“Experimental study of the effectiveness and exergetic efficiency of counter-rotating screw heat exchanger in a prebaked anode production plant” M. A. Mohamed 1,2 , C.K Tan 2,* , A. A. Abd El-Rahman 5 , S. S. Wahid 3 , M. Attalla 1 , S. A. Ahmed 4 1 Department of Mechanical Engineering, South Valley University, Qena, Egypt 2 School of Engineering, University of South Wales, Pontypridd, UK 3 Depaerment of Mechanical Engineering, Minia University, Minia, Egypt 4 Departmentof Mechanical Engineering, Beni-Suef University, Beni-Suef, Egypt 5 The Aluminum Company of Egypt, Nag Hammadi, Qena, Egypt *Corresponding Authors: email: [email protected]Abstract: The current paper aims to carry out an exergy and energy analyses of the counter-rotating screw heat exchanger currently used in the prebaked anode production plant (Green Carbon Anode Plant) in Nag Hammadi Aluminum Factories. In the experiments, the flow rate of thermal oil into the coke preheater was varied from 90 m 3 /h, 100 m 3 /h, 110 m 3 /h and 120 m 3 /h and its temperature was changed from 230 C° to 260 C° with step of 10C° each test. First- and Second-law analyses were employed to evaluate the thermodynamic efficiency of the system. Studies of the impact of measurement uncertainties revealed that while the calculated effectiveness of the heat exchanger is less sensitive to that, the dimensionless exergy destruction and exergetic efficiency are more significantly affected. Further sensitivity analysis also concluded that there is not much economical value in improving the accuracy of the measurements due to the potential cost of more accurate measuring devices. Further results also showed that the effectiveness and exergetic efficiency generally improved at higher mass flow rate and inlet temperature of the hot stream. The study also found that between 54% to 74% of the available energy (i.e. exergy) of the hot stream could be potentially wasted. A large proportion of exergy destruction was due to the relatively high mean temperature difference between the hot and cold streams, which was necessary to 1
Transcript
“Experimental study of the effectiveness and exergetic efficiency of counter-rotating screw heat exchanger in a prebaked anode production plant”
M. A. Mohamed1,2, C.K Tan2,*, A. A. Abd El-Rahman5, S. S. Wahid3, M. Attalla1, S. A. Ahmed4 1Department of Mechanical Engineering, South Valley University, Qena, Egypt
2School of Engineering, University of South Wales, Pontypridd, UK3Depaerment of Mechanical Engineering, Minia University, Minia, Egypt
4Departmentof Mechanical Engineering, Beni-Suef University, Beni-Suef, Egypt5The Aluminum Company of Egypt, Nag Hammadi, Qena, Egypt
Abstract:The current paper aims to carry out an exergy and energy analyses of the counter-rotating screw heat exchanger currently used in the prebaked anode production plant (Green Carbon Anode Plant) in Nag Hammadi Aluminum Factories. In the experiments, the flow rate of thermal oil into the coke preheater was varied from 90 m3/h, 100 m3/h, 110 m3/h and 120 m3/h and its temperature was changed from 230 C° to 260 C° with step of 10C° each test. First- and Second-law analyses were employed to evaluate the thermodynamic efficiency of the system. Studies of the impact of measurement uncertainties revealed that while the calculated effectiveness of the heat exchanger is less sensitive to that, the dimensionless exergy destruction and exergetic efficiency are more significantly affected. Further sensitivity analysis also concluded that there is not much economical value in improving the accuracy of the measurements due to the potential cost of more accurate measuring devices. Further results also showed that the effectiveness and exergetic efficiency generally improved at higher mass flow rate and inlet temperature of the hot stream. The study also found that between 54% to 74% of the available energy (i.e. exergy) of the hot stream could be potentially wasted. A large proportion of exergy destruction was due to the relatively high mean temperature difference between the hot and cold streams, which was necessary to ensure satisfactory heating of the coke by conduction. To increase the exergetic efficiency of this heating process, it is proposed that the heat losses from the heat exchanger to the environment be minimized and that the mean temperature difference (for the same rate of heat transfer) should be reduced by redesigning the heat exchanger.
