Nowadays, it is well established and known that fossil fuels represents the principal source of energy in the world. It is reported that until 2009, 88% of the global primary energy was produced from the three leading fuels; oil accounting for 35%, natural gas with 24% and coal with 29% [Lin 2011]. However, the availability of these natural resources has decreased in the last decades as consequence of the increasing consume and depletion of them; as a result, fuel price have been increasing in the last 20 years. It is expected that the nominal price of oil, natural gas and coal will continue to increase in the future [Shafiee 2010].
In the process industry, important quantity of energy is used in equipment such as compressors, boilers, pumps, dryers, furnaces, gas turbines, ovens, heaters and more. Thermal efficiency is a key parameter in each of these devices since it represents the energy loss regarding the energy input into the system. For instance, Galanis [2008] reports that within the eight largest Canadian manufacturing sectors which account for 66% of the energy used by Canadian industries, 70% of the input energy is wasted to the environment. Kapil [2011] states that low grade heat is often wasted without being recovered even though the economic benefits can be significant. This historical behaviour is due to the fact that in the past, fuel prices were lower and therefore investing in recovering low grade heat was not economically convenient. However, in recent years, the interest in low grade heat recovery has changed due to economic factors [Ammar 2012].
Some of the low grade heat sources reported in the literature are condensers, combustion engines, boilers blowdown, gas turbine exhausts, flue gases, and furnaces, among other equipment and sources [He 2011, Sukamongkol 2010, Zhelev 2004]. Each of the sources has different characteristics in terms of its composition, temperature, quantity, accessibility and state of matter. Boilers are widely used in the industrial sector and it represents a major consumer of energy (coal, gas, oil, biomass), a typical heat balances in a boiler is shown in Figure 1.1.
Figure 1.: Typical heat balance of in a boiler [Saidur 2010].
From Figure 1.1 it can be noticed that flue gases represent a significant source of heat available in the process industry and utility systems, in fact, the flue gas is used to preheat the boiler intake air as well as the boiler feedwater [Wang 2012, Smith 2005]. However, flue gases are usually discharged to the atmosphere at temperatures higher than 120°C and therefore, still represent a valuable source of energy that can be used in low grade heat applications.
Low grade heat recovery can be used for power generation, combustion air and boiler feedwater preheating, furnace load preheating, absorption refrigeration systems, and space heating among others. For power generation, one option is to use an Organic Rankine Cycle (ORC) [Lee 1988]. The ORC is similar to the conventional steam Rankine Cycle with the difference that an organic fluid is used instead of water as working fluid and one of the advantages over conventional Rankine Cycle is its better performance in the use of heat from low temperature sources [Yamamoto 2001]. Absorption refrigeration systems are heat-operated and therefore waste heat can be used as the source of energy, thus contributing to reduce energy consumption from fossil fuel [Fernandez-Seara 1997], besides, it operates with environmental friendly mixtures, such as ammonia-water and lithium bromide-water [Pratihar 2010].
Due to current economic and environmental issues, heat recovery at any grade, especially at low temperatures from flue gases streams, represents an attractive path to achieve higher efficiencies (or higher economic benefits) in different types of processes and comply with the environmental laws and restrictions.
Objective
The main objective of this dissertation is to evaluate thermodynamically the potential of low grade heat sources to produce power or enhance the Brayton and steam Rankine Cycle to produce more power. Three different options are assessed and compared regarding the low grade heat source, namely, a flue gas stream:
Organic Rankine Cycle to produce power.
Absorption refrigeration system to produce cooling capacity to cool down the intake air temperature in the Brayton Cycle.
Absorption refrigeration system to produce cooling capacity to reduce the outlet pressure of the steam turbine in the steam Rankine Cycle.
In order to compare the feasibility of each option and achieve the objective of this dissertation, modelling and simulations are necessary. Thus, Aspen HYSYS is required [Aspen Technology, Inc. 2006]. It is not included in the objective of this work neither an economic analysis nor any economic assessment of the processes to recover energy.
Scope of the Report
Chapter 1 of this report includes the introduction and motivation of the dissertation. The objective is presented in this chapter as well.
Chapter 2 presents the literature review used as background regarding the definition and characteristics of low grade heat, how to recover it and describes three available possibilities to use the recovered energy from the flue gas.
