Design methodology of organic Rankine cycle for waste heat recovery in cement plants

Design methodology of organic Rankine cycle for waste heat recovery in cement plants

Accepted Manuscript Research Paper Design Methodology of Organic Rankine Cycle for Waste Heat Recovery in Cement Plants Awais Ahmed, Khaled Khodary Es...

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Accepted Manuscript Research Paper Design Methodology of Organic Rankine Cycle for Waste Heat Recovery in Cement Plants Awais Ahmed, Khaled Khodary Esmaeil, Mohammad A Irfan, Fahad Abdulrahman Al-Mufadi PII: DOI: Reference:

S1359-4311(17)33317-3 ATE 11217

To appear in:

Applied Thermal Engineering

Received Date: Revised Date: Accepted Date:

14 May 2017 1 October 2017 4 October 2017

Please cite this article as: A. Ahmed, K. Khodary Esmaeil, M.A. Irfan, F. Abdulrahman Al-Mufadi, Design Methodology of Organic Rankine Cycle for Waste Heat Recovery in Cement Plants, Applied Thermal Engineering (2017), doi:

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Design Methodology of Organic Rankine Cycle for Waste Heat Recovery in Cement Plants Awais Ahmed a, Khaled Khodary Esmaeil a,b , Mohammad A Irfan a,c, Fahad Abdulrahman Al-Mufadi a a

Department of Mechanical Engineering, Qassim University, KSA. Mechanical Power Engineering Department, Faculty of Engineering, Tanta University, Egypt. c Department of Mechanical Engineering, University of Engineering and Technology, Peshawar, Pakistan. b

Corresponding Author: Mohammad A Irfan ([email protected]) PO Box 6677, College of Engineering, Qassim University, Saudi Arabia

Abstract An organic Rankine cycle (ORC) is similar to a conventional steam cycle energy conversion system, but uses organic fluid, such as refrigerants and hydrocarbons, instead of water. A renewed research interest in ORC focuses on its progressive adoption as a premier technology for converting low and medium temperature i.e. 80°C < T < 300°C heat resources into power. Available heat resources are solar energy, geothermal energy, biomass products, surface seawater, and waste heat from various thermal processes. This study presents a design methodology for an ORC. The design is conducted based on the actual data from a local cement factory. The working fluid directly affects the efficiency of the cycle. The fluid choice is fundamental for a good cycle performance because the optimal thermos physical properties depend on the source temperature. This study illustrates the results of an organic Rankine cycle combined with a gas turbine to convert the gas turbine waste heat into electrical power. The R134a working fluid is chosen for the design. Consequently, approximately 1 MW power can be generated using an ORC. Exergy analysis is performed using actual data from the industry, showing that most of the exergy loss is in the working part of the turbine. Keywords: Design of ORC; R134a, Heat Transfer Coefficients, Exergy NOMENCLATURE

ORC Q h m U A D K NuD Pr µ ρ V λ SQ

Organic Rankine cycle heat rate (kW) enthalpy (kJ/kg) mass flow rate (kg/s//) overall heat transfer coefficient (kW/m2.K) area (m2) diameter (m) thermal conductivity of pipe (W/mK) Nusselt number Prandtl number dynamic viscosity (N.s/m2) density (kg/m3) velocity (m/s) parametric length of tube (m) transverse tube spacing (m)

NTU f L W Re E

number of transfer units Darcy coefficient of friction length (m) work done (kW) Reynolds number exergy (kJ)

Subscripts g gas i inlet l liquid phase o outlet cold (c) fluid at cold side


SL a b r N n

longitudinal tube spacing (m) transverse pitch ratio (1/m) longitudinal pitch ratio (1/m) radius (m) number of tubes number of fins on unit pipe

