Applied Thermal Engineering 120 (2017) 402–415 Contents lists available at ScienceDirect Applied Thermal Engineering journal homepage: www.elsevier.com/locate/apthermeng Research Paper Experiment and optimization of a new kind once-through heat recovery steam generator (HRSG) based on analysis of exergy and economy Jinbo Li, Kunyu Wang, Lin Cheng ⇑ Center of Thermal Science and Technology, Shandong University, Ji’nan 250061, Shandong, China h i g h l i g h t s A novel U-typed once-through HRSG is presented for the low temperature heat source. Heat utilization and exergy of HRSG are analyzed by experiments in a cement plant. HRSG-flash system is modeled and validated to find the optimal system solution. Exergy and economic research are conducted and contrasted for different system. a r t i c l e i n f o Article history: Received 14 January 2016 Revised 11 January 2017 Accepted 6 April 2017 Available online 8 April 2017 Keywords: Heat recovery steam generator (HRSG) Exergy efficiency Flash system Operating optimization Thermal equilibrium analysis a b s t r a c t In this work, we proposed a novel once-through heat recovery steam generator (HRSG) which could be used for low temperature heat resource recovery. Experiments have been done in a cement plant under different conditions to study its thermal performance. Exergetic and economic analyses of the HRSG have been performed. We also built a mathematical model based on the energy and mass balance equations. Moreover, a flash tank is implemented in this study and optimized researches are carried out to find the best exergy efficiency of the HRSG. Multiple sets of optimized results are compared and experiments have been done to show the rationality of the model. Results show that the HRSG is highly efficient in recovering energy from low temperature heat source. Higher feed water flow rate is more suitable for the working conditions than the lower one. When the flow rate is around 16 m3/h and the diversion water temperature is around 413 K, the highest exergy efficiency of the generating system and the HRSG can be up to 44.43% and 54.61%, respectively. The economic analysis result shows that annual income can be 121,242 € and the construction cost can be recovered in 1.1 years. Ó 2017 Elsevier Ltd. All rights reserved. 1. Introduction Energy shortage is a serious problem in social development. Energy recycling has become the focus of the energy industry. Heat recovery steam generator (HRSG) is one of the most important parts in recovering waste heat. Analysis and optimization of HRSG is a significant subject due to the increasing fuel prices and decreasing fossil fuel resources [1,2]. Increasing thermal efficiency and steam quality of HRSG has interested numerous researchers. The optimum application of energy and the energy consumption management methods are also very important. Hence, the thermodynamic , exergoeconomic , exergetic [5,6] and exergoenvironmental  analysis and optimizations have been used widely in thermal systems. Con- ⇑ Corresponding author. E-mail address: [email protected] (L. Cheng). http://dx.doi.org/10.1016/j.applthermaleng.2017.04.025 1359-4311/Ó 2017 Elsevier Ltd. All rights reserved. sidering minimizing the generating cost or maximizing the annual cash flow of the plant, the structure of heating surface [1,8] is also analyzed and optimized. Soft computing methods can be used to optimize model parameters over a full range of input and output. For design and operation of HRSG and other heat exchangers, genetic algorithm (GA) , particle swarm algorithm (PSO)  are the most widely used as optimization methods. For the optimization of the heating surface and the specific components of the HRSG, some scholars have put forward some new models or methods. Feng et al.  have presented a new algorithm model of multi-pressure HRSG. They analyzed the model characteristics by changing different feed water/steam flowing route and optimizing heat exchangers layout. Hanafizadeh et al.  have presented part elimination lattice search (PELS) method to optimize the HRSG inlet duct geometry. They also applied this method to improve a 5 MW HRSG inlet duct and enhanced the flow uniformity and efficiency. In and Sang  presented a new J. Li et al. / Applied Thermal Engineering 120 (2017) 402–415 403 Nomenclature c C CE _ Ex h K _ M P s T vp _ W gex gex, HRSG gpump specific heat capacity cost (€) unit price of electrical energy exergy flow rate (kW) specific enthalpy (kJ/kg) average investment cost of HRSG equipment mass flow rate (kg/s) pressure (MPa) specific entropy temperature (K) specific volume (m3/kg) net power (kW) exergy efficiency of the generating system (%) exergy efficiency of HRSG efficiency of the pump approach to find the optimum design parameter of a single pressure HRSG. It could maximize the exergy efficiency by minimizing the unavailable exergy. Sharma and Singh  have done the exergy analysis of the dual pressure HRSG for varying experimental steam conditions and dead states. They concluded the HP evaporator was major source of irreversibilities, which were useful to find the thermodynamic states and reduce the exergy destruction. Mansouri et al.  performed an economic and exergetic study for a double pressure and two triple pressure HRSGs (with and without reheat) in a gas-steam combined cycle power plant. They estimated the exergy loss of each component of the HRSGs. The result showed that an increase in the number of pressure levels caused a tangible increase in exergy efficiency of the system. Dumont and Heyen  built a mathematical model of an advanced once-through HRSG. They used this model to study the thermodynamic characteristics under high temperature and pressure conditions. Rovira et al.  described a one-dimensional model to simulated the performance of an once-through HRSG at supercritical pressure. The model toke account the strong variation of some thermal and transport properties of fluid and discussed their influence. Vandani et al.  added a flash tank to recover the energy content of blowdown water and made optimization for extraction pressure. The results indicated that using blowdown recovery technique, the net generated power and exergy efficiency increased. Baig et al.  built a once-through multi-stage flash distillation system. They have analyzed the effect of various design and operating conditions on the performance ratio. Most of the research and optimization of the HRSG were based on theoretical studies and model calculation. Because of the fluctuation of the working conditions in the actual production, the input parameters are difficult to match the optimized ones, and it is almost impossible to achieve the optimal results. Hence, analysis and optimization should be based on experiments and operating