Chapter 3 provides the methodology used to model a flue gas stream and its characterisation in terms of enthalpy and exergy. It also provides the methodology developed to assess the energy recovery and its integration in; i) the ORC to produce power, and ii) the ARS to generate cooling capacity to increase the efficiency in the Steam Rankine Cycle and the Brayton Cycle. All modelling were carried out in Aspen HYSYS.
Chapter 4 delivers the obtained results for the simulations and its discussion along with the sensitivity analysis of the modelling.
Chapter 5 conveys the final conclusion of the dissertation project and suggestions for future research.
Literature Review
Low Grade and Waste Heat
In the industrial sector, whether it is chemical, mechanical, biological or utility system, the concept of waste heat can be described as the energy that is produced within a site or in a particular process which has the potential to be used as energy source, but is often discharged to the atmosphere without making practical use of it. Low grade heat is essentially the same as waste heat; the term low grade is used to denote the low temperature level of the source. Galanis [2009] suggests that the temperature range of a low grade heat source is between 80 and 150°C and intermediate grade heat has temperatures between 150°C up to 350°C. However, this definition is not universal, for instance the U.S DOE [2008] states that heat sources below 232°C are considered low grade heat sources. Perhaps, it is not possible to define a unique interval of temperatures for different sources of waste heat, in fact, according to Markides [2012] the low temperature concept should be established in relation to the temperature level of the source; nevertheless, diverse reports agree in classifying an energy source as a low grade heat source for temperature levels below 250°C [U.S. DOE 2008, Galanis 2009, Saidur 2010, Wang 2012, Chen 2012].
In addition, the amount of energy wasted in industry is not accurately quantified, several studies estimate that as much as 20 to 50% of industrial energy consumption is discharged as waste heat [U.S. DOE 2008, Galanis 2009, Ammar 2012]. In the United States of America it has been estimated that between the 20 and 30% of the energy input in to the industry sector, which accounts for 33% of the total energy consumed in the country, is wasted in the form of heat by hot exhaust gases, cooling water, heat lost from process equipment surfaces and heated final products [U.S. DOE 2008].
Sources of Waste Heat
There are many sources of waste heat available at different temperature levels in industry; Table 2.1 shows some of the typical waste heat sources at low, medium and high temperature levels.
Table 2.: Sources of waste heat, adapted from U.S. DOE [2008].
T. Range
Sources
T (°C)
Nickel refining furnace
1370 - 1650
Steel electric arc furnace
1370 - 1650
Basic oxygen furnace
1200
Aluminium reverberatory furnace
1100 - 1200
High
Copper refining furnace
760 - 820
[> 650°C]
Steel heating furnace
930 - 1040
Copper reverberatory furnace
900 - 1090
Hydrogen plants
650 - 980
Fume incinerators
650 - 1430
Glass melting furnace
1300 - 1540
Coke oven
650 - 1000
Iron cupola
820 - 980
Steam boiler exhaust
230 - 480
Gas turbine exhaust
370 - 540
Medium
Reciprocating engine exhaust
320 - 590
[230 - 650°C]
Heat treating furnace
430 - 650
Drying & baking ovens
230 - 590
Cement kiln
450 - 620
Exhaust gases exiting recovery devices in gas-fired boilers, ethylene furnaces, etc.
70 - 230
Process steam condensate
50 - 90
Cooling water from:
furnace doors
30 - 50
Low
annealing furnaces
70 - 230
[<230°C]
air compressors
30 - 50
internal combustion engines
70 - 120
air conditioning and refrigeration condensers
30 - 40
Drying, baking, and curing ovens
90 - 230
Hot processed liquids/solids
30 - 230
Regarding low temperature sources, Table 2.2 presents more detailed information and some specific characteristics reported of these sources.
Table 2.: Low temperature sources [Alves 2004, Sass 2005, U.S. DOE 2008].
Source
Components
Phase
Temp. (°C)
Enthalpy (kJ/kg) or Cp (kJ/kg K)
Recovery
Constrains/
Limits
Boiler Blowdown
H2O, solids
Liquid
Saturated
Saturated Liquid
Flash Steam
BFW
pre-heat
Solid impurities
Exhaust gases exiting recovery devices
CO2, NOx, SOx, H2O, O2, N2, CO, CxHy
Gas
70 - 230
Cp = 1.0 - 1.3
Heat Recovery
LP-MP Steam Generation
Stack T. determined by sulphur content. Typically
(110-180°C)
Cooling water - annealing furnaces
H2O
Liquid
Saturated
Saturated Liquid
Flash Steam
BFW
pre-heat
Condenser cooling water
H2O
Liquid
35 - 45
Sub Saturated Water
Usage of available heat
Very low temperature
Quality of the Heat Source
Every different source must be assessed in order to determine the feasibility of recovering energy from it. Although the First Law of the Thermodynamics is normally employed to analyse the energy utilization, exergy analysis (Second Law of the Thermodynamics) is more suitable to describe the quality aspect of energy and it allows the determination of the useful work potential of a certain quantity of energy in a particular state [Saidur 2010]. The description of any low grade or waste heat source should include the following parameters [U.S. DOE 2008]:
The amount and purity/quality of heat.