hot (h) fluid at hot side b p t s1 s2

bleeding pump turbine steam of high-pressure cycle steam of low-pressure cycle

1. Introduction 1.1.


A large amount of industrial waste heat is directly released to the environment (Hiba Khodr, Oct 2013). The aluminum smelters in Canada produce 80 PJ of waste heat annually (L. Wei, 18 Feb 2011). Canadian aluminum producers could save up to 96 million dollars per year and reduce greenhouse gas emissions by approximately 0.45 megatons per year (Roy P, 2010) if only 10% of this heat is recovered to produce useful work (i.e., waste heat recovery). The cement manufacturing process is one of the most energy-intensive industries in the world. The energy required to produce one ton of cement is between 3 and 5 GJ/ton (Sogut Z, 2010). Cement production has steadily increased with the economic growth experienced in the developing countries. China alone produced approximately 1388 million metric tons (Mt) in 2008, which accounted for nearly half of the world’s total cement production (Hasanbeigi A, 2010). The cement manufacturing process is well established. A critical step is the clinker production that consumes approximately 80% of the total energy (Rasul MG, 2005). Clinker is produced by burning a mixture of materials, mainly limestone, silicon oxides, aluminum, and iron oxides. The exit gases from the kilns are exhausted to the atmosphere at approximately 300–350 °C in four stages preheater and at 200–300 °C in case of five–six stages preheater. The clinker coming out of the kiln is at approximately 1000 °C and cooled to 100–120 °C using ambient air. This process generates hot air of approximately 200–300 °C. Heat from the hot air and gases exhausted to the environment can be recovered using a well-proven waste heat recovery steam technology pioneered by the Japanese or by adopting low-temperature organic Rankine cycles (Bundela PS, 2010). Organic Rankine cycles offer greater modularity and lower investment and maintenance costs over steam cycles and could be a technology of choice for energy efficiency in the cement industry. 1.2.

ORC review

A Rankine cycle using organic fluids as working fluids is called an organic Rankine cycle (ORC). The ORC is potentially feasible in recovering heat sources containing low enthalpy. An efficient operation of the ORC heavily depends on two factors: working conditions of the cycle and thermodynamic properties of the working fluids (Hung, 2001). For the past 20 years, extensive research has been performed on the organic Rankine cycle (ORC) mainly to convert low temperature heat (80 °C
Bahaa and Guo showed a selection of suitable working fluids for high (300 °C) and moderate temperature (below 200 °C) heat sources (Bahaa S, 2007) (Guo T, 2011). Most working fluids do not work well over a wide temperature range (300–100 °C) of the heat source. Zhang and Karellas suggested to replace subcritical Rankine cycles with supercritical Rankine cycles to obtain a better thermal performance (Zhang S, 2011) (Karellas S, 2008), while Chen and Wang recommended to replace pure working fluids with zeotropic mixture working fluids to acquire the same (Chen H, 2011) (Wang JL, 2010). Hung categorized low-grade waste heat recovery and analyzed ORC efficiency using cryogens, such as benzene, ammonia, R11, R12, R134a, and R113 as working fluids (Hung TC S. T., 1997). Maizza V and Maizza A investigated the thermodynamic and physical properties of 20 unconventional fluids used in organic Rankine cycles supplied by waste energy sources (Maizza V, 2001). Stappato and Quoilin stated that selecting the working fluids is very important for the ORC system performance and economy because it effects the system efficiency, system component size, expansion machine design, system stability, and safety and environmental concerns (Stoppato, 2012) (Quoilin S D. S., 2011). The working fluid selection of the ORC system is a complicated task because of the two following reasons: i.


The heat source types and the working conditions of the ORC widely vary from a lowtemperature heat source of 80 °C (e.g., geothermal, plate type solar collector) to a hightemperature of 300 °C heat source (e.g., biomass). Except for some substances whose critical temperatures are too low or too high, hundreds of substances can be used as working fluid candidates of ORC, including hydrocarbons, aromatic hydrocarbons, ethers, perfluorocarbons, alcohols, siloxanes, and inorganics (which should inherently not be an ORC, but was included because of the similarity with ORC).

Table 1 shows the ORC installed in various countries along with a characteristic, such as the power produced using different working fluids. The working fluid used in the ORC is also mentioned. The expansion devise used in ORC and their type is an important parameter. Table 1 Main characteristics of the commercial ORC system. Manufacturer W (kW) T (°C) Working fluid Expansion devices GMK 50– 120–350 GL-160, WL-220 Turbine multistage, (Germany) 5000 axial Infinity turbine 250– >220 R134a,R245fa Turbine (USA) 1200 Turboden 200– 100–300 OMTS, Solkatherm Turbine (two-stage (Italy) 2,000 axial) Ormat (USA) 200– 150–300 n-pentano, other Turbine (two-stage 70,000 axial) Enertime 300– 200 HFC Turbine (France) 5,000

Reference (Bernardo Peris, 2015) (Bernardo Peris, 2015) (Bernardo Peris, 2015) (Bianchi M, 2011) (Enertime, 2016)


