). Details of a more recent version of the FURNACE code are described elsewhere (Woodfield, 2001). Much of the bagasse boiler related research and consulting carried out over the past few years has made extensive use of the FURNACE model. In this work, the FURNACE code has been modified to better predict the heat transfer performance of convection banks. Each bank of tubes can now be represented as a porous region (Mann et al., 2001) and empirical correlations (Zhukauskas and Ulinskas, 1988) are used to predict heat transfer coefficients and therefore heat transfer rates. These expressions are of the form: h k d m n = C Re Pr (1) where h is the heat transfer coefficient, k is the thermal conductivity of the flue gas, d is the tube diameter, Re is the Reynolds number and Pr is the Prandtl number of the flue gas. C, m and n are empirical constants that depend on the arrangement and geometry of the tube bank, the gas velocity, the flow angle and the level of turbulence in the flow. The thermal conductivity (k) and Prandtl number (Pr) of the flue gas are calculated from published correlations (Verbank, 1997). Note that the correlations for the heat transfer coefficient (h) are based on experiments performed on complete banks of tubes using the thermal properties (k, Pr) of the carrier fluid at the entrance of the tube bank. In this work the correlations are used to calculate the heat transfer coefficient in each flow cell that is part of the porous region using local values of k and Pr. If k and Pr vary significantly through the bank the calculated heat transfer rates could be in error. The constant C in equation (1) includes correction factors for flow angle (C β ) and row position (C z ). Figure 1 shows the variation of the heat transfer coefficient correction factor (C β ) with flow angle (β) used in the FURNACE code. Figures 2 and 3 show the variation of the row position correction factor (C z ) with tube row number for in-line and staggered banks.

4 y = -8E-05x x R 2 = C β Fig. 1 Correction factor (C β ) for variation of the heat transfer coefficient with flow angle (β) used in the FURNACE code. Effect of row number - in line Cz y = x x x R 2 = Row number Fig. 2 Correction factor (C z ) for the variation of the heat transfer coefficient with row position in the bank for an in-line tube arrangement used in the FURNACE code. Effect of row number - staggered Cz y = x x R 2 = Row number Fig. 3 Correction factor (C z ) for variation of the heat transfer coefficient with row position in the bank for a staggered tube arrangement used in the FURNACE code.

5 Having determined the heat transfer coefficient (h), the local heat transfer rate (q) for a given area of tube surface (A) can be determined from: q = ha T gas T ) (2) ( wall where T gas is the local temperature of the flue gas and T wall is the temperature of the tube wall. As the steam flowing through convection bank tubes is saturated, T wall will be essentially constant. However, superheaters and economiser tubes will not have constant wall temperatures. In these cases, an average value of T wall can be used to calculate an approximate heat transfer rate. Modelling procedure To test the previously described modifications to the FURNACE code, the predicted changes in convection bank gas exit temperature resulting from convection bank modifications were compared with the actual changes for two boilers, the Union Saint Aubin boiler in Mauritius (29 bar) and the Kalamia No. 1 boiler in Australia (17.5 bar). Both boilers have suffered severe tube erosion problems and, in both cases, modifications recommended by SRI have extended the life of the convection bank tubes. At the request of the respective factories, minimal modifications were made to the convection banks. The priority was to reduce tube failures at minimum cost. In both cases, the modifications achieved the aim of no tube failures in the next crushing season at the required steam load. However, a consequence of the reduced convection bank tube wear was an increase in the convection bank gas exit temperature, causing a reduction in boiler efficiency. The FURNACE code predicted an increase in convection bank exit gas temperature for both boilers but significantly under predicted the magnitude of the temperature change. In this work, both boilers were modelled using porous regions to represent the convection bank and superheater tube arrays. Local heat transfer coefficients were calculated using correlations based on equation (1). These local heat transfer coefficients were multiplied by a scaling factor so that the predicted convection bank exit temperatures of both boilers before any convection bank modifications would match the measured values. The scaling factor takes into account all other effects such as radiation in the convection bank. Both boilers were then modelled with the modified convection bank configurations using the same scaling factor for the local heat transfer coefficients. The predicted and measured convection bank exit temperatures with the modified convection bank geometries were then compared. The Union Saint Aubin boiler was modelled using a three-dimensional flow grid with approximately cells. An embedded fine region was used to give increased resolution of the flow patterns in the convection bank. The superheater tube bank and the front and rear sections of the convection bank were represented by porous regions. A side elevation view of the flow grid used to model the Union Saint Aubin boiler showing the locations of the porous regions used for the superheater and convection bank tubes is shown in Figure 4. The radiation grid included only the furnace and had 8640 cells. Lagrangian tracking of trajectories was used to represent the motion of bagasse and ash particles through the boiler. The boiler was modelled for operation at a steam flow of 74 t/h at 22% excess air with 48% bagasse moisture content.

