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Fig. 10. Effect of flow maldistribution on the thermal performance of a plate condenser.

Figure 10 Effect of flow maldistribution on the thermal performance of a plate condenser.

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Substituting ¢,,; from Eq. (6) into Eqs. (4) and (5) gives
Substituting ¢,,; from Eq. (6) into Eqs. (4) and (5) gives
Fig. 1. Control volume for the channels.
Fig. 1. Control volume for the channels.
Fig. 2. The grid structure employed in the finite difference analysis of a single pass PHE. Using a numerical technique, the system of first order ordinary differential equations (12)-(14) can be solved. A The finite difference technique with a semi implicit proce- dure is used here. The plate heat exchanger channels are divided into M elements of equal size as shown in Fig. 2. rT he convergence criterion chosen for the termination of computation was | x 10~° in the non-dimensional temper- For two-phase flow channels, the modified form of the Boyko and Kruzhilin [36] correlation, suggested for vapor condensation in the PHE channels by Wang et al. [3],
Fig. 2. The grid structure employed in the finite difference analysis of a single pass PHE. Using a numerical technique, the system of first order ordinary differential equations (12)-(14) can be solved. A The finite difference technique with a semi implicit proce- dure is used here. The plate heat exchanger channels are divided into M elements of equal size as shown in Fig. 2. rT he convergence criterion chosen for the termination of computation was | x 10~° in the non-dimensional temper- For two-phase flow channels, the modified form of the Boyko and Kruzhilin [36] correlation, suggested for vapor condensation in the PHE channels by Wang et al. [3],
An iteration process has been continued after calculat- ing the dryness fraction of the two-phase flow. The dryness fraction of the two-phase flow has been calculated by taking an energy balance at the end of each two-phase flow channel. Eq. (19) has been used for calculating the heat transfer coefficient for the single phase flow channels. Eq. (17) can be modified for negligible variation in the saturation pres- sure of the two-phase flow from channel to channel for evaluating the heat transfer coefficient as h=f{ Re) and hence Eq. (14) can be simplified as
An iteration process has been continued after calculat- ing the dryness fraction of the two-phase flow. The dryness fraction of the two-phase flow has been calculated by taking an energy balance at the end of each two-phase flow channel. Eq. (19) has been used for calculating the heat transfer coefficient for the single phase flow channels. Eq. (17) can be modified for negligible variation in the saturation pres- sure of the two-phase flow from channel to channel for evaluating the heat transfer coefficient as h=f{ Re) and hence Eq. (14) can be simplified as
Fig. 3. Validation of the present model against the thermal performance of an ideal two-phase flow heat exchanger. the mass flow rate are the main operating parameters. In this analysis, the term, msr, is the mass flux ratio of the cooling fluid (water) and vapor is used wit QoUu ondition of the p hase fluid with oUuU'U f an ideal two- an) ids is zero (R=0 ate condenser. T he process stream. In practice, the process h negligible superheat before entering the o demonstrate the present model, the inlet rocess vapor has been considered as a two- he dryness fraction of 0.90 at 1 bar. The resent results were validated against the e - NTU value phase flow heat exchanger, ie., ¢=1— ~NTU when the heat capacity flow rate ratio of the two flu- , aS shown in Fig. 3. This has been carried out by setting the maldistribution parameter, m* = 0 for a total number of the channels equal to 98, the saturation pres- Parametric study has been carried out to observe the behavior of the heat exchanger for the different conditions. Even though the effectiveness is an indicator of the temper- ature rise of the cooling fluid, in practical applications, the value of the actual dimensional temperature is also important. It gives an idea about the practical ranges of temperatures over which the heat exchanger should operate which in turn decides about the secondary effects such as axial conduction and heat loss from the heat exchanger.