Phoenix (Australia)

10– 5,000


Not specified (scroll expander, turbine)

80, >140

R245fa, Novec649, Cyclohexano R245fa, other

Rank (Spain)


Bosch KWK (Germany)






(Phoenix. Port Melbourne, 2016) (Rank®. Castellon, 2016) (Quoilin S B. M., 2013)

1.3. ORC for waste heat The steam Rankine cycle is one of the most important methods used to transform large- or medium-scale thermal energy into useful power. Examples are nuclear and coal-fired power plants. The main components of a steam power cycle are an economizer, evaporator, condenser, pump, turbine, and the working fluid (steam). The advantages presented by water as the working fluid are as follows:     

good thermal/chemical stability (no decomposition); low viscosity (pumping work reduces); good energy carrier (high latent and specific heat); non-flammable, non-toxic, and no threat to the environment; and cheap and abundant.

However, many problems, such as the following, are encountered when using water as the working fluid (Wali, 1980):   

superheating is needed to prevent condensation during expansion; erosion of turbine blades; and complex and expensive turbines

Water is more suitable for high-temperature applications and large integrated systems because of the abovementioned reasons. Looking for small- and medium-scale power plants, the problems encountered with water are related to checking the water feasibility. Organic compounds characterized by higher molecular mass and lower critical temperature than water have been suggested in “organic Rankine cycles.” An ORC has several advantages over a conventional steam power plant:  The evaporation process needs less heat.  The evaporation occurs at lower pressure and temperature.  The expansion process is isentropic and terminates in vapor region.  Blade erosion is avoided. The utilization of renewable and low-grade heat sources for power generation has received significant attention in the past decade in view of increasing concerns over climate change and high energy prices (Hiba Khodr, Oct 2013). Closed power cycles, mainly ORC (organic Rankine cycles) and supercritical CO2 cycles, have been studied, developed, and implemented at 4

industrial scales to achieve this aim. While supercritical CO2 cycles are still an object of research and development as well as tests (Conboy T, 2012), ORCs are nowadays widely adopted for small- (100 kW) to medium-scale (1e10 MW) power generation (e.g., ORC installed by Turboden (Turboden, 2015) and Triogen (Triogen, 2015)). Their main advantages over conventional steam cycles include higher cycle and turbine efficiencies at small scales (<1 MW) and/or low heat source temperatures (<200 °C), cheaper turbine, no blade erosion caused by the adoption of dry-expansion fluids, and no need for water demineralization. Compared to internal combustion engines and micro turbines, the ORCs can use a wide variety of heat sources, including solid fuels, such as wood and straw, concentrated solar energy, geothermal heat, and waste heat made available by industrial processes. The design criteria for the ORCs are still an object of study because of the following factors:   

the large number of available working fluids; the influence of the thermodynamic properties of the working fluid on the optimal cycle configuration, operating variables, and plant cost; the wide range of possible applications (e.g., biomass-fired combined heat and power plants, concentrated solar plants, binary geothermal plants, waste heat recovery, etc.) with peculiar specifications; and the ongoing research and development on turbines and heat exchangers.

These reasons lead to more than four hundred published papers on the topic of orc optimization. Most of these studies have been reported in the last four years (Elsevier. Scopus, 2015). Gases are cooled from 300°F [150 °C] to 140°F [60 °C]. After which, the facility can then obtain a 3% efficiency increase. Cooling gases further to 100°F [38 °C] captures a portion of the latent heat and provide an 11% efficiency increase. Cooling the flue gas further could significantly increase heat recovery by allowing the latent heat of vaporization to be recovered (Ilona Johnson, March 2008). R134a working fluid was used to design ORC for this case study. The only problem in selection of the organic fluid is their higher Ozone depletion potential (ODP) and global worming potential (GWP), but in case of R134a the ODP is zero and the GWP is 1300. Figure 1 shows the schematic view of an ORC power plant. The diagram represents a typical waste heat recovery system. The main components are Heat exchanger, Turbine, Compressor, Condenser. Figure 2 shows the TS diagram for R134a for complete cycle. At the end of condenser, the fluid is in liquid state and is represented by 1. The pressure is increased by the pump from 1 to 2 by the external work Win. The high pressure fluid passes through the economizer from 2 to 2a, through the evaporator from 2a to 2b and finally through the superheater 2b to 3. In these three processes heat Qin is supplied to the fluid. At 3 the working fluid is in superheated state, and is expanded through turbine from 3 to 4. During this process the entropy increases. Fixed isentropic efficiency is considered for this process. During this process mechanical work Wout is 5

obtained. At the end of the expansion the working fluid is still in the superheated state at low


pressure. The fluid is the cooled at constant pressure form 4 to 4a outside the dome and then from 4a to 1 inside the dome. During 4 to 4a heat is rejected and temperature decreases, whereas from 4a to 1 the heat rejection is at constant temperature and the working fluid converts into liquid state. This completes the whole cycle.