6 Fig. 4 Side elevation view of the flow grid used to model the Union Saint Aubin boiler showing the positions of the superheater and convection bank porous regions. The three-dimensional flow grid used to model the Kalamia No. 1 boiler had approximately cells with an embedded fine region for increased flow resolution in the convection bank. Porous regions were used to represent the tubes in the convection bank and superheaters. Figure 5 shows the flow grid and porous regions used to model the Kalamia No. 1 boiler. The radiation grid had cells, and particle tracks were used to represent the flow of bagasse and ash particles through the boiler. Modelled operating conditions for the Kalamia No. 1 boiler were 150 t/h steam flow, 44% excess air and 47.2% bagasse moisture content. Fig. 5 Side elevation view of the flow grid used for the Kalamia No. 1 boiler showing the positions of the superheater and convection bank porous regions.

7 Results Figure 6 shows a side elevation view of the predicted gas velocity distribution through the Union Saint Aubin boiler before the convection bank modifications. High tube wear rates were recorded where the flow changes direction at the bottom of the centre convection bank baffle, at the end of the convection bank exit baffle, and on the tubes near the convection bank exit just below the steam drum. All these high wear rate locations correspond to areas of high local gas velocity in Figure 6. To reduce tube-wear rates, a small baffle was placed just above and on the upstream side of the bottom of the centre convection bank baffle, a baffle was added to the rear wall of the convection bank, and the furnace exit baffle was slightly shortened. The predicted gas velocity distribution through the boiler with these modifications is shown in Figure 7. The location of high gas velocity at the bottom of the centre baffle has been shifted down so that the gas crosses the tubes of the bank at near right angles. Peak gas velocities near the convection bank exit baffle were predicted to decrease. These changes caused a reduction in tube wear rates. Figure 7 also shows that on the lee side of the small baffle added to the bottom of the centre convection bank baffle there is very little flow. There is also very little flow over the sections of tubes above the baffle added to the rear wall of the convection bank. The reduced utilisation of convection bank tube surface area would be expected to reduce the heat transfer performance of the convection bank. Fig. 6 Side elevation view of the gas velocity (m/s) distribution through the Union Saint Aubin boiler before the convection bank baffle modifications.

8 Fig. 7 Side elevation view of the gas velocity (m/s) distribution through the Union Saint Aubin boiler after the convection bank baffle modifications. Figures 8 and 9 show side elevation views of the predicted gas velocity distribution through the Kalamia No. 1 boiler before and after convection bank baffle modifications. High tube-wear rates occurred at the bottom of the centre convection bank baffle and near the end of the convection bank exit baffle. These high wear rates correspond to the high local gas velocities and sharp changes in flow direction observed in these locations (Figure 8). By installing a new baffle on the rear wall of the convection bank and removing the furnace exit baffle, tube-wear rates near the furnace exit baffle were reduced. The baffle at the bottom of the convection bank baffle was shortened to shift the location of highest wear. Note that there is predicted to be much less cross flow over the convection bank tubes after the baffle modifications. Heat transfer rates are much higher for a gas flowing over tubes in cross flow than parallel flow due to the increased level of turbulence. The baffle modifications would therefore be expected to reduce convection bank heat transfer rates.

9 Z Y X X > Y Z > Fig. 8 Side elevation view of the gas velocity (m/s) distribution through the Kalamia No. 1 boiler before the convection bank baffle modifications. Fig. 9 Side elevation view of the gas velocity (m/s) distribution through the Kalamia No. 1 boiler after the convection bank baffle modifications. The predicted (using the upgraded FURNACE code) and measured convection bank gas outlet temperatures for both the Union Saint Aubin and Kalamia No. 1 boilers before and after the convection bank baffle modifications along with the scaling factors used in the predictions are shown in Table 1.