Highlights:1. Higher mass flow rate and inlet temperature of hot stream lead to better performance.2. Large logarithmic mean temperature difference contributes to large exergy destruction. 3. Between 54% to 74% of exergy of the hot stream could be potentially wasted.4. Measurement errors are acceptable relative to higher cost of more accurate instruments.
Nomenclaturec constant-pressure specific heat, kJ/
kgKGreek symbols
C heat capacity flow rate, kW/k uncertaintym mass flow rate, kg/s effectiveness of heat exchangerQ rate of heat transfer, kW dimensionless exergy destructionQactual actual rate of heat transfer, kW ex exergetic efficiency
Qloss unaccounted heat losses, kW density
s entropy, kJ/kgK
s rate of entropy transfer, kW/K Subscriptssgen rate of entropy generation, kW/K max maximumT temperature, K min minimumT o environmental temperature K oil,in oil inlet streamT oil oil temperature, K oil,out oil outlet streamT coke coke temperature, K coke,in coke inlet stream∆ T lm ,CF Logarithmic mean temperature
difference, Kcoke,out coke outlet stream
T s outer-casing temperature, Kx specific exergy, kJ/kgX dest rate of exergy destruction, kWV oil volume flow rate of oil, m3/h
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1.0 IntroductionHeat exchangers are widely used in various industries as a mean to transfer heat from one medium to another. In prebaked anode production plant, for example, counter-rotating screw heat exchangers (see Figure 1) are widely used to preheat green coke aggregates, which are then fed to kneader for subsequent forming of anode block. The preheating of green coke is an energy intensive process and thus appropriate design and operation of the heat exchanger is essential to ensure not only high heat transfer effectiveness but also to reduce energy wastage due to process irreversibility. First- and Second-law analyses of heat exchanger’s performance have been the subject of research for a number of years. In term of First-law analysis, estimation of effectiveness in heat transfer has been a widely accepted approach in designing and analysing the thermal performance of heat exchangers in general; see for examples [1-3]. However, Second-law analysis can be particularly useful to quantify energy wastage due to irreversibility and hence allows further optimization of heat exchanger’s characteristics to reduce sources of inefficiency in existing system. A detailed review has been conducted recently [4], which provided useful insight from many studies relating to exergy analysis and entropy generation in various heat exchanger applications. Although the review was limited to heat exchangers with gas, liquid and gas-liquid two-phase flows, they highlighted the usefulness of Second-law analysis in complementing so-called Constructal theory for optimization of heat exchanger’s design. Perhaps, the more significant finding from the review was due to the early work published by Bejan [5] in 1977. In the study, the author analyzed entropy generation in a balanced counter-flow gas-to-gas heat exchanger with negligible pressure drop and discovered non-monotonic relationship between entropy generation and effectiveness of the heat exchanger. The results showed that entropy generation (irreversibility) first increases with increasing effectiveness and peaks at an effectiveness of 0.5 after which it began to decrease. In 2005, a paper published by Shah and Skiepko [6], which investigated theoretically 18 different heat exchanger flow arrangements, further highlighted the complex relationship between irreversibility and effectiveness of heat exchangers in general. Particularly, the authors highlighted the ambiguity on the association of minimum entropy generation with maximum effectiveness in heat exchanger analysis such that heat exchanger effectiveness can be maximum, having an intermediate value or minimum at the maximum irreversibility operating point depending on the flow arrangement of the two fluids. This phenomenon is commonly known in the literature as the “entropy generation paradoxes”.