Figure 1: Schematic Flow Diagram of ORC

Figure 2. T–S diagram for R134a


2. Fundamentals of the heat recovery system (HRS) The heat recovery generator is a heat exchanger comprising three heat exchangers: economizer, evaporator, and superheater. The heat exchanger is in the form of a bundle of tubes placed in a staggered arrangement to increase the heat transfer coefficient. The flow of the working fluid (organic fluid) in the pipes is horizontal, while that of the exhaust gasses is vertical, making each heat exchanger in the system a cross-flow heat exchanger. The feed water heat is used to pre-heat the organic fluid before entering the heat exchanger. The expansion occurs in the turbine. The exhaust gases of the gas turbine have a low heat transfer coefficient compared with the heat transfer coefficients of the working fluid in the evaporator and the economizer. Therefore, finned tubes are used at the outer circumference of the heat exchanger pipes. The most important parameters of the HRS are the pinch and approach points, which affect the effectiveness of heat exchangers. The pinch point is the difference between the gas temperature leaving the evaporator section of the system. The approach point is the difference between the saturation temperature of the fluid and the inlet temperature of the fluid in the evaporator. The selection of these two factors also affects the size of the superheater, evaporator, and economizer. The pinch and approach points for unfired HRSGs based on the sizes of the evaporators that can be economically constructed and transported are usually in the range of 10 °C to 15 °C (Ganapathy V. , 1996). Variables that directly affect steam generation and the steam and gas temperature profiles are the approach and pinch point as shown in Figure 3. The designer selects the approach point -------- Flue Gas Line

Tg 1


_____ Water/Steam Line



Pinch point = Tg3 – TS


Te m pe rat ur e

Tg4 Approach point = Ts – TTws w


Tw1 Super heater



of the heat exchanger between 0.5oC to 2oC. The approach point saves pre-boiling of the fluid in


the economizer. Heat transfer is dependent on the mass flow of the hot gases, mass flow of the water or steam, the temperature difference, and the surface area. Figure 3: Pinch and approach points 3. Design approach 3.1. Design of the logic flow diagram First, the magnitude (energy flow) and the temperature of the flue gases must be determined. The most reliable method of quantifying the flows is by making heat and mass balances. In simple cases, suitable heat and mass balances can be made by hand calculation. Beyond that, spread sheet balances or specialist flowsheet simulation software may be the best way forward. Different calculation platforms have advantages and disadvantages depending on the context. Figure 4 represents the logical flow process used in the HRS design. The flow is distributed into different steps.

Figure 4: Logic flow chart for the ORC 3.2. Assumptions 9

Some of the main assumptions to derive the mathematical model are listed as follows:         

Unfired HRS Steady state condition exists for both fluids i.e. working fluid and the exhaust gas No heat loss from the heat exchangers; To avoid dew point, the stack temperature is more than 100 °C (Ganapathy V. , 2003); Forced circulation is considered for exhaust flue gases using fans; Staggered tube arrangement with fins is considered in heat exchanger Mass flow rate of the working fluid remain same in the economizer, evaporator, and superheater Input data of flue gas condition i.e. mass flow rate and temperature are obtained from local industry. Selection of Pinch and approach point.