conditions. Hanafizadeh et al.  have done experiments to optimize the HRSG inlet duct design. They also performed numerical simulation to validate the experimental study and presented the optimized inlet duct angle and movable plate length. The cement industry is an industry that consumes a considerable quantity of resources and energy and has a very large influence on the efficient use of global resources and energy . In typical cement production process, the waste heat for the HRSG comes from the grate cooler. It is the air mixed with cement clinker particles after clinker cooling. The heat source can be divided into three sections: high temperature air, intermediate temperature air and low temperature air. For high temperature air, the temperature is about 500–900 °C, and it is mainly used as the secondary Subscripts 0 environmental conditions C condensate removal pump dp diversion point flash flash system HRSG HRSG system in inlet out outlet r rated parameter s steam w water air and tertiary air respectively into the rotary kiln and calciner [20,21]. For intermediate temperature air, the temperature is about 300–500 °C, and it is mainly used in cement plant double-pressure and multi-pressure HRSG for waste heat power generation . For the low temperature air in our experiment cement plant, in addition to low temperature (less than 300 °C), it also has the characteristics of small flow rate (less than 30,000 N m3/h) and high dust content. Mixing it with intermediate temperature air will reduce the quality of waste gas. Designing a single pressure HRSG individually comes with high cost, poor economy and low efficiency. Hence, this exhaust air in cement companies is directly discharged into the atmosphere which results in low temperature heat loss. It is found from the literature survey that researchers focused on the analysis and modeling of multi-pressure HRSG and supercritical once-through HRSG. Once-through HRSG which is suitable for low temperature conditions is in the blank stage. The present work has been carried out to fill this gap by designing and building a novel once-through HRSG-flash system. In the previous paper, the author has analyzed the advantage of U-type HRSG heating surface in inhibiting the particle deposition . In this article, we will analyze its experimental performance in terms of economy and exergy and do operating optimization under low temperature conditions. The following are the specific contributions of this paper in the subject matter area: To present a U-typed once-through HRSG for the low temperature heat source. To verify its feasibility and analyze the utilization of waste heat and heat exchange on different heating surfaces by experimental researches in a cement plant. To construct HRSG-flash system model and optimize system configurations to find the optimal cogeneration system solution. To conduct the exergy and economic aspects of the research and contrast the results of different constructions. 2. Energy analysis and exergy analysis The temperature profile of the HRSG and the energy consumption of each heating element are estimated to study the HRSG performance. For one heating surface, the energy balance equation of exhaust gas and feed water are given in Eq. (1) : _ w ðhw;out hw;in Þ _ g cg ðT g;in T g;out Þ ¼ M M ð1Þ 404 J. Li et al. / Applied Thermal Engineering 120 (2017) 402–415 _ T and h are the mass flow rate, absolute temperature and where M, specific heat, respectively. Subscript (w) and (g) refer to water and gas. It must be mentioned that the exhaust gas for the HRSG is the hot air from grate cooler. Hence, the specific heat at constant pressure, cg is considered as temperature-related variable as follows : 3:8371 9:4537 2 cg ðTÞ ¼ 1:04841 T þ T 107 104 5:49031 3 7:9298 4 þ T T 1010 1014 ð2Þ Exergy is defined as the maximum theoretical useful work, which means the work obtained by heat source is completely cooled to ambient temperature. It is defined as follows : _ ¼M _ ½ðh h0 Þ T 0 ðs s0 Þ Ex ð3Þ where s is the specific entropy and subscript (0) refers to environmental conditions. In reversible systems, it is also the theoretical work done by the temperature of waste heat completely cooled to the ambient temperature T0, which is obtained from infinite number of small Kanocycle cumulative, as Eq. (4): Z Z _ g T T 0 dT cg M T T0 T0 _ g T g;in T 0 T 0 ln T g;in ¼ cg M T0 _ ¼ Ex T g;in dEx ¼ T g;in ð4Þ Exergy efficiency of the generating system is given by: X gex ¼ n _n W ð5Þ _ Ex P _ n W n is the net power, as the following: X X _n¼ _ n ðhn hc Þ W M n ð6Þ n To calculate the exergy efficiency of HRSG, the following Eq. (7) is given by: _ Ex _ Ex gex;HRSG ¼ _ s;out _ w;in Exg;in Exg;out ð7Þ where subscript (s) refers to steam. 3. Performance experiment of the U-type once-through HRSG 3.1. Introduction for experiment According to the basic characteristics of the cement plant waste heat, our research group has designed and installed a novel oncethrough HRSG with U-type structure. It is consist of 8 heating surfaces, named as Evaporator (EVA) 1–3 and Economizer (ECO) 4–8. The experimental apparatus picture and the schematic diagram of the heat-recovering system are shown in Fig. 1. The size of a single heating surface is 2.5 m 2.5 m 2.5 m. It consists of 48 serpentine tube screens (£38 mm 3.5 mm each) and 2 headers. The exhaust gas is extracted by an induced draft fan. During the experiment, the gas flows into HRSG from the flue entrance near Evaporator 1 and continues going downward vertically until reaching the ash bucket at the bottom. After that, the movement direction of gas is changed to upward. Then it flows through Economizer 4–8 and dust collector. At last, it is discharged into the atmosphere. Feed water is supplied by two centrifugal pumps from a 32 m3 cylindrical tank and the rated flow is 32 m3/h. The flow direction of water is opposite to that of exhaust gas to ensure contraflow heat transfer. Therefore, the feed water flows from Economizer 8 to Evaporator 1. After heated by 8 heating surfaces, it flows into the steam-water separator. Then saturated steam is used to perform work in low-pressure cylinder of steam turbine, while remaining water returns to tank for a new cycle of the experiment. There are 6 platinum resistance thermometers configured in the upper and lower parts of each heating surface to measure temperature of gas, water and steam. In order to ensure the accuracy of temperature measurement, the average value is taken as the experimental temperature. A turbine flow meter is installed at the feed water entrance of boiler and a vortex street flow meter is installed at the outlet of steam-water separator to measure the feed water and steam production, respectively. Pressure sensors are installed on each entrance of evaporative heating surfaces to measure the pressure of saturated water. Detailed parameters of the measurement instruments are summarized in Table 1. 