More recently, research work has also focused on the impact of experimental uncertainties in the calculated entropy generation [7]. Particularly this work highlighted the potential existent of negative (i.e. invalid) entropy generation anomalies calculated from actual measured data, which could be effectively (albeit 82% success rate) identified by the so-called heat balance error. In their work, the authors defined the heat balance error relative to an ideal heat exchanger experiment with no leakage of heat to the environment and absolutely zero uncertainty in process measurement. Later, an improved method based on the concept of “virtual entropy generation” was further developed by the authors [8], which was demonstrated to be promising in identifying experimental anomalies and dealing with uncertainty in measurements. Other researchers [9-12] also considered exergy analysis as a measure of irreversibility in heat exchangers. This is particularly meaningful as the Second-law performance of a heat
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exchanger can be viewed as the ability of the cold fluid stream to recover exergy (i.e. available energy) from the hot fluid stream. However, relatively few studies [13] actually consider the impact of exergy losses to the surroundings (due to heat losses) on the performance of the heat exchanger, which can be significant if the exchanger operates at temperatures further away from the ambient, as in cryogenic or high temperature processes. It is also worth highlighting that many different performance indicators associated with entropy and exergy analyses have evolved over the years, such as, fractional exergy loss, dimensionless entropy generation, the number of entropy generation units, the number of irreversibility units, exergetic/Second-law efficiency, exergetic-transfer effectiveness, dimensionless exergy destruction, exergy destruction number, or specific irreversibility, etc. However, many of these are either identical or provided similar trend in comparison [14].
The prior literature review indicated that the exergy efficiency is much lower than energy efficiency in all investigated cases and exergy analysis is particularly valuable tool for identifying sources of irreversibility in the heat exchanger systems, which could help improving their design and operation and thus significant saving in operating cost. However, from the literature review conducted so far, none of the research carried out that utilizes the First- and Second-law analyses has focused on this particular type of rotating screw heat exchanger where conduction heat transfer is also the dominant mode of heat transfer between the heat transfer surfaces of the rotating screw and the solid aggregates. This was the primary motivation for the current study to evaluate the First- and Second-law performances in order to obtain a detailed understanding of the heat transfer processes and sources of inefficiency of the heat exchanger used in the prebaked anode production plant.
2.0 Aluminum Production:In Egypt, aluminum industry is one of the largest electrical energy consumers among all other industries. It consumes nearly 3% of total electricity production in Egypt. Typically, aluminum is produced from alumina ore by electrolysis process, which is known as Hall-Héroult process [15-17]. In this process, the alumina is dissolved in an electrolytic bath of molten cryolite (sodium aluminum fluoride), then an electric current is passed through the electrolyte and flows between the anode and cathode. Molten aluminum is produced, deposited at the bottom of the electrolytic cell “pot”, and finally it is periodically siphoned off and transferred to a reverberatory holding furnace. In aluminum industry, there are two types of technologies used for aluminum production; first type by using self-baking anodes (Söderberg anodes) [18] and second type by using pre-baked anodes [19]. The Söderberg anodes are made in-situ from a paste of Calcined petroleum coke (or green coke) and coal tar pitch, and are baked by the heat from the molten electrolytic bath. As the anode is consumed, more paste descends through the anode shell in a process that does not require anode changes. Alternatively, prebaked anodes that are manufactured directly from dedicated manufacturing plant (so-called anode plant) can be used. Since the anodes are not consumed completely in the aluminium production cell, as the yoke stubs would contaminate the metal and be irrecoverable, the anode butts are recovered and used as a valuable recycled material for anode manufacture.
2.1 Description of the Anode Plant and screw-preheater:The anode plant is owned by Egyptalum located in Nag Hammadi. In the electrolytic production of aluminum, 415 Kg of carbon is used for each metric ton of aluminum metal production. The process of Anode production mainly consists of seven sub-processes that can be described as follows:
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1. Classification of feedstock materials: Green coke is received as mixture of coke particles ranging from fine dust to sizes of 2 – 3 cm in diameter. The coke is screen and sorted into 3 different sizes (coarse, medium and fine). The oversize particles are again passed through an impact crusher, screened and classified accordingly.
2. Proportioning of dry-mix: Measured quantities of the classified aggregates are subsequently mixed to produce a dry-mix that enter a screw preheater.
3. Preheating of dry-mix: A Holo-Flite indirect heat exchanger (described in more details below) is used to preheat the dry-mix to required temperature of between 150C and 200C.
4. Anode paste preparation and block forming: The hot dry-mix from the heat exchanger enters a kneader where it is further mixed with liquid coal tar pitch (served as binding agent) to produce a homogeneous paste for subsequent forming of anode block by hydraulic press.