3.3. Mathematical modeling A section of the tube of the heat exchanger with annular fins is selected as shown in Fin Height

Pipe diameter

Outer Pipe Diameter

Fin Distance

figure 5 for the mathematical modeling, by considering the assumptions listed above. The mathematical model consists of temperature profile, overall heat transfer coefficient, area, pressure loss, pump and turbine work condenser and feed water heater. Figure 5. Geometrical variables of the pipe 3.3.1. Temperature profile The temperature profile is the first design step. The temperatures of the working fluid and the exhaust gases at the inlet and the outlet for each section are calculated using energy balance. A suitable pinch point is also selected. Figure 6 shows the temperature distribution of the resulting dual pressure heat exchanger. The arrangement is presented in such a manner that the exit temperature of the chimney should be less and the maximum heat can be recovered. The following equation used to develop the temperature profile is the heat balance equation: (1)


Super Heater






Temperature (oC)

170 150

150 160

130 104



90 86




50 1




Heat Exchanger Location

Figure 6: Temperature profile for the cross-flow heat exchanger

3.3.2. Calculation over all the heat transfer coefficients The designed method in this study is based on the heat transfer coefficients used to obtain the heat transfer area for each heat exchanger in high- and low-pressure level of the HRSG. The overall heat transfer coefficient, U, by the total heat exchange area is calculated as follows: (2) Practically, heat exchanger tubes are made of high-conduction material with small thickness, which permits a reasonable assumption for neglecting the third term. If the term is not included in the equation the error is about 5%.

Hence, the overall heat transfer coefficient is modified to: (3) The inner heat transfer coefficient for the steady-state incompressible flow inside the tube with a uniform cross-section area is calculated as follows using the flow intensity calculated by Reynolds number: (4) 11

The present calculation shows that the Reynolds number is always greater than 3000, indicating that the flow is turbulent. The Dittus-Boelter equation (for turbulent flow) is an explicit function for calculating the Nusselt number, which is easy to solve. Therefore, the Nusselt number (Nu) for the turbulent flow in a smooth circular tube for a single phase can be obtained using the Dittus–Boelter equation (Winterton R. H. S., 1998). For heating n = 0.4 For cooling n = 0.3 (5) The internal convection heat transfer coefficient can be calculated as follows: (6) The following equation is used to find the convection coefficient for over the bank of tubes for

flue gases: (7) Figure 7: Staggered tube arrangement Figure 7 shows the staggered tube arrangement. As heat transfer occurs from a single row of tubes, that situation is dealt with by Gnielinski who treated the problem as an extension of his method for a single tube in cross flow. He suggests that the average Nusselt number of a single row of smooth tubes in cross flow can be calculated using the equation for the Nusselt number for a single tube in cross flow, but with the Reynolds number with in a range of (Gnielinski, 1975) 10 < Reλ < 107 and 0.6 < Pr < 1000. (8)


The convection coefficient for the flue gases flow outside the bank tubes is presented as follows: (9) A temperature correction factor now exists (Sieder, 1936). (10) All the heat transfer coefficients are calculated by the end of this section. The overall heat transfer coefficient will also be known. 3.3.3. Area calculation The number of transfer units (NTU) method is used to calculate the rate of heat transfer in heat exchangers, especially counter current exchangers, when there is insufficient information to calculate the log-mean temperature difference (LMTD). In the heat exchanger analysis, the LMTD method can be used if the fluid inlet and outlet temperatures are specified or can be determined by a simple energy balance. However, the NTU or the effectiveness method is used when these temperatures are not available. The area for each section of the HRSG (i.e., superheaters, evaporators, and economizers) is calculated based on the overall heat transfer coefficient U. The area of each section can be calculated using the following correlations: (11)

Fined tubes are used for heat transfer. Hence, the length of the fin must be calculated as follows: (12) 3.3.4. Pressure drop and pump work The pressure drop is calculated by Darcy formula (13) Where friction factor f for the internal flow is given by (Petukhov, 1970.) (14) The pressure drop has to be included in the calculation of net pressure rise for highpressure pump feeding the organic fluid in the cycle. The work done required by the is calculated as, (15) 3.3.5. Feed water heater 13

A feed heater is a power plant component used to pre-heat the working fluid and deliver the working fluid to the next level. In an open feed heater, the hot working fluid is directly mixed with the cold fluid. Hence, the mixture proceeds to the next step. The only important thing in an open feed water is that both mixing fluids should have the same pressure. The mathematical analysis of the feed water heater is based on the energy balance: (16) 3.3.6. Power generated by the turbine From Figure 3, the expansion in turbine is the net work done available. In ideal condition the obtained power is more but due to irreversibly the net power is reduced having the fixed isentropic efficiency. The power generated for the turbine is calculated as follows: (17) If the working fluid is bled from the turbine, then the net work done is obtained by subtracting Wb from Wt. (18) (19) 3.3.7. Condenser A condenser is a device or unit used to condense a substance from its gaseous to liquid state by heat rejection. The latent heat is given up by the substance and will transfer to the condenser coolant. (20) (21) (22) 4. Case study The HRS design depends on the thermodynamic parameters of the exhaust gases from the gas turbine in the simple cycle mode, such as temperature, pressure, and mass flow rate of gases. The above design method is applied to a case study taking data from a local cement factory. The date is collected from the chimney after the pre-heater of clinker. Table 2 shows the basic parameters. Table 2 Technical data and operation parameters of the kiln at 29 °C. Parameters Exhaust gas flow from the chimney Exhaust gas temperature from the chimney Fuel consumption in the kiln