3.2. Analysis of HRSG pitch point and approach point The pinch point temperature is the difference between the saturation temperature and the gas temperature at the economizer inlet. We have chosen two experimental conditions to reflect the pitch point temperature. The temperature of U-type HRSG heating surfaces is shown in Table 2 and Fig. 2(a) and (b). It can be seen from the figures that the pinch point temperature of the U-type once-through HRSG is 26.5 °C. The once-through HRSG is different from the drum HRSG. The feed water gradually completes the phase change in different heating surfaces instead of drum. When the feed water starts to change from simple fluid to steam/water mixtures, the transition curve is smooth without obvious inflection point. It is shown as the curves between the fourth and fifth red point in Fig. 2. When the pinch point temperature decreases, the outlet gas temperature of the HRSG will decline. Then, the recovery of the waste heat from the exhaust gas will increase, as well as the output of power of the steam turbine and the HRSG efficiency. However, it will increase the heat exchange area of the HRSG and reduce the economic performance. Hence, the pinch point temperature is not allowed to be equal to 0 °C. Otherwise, the heat exchange area of the HRSG will be infinite. Approach point temperature refers to the difference between the water temperature at the economizer outlet and the saturated water temperature. If the approach point temperature increases, the total heat transfer area of the HRSG will increase, and it will lead to increase of the investment cost. If the temperature is too small, under the low load condition or during the start-up period, the feed water may vaporize in the economizer, which is not allowed for the single pressure or multi-pressure HRSG. Hence, the selection of the approach point temperature is a difficult task, because it can lead to economic or operational problems. The drum HRSG is not suitable for low load conditions. For our working conditions in the cement plant, the exhaust gas temperature is 200–300 °C, and mass flow rate is 20,000– 35,000 N m3/h, which are much smaller than the input parameters of the HRSG in combined cycle power plant. It is a typical low load conditions. Therefore, our research group designed a once-through HRSG, which could avoid the choice of approach point temperature. For the once-through HRSG, the feed water is gradually changed into steam through the heating of the exhaust gas in the heating surface. There is no obvious boundary between the economizer and the evaporator. Hence, there is no obvious approach point temperature. 3.3. Experimental data and working process of the U-type oncethrough HRSG In this article, thermal performance of the HRSG is studied under different working conditions shown in Table 3. 405 J. Li et al. / Applied Thermal Engineering 120 (2017) 402–415 P F Steam turbine T T T P Steam-water separator P EVA 1 T ECO 8 EVA 1 EVA 2 T ECO 7 EVA 3 T ECO 6 P Condenser Condensate pump P Water tank ECO 5 ECO 4 Steam flow Water flow Gas flow F P Pressure transmitter T Temperature thermometer F P T Feed water pump Flow meter Fig. 1. Experimental apparatus picture and schematic diagram of the waste-heat-recovering system. Table 1 Model, range and precision of experiment instrument. Equipment Model Range Unit Precision Thermal resistance Turbine flow meter Vortex street flow meter Pressure sensor Pt100 LWGY-80 DY100 511 0–450 0–100 0–8.3 0–1.6 °C m3/h t/h MPa ±0.1 °C ±0.5% ±0.75% ±0.3% Table 2 Inlet parameters of different operating conditions. Experimental condition in Exhaust gas inlet temperature, Tg, (°C) _g Exhaust gas mass flow rate, M (N m3/h) Feed water inlet temperature, Tw, in (°C) _w Feed water mass flow rate, M (t/h) A B 286.50 333.15 23,800 23,900 96.6 74.3 4.56 5.17 406 J. Li et al. / Applied Thermal Engineering 120 (2017) 402–415 300 350 280 exhaust gas feed water 260 Temperature T (°C) Temperature T (°C) 240 exhaust gas feed water 300 220 200 180 160 250 200 150 140 100 120 100 50 0 200 400 600 800 1000 1200 1400 0 500 1000 1500 2000 Total heat transfer rate (kW) Total heat transfer rate (kW) (a) (b) Fig. 2. Temperature of feed water and exhaust gas on different heating surfaces: (a) condition A; and (b) condition B. On the other, through the study in experiment, the measure curves of temperature and steam production are obtained in Figs. 3–5. Fig. 3 shows that exhaust gas temperature of different heating surfaces varies over time under test conditions. It can be seen that temperature fluctuation of exhaust gas is obvious, especially for the gas inlet temperature (EVA 1 Inlet). The reason is that the exhaust gas is from the grate cooler. Temperature of the cement clinker changes large, and exhaust gas temperature has the same trend. Fig. 4 shows the temperature of water/steam varies over time. The comparison among feed water temperature turns out that the stable temperature decreases from Evaporator 3 to Evaporator 1, while flue gas temperature and the quantity of heat exchange have the opposite change. The reason is that in these three heating surface, steam/water mixtures are the final products, whose temperature is determined by steam pressure. For single-phase heating surface (ECO 4–8), the pressure loss can be ignored. However, flow resistance of steam is much greater than that of feed water. In evaporative heating surfaces, the pressure drop is obvious, which leads to the temperature decrease rather than increase, even though the feed water has absorbed more heat and gas temperature is higher than that in economizers. Steam production increases in spite of the gradual lowering of the water-vapor temperature from Evaporator 3–1. For the two operating conditions 1 and 2, the feed water pressure is 0.65 MPa, but there is a big gap between the temperature and pressure of the final steam. The reason is that when the flow rate is around 4 m3/h, the feed water reaches saturated state in Economizer 5 and then the HRSG begins to produce steam. With the gradual increase of the steam/water mixture, the pressure drop caused by the steam is gradually increased (0–0.23 MPa). When the flow rate is around 16 m3/h, steam/water will not be generated until in Evaporator 3, the pres- Table 3 Average flow rate of exhaust gas, feed water and steam of typical working conditions. Parameters Unit Condition 1 Condition 2 _g M _w M N m /h 32,222 29,000 m3/h 15.87 4.02 _s M t/h 2.44 1.82 3 sure drop is less than that under condition 2. The final steam temperature is decided by the pressure. Fig. 5 shows the flow rate changing of feed water and steam over time. It shows that the volume of feed water and steam is negatively correlated. This is because the flow rate and the temperature of the waste heat are low and certain. The more feed water, the more waste heat can be utilized and the thermal efficiency can be higher. But the more feed water, the less steam can be generated. When the flow rate is around 16 m3/h, the mass flow of steam is around 2.5 t/h. When the flow rate is around 4 m3/h, the mass flow of steam is around 1.8 t/h. For the two operating conditions 1 and 2, the flow rate and inlet temperature of exhaust gas are almost the same, but the final steam production have a great difference. This is because less flow rate and lower velocity under condition 2 reduce the convective heat transfer coefficient, which makes the waste heat in the exhaust gas poorly utilized. It is obvious that temperature and flow rate tend to be stable after 325 min. Averaging all data during 325–375 min and the result is given in Tables 4 and 5. 