5. Finally the anode blocks are baked in furnace to obtain the final desirable physical and thermal properties.
Figure 1. Holo-Flite Heat Exchanger.
Figure 2 preheating screws.
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In the Anode Plant a Holo-Flite counter-rotating screw heat exchanger, which is effectively an indirect heat exchanger as shown in Figure 1, was used to preheat the dry-mix to a target temperature of around 150C to 200C. It consists of a trough (1), which has volume of approximately 20 m3 with a double jacket through which preheated hot thermal oil circulates. Four counter-rotating screws (2) which are approximately 6.5 m long, are driven by an electric-motor drive system (3). As illustrated in Figure 2, each rotating screw consists of a hollow spiral (1), which allows the circulation of the heat transfer fluid entering it and a hollow shaft (2), which allows the heat transfer fluid to returns to the heating circuit. The heat exchange surface area of the four screws and the trough is approximately 160 m2. Heat is transferred by conduction from preheated thermal oil (230C - 260C at inlet to the heat exchanger) via the contact surfaces to the dry-mix as it is continuously transported along the axial direction by the rotating screw.The outlet temperature of the dry-mix is directly affected by many factors. The first factor being the heat exchange surface area, which can be controlled by adjusting the height to which the trough is filled. A level too low would give inadequate heat exchange area, resulting in reduction in the rate of heat transfer. On the other hand, level to high would results in the formation of cold zones, causing non-uniform heating of the dry-mix and may even lead to clogging. In practice, the indicative filling value that corresponds to the optimum height of material is to leave a few centimeters of the spiral visible. The second factor is the residence time, which must be sufficiently long (approximately 10 to 15 minutes) to allow the dry-mix to reach the required homogeneous temperature. This time is controlled by the velocity of the rotating screws which also directly impact on the production rate. The third factor is the amount of heat provided by the rotating screws and the double jacket, which can be controlled by increasing or decreasing the inlet temperature and flow rate of the thermal oil.
3.0 Description of the experiments
Figure 3 Schematic representation of the heat exchanger system
The primary aim of the experiments was to characterize the performance of the Holo-Flite heat exchanger (Figure 3) through energy and exergy analyses of the associated heat transfer processes. In the experiments, MOBILTHERM 605 oil was used as the thermal oil and its temperature-dependent density and specific heat capacity are given in Table 1. Four different flow rates of the thermal oil (90 m3/h, 100 m3/h, 110 m3/h and 120 m3/h) were investigated and at each of the flow rates, the inlet temperature of the thermal oil, T oil ,∈¿ ¿, was varied between 230C and 260C, resulting in a total of 16 test cases as illustrated in Table 2. Throughout the experiment, the production of dry-mix is maintained at a constant rate of 28000 kg/h.
Upon attaining a steady operation during each test, the outlet temperature of the thermal oil, T oil , out, was measured using PT100 resistance thermometer. The average inlet and outlet temperatures (T coke ,∈¿¿ and
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Outer casing
Coke (dry-mix)
Thermal oil
T coke ,∈¿¿T coke ,out
T oil ,∈¿ ¿ T oil , out
Qloss , To
T coke ,out) of the dry-mix (albeit surface temperature) was measured by Raytek MiniTemp MT2 infrared thermometer. The same infrared thermometer was also used to measure the average outer-casing temperature, Ts, of the heat exchanger, which varies between 68C and 74C. The constant average specific heat capacity of the dry-mix, ccoke, for the range of temperature experienced was taken as 0.879 kJ/kgK. For calculation of the mass flow rate and enthalpy change of the thermal oil passing through the heat exchanger, the arithmetic mean of the inlet and outlet temperature of the oil was used when evaluating its average density, ρoil and specific heat, coil. This assumption is justified since the temperature change of the oil is relatively small between 17C - 24C.
Table 1: Properties of MOBILTHERM® 605 thermal oil.