Base load 28 200 42

Units Kg/s °C Kg/s


5. Exergy This section describes the method used to estimate the energy and exergy use, energy and exergy efficiencies for the heat exchanger, and other items. Exergy is defined as the maximum useful work which can be extracted from a system as it reversibly comes into equilibrium with its environment. According to the data, the R134a working fluid is circulated by the pump, enters the evaporator as compressed liquid at 30 bar, and leaves as a saturated vapor at an evaporator pressure and it makes it a close cycle. The working fluid leaves the turbine as a superheated vapor at a condenser pressure of 14 bar and leaves the condenser as a saturated liquid. The exergy balance for a close cycle can be written as follows: (23) Analysis (24) (25) (26) (27) Figure 8 shows the exergy distribution of each component of the Heat Recovery System. The maximum available exergy from the flue gases is divided into two: one part is utilized in the heat exchanger, while the other is lost. The heat exchanger provides the exergy to the turbine, which is then converted to useful work and for per heating of steam in an open feed water heater. The turbine outlet is collected in the condenser, which enables the open feed water heater to gain some exergy. The fluid is then pumped to the desired design pressure by externally adding exergy by pumps. After which, the fluid again circulates through the heat exchanger, and the cycle continuous.


Figure 8: Exergy distribution of the Heat Recovery System Figure 9 represents the exergy losses in the HRSG as a percentage of the exergy input. The exergy input is the available exergy from the flue gases. A large amount of exergy loss occurs in the flue gases. The second larger exergy loss occurs in the heat exchanger because of the evaporator, where the phase change happens.

Turbine 26%

Condenser [PERCENTAGE] 0% Flue Gases 69%

Pump 3%

Figure 9: Component Exergy losses as a percentage of total exergy loss


6. Results and discussion Figure 10 depicts the effect of the design pressure on the effectiveness of the heat exchanger (WHRS) for a constant mass flow rate of the working fluid. The x axis shows that the possible design pressure varies from 22 to 35 bar. The increase in the design pressure leads to an increase in the effectiveness of the heat exchanger because more heat transfer occurs in a higher pressure, thereby leading to an increase in the effectiveness for the constant mass flow rate. At the same time, the area of the heat exchanger becomes larger, which is a compromise between pressure and the area of the heat exchanger. The increase in preheating decreases in effectiveness because the amount of heat transfer remains the same as the mass flow rate of the heat exchanger



Pre Heating 70C


Pre Heating 60C


Pre Heating 50C

0.9 0.87 0.84 0.81

0.78 20










Figure 10: Effect of the Design Pressure on the effectiveness of Heat Exchanger is kept constant.

Figure 11 shows the effect of the design pressure on the stack temperature of the heat exchanger (WHRS) when the mass flow rate is kept constant. The x axis shows that the possible design pressure varies from 22 to 35 bar. The increase in the design pressure leads to increase in cp, thus improving the heat transfer (Q = m cp ΔT), and hence the exhaust gasses transfer more heat to the working fluid, resulting in decrease in the stack temperature. The effect of preheating is the increase in per-heating increase in effectiveness caused by the lesser temperature difference in the inlet and the outlet of the heat exchanger.




365 360 355

pre heating 70C


pre heating 60C pre heating 50C

345 20







Figure 11: Effect of the Design Pressure on the stack temperature Figure 12 presents the effect of the design pressure and the specific work output of the working fluid. The x axis shows that the possible design pressure varies from 20 to 40 bar. The increase in the design pressure results in the increase in the specific work of the working fluid because of the higher heat available at a higher pressure leading to an increase in the specific work of the working fluid.