3.4. Experimental results and analysis Figs. 6 and 7 show released heat of exhaust gas and exergy destruction in different heating surfaces. For both condition 1 and 2, these two parameters of the heating surfaces show the similar trend. They both decrease gradually from Evaporator 1 to Economizer 8. The reason is that in the process of heat transfer, the temperature difference between exhaust gas and feed water decreases gradually, which is shown in Table 4. But for Economizer 8 under condition 2, the heat transfer and exergy destruction are higher than those of Economizer 7. The water flow rate is around 4 t/h, and inlet water temperature is 90.61 °C. The inlet temperature difference between exhaust gas and feed water of Economizer 8 is 37.89 °C which is higher than that under condition 1 (9.75 °C). Hence, the parameters are higher. From Fig. 6, it can be seen that the highest heat exchange rate appears in Evaporator 1. From Fig. 7, exergy loss in evaporator 1–3 is significantly higher than that in the economizers. This is because feed water is in saturated state in evaporators. The temperature difference between exhaust gas and feed water is large, so as the irreversibility of the heat exchange surface. 407 J. Li et al. / Applied Thermal Engineering 120 (2017) 402–415 350 350 300 300 250 250 200 200 150 150 100 EVA 1 Inlet EVA 2 Inlet EVA 3 Inlet ECO 4 Inlet ECO 5 Inlet HRSG Outlet 50 0 0 50 100 150 200 250 300 350 400 450 500 100 EVA 1 Inlet EVA 2 Inlet EVA 3 Inlet ECO 4 Inlet ECO 5 Inlet HRSG Outlet 50 0 0 50 100 150 Time (min) 200 250 300 350 400 450 400 450 Time (min) (a) (b) Fig. 3. Temperature curves of heat source under typical operating conditions: (a) condition 1; and (b) condition 2. 180 180 160 160 140 140 120 120 100 100 EVA 1 Outlet EVA 2 Outlet EVA 3 Outlet ECO 4 Outlet ECO 5 Outlet HRSG Inlet 80 60 EVA 1 Outlet EVA 2 Outlet EVA 3 Outlet ECO 4 Outlet ECO 5 Outlet HRSG Inlet 80 60 40 40 0 50 100 150 200 250 300 350 400 450 500 Time (min) (a) 0 50 100 150 200 250 300 350 Time (min) (b) Fig. 4. Temperature curves of water and steam: (a) condition 1; and (b) condition 2. Table 6 shows the exhaust parameters of steam turbine which are measured in the cement plant. Available steam power can be given according to the result calculated in Eq. (8). Exergy efficiency of generating system and HRSG are defined as the following Eqs. (9)–(11). The result is shown in Table 7. From the experimental results, low temperature waste heat can be recovered by the U-type once-through HRSG. Under the experimental condition, the power generated by the steam cycle is 327.35 kW, and the exergy efficiency of generating system can be 31.05%. When the feed water flow rate is low (condition 2), steam production and available steam power are both low. Hence, utilization of waste heat is poor. 4. U-type once-through HRSG-flash system model equation and optimization Through the analysis of pitch point and approach point temperature, it is concluded that for the condition of low mass flow rate and low temperature exhaust gas, the once-through HRSG can be a good choice. For the once-through HRSG, the feed water is gradually changed into steam through the heating of the exhaust gas in the heating surface. During the low water flow experiment, steam/ water mixture can be produced in Economizer 5 or 6. When the mixture flows through Economizer 4 and Evaporator 3–1 in turn, it will have large pressure loss, which leads to low vapor pressure and poor steam quality. Besides, the low water flow rate can also lead to low convective heat transfer coefficient and low waste heat utilization efficiency. Hence, in the experiment, we choose high feed water flow rate. In the Economizer 4–8, the feed water can fully utilize waste heat and steam/water mixture will be generated in evaporating heating surfaces (Evaporator 1–3). In this way, large vapor pressure loss is avoid and steam quality is improved, while high temperature water contains much heat which is not fully utilized. In order to utilize this part of the waste heat and improve exergy efficiency, a HRSG-flash system is designed. In this paper, a bypass valve is installed and different water flow can be dispensed into the heat recovery steam generator and the flash evaporator respectively with energy balance discipline. The water flow 408 J. Li et al. / Applied Thermal Engineering 120 (2017) 402–415 50 3.0 7 45 2.5 6 2.0 5 2.5 1.5 Feed water Steam 30 25 1.0 20 0.5 2.0 1.5 4 3 1.0 Feed water Steam 2 Steam flow (t/h) 35 Steam flow (t/h) Feed water flow (m3/h) Feed water flow (m3/h) 40 0.5 1 15 0.0 10 0 50 100 150 200 250 300 350 400 450 500 0.0 0 0 50 100 150 Time (min) 200 250 300 350 400 450 Time (min) (b) (a) Fig. 5. Flow curves of water and steam: (a) condition 1; and (b) condition 2. Table 4 Exhaust gas and water/steam properties of different heating surface. Heating surface Ta,in (°C) Ta,out (°C) Absorbed heat (kW) Exergy destruction rate (kW) Tw,out (°C) Tw,in (°C) Pw (MPa) Condition 1 1 2 3 4 5 6 7 8 311.61 232.12 199.96 174.76 156.66 139.63 127.43 118.11 232.12 199.96 174.76 156.66 139.63 127.43 118.11 109.95 955.42 382.34 298.11 213.40 200.25 143.16 109.21 95.52 431.79 149.18 105.08 69.33 58.56 38.19 26.97 21.99 152.58 158.73 161.82 147.19 136.52 125.93 118.85 113.35 158.73 161.82 147.19 136.52 125.93 118.85 113.35 108.36 0.51 0.6 0.65 0.65 0.65 0.65 0.65 0.65 Condition 2 1 2 3 4 5 6 7 8 292.85 239.93 219.05 193.65 179.25 166.42 152.75 142.55 239.93 219.05 193.65 179.25 166.42 152.75 142.55 128.50 536.99 212.38 257.23 146.12 130.39 138.50 103.50 142.57 273.87 85.61 96.37 50.86 42.83 42.68 29.89 38.20 145.21 152.40 154.43 157.52 161.01 156.19 132.82 114.82 152.40 154.43 157.52 161.01 156.19 132.82 114.82 90.61 0.42 0.51 0.54 0.58 0.63 0.65 0.65 0.65 Table 5 Ambient temperature and pressure parameters. Parameters Unit Value T0 P0 h0 s0 °C MPa kJ/kg kJ/(kg K) 25 0.10 104.93 0.3672 to the heat recovery steam generator is completely converted to high temperature saturated steam, while the other to be low temperature saturated steam through flash. The once-through HRSGflash optimization model is constructed taking exergy efficiency of system as the optimization goal based on mass and energy conservation equations. 