Heat temperatures and the minimum permissible temperatures (e.g. stack temperature).
The composition (e.g. vapour, gas, liquid).
Operational availability of the heat source and logistics (e.g. distance from heat sink).
In addition to these parameters, site profiles could be useful tools to evaluate and identify different options for heat recovery within a total site. Individual waste heat streams from groups of processes can be compared using the total site profiles analysis in order to evaluate the overall potential for heat recovery [Smith 2005]. Nevertheless, it should be noted that site profiles include different streams or sources and therefore it is possible that one source alone is not able to provide useful heat.
Equipment for Heat Recovering
The most common devices used to recover heat are heat exchangers, they are used depending on the nature of the source along with the characteristics of the working fluid or solid that receives the heat transfer from the source. Thus, heat transferring occurs between gasses and/or liquids (e.g. combustion air preheating) and also between a liquid or gas stream with a solid that requires energy to be heated up (e.g. batch/cullet preheating in glass furnaces).
Heat Exchangers
A summary of different types of heat exchangers used for heat recovering is shown Table 2.3. Devices such as recuperators, regenerators, air preheaters, regenerative burners and economisers are among the most commonly used [Turner 2007, U.S. DOE 2008].
Recuperators: This type of equipment can recover waste heat from exhaust gases in the range of medium to high temperatures in gas-to-gas applications. For instance, it is commonly used for the combustion air preheating in gas incinerators, melting furnaces, afterburners and reheat furnaces among others. The operation mechanism is based on radiation, convection, or combination of both and its simplest configuration consists in two concentric tubes, as shown in Figure 2.1 (a), with the same length. Another common configuration is the convective recuperator (tube-type), where the cold air passes through a shell and the hot gases are conducted through small parallel tubes, Figure 2.1 (b).
Figure 2.: (a) Metallic radiation recuperator, (b) combined radiation and convective recuperator [Turner 2007].
Regenerators (heat wheel): The heat wheel is a rotatory regenerator used in the range of low to high temperatures. Turner [2007] reported that this type of heat exchanger has been used in the range of 20 to 1093°C. Some of their usual applications are for curing, drying ovens, space heating and heat-treat furnaces. According to Turner [2007], at least four types of heat wheels are available: i) A metal frame packed with a core of knitted mesh stainless steel, brass, or aluminum wire, ii) laminar wheel fabricated from corrugated materials, ii) laminar wheel constructed from a high-temperature ceramic honeycomb and iv) laminar wheel constructed of a fibrous material coated with a hygroscopic substance. Figure 2.2 presents a schematic of the heat wheel.
Figure 2.: Rotary regenerator (heat wheel) [U.S. DOE 2008]
Passive Air Preheaters: These devices are gas-to-gas regenerators where both streams, hot and cold, are maintained apart from each other, i.e. there is not cross-contamination. Passive air preheaters are used in the range of low to medium temperature, including baking ovens, drying, steam boilers, and air dryers, among others. There are two typical configurations used in these regenerators; the plate-type and the heat type, Figure 2.3 shows both types. The plate-type is made of alternate channels which separate both streams, thus, cross-contamination is avoided. However, they are bulkier, heavier and more expensive when compared with heat wheels. On the other hand, the heat type is assembled into arrays which are used as compact and efficient passive gas-to-gas heat exchangers. Its major drawback is the temperature range of the working fluid within the heat pipes, for instance, water can be used between 4 and 218°C [Turner 2007].
Figure 2.: Passive gas-to-gas regenerators, (a) the plate-type and (b) the heat pipe [Turner 2007].
Regenerative/Recuperative Burners: Burners that combine regenerative o recuperative configurations are simpler and more compact, regarding its design and construction, than independent regenerative furnaces or recuperators. In order to use the heat available in the flue gas stream, a self-recuperative burner incorporates heat exchanger surface as part of the burner body design. Normally, recuperative burner systems have less heat exchange area than regenerative burner systems and therefore their energy recovery is lower. However, they have lower related costs and its ease of retrofitting make them a feasible option for energy recovery [U.S. DOE 2008]. Figure 2.4 shows an example of a self-recuperative burner.