Specfic Work (KJ/Kg)

25 20 15 10 5 0 15







Pressure (bar)

Figure 12: Effect of the Design Pressure on specific work

Figure 13 shows the effect of pinch point on the efficiency of the ORC, for a certain range of temperature, when the mass flow rate and the pressure are kept constant. The x axis shows that the flue gas temperature varies from 420 K to 500 K. The pinch point has a major effect on the efficiency of the ORC (η = Wout/Qin). The lesser the pinch point, the higher the 18

efficiency. For example, at a temperature of 460 K the efficiency of the ORC cycle is 15.5 % if the pinch point is set to be Δ15 K, for the same temperature if the pinch point is considered Δ5 K the efficiency increases to 18%. 0.21

Pinch point (∆5 K)

Pinch point (∆15 K)



0.19 0.18 0.17 0.16 0.15 0.14 400








Figure 13: Effect of the turbine inlet temperature on ORC cycle efficiency

7. Conclusion The performance and the characteristics of the Organic Rankine cycle heat exchanger using working fluids R134a were investigated in this study. A case study for Heat Recovery from a Cement Plant was analyzed. The best operating conditions for R134a existed when the turbine inlet temperature was as low between 120 °C and 220 °C. The effectiveness of the ORC Heat exchanger using the R134a working fluid can increase up to 93%. The ORC can be applied to low-grade heat sources, and R134a is able to significantly improve the ORC performance. The exergy losses in turbine are 20%, which was less than the exergy loss of a steam turbine. The organic fluid had a low boiling point. Hence, the wetness of the organic fluid did not happen, and the turbine efficiency increased.

Acknowledgment The authors are indebted to the King Abdulaziz City of Science and Technology — National Science and Technology Initiative Plan (KACST-NSTIP) research grant number 13-ENE-170309 for supporting this research. The help and cooperation extended by the local factory in collecting the data are also gratefully acknowledged.


References Bahaa S, K. G. (2007). Working fluids for lowtemperature organic rankine cycles. Energy , 1210–21. Bernardo Peris, J. N.-E.-B. (2015). Experimental study of an ORC (organic Rankine cycle) for low grade waste heat recovery in a ceramic industry. Energy 85, 534-542. Bianchi M, P. A. (2011). Bottoming cycles for electric energy generation: parametric investigation of available and innovative solutions for the exploitation of low and medium temperature heat sources. . Appl Energy , 1500-1509. Bundela PS, C. V. (2010). Sustainable development through waste heat recovery. . American Journal of Environmental Sciences, 83–89. Chen H, G. D. (2011). A supercritical rankine cycle using zeotropic mixture working fluids for the conversion of low-temperature heat into power. . Energy , 549–55. Conboy T, W. S. (2012). Performance characteristics of an operating supercritical CO Brayton cycle. . J Eng Gas Turbines Power, 134:111703.2. Dai Y, W. J. (March 2009). Parametric optimization and comparative study of organic Rankine cycle (ORC) for low grade waste heat recovery . Energy Conversion and Management Volume 50, Issue 3, 576–582. Elsevier. Scopus. (2015). Literature Search URL, Retrieved from Abstract and citation database of peer-reviewed literature.: Enertime, P. F. (2016, 03 25). Retrieved from Available from: Ganapathy, V. (1996). Heat-Recovery Steam Generators: Understand the Basics. CHEMICAL ENGINEERING PROGRESS, 32-45. Ganapathy, V. (2003). Industrial boilers and heat steam generators. Design applications and calculations" New York: Marcel Dekker,. Gnielinski, V. (1975). "Berechnung mittlerer Warme- und Stoffubergshoeffizienten an laminar und turbulent uberstromten Einzellkorpern mit Hilfe einer einheitlichen Gleichung," . Forschung im Ingenierwesen, Vol. 41,, 145-153. Guo T, W. H. (2011). Selection of working fluids for a novel lowtemperature geothermally-powered ORC based cogeneration system. . Energy Convers Manage , 2384–91. Hasanbeigi A, P. L. (2010). Analysis of energy-efficiency opportunities for the cement industry in Shandong Province China: a case study of 16 cement plants. . Energy , 3461–3473. Hiba Khodr, I. R. (Oct 2013). Energy Policies and Domestic Politics in the MENA Region in the Aftermath of the Arab Upheavals: The Cases of Lebanon, Libya, and KSA. Politics & policy, 656-689. Hung TC, S. T. (1996). A review of organic rankine cycles (ORCs), for the recovery of low-grade waste heat. Energy, 661–7. 20