4.1. Objective function, parameters design and constraints In this paper, a general calculation for thermodynamic performance of HRSG-flash system composed of the 8 heating surfaces is designed based on the existing experimental data. For given _ g Þ, water/ exhaust gas inlet temperature (Tg,in), mass flow rate ðM steam temperature and pressure (Tw,in, Ts,out and Pw), the main characteristics and assumptions are made as the followings: _ r Þ of steam is certain. Steam outFor the HRSG, the rated flow ðM _ s;HRSG Þ of the HRSG in the model is less than Dr. The pressure put ðM of flash steam (Ps,flash) is equal to that of supplementary steam. For the HRSG-flash system, even though the temperature of the feed water varies with position and condition of heating surfaces, the water temperature of diversion point (Tw,dp) is less than the saturated temperature (TP) corresponding with feed water pressure and more than the temperature of flash steam, which is calculated through thermodynamic properties of water by IAPWS-IF97. The feed water is diverged at the diversion point. Part of the water, after the shunt valve, re-enters the next heating surface of the HRSG, and is heated to high pressure saturated steam. A certain amount of water flows into the flash tank through the shunt valve and nozzle, resulting in low pressure saturated steam. The remaining water after flash is conveyed to the water tank by the pump at the bottom of the flash tank, and continues the next round of cycle. Ignore water-side pressure drop. The steps to adjust the results are given as: 1. Select the diversion temperature point and apply the energy balance principle for the exhaust gas and water/steam in the exchanger from the point to Economizer 8. Next, take the diversion 409 J. Li et al. / Applied Thermal Engineering 120 (2017) 402–415 1000 1000 900 900 800 800 Absorbed heat (kW) 1100 Absorbed heat rate (kW) 1100 700 600 500 400 300 700 600 500 400 300 200 200 100 100 0 1 2 3 4 5 6 7 0 8 1 2 3 4 5 Heat surface No. Heat surface No. (a) (b) 6 7 8 Fig. 6. Absorbed heat rate in HRSG components: (a) condition 1; and (b) condition 2. 400 Exergy destruction rate (kW) Exergy destruction rate (kW) 400 300 200 100 300 200 100 0 0 1 2 3 4 5 6 7 8 1 2 3 4 5 6 7 8 Heat surface No. Heat surface No. (a) (b) Fig. 7. Exergy destruction rate for HRSG components: (a) condition 1; and (b) condition 2. point of water temperature Tw,dp as a temperature reference to calculate the gas temperature Tg,dp at this point. It is shown in Eq. (8): Table 6 Exhaust steam parameters of steam turbine. Parameter Value Unit _ w ðhw;dp hw;in Þ ¼ M _ g cg ðT g;dp T g;out Þ M Steam turbine vacuum degree, Pv Internal efficiency of steam turbine, gt Pump efficiency, gpump Steam turbine exhaust temperature, Tk Steam turbine exhaust enthalpy, hks 9 85 85 49.4 2265.58 % % % °C kJ/kg 2. Eq. (9) is based on the energy balance of the heat exchanger from Evaporator 1 to the diversion point to calculate the steam _ s1 of HRSG. production M ð9Þ Then, quantity of water to the flash system can be estimated as follows: Table 7 Available steam power and exergy efficiency of different conditions. Parameter Condition 1 Condition 2 Unit _ Available steam power, W Exergy efficiency of generating system, 327.35 240.38 kW 31.05 30.07 % gex _ s;HRSG hs;out hw;dp ¼ M _ g cg T g;in T g;dp M ð8Þ _ w;flash ¼ M _ wM _ s;HRSG M ð10Þ 3. Applying the first law of thermodynamic to the flash tank as an adiabatic control volume, calculation formula of steam produced by flash system is shown in Eq. (11). _ w;flash hw;dp ¼ M _ s;flash hs;flash þ ðM _ w;flash M _ s;flash Þhw;flash M ð11Þ 410 J. Li et al. / Applied Thermal Engineering 120 (2017) 402–415 Steam power and exergy efficiency are calculated based on the quantity of high pressure steam generated by HRSG and low pressure steam from flash system. And optimal water temperature of diversion point corresponding to the highest exergy efficiency can be given in exhaustive method. Start 43.0 Exergy efficiency (%) Flow diagram and the calculation results are shown in Figs. 8– 10 through FORTRAN programming calculation. It is obtained from the following figures that: under a certain exhaust flow, different selection of the feed water flow rate and water temperature at diversion point cause different system exergy efficiency. There is an optimal value. As shown in Figs. 9 and 10, the system exergy efficiency is the highest when the flow rate is around 16 m3/h with the diversion point water temperature about 413 K. Fig. 9 shows the relationship between the flow rate of feed water and system exergy efficiency. The increase of flow rate means increase of velocity of flow and the heat transfer coefficient. In this way, the waste heat can be utilized better and the outlet temperature of the waste gas decreases. More saturated steam can be made, and the system exergy efficiency increases. When the flow rate is around 16 m3/h, the system exergy efficiency is the highest. Then when the flow rate is more than 16 m3/h, the influence of heat transfer coefficient is very little. But the waste heat is not enough and saturated vapor reduces. As shown in Fig. 10, exergy efficiency presents a linear rise followed by a downward trend, and the highest point (corresponding water temperature of diversion point 413 K) is 43.4%. The reason is that: when the temperature of diversion point increases to 413 K, the feed water can be heated to a higher temperature before it is diverged. Water flow rate and water temperature of the flash sys- 43.2 42.8 42.6 42.4 42.2 42.0 10 12 14 16 18 20 Water flow rate (t/h) Fig. 9. Optimal exergy efficiency of HRSG-flash system on different water flow rate. 44 43 System exergy efficiency (%) 4.2. Results and analysis of model calculation 43.4 42 41 40 39 38 37 36 35 380 Energy consumption calculation of HRSG 390 400 410 420 430 440 Temperature (K) Fig. 10. Exergy efficiency changing with diversion point water temperature. Diversion point water temperature Tw,dp tem will increase, which leads to a higher system exergy efficiency. When the temperature of diversion water is over 413 K, with the increase of temperature, the water flow rate of the HRSG will reduce and main steam output will fall. Most of the feed water flows into the flash tank and becomes low pressure steam, which reduce the quality of steam and system exergy efficiency. Exergy efficiency calculation Tw,dp >Tp No Changing diversion point water temperature Tw,dp=Tw,dp+1 Yes Calculation result End Fig. 8. Flow chart of HRSG-flash system calculation. 