Figure 2.: Double P-tube with self-recuperative burner [Wuenning 2006]
Economisers (finned tubes): These types of equipment are usually used to heat up liquid streams using exhaust gases as heat source. They operate in range of low to medium temperatures applications such as boiler feedwater preheating, hot water for space heating and domestic hot water. The exchanger is constructed as a bundle of finned tubes where the liquid streams passes through them and the gas stream flows across the tubes. Figure 2.5 shows a finned tube heat exchanger where water is heated up by fuel gases from the boiler. This configuration is commonly known as boiler economisers [Tuner 2007].
Figure 2.: Finned Tube Exchanger/ boiler economisers [U.S. DOE 2008]
Table 2.: Heat Exchangers for Heat Recovering [Turner 2007, U.S DOE 2008, Chen 2012].
Type
Subtype/Mechanism
Temp. Level
Characteristics/Use/Heat transfer
Recuperators (e.g. Shell and tubes HE)
Radiation, convection or combination of both
Medium to high
Air preheating in annealing ovens, afterburners, gas incinerators, etc.
Regenerator (storing heat in a porous media)
Furnace Regenerator
High
Air preheating. Used with dirty exhausts.
Rotary Regenerator/Heat Wheel
Low to medium
Useful in air conditioning applications.
Passive Air Preheaters
Plate type and heat pipe
Low to medium
Gas to gas. No cross-contamination.
Regenerative/Recuperative Burners
High
Incorporates heat exchange surfaces as part of the burner body.
Finned Tube Heat Exchangers / Economizers
Liquid through tubes -> receive heat from hot gases
Low to medium
Gas to liquid. Heating liquid (BFW,process stream, etc.)
Waste Heat Boilers
Water tube boilers (steam gen.)
Medium to high
Gas to liquid-gas. Auxiliary burners or an afterburner can be added. Superheated steam requires additional superheater.
Low Temperature Heat Exchange
Deep Economizers
Cool exhaust gas to 65°C - 71°C
Low
Stainless steel tubes in the cold end. Glass tubed heat exchangers or other materials like Teflon.
Indirect Contact Condensation Recovery
Cool exhaust gas to 38°C - 43°C in a Shell and Tube HE
Low
H2O vapour in gases condenses almost completely. Stainless steel, glass, Teflon, or other advanced materials.
Direct Contact Condensation Recovery
Mixing process stream and cooling fluid.
Low
Contamination generated by substances in the flue gas.
Transport Membrane Condenser
Capturing H2O (latent heat) from the gas exhaust streams
Low
It avoids contamination; however, it requires more research.
There are other methods to recover energy, for instance, in boilers operating in the range of medium to high pressures and where large mass flows of steam are generated, the boiler blowdown can be used to produce flash steam at lower pressures. Smith [2005] reported that in large high-pressure boilers the blowdown rate is between 2 and 5%. Bahadori [2010] reported a method to calculate the energy that could be recovered from boiler blowdowns as a function of the pressure of the boiler and the flash drum. However, this stream it is not pure liquid since it contents impurities and undissolved solids, and the author does not mention the feasibility of the method in terms of these physical constraints. The method should consider filters or membranes to remove the solids and then obtain the full amount of energy that can be recovered.
Another potential heat recovery source are steam condensers, which use water as cooling medium. After condensing the steam coming from the steam turbine, the heated water needs to be cooled down to be used again. This represents a loss in efficiency in the process. Atmospheric air which is essentially free, could be used instead of water thus saving money in cooling water, nevertheless, there are physical constraints in such heat exchanger working with air and Tawney [2005] suggested that air cooled condensers (ACC) should be used in sites with scarce water resource. However, with the increasing conscious in water foot print it is worth to re-evaluate the use of air instead of water in condensers, not only in terms of the economics but also taking into account future environmental restrictions.
Wang [2012] reports the application of a low pressure economizer which heats the condensate exiting the condenser, before they passed to the deaerator, using the exhaust flue gas on a 600 MW power plant. Even though the configuration of such heat exchange is complex, it can be achieved and the results of the simulation indicate that fuel, coal, and make up water are saved; 2 to 4 g/kWh and 25 to 35 t/h respectively. However, the temperature of the exhaust gas at the end of the process is 98.2°C and the acid dew point is 98.5°C. The outlet temperature is around the limit of condensation of acids that could cause corrosion [Huijbregts 2004, ZareNezhad 2010].