Hung TC, S. T. (1997). A review of organic Rankine cycles (ORCs) for the recovery of low-grade waste heat. Energy, 661-667. Hung, T.-C. (2001). Waste heat recovery of organic Rankine cycle using dry Fluids . Energy Conversion and Management 42 , 539±553. Ilona Johnson, W. T. (March 2008). Waste Heat Recovery:Technology and Opportunities in U.S. Industry. Agency of the United States. Karellas S, S. A. (2008). Supercritical fluid parameters in organic rankine cycle applications. Int J Thermodyn, 101–8. L. Wei, Y. Z. (18 Feb 2011). Efficiency Improving Strategies of Low-temperature Heat Conversion Systems Using Organic Rankine Cycles: An Overview. Energy , 869-878. Larjola, J. (1995). Electricity from industrial wast heat using high-speed organic rankine cycle (ORC). Int J Prod Econ, 41–35. Lin S, L. X. (2011). Synthesis of multipass heat exchanger networks based on pinch technology. . Comput Chem Eng, 1257–64. Maizza V, M. A. (2001). Unconventional working fluids in organic Rankine-cycles for waste energy recovery systems. . Appl Therm Eng, 21:381e90. Petukhov, B. S. (1970.). Advances in Heat Transfer. New York, : Vol. 6, Academic Press,. Phoenix. Port Melbourne, A. (2016, 03 25). Retrieved from Available from:: Quoilin S, B. M. (2013). Techno-economic survey of organic Rankine cycle (ORC) systems. Renew Sustain . Energy, 168-86. Retrieved from 168-86. Quoilin S, D. S. (2011). Thermo-economic optimization of waste heat recovery organic Rankine cycles. Applied Thermal Engineering , 2885–93. Rank®. Castellon, S. (2016, 2015 03). Retrieved from Available from:: Rasul MG, W. W. (2005). Assessment of the thermal performance and energy conservation opportunities of a cement industry in Indonesia. . Applied Thermal Engineering, 2950–2965. Roy P, D. M. (2010). Thermodynamic analysis of a power cycle using a low-temperature source and a binary NH3–HO mixture as working fluid. . Int J Therm Sci , 49-58. Sieder, E. N. (1936). Ind. Eng. Chem. Sogut Z, O. Z. (2010). Mathematical modeling of heat recovery from a rotary kiln. . Applied Thermal Engineering , 817–825. Stoppato, A. (2012). Energetic and economic investigation of the operation management of an organic Rankine cycle cogeneration plant. Energy, 3–09. Triogen. (2015, 31 05). Retrieved from List of Triogen ORCs installed throughtout the world.: 21

Turboden. (2015, 5 31). Retrieved from List of Turboden ORCs installed throughout the world. 2015: Wali, E. (1980). Optimum working fluids for solar powered Rankine cycle cooling ofbuildings. Solar Energy, 235-341. Wang JL, Z. L. (2010). A comparative study of pure and zeotropic mixtures in low-temperature solar rankine cycle. Appl Energy , 3366–73. Winterton R. H. S. (1998). Heat Mass Transfer. Int. J. Zhang S, W. H. (2011). Performance comparison and parametric optimization of subcritical organic rankine cycle (ORC) and transcritical power cycle system for low-temperature geothermal power generation. . Appl Energy , 2740–54.

List of Figures Figure 1: Schematic Flow Diagram of ORC


Figure 2: T-S diagram for R134a

Figure 3: Pinch and Approach Points

Figure 4: Logic Flow chart for ORC

Figure 5. Geometrical Variables of the pipe Figure 6: Temperature Profile for cross-flow Heat Exchanger Figure 7: Staggered tube arrangement

Figure 8: Exergy distribution of Heat Recovery System

Figure 9: Component Exergy losses as a percentage of the total exergy loss

Figure 10: Effect of Design Pressure on the Effectiveness of HX

Figure 11: Effect of Design Pressure on the Stack Temperature

Figure 12: Effect of Design Pressure on Specific work

Figure 13: Effect of the turbine inlet temperature on ORC cycle efficiency

List of Tables

Table 1 Main characteristics of commercial ORC system 23

Table 2 Technical data and operation parameters of Kiln at 29°C


Highlights:   

A complete design methodology for waste heat recovery using and Organic Rankine Cycle is presented. The design methodology is applied for waste heat recovery from the exhaust gasses of a cement plant. The ORC effectiveness, using the R134a working fluid, can increase up to 95%.