10 Table 1 Measured and predicted (with the upgraded FURNACE code) convection bank gas exit temperatures ( C) for the Union Saint Aubin and Kalamia No. 1 boilers before and after the convection bank modifications. Convection bank arrangement Saint Aubin original Saint Aubin modified Kalamia No. 1 original Kalamia No. 1 modified Measured convection bank exit temperature ( C) Predicted convection bank exit temperature ( C) Scaling factor After the convection bank baffle modification the convection bank gas exit temperatures increased by 36 C for the Union Saint Aubin boiler and by 92 C for the Kalamia boiler. As the heat transfer in the convection banks decreases, the boiler control system increases the bagasse and air flow rates to the furnace to maintain constant steam output. Furnace and superheater heat transfer increases to compensate for the reduced heat transfer in the convection bank. Both boilers have air heaters downstream of the convection bank. As the convection bank exit temperature increases, the heat transfer to the air heater and therefore the grate air temperature increases. The Union Saint Aubin boiler has a large superheater tube bank. The increased heat transfer to the superheater tubes caused an excessive increase in the steam outlet temperature. Water sprays were needed to control the final steam temperature. From Table 1 the predicted increase in convection bank exit gas temperature for the Union Saint Aubin boiler after the baffle modifications was 33 C. For the Kalamia No. 1 boiler the predicted convection bank gas exit temperature increase was 76 C. With the old FURNACE code less than half the actual temperature rise was predicted. In both cases, the upgraded FURNACE code has predicted well the temperature rise caused by the changed convection bank geometry. This is encouraging given the uncertainties associated with modelling heat transfer in bagasse-fired boilers. For example, there is considerable uncertainty about the relative contributions of radiation (which depend on the soot concentration and the emissivity of the furnace gases) and convection to furnace heat transfer. The k-ε turbulence model, used by the FURNACE code, is known to be deficient in strongly swirling flows. This could affect the prediction of flow patterns near the ends of the convection bank baffles, where there are high gas velocity gradients. As mentioned previously, using heat transfer correlations based on the properties of the oncoming flue gas steam to evaluate local heat transfer coefficients based on local flue gas properties could also be a source of error. Conclusions In this work the ability of the FURNACE code to predict the changes in convection bank heat transfer performance caused by baffle modifications has been improved significantly. The upgraded FURNACE code successfully predicted most of the increase in convection bank exit gas temperature caused by convection bank baffle modifications in two boilers. Further testing of the model will be performed using data from other boilers with a wider range of convection bank baffle configurations. The model will be applied in the future towards achieving the combined aims of improved heat transfer and reduced wear in convection banks and other boiler components.

11 Acknowledgments This work is funded by the Queensland Department of State Development and the Federal Government under the Sugar Industry Renewable Energy (SIRE) program. The authors would also like to thank Raymond Rivalland of Union Saint Aubin sugar factory and CSR Limited for giving permission to publish details of previous consulting work. REFERENCES Anon. (1992). Steam, its Generation and Use. Babcock and Wilcox Company, Barberton, Ohio, USA. Boyd, R.K and Kent, J.H. (1986). Three-dimensional furnace computer modelling, 21 st Symposium (International) on Combustion. The Combustion Institute, 21: Dixon, T.F. and Plaza, F. (1995). Prediction of erosion in the convection banks of boilers. Proc. Aust. Soc. Sugar Cane Technol., 17: Luo, M.C. (1993). Combustion of bagasse in a sugar mill boiler. Ph.D. thesis. Univ. of Qld., Australia. Mann, A.P., Pennisi, S.N., Dixon, T.F., Novozhilov, V., Kirkpatrick, M., Kent, J.H., Sazonov, V. and Zhang, L. (2001). Modelling of boiler tube erosion. Proc. Aust. Soc. Sugar Cane Technol., 23: Moir, M.K. and Mason, V. (1982). Tube wear in sugar mill boilers. Proc. Aust. Soc. Sugar Cane Technol., 4: Plaza, F., Dixon, T.F., Dickenson, N.L., Fitzmaurice, A.L. and Owens, M. (1999). Performance of baffled boilers with redesigned convection banks. Proc. Aust. Soc. Sugar Cane Technol. 21: Verbank, H. (1997). Development of a mathematical model for watertube boiler heat transfer calculations. Proc. S. Afr. Technol. Assoc., 71: Woodfield, P.L. (2001). Combustion instability in bagasse-fired furnaces. Ph.D. thesis. Univ. of Sydney, Australia. Woodfield, P.L., Kent, J.H. and Dixon, T.F. (1998). Computational modelling of a bagasse-fired furnace effects of fuel moisture. Proc. Aust. Soc. Sugar Cane Technol., 20: Zhukauskas, A.A. and Ulinskas, R. (1988). Heat Transfer in Tube Banks in Cross-flow, Springer, New York.