5. U-type once-through HRSG-flash system experiment validation In order to validate the optimization scheme of different conditions, our group have designed and built a flash tank and connected it with the HRSG. The high temperature and high pressure spray flash experimental system is shown in Fig. 11. The main body is a cylindrical cylinder with a standard ellipsoidal head. It has a diameter of 2.5 m and a total height of 5.33 m. Its material is Q345R, and the wall thickness is 30 mm, which can withstand pressure up to 3 MPa. The tank and pipe are covered by insulation material with a thickness of 100 mm. There are two water supply pipelines of the flash tank, which are symmetrically arranged on upper and lower of the cylinder body. Hence, it can meet the liquid downward or upward jet requirements. 411 J. Li et al. / Applied Thermal Engineering 120 (2017) 402–415 LP cylinders F Flow meter P Pressure transmitter T Temperature thermometer T P Electric valve Pump To HRSG F P T P T Shunt valve Nozzle Flash Evaporator HRSG Valve Water tank Fig. 11. Schematic diagram and photo of spray flash evaporation experimental system. The specific working process is as follows: The feed water is heated in the HRSG to the target temperature. Through the calculation of the amount of exhaust gas, water supply and temperature parameters, we select the appropriate heating surface, and change the opening of the valve between the heating surfaces. When the feed water flows out from this heating surface, it does not enter the next heating surface directly, but flows through the shunt valve. Part of the water, after the shunt valve, re-enters the next heating surface of the HRSG, and is heated to saturated steam. A certain amount of water flows into the flash tank through the shunt valve and nozzle. Then flash occurs. Flash steam is transferred through the pipeline at the top of the tank. And it is used as supplementary steam in the low pressure cylinder of steam turbine. The remaining water is conveyed to the water tank by the pump at the bottom of the flash tank, and continues the next round of cycle. The connection mode and quantity of heating surfaces in the experiment are determined based on the theoretical calculation results and experimental conditions. The alternative locations of diversion point are displayed as follows: 2# experimental system: diversion point is between Evaporator 3 outlet and Evaporator 2 inlet. 3# experimental system: diversion point is between Economizer 4 outlet and Evaporator 3 inlet. 4# experimental system: diversion point is between Economizer 5 outlet and Economizer 4 inlet. Experiments are carried out based on the three improved schemes above. Different feed water flow rates and diversion point temperature are selected for the experiment validation. Production of high and low pressure saturated steam is measured. The exergy of system and the HRSG are calculated and contrasted with each other. Table 8 describes the different conditions of the flow rate and diversion point temperature. Optimized architecture of HRSG-flash system is illustrated in Fig. 12. Fig. 13 shows comparison of system exergy efficiency between different feed water flow rate and diversion point temperature. It is evident that system exergy efficiency has an upward trend when the flow rate changes from 11.70 m3/h to 15.90 m3/h; but when the flow rate changes from 15.90 m3/h to 17.90 m3/h and 19.70 m3/h, the system exergy efficiency decreases. Furthermore, for a Table 8 Water flow rate and diversion point temperature of different experimental system. Water flow rate (m3/h) Temperature of diversion point (K) 2# experimental system 3# experimental system 4# experimental system 11.70 15.90 17.90 19.70 426.25 434.97 424.43 418.46 418.96 420.34 409.05 402.95 407.08 409.67 400.22 395.19 single flow rate condition, the result shows that there is an optimal diversion point temperature corresponding to the highest system exergy efficiency. The experiment results are in good agreement with the simulation ones. The result also shows that the optimal diversion point temperature decreases when the water flow rate increases. By means of analyzing the experiment condition of 15.9 m3/h feed water using different experiment schemes, results of steam and exergy efficiency are shown in Tables 9–11, separately. 6. Result and discussion 6.1. Exergy analyses results As shown in Tables 4, 9 and 10, the low temperature waste gas from grate cooler which was discharged into the atmosphere directly is efficiently utilized through the U-type once-through heat recovery steam generator. High-temperature saturated steam production rate is 2.44 t/h, and the generated steam power is 327.35 kW, under working condition of feed water flow 15.9 m3/ h, hot air flow 32,222 N m3/h, inlet hot air temperature 311.61 °C. The exergy efficiency of generation system and HRSG are 31.05% and 33.64%, respectively. Available steam power and exergy efficiency have both substantial increase with configuration optimization based on numerical model. As shown in Fig. 14, comparison is made between 3 schemes. From schemes 2 to 4, the temperature of the diversion point decreases gradually. Available steam power of the HRSG steam increases, while that of flash steam decreases. Based on Tables 10 and 11 and Fig. 15, the overall quality of steam is relatively improved (due to increased high pressure 412 J. Li et al. / Applied Thermal Engineering 120 (2017) 402–415 High pressure HRSG steam High pressure HRSG steam ECO 8 EVA 1 Low pressure falsh steam EVA 1 EVA 2 Low pressure falsh steam ECO 7 EVA 3 ECO 6 ECO 4 ECO 5 Shunt valve EVA 1 ECO 8 EVA 1 EVA 2 ECO 7 EVA 3 ECO 6 ECO 4 ECO 5 Flash evaporator Shunt valve Condensated water Flash evaporator Condensated water Water tank Water tank Water flow Water flow Steam flow Gas flow Steam flow Gas flow Feed water pump Optimized cogeneration system 2# Feed water pump Optimized cogeneration system 3# High pressure HRSG steam Low pressure falsh steam Shunt valve Flash evaporator Condensated water EVA 1 ECO 8 EVA 1 EVA 2 ECO 7 EVA 3 ECO 6 ECO 4 ECO 5 Water tank Water flow Steam flow Gas flow Feed water pump Optimized cogeneration system 4# Fig. 12. Optimized cement kiln waste-heat-recovering system. steam) with the decreasing temperature of the shunt. Available steam power, as well as exergy efficiency of system and HRSG all increase and the optimal value is achieved in 4# experiment schemes, being 44.43% and 54.61% respectively. 