Fig. 3. Validation of the present model against the thermal performance of an ideal two-phase flow heat exchanger. the mass flow rate are the main operating parameters. In this analysis, the term, msr, is the mass flux ratio of the cooling fluid (water) and vapor is used wit QoUu ondition of the p hase fluid with oUuU'U f an ideal two- an) ids is zero (R=0 ate condenser. T he process stream. In practice, the process h negligible superheat before entering the o demonstrate the present model, the inlet rocess vapor has been considered as a two- he dryness fraction of 0.90 at 1 bar. The resent results were validated against the e - NTU value phase flow heat exchanger, ie., ¢=1— ~NTU when the heat capacity flow rate ratio of the two flu- , aS shown in Fig. 3. This has been carried out by setting the maldistribution parameter, m* = 0 for a total number of the channels equal to 98, the saturation pres- Parametric study has been carried out to observe the behavior of the heat exchanger for the different conditions. Even though the effectiveness is an indicator of the temper- ature rise of the cooling fluid, in practical applications, the value of the actual dimensional temperature is also important. It gives an idea about the practical ranges of temperatures over which the heat exchanger should operate which in turn decides about the secondary effects such as axial conduction and heat loss from the heat exchanger.
Fig. 4. Effect of the number of transfer units (NTU) on the cold fluid temperature rise (inlet to outlet) for a given plate condenser at different steam saturation pressure.
Fig. 4. Effect of the number of transfer units (NTU) on the cold fluid temperature rise (inlet to outlet) for a given plate condenser at different steam saturation pressure.
Fig. 5. Effect of the number of transfer units (NTU) on the condensation process for a given plate condenser at different steam saturation pressures t also plays a major role in deciding the ranges of dry wall and de-superheating in practical applications where the superheated vapor enters the condenser. For this reason, he dimensional temperature rise of the coolant versus NTU for the range of steam pressures from 1 to 10 bar has been separately plotted in Fig. 4. This figure shows that as the saturation pressure increases the temperature differ- ence from inlet to outlet of the cold fluid increases asymp- otically with NTU. Higher saturation pressure gives higher heat transfer and thus greater performance indi- cated by higher temperature rise of the coolant. At the same conditions, results are plotted to observe the outlet dryness fraction of the steam as a function of the NTU. Fig. 5 confirms that the exit vapor quality decreases with increasing NTU and with increasing saturation pressure of the steam. It has been observed that the variation of steam quality in the condensation process is very small beyond NTU = 3.0. This is due to formation of a thick water film which reduces the heat transfer rate. Further to be observed is the behavior of the condensation of the vapor, due to the mass flux ratio of the fluids (msr) for the different saturation pressures of the steam and these results are depicted in Fig. 6. As the mass flux ratio of the two fluids increases the condensation process increases steeply. This is also true for increasing the saturation pres- sure of the steam. Hence, these results reveal that in
Fig. 5. Effect of the number of transfer units (NTU) on the condensation process for a given plate condenser at different steam saturation pressures t also plays a major role in deciding the ranges of dry wall and de-superheating in practical applications where the superheated vapor enters the condenser. For this reason, he dimensional temperature rise of the coolant versus NTU for the range of steam pressures from 1 to 10 bar has been separately plotted in Fig. 4. This figure shows that as the saturation pressure increases the temperature differ- ence from inlet to outlet of the cold fluid increases asymp- otically with NTU. Higher saturation pressure gives higher heat transfer and thus greater performance indi- cated by higher temperature rise of the coolant. At the same conditions, results are plotted to observe the outlet dryness fraction of the steam as a function of the NTU. Fig. 5 confirms that the exit vapor quality decreases with increasing NTU and with increasing saturation pressure of the steam. It has been observed that the variation of steam quality in the condensation process is very small beyond NTU = 3.0. This is due to formation of a thick water film which reduces the heat transfer rate. Further to be observed is the behavior of the condensation of the vapor, due to the mass flux ratio of the fluids (msr) for the different saturation pressures of the steam and these results are depicted in Fig. 6. As the mass flux ratio of the two fluids increases the condensation process increases steeply. This is also true for increasing the saturation pres- sure of the steam. Hence, these results reveal that in
Fig. 6. Effect of the mass flux ratio on the condensation process for a given plate condenser at different steam saturation pressures
Fig. 6. Effect of the mass flux ratio on the condensation process for a given plate condenser at different steam saturation pressures