T (℃ ) 0 20 40 60 80 100 120 140 160
ρ (kg /m3) 871 859 847 834 823 811 798 786 774
c (kJ /kg K ) 1.81 1.89 1.96 2.03 2.11 2.18 2.25 2.32 2.40
4.0 First- and Second-law analysesThe energy efficiency of the heat exchanger can be evaluated by mean of its effectiveness in heat transfer. By definition, the effectiveness of the heat exchanger is the ratio of the actual heat transfer to the maximum possible heat transfer to the dry-mix, given by Equations (1) and (2):
Qactual=mcoke ccoke ¿and,
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Qmax=Cmin¿
where Cmin is the minimum of the heat capacity flow rates of the oil and coke streams, min[moil coil ,mcoke ccoke]. Inspection of the mass flowrates and specific heats of the two streams suggested that Cmin=mcokeccoke. Hence the effectiveness of the heat exchanger is:
ε=mcoke ccoke¿¿
Under ideal situation, the rate at which heat is rejected from the hot stream (thermal oil) should be equal to that gained by the cold stream (coke). However, in the current situation, the heat exchanger was not perfectly insulated so that heat losses to the surrounding is inevitable. This, coupled with uncertainty in the thermodynamic properties of materials and measuring devices, gives rise to heat balance error as described in [7]. However, in current situation, since the magnitude of heat losses (due to imperfect insulation) dominates over other factors, there is always a net positive loss as described in Equation (4) as unaccounted heat losses. This treatment is consistent with the approach adopted in [13].
Qloss=moil coil¿
In Equations (1) – (4), the temperatures used may be in Kelvin (K) or degree Celsius (C). It is well known that the second law of thermodynamics defines the efficiency of process equipments more restrictively than the first law. Principally, when considering heat exchangers, First-law presents the effectiveness with which energy is transferred from the hot stream to the cold stream. Analysis in terms of the second law of thermodynamics more closely describes the effectiveness with which available energy (i.e. exergy) in the hot stream is utilized. As far as the process of heating the dry-mix is concerned, the amount of entropy transfer to and from the heat exchanger system can be attributed to the mass flow of the coke and thermal oil and the heat transfer between the exchanger and its surroundings as detailed below. Thus according to standard definition, neglecting changes in the kinetic and potential energy, the entropy balance for the steady-flow heat exchanger is:
moil¿where,
soil ,∈¿−s oil, out=coil ln¿ ¿¿
and,scoke , out−scoke ,∈¿=ccoke ln ¿¿ ¿
Substituting Equations (6) and (7) into Equation (5) and rearranging gives:
Sgen=mcoke ccoke ln ¿¿
Thus, the exergy destruction can be calculated from:
X dest=T o Sgen(9)
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The second-law or exergetic efficiency [20-21] of the heat exchanger is then given by:
❑ex=1−(10)where,
¿Xdest
moil¿¿
is known as the dimensionless exergy destruction similar to the one defined in [11]. In Equation (12), the exergy change per unit mass of the thermal oil is calculated from:
xoil ,∈¿−xoil ,out=coil¿¿
For the Second-law analysis in Equations (5) – (12), all temperatures used must be in Kelvin (K). The average environment temperature T o was taken to be 303 K (30C).
5.0 Experimental Uncertainty Analysis:This section presents the assessment of uncertainty in the measured data used to calculate the First- and Second-law performance indicators described in Section 4.0. This follows the standard procedure described in [22]. If a variable Z depends on two or more independent variables (e.g. A and B), each with its own uncertainty ∆ A and ∆ B respectively, then the uncertainty in Z can be calculated based on a set of formulas for propagating the uncertainty from the independent variables. For brevity only functional forms that are relevant to the current study are listed in Table 3 here.
Table 3: Formulas for propagating uncertainty.
Functional form Formula Uncertainty formulaProduct or Quotient Z=AB or Z=A / B (∆ Z /Z )2=( ∆ A / A )2+(∆ B/B )2
Sum or Difference Z=A+B or Z=A−B (∆ Z )2=(∆ A )2+( ∆ B )2
Constant multipliers Z=nA ∆ Z=n (∆ A )Logarithmic function Z=ln A ∆ Z=∆ A / A
For more complicated functional forms, e.g. Equation (8), the uncertainty in the dependent variable Sgen can be determined by combination of the rules in Table 3. In addition, all measurement uncertainties associated with independent variables relevant for the analysis are listed in Table 4 here.