6.2. Economic analyses results Economic analyses are important to determine whether heat transfer equipment makes sense. As mentioned above, it is necessary to do calculation on economic aspects of generation system to have a better evaluation about the new once-through heat recovery steam generator and its different configurations. Cost estimation can have a major impact both on project profitability and influences of the technical solution. The total capital investment of the once-through HRSG-flash experimental system Ctot contains three parts: the tube bundles cost Ct, the shell cost Cs and the insulation cost Cins. The capital investment can be expressed as: C tot ¼ C t þ C s þ C ins ð12Þ The size of a single heating surface is 2.5 m 2.5 m 2.5 m. The tube bundles consist of 48 serpentine tube screens (£38 mm 3.5 mm each) and 2 headers (£89 mm 6.0 mm each). Hence, the tube bundles cost is calculated as follows: 413 J. Li et al. / Applied Thermal Engineering 120 (2017) 402–415 Table 11 Steam exergy and exergy efficiency of HRSG. System exergy efficiency (%) 44.0 11.7 m3/h 15.9 m3/h 17.9 m3/h 19.7 m3/h 43.5 43.0 Experiment schemes High-temperature saturated steam (kJ/kg) Flash saturated steam (kJ/kg) HRSG exergy efficiency (%) 1# 2# 3# 4# 302.70 258.72 340.69 396.67 0.00 201.03 138.02 94.73 33.64 51.09 53.20 54.61 42.5 550 42.0 500 450 41.5 HRSG steam Flash steam Total 41.0 395 400 405 410 415 420 425 430 435 Temperature (K) Fig. 13. System exergy efficiency of different flow rate and diversion point temperature. Steam power (kW) 400 350 300 250 200 150 100 Table 9 Steam production of different experiment scheme. 50 Experiment schemes HRSG saturated steam (t/ h) Flash saturated steam (t/ h) 2# 3# 4# 2.22 2.63 2.91 1.33 0.91 0.63 0 1 2 3 4 Experiment scheme No. Fig. 14. Available steam power of different optimized cogeneration system. 60 HRSG HRSG generation system ð13Þ where Cpm1 is the price of the tube structural material, wwt and whd represent the tube weight per unit length of the serpentine tube screens and headers, respectively. Nwt and Nhd are the number of those. Ns is the heating surfaces number of the HRSG, Ns = 8. For the shell and insulation material cost, it includes the following parts: heating surfaces, flue duct, water tank and flash tank. Shell cost and thermal insulation material cost formulas are as follows: C s ¼ C pm2 qm2 ðNs Ass tss þ Afd tfd þ Afl tfl þ Atk t tk Þ ð14Þ C ins ¼ C pins tins ðNs Ass þ Afd þ Afl þ Atk Þ ð15Þ Exergy efficiency (%) C t ¼ C pm1 ðwwt lwt Nwt þ whd lhd Nhd ÞNs 50 40 30 where Cp and q are the unit price and density of different materials, respectively. Subscript (m2) and (ins) are the metal and insulation. A represents the total area and t is the thickness of shell metal or insulation layer. The parameters of the waste-heat-recovering system structure and materials for the economic analysis are shown in Table 12. The total capital investment of the heat recovery system are 45346.23 € (without flash system) and 52153.10 € (with flash system), respectively. The annual operational cost can be devised into Cpw and Cpg. Cpw comes from pump work and it is expressed as: 1 2 3 4 Experiment scheme No. Fig. 15. Exergy efficiency of different optimized cogeneration system. C pw ¼ _ w DP w M C t q gpump e p ð16Þ where q is the fluid density and DPw is the pressure drop of feed water in the whole cycle. gpump is the pump efficiency. tp is the operating time per year. Ce is the price of the electrical energy. Table 10 Available steam power and exergy efficiency of system. Experiment schemes HRSG high-pressure saturated steam (kW) Flash saturated steam (kW) Total steam power (kW) System exergy efficiency (%) 1# 2# 3# 4# 327.35 297.83 352.84 390.40 0.00 157.94 108.44 74.43 327.35 455.77 461.27 464.83 31.05 43.23 43.75 44.43 414 J. Li et al. / Applied Thermal Engineering 120 (2017) 402–415 Cpg is caused by the induced draft fans and it is expressed as: C pg D_ g DPg ¼ K C e tp 36; 00; 000 gf ð17Þ where DPg is the exhaust gas pressure drop of HRSG. g is the total pressure efficiency. K is the motor capacity factor. Whether only HRSG or HRSG-flash system, the final product is saturated steam used for electricity generation in the low pressure cylinder of steam turbine. Hence, the system profit, Cga is the economic benefits generated by waste heat power generation, which is expressed in Eq. (18). Table 13 shows parameters about system operational cost and profit calculation. C ga ¼ ðW b þ W f Þ gt gg C e t p ð18Þ where Wb and Wf are the net power done by boiler steam and flash steam. gt and gg are the steam turbine efficiency and generator efficiency, respectively. The experiment measurements system is shown in Table 1. The measurement data includes the mass flow rate and the pressure drop of the HRSG and flash system in Tables 3–5. Taking the above experimental analysis in Section 3 as the representative average operating conditions of the waste heat recovery system for the year analysis, cost recovery is calculated. Assuming that it takes n years to start profit, and investment capital annual interest rate i is15%. The condition for achieving profitability is shown in Eq. (19) and the economic analysis result of different systems is shown in Table 14 . n ðC ga C pw C pg Þ P ið1 þ iÞ C tot n ð1 þ iÞ 1 ð19Þ Table 12 Waste-heat-recovering system structure and materials parameters. Parameters Value Unit Cpm1 Cpm2 Cins qm1, qm2 tss, tfd ttk tfl tins lwt lhd Ass Afd Atk Afl 0.47 0.54 53.98 7850 10 12 30 100 56.64 3.1 25 30.91 71.97 51.33 €/kg €/kg €/m3 kg/m3 mm mm mm mm m m m2 m2 m2 m2 Table 13 The annual operational cost parameters of exhaust gas and feed water. Parameters Value Unit Ce tp 0.09 4800 85 1.1 95 €/(kW h) h % – % gf K gf As it is shown in Table 14, with this heat recovery system, 200– 300 °C low temperature waste heat of the cement industry can be recovered well with an annual income of 85,383 €. The construction cost can be recovered in 2.7 years. After flash system optimization, the optimal results show that annual income can be 121,242 € and only 1.1 years is needed to recover construction cost. The U-type once through HRSG can make effective use of low temperature waste heat from grate cooler to produce saturated steam and to generate electricity. After incorporating flash tank into waste heat system and configuration optimization of different shunts, the corresponding exergy efficiency of system and HRSG achieve the optimal values. 7. Conclusion Based on the above-mentioned results, related conclusions are summarized as the following: 1. A novel U-type heat recovery steam generator is proposed, which is efficient in recovering the low temperature waste heat originally emitted into the atmosphere. Experiments results indicate that the heat transfer performance and exergy loss of each heating surface in the heat recovery steam generator system go down with the decrease of heat source temperature. Exergy loss of the remaining heat transfer surface, except Evaporator 1 is small, which indicates low irreversibility of the heat surface. The exergy efficiency of the system and heat recovery steam generator can be 31.05% and 33.64% respectively in typical experimental condition. 