Fig. 7. Effect of msr on the condensation process with flow maldistribution in a plate condenser The above results have been obtained for uniform flow distribution (m?=0) to observe how the present model behaves against the parameters of which the influence can be initiatively guessed. The effect of flow maldistribution on the outlet vapor guessed quality of the process condens- ing steam with variation of the mass flow rate ratio of the cold water and steam has been analysed for a given inlet conditions of the process steam at the saturated pressure, Ps, = 10 bar and 1 and inlet vapor quality, Xin = 0.90, NTU = 2.0 (based on cold fluid) and the inlet temperature of cold fluid is 15 °C. It has been observed that there is a very small difference in outlet quality of steam both at high and low pressure when there is maldistribution from the port to the channels as compared with the uniform flow in each channel as shown in Fig. 7. This small difference justifies neglecting the properties variation when maldistri- bution is taking place from port to the channels. Fig. 8 show the effect of flow maldistribution on the temperature difference of the cold fluid against the variation of NTU for a given plate condenser of having mass flow rate ratio of 3.0 and the same inlet conditions as mentioned earlier. The results indicate that due to the non-uniform flow in each channel, the net enthalpy drop of steam in the plate
Fig. 7. Effect of msr on the condensation process with flow maldistribution in a plate condenser The above results have been obtained for uniform flow distribution (m?=0) to observe how the present model behaves against the parameters of which the influence can be initiatively guessed. The effect of flow maldistribution on the outlet vapor guessed quality of the process condens- ing steam with variation of the mass flow rate ratio of the cold water and steam has been analysed for a given inlet conditions of the process steam at the saturated pressure, Ps, = 10 bar and 1 and inlet vapor quality, Xin = 0.90, NTU = 2.0 (based on cold fluid) and the inlet temperature of cold fluid is 15 °C. It has been observed that there is a very small difference in outlet quality of steam both at high and low pressure when there is maldistribution from the port to the channels as compared with the uniform flow in each channel as shown in Fig. 7. This small difference justifies neglecting the properties variation when maldistri- bution is taking place from port to the channels. Fig. 8 show the effect of flow maldistribution on the temperature difference of the cold fluid against the variation of NTU for a given plate condenser of having mass flow rate ratio of 3.0 and the same inlet conditions as mentioned earlier. The results indicate that due to the non-uniform flow in each channel, the net enthalpy drop of steam in the plate
Fig. 8. Effect of NTU on the condensation process with flow maldistribution in a plate condenser.
Fig. 8. Effect of NTU on the condensation process with flow maldistribution in a plate condenser.
Fig. 9. Effect of NTU on the thermal effectiveness of a plate condenser with flow maldistribution. Wang and Zhao [33] have carried out an analysis of the performance of steam condensation in plate heat package will be decreasing with increase of flow maldistri- bution and results in the decrease in temperature difference of the cold fluid due to less heat transfer between the pro- cess and the cooling fluid. It is interesting to observe that this effect increases with pressure and is significant in the usual operating range of NTU for plate condenser (2-4). This effect will be even higher when the changes of the properties are taken into the consideration. Fig. 9 com- pares the effectiveness of the plate condenser with and with- out flow maldistribution from the port to the channel, i.e., m > 0. The flow maldistribution parameter, m has been varied from zero to 6.0 and pressure from | to 10 bar. It has been observed from Fig. 9 that when the maldistribu-
Fig. 9. Effect of NTU on the thermal effectiveness of a plate condenser with flow maldistribution. Wang and Zhao [33] have carried out an analysis of the performance of steam condensation in plate heat package will be decreasing with increase of flow maldistri- bution and results in the decrease in temperature difference of the cold fluid due to less heat transfer between the pro- cess and the cooling fluid. It is interesting to observe that this effect increases with pressure and is significant in the usual operating range of NTU for plate condenser (2-4). This effect will be even higher when the changes of the properties are taken into the consideration. Fig. 9 com- pares the effectiveness of the plate condenser with and with- out flow maldistribution from the port to the channel, i.e., m > 0. The flow maldistribution parameter, m has been varied from zero to 6.0 and pressure from | to 10 bar. It has been observed from Fig. 9 that when the maldistribu-
Fig. 11. Effect of flow maldistribution on the temperature rise of the cold fluid for a plate condenser.
Fig. 11. Effect of flow maldistribution on the temperature rise of the cold fluid for a plate condenser.
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EngineeringTwo Phase FlowThermal PerformanceChannel FlowProcess Industries
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