Table 4: Uncertainty in sensor measurement.
Variable Measuring device UncertaintyT oil PT100 RTD (DIN 1/10) ± 1/10(0.3+0.005 T ), T in CT coke Ray tek MiniTemp infrared thermometer ± 2 % of reading (in C) or
± 2℃ whichever is greaterV oil Fischer DE03 flow meter ± 1% of reading
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mcoke Pfister TRW-S rotor weighfeeder ± 1% of reading
Based on these information, the uncertainties in the calculated value of effectiveness (❑/❑), exergetic efficiency (❑ex /❑ex) and dimensionless exergy destruction (❑/❑) for the 16 test cases are listed in Table 5. It can be seen that the impact of measurement uncertainties on the uncertainty of the calculated effectiveness of the heat exchanger is no more than 3.2% in all 16 test cases while this was more significant for the second-law parameters (16.2% for in test case 13 and 31.7% for ex in test case 4). Thus further sensitivity analysis of the measurement uncertainty was investigated to provide a more general picture of the impact of the quality of measurements on the calculated Second-law parameters.
Table 5: Uncertainties in calculated First- and Second-law parameters.
For brevity, only comparisons for the dimensionless exergy destruction is presented here, as ratio (∆ /❑)new / (∆ /❑ )o, where (∆ /❑)o is the current uncertainty in the calculated values of ( as given in Table 5) and (∆ /❑)new corresponds to the new uncertainty values due to changes in uncertainty in the measured data. In this study, only uncertainty of one measured parameter was varied while the others remain unchanged. Firstly, Figure 4(a) illustrates the impact of changes in uncertainty in T oil alone, where it is clear that any degradation in accuracy in the measured T oil beyond 1.5 times (denoted by 1.5x) the current values would result in significant increase in the uncertainty in the calculated values of . It is also interesting to note that any improvement in accuracy (0.9x and below) in the measured T oil would not have resulted in huge improvement in the accuracy of the calculated values of as, under these situations, the ratio (∆ /❑ )new / (∆ /❑)o is closed to unity. Similar conclusions can be made with regard to the impact of changes in the measurement uncertainty in T coke, V oil and mcoke as illustrated in Figures 4(b) to 4(d). Note that following the same sensitivity analysis it was found that the uncertainty in the calculated values of the effectiveness, , was almost insensitive to changes in the measurement uncertainties.
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(a) (b)
(c) (d)
Figure 4 Impact of measurement uncertainty on the uncertainty in calculated values of dimensionless exergy destruction when, (a) uncertainty in measured T oil is changed, (b) uncertainty in measured T coke is changed, (c) uncertainty in measured V oil is changed and (d) uncertainty in measured mcoke is changed
6.0 Results and Discussions:Figures 5(a) and 5(b) show the relationships between effectiveness of the heat exchanger with respect to the inlet temperature of the thermal oil (hot stream) and outlet temperature of the coke (cold stream) for a range of mass flow rates of the thermal oil from 90 m3/h to 120 m3/h. The effectiveness is observed to be generally improved at higher mass flow rate of the hot stream. Firstly, this can be attributed to improvement in the Number of Transfer Units (NTU) as explained here. As the minimum heat capacity flow rate (due to the cold stream) is constant, the NTU of the exchanger is directly proportional to the overall heat transfer coefficient which improves at higher mass flow rate of the hot stream. Secondly, as the ratio of heat capacity flow rates Cmin/Cmax decreases at higher mass flow rate of the hot stream, the effectiveness also improved. On the other hand, although effectiveness generally improves with respect to inlet temperature of the hot stream and outlet temperature of the cold stream, however, at the highest mass flow rate of 120 m3/h, the effect was hardly noticeable above 250C. This implied that the heat exchanger has reached its heat transfer limit for the current coke production rate. This situation is also
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evident from the heat loss data presented in Figure 5(c), where lower heat loss from the heat exchanger to the environment did not result in improvement of the effectiveness.