2. HRSG-flash system model is applicable and favorable to achieve system exergy efficiency optimization based on mass and energy balance equation. Optimized system schemes are studied by numerical simulation and experiments. Exergy efficiencies of generation system and HRSG of the optimized HRSGflash system are both higher than those when the HRSG separately operates. The corresponding optimal efficiency is 44.43% and 54.61%, respectively. It is obviously that the utilization efficiency of waste heat is enhanced. 3. The economic analysis results show that the application of the HRSG makes an annual income of over 85,383 € and it takes no more than 3 year to recover the cost of building the boiler. After optimization, the annual income of the combined HRSGflash system can be 121,242 € and the construction cost can be recovered in 1.1 years. With the rapid development of the industrial production, the quantity of the low temperature waste gases will increase and the quality of the low temperature waste gases will decrease. Hence, related researches in this paper provide a reference for once-through HRSG in low temperature conditions, which is of great significance and economic value in the fields of cement, steel and other industries. Acknowledgements Funding for this research was provided by the National Basic Research Program (973 Program) of China (Project No.: 2013CB228305). References Table 14 Economic performance and payback period of the heat recovery system. Experimental program Ctot (€/yr) Cpw (€/yr) Cpg (€/yr) Cga (€/yr) n (yr) HRSG system HRSG-flash system 45346.23 52153.10 1434.24 1434.24 62117.28 62117.28 85383.36 121242.5 2.7 1.1  A. Behbahani-Nia, M. Bagheri, R. Bahrampoury, Optimization of fire tube heat recovery steam generators for cogeneration plants through genetic algorithm, Appl. Therm. Eng. 30 (2010) 2378–2385.  P. Ahmadi, M.A. Rosen, I. Dincer, Greenhouse gas emission and exergoenvironmental analyses of a trigeneration energy system, Int. J. Greenhouse Gas Control 5 (2011) 1540–1549. J. Li et al. / Applied Thermal Engineering 120 (2017) 402–415  M. Mohagheghi, J. Shayegan, Thermodynamic optimization of design variables and heat exchangers layout in HRSGs for CCGT using genetic algorithm, Appl. Therm. Eng. 29 (2009) 290–299.  M. Ghazi, P. Ahmadi, A.F. Sotoodeh, A. Taherkhani, Modeling and thermoeconomic optimization of heat recovery heat exchangers using a multimodal genetic algorithm, Energy Convers. Manage. 58 (2012) 149–156.  V. Cenu, A. Badea, M. Feidt, R. Benelmir, Exergetic optimization of the heat recovery steam generators by imposing the total heat transfer area, Int. J. Thermodyn. 7 (2004) 149–156.  H. Hajabdollahi, P. Ahmadi, I. Dincer, An exergy-based multi-objective optimization of a heat recovery steam generator (HRSG) in a combined cycle power plant (CCPP) using evolutionary algorithm, Int. J. Green Energy 8 (2011) 44–64.  A.G. Kaviri, M.N.M. Jaafar, T.M. Lazim, H. Barzegaravval, Exergoenvironmental optimization of heat recovery steam generators in combined cycle power plant through energy and exergy analysis, Energy Convers. Manage. 67 (2013) 27– 33.  M.D. Durán, M. Valdés, A. Rovira, E. Rincón, A methodology for the geometric design of heat recovery steam generators applying genetic algorithms, Appl. Therm. Eng. 52 (2013) 77–83.  A.M.K. Vandani, M. Bidi, F. Ahmadi, Exergy analysis and evolutionary optimization of boiler blowdown heat recovery in steam power plants, Energy Convers. Manage. 106 (2015) 1–9.  H. Feng, W. Zhong, Y. Wu, S. Tong, Thermodynamic performance analysis and algorithm model of multi-pressure heat recovery steam generators (HRSG) based on heat exchangers layout, Energy Convers. Manage. 81 (2014) 282– 289.  P. Hanafizadeh, S. Falahatkar, P. Ahmadi, M.M. Siahkalroudi, A novel method for inlet duct geometry improvement of heat recovery steam generators, Appl. Therm. Eng. 89 (2015) 125–133.  J.S. In, Y.L. Sang, Optimization of heat recovery steam generator through exergy analysis for combined cycle gas turbine power plants, Int. J. Energy Res. 32 (2008) 859–869.  M. Sharma, O. Singh, Exergy analysis of dual pressure HRSG for different dead states and varying steam generation states in gas/steam combined cycle power plant, Appl. Therm. Eng. 93 (2016) 614–622. 415  M.T. Mansouri, P. Ahmadi, A.G. Kaviri, M.N.M. Jaafar, Exergetic and economic evaluation of the effect of HRSG configurations on the performance of combined cycle power plants, Energy Convers. Manage. 58 (2012) 47–58.  M.N. Dumont, G. Heyen, Mathematical modelling and design of an advanced once-through heat recovery steam generator, Comput. Chem. Eng. 28 (2004) 651–660.  A. Rovira, M. Valdés, M.A.D. Durán, A model to predict the behaviour at part load operation of once-through heat recovery steam generators working with water at supercritical pressure, Appl. Therm. Eng. 30 (2010) 1652–1658.  H. Baig, M.A. Antar, S.M. Zubair, Performance evaluation of a once-through multi-stage flash distillation system: impact of brine heater fouling, Energy Convers. Manage. 52 (2011) 1414–1425.  P. Hanafizadeh, M.M. Siahkalroudi, P. Ahmadi, Experimental and numerical investigation of optimum design of semi industrial heat recovery steam generator inlet duct, Appl. Therm. Eng. 104 (2016) 375–385.  X. Sui, Y. Zhang, S. Shao, S. Zhang, Exergetic life cycle assessment of cement production process with waste heat power generation, Energy Convers. Manage. 88 (2014) 684–692.  M.L.G. Reno, F.M. Torres, R.J.D. Silva, Exergy analyses in cement production applying waste fuel and mineralizer, Energy Convers. Manage. 75 (2013) 98– 104.  A. Atmaca, R.æ.K. Yumrutasß, Thermodynamic and exergoeconomic analysis of a cement plant: part II – application, Energy Convers. Manage. 79 (2014) 799– 808.  S. Karellas, A.D. Leontaritis, G. Panousis, E. Bellos, E. Kakaras, Energetic and exergetic analysis of waste heat recovery systems in the cement industry, Energy 58 (2013) 147–156.  J. Li, W. Du, L. Cheng, Numerical simulation and experiment of gas-solid two phase flow and ash deposition on a novel heat transfer surface, Appl. Therm. Eng. 113 (2017) 1033–1046.  P. Ahmadi, I. Dincer, Thermodynamic analysis and thermoeconomic optimization of a dual pressure combined cycle power plant with a supplementary firing unit, Energy Convers. Manage. 52 (2011) 2296–2308.  H.B. Avval, P. Ahmadi, A.R. Ghaffarizadeh, M.H. Saidi, Thermo-economicenvironmental multiobjective optimization of a gas turbine power plant with preheater using evolutionary algorithm, Int. J. Energy Res. 35 (2011) 389–403.