(a) (b)
(c)
Figure 5 Relationships between effectiveness of the heat exchanger and (a) hot stream inlet temperatures (b) cold stream outlet temperatures and (c) unaccounted heat losses
Similarly, Figures 6(a) and 6(b) shows the relationships between exergetic efficiency of the heat exchanger with respect to the inlet temperature of the hot stream and outlet temperature of the cold stream for the same range of mass flow rates of the thermal oil from 90 m3/h to 120 m3/h. Clearly, the exergetic efficiency generally improved at higher mass flow rate and inlet temperature of the hot stream, which resulted in higher rate of heat and exergy transfer to the cold stream, as illustrated in Figure 6(a). The higher exergetic efficiency experienced at higher flow is also attributed to the lower exergy destruction due to the lower mean temperature difference (albeit, logarithmic mean temperature difference assuming true counter-flow configuration) between the two streams, as illustrated in Figure 7(b). Figure 6(c) also shows a strong negative correlation between the heat losses from the exchanger and its energetic efficiency. This suggests that the heat losses from the heat exchanger to the environment should be minimized as much as possible.
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(a) (b)
(c)
Figure 6 Relationships between exergetic efficiency of the heat exchanger and (a) hot stream inlet temperatures (b) cold stream outlet temperatures and (c) unaccounted heat losses
Overall, there is a strong positive correlation between the effectiveness and exergetic efficiency of the heat exchanger as illustrated in Figure 7(a), however, as agreed with other research, the exergetic efficiency is much lower than the effectiveness of the heat exchanger. In this case, the best effectiveness and exergetic efficiency of around 72% and 46% were achieved respectively, at 120 m 3/h flow rate and 250C-260C inlet temperature of the hot stream. The comparatively higher values in effectiveness suggested its inadequacy alone in accounting for energy wastage of the heat exchanger. The calculated exergy destruction ranges between 255 kW to 353 kW, which represented between 54% to 72% of the available energy of the hot stream. A large proportion of the exergy destroyed could be well attributed to the relatively high mean temperature differences (∆ T lm ,CF ranged from 114C to 143C) experienced, which was necessary to promote conductive heat transfer from the outer heat exchange surfaces to the coke. Finally, Figure 7(c) further confirm (albeit, partially) the “entropy generation paradoxes” for effectiveness greater than 0.5, as the effectiveness increases with decreasing exergy destruction (or entropy generation).
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(a) (b)
(c)
Figure 7 Graphs of (a) exergetic efficiency vs effectiveness (b) dimensionless exergy destruction vs logarithmic mean temperature difference (c) dimensionless exergy destruction vs effectiveness
7.0 Conclusions and Recommendations:Exergy and energy analysis have been carried out for the Holo-Flite counter-rotating screw heat exchanger used in Nag Hammadi Green Carbon Anode Plant. The effect of varying the inlet temperature of the thermal oil and its volume flow rate on effectiveness and exergy efficiency of the heat exchanger have been studied. In the experiments, the flow rate of thermal oil into the heat exchanger was varied from 90 m3/h, 100 m3/h, 110 m3/h and 120 m3/h and its temperature was changed from 230C to 260C with step of 10C each test. The current studies showed that the second-law parameters are more sensitive to the measurement uncertainties than that of the First-law. Further sensitivity studies also revealed that there is not much economical value in improving the accuracy of the current measurements due to the potentially high cost of more accurate measuring devices. Further results also showed that effectiveness and exergetic efficiency generally improved at higher mass flow rate and inlet temperature of the thermal oil, with the maximum values of around 72% and 46% respectively. This performance could be achieved by operating at 120 m3/h flow rate and 250C-260C inlet temperature of the hot stream. Heating up the coke from ambient temperature of around 30C to the required processing temperature (150C-200C)
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resulted in exergy destruction of between 255 kW to 353 kW, which represented around 54% to 72% of available energy of the hot stream. To increase the exergetic efficiency further, it is proposed that the heat losses from the heat exchanger to the environment be minimized and that the mean temperature difference (for the same rate of heat transfer) should be reduced by redesigning the heat exchanger.
Acknowledgement:The authors would like to thank the financial support by the Aluminum Company of Egypt, Nag Hammadi, Qena, Egypt for conducting the experimental work of this research.
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