Garimella [2012] reported an investigation to recover waste heat from a large (945 kg/s) exhaust flue gas flow at 120 °C. It is informed that through an absorption refrigeration system, 2.26 MW of heat available in the flue gas generate 1.28 MW of cooling capacity at 7 °C and 3.57 MW of heating capacity at 54 °C, where the released temperature of the flue gas is 117.6°C. It is not mentioned information concerning the stack temperature, thus, it can be assumed that 117.6 °C for this case is high enough to avoid condensation. Nevertheless, data regarding the stack temperature should always be included since it represents a physical constraint to recover energy from exhaust gases.
It can be found in the literature that many authors have conducted investigations regarding low grade heat recovery from flue gases, which denotes the importance of this heat source [Chawla 1999, Pintaric 2002, Garimella 2012, Wang 2012, Firdaus 2012]. They all reported their results, regarding energy recovery, considering the outlet temperature of the flue gas above the dew point. This is sensible and makes sense in terms of avoiding corrosion and destruction of the surfaces [Huijbregts 2004]. However, this fact limits the available heat content of the flue gases and consequently the thermal efficiency. If temperatures are reduced below the dew point then more energy could be recovered. The issue is to design equipment constructed with material capable to resist corrosive compounds. In this regards, condensing boilers are a suitable option.
Condensing Boilers
Cooling down flue gases below the dew point offers a good opportunity to recover latent (water) and sensible heat from these streams. Thus, thermal efficiency can be significantly improved, depending on other conditions such as excess air ratio, relative humidity of air and fuel type [Chen 2012]. Nevertheless, it is well established that the presence of corrosive compounds in flue gases produces failures in the heat exchanger's surfaces, for instance, the sulphur content in fossil fuels leads to the formation of SO2 and SO3 and these compounds, in addition to water vapour, combine to form sulphuric acid, which causes corrosion and destruction of the surfaces when it condensates [ZareNezhad 2010]. Corrosion failures generated by condensing flue gases containing SOx, NOx and H2O can be of different types: general corrosion, pitting attack and stress corrosion cracking [Huijbregts 2004].
There are materials that are suitable to resist corrosive compounds (e.g. PTFE, Aluminium, Stainless steel etc.) [Chen 2012], nonetheless, its high cost when compared with carbon steel represents the major drawback of these materials. However, with increased energy prices and legislative constrains they are likely to become more established practices. For instance, in modern condensing gas fired domestic boiler systems which are made of corrosive resistance materials, energy can be recovered from flue gases by cooling them down below the dew point, and efficiency can be increased from 75% to 90% [Comakli 2008]. In the case of industrial sites, some examples can be found in the literature, in fact, Chen [2012] report a few of these. Table 2.4 summarizes the examples reported.
Table 2.: Industrial examples of condensing boilers [Chen 2012].
Plant
Capacity
Capacity of condensers / heat pumps
Sodra Nas Vimmerby Energi AB Biomass District Heating, Sweden
10 MW
2 MW condenser
Kraftvarmeværk Waste Incineration, Denmark
3.3 MWe / 10.6 MWth
1 MW conden. and 17 GW h/y absorp. heat pumps
Davamyran Heat and Power, Sweden
350 GW h(th) / 80 GW h(e) per annum
11 MW condenser and 11.4 MW heat pumps
Vestforbranding Waste to Energy, Denmark
400 GW h(th) / 140 GW h(e) per annum
13 MW condenser
Chen [2012] reported a case study to evaluate the integration of a condensing boiler into a heat system, where the flue gas comes from a wood chip boiler with 40 MW of capacity and the temperature of the flue gas is around 150°C, an schematic of the configuration of the system is shown in Figure 2.6.
Figure 2.: Condensing boiler system adapted from Chen [2012].
They selected an E type shell (TEMA) single tube pass shell-and-tube heat exchanger for the design of the condensing boiler (Figure 2.6), and two corrosion resistant materials were evaluated in order to assess the economics; stainless steel 316 and carbon steel (<0,5% C) with a polypropylene coating in the outer surface of the tubes and in the inner surface of the shell to avoid corrosion. It was concluded that even though the addition of the polypropylene coating increases the required heat transfer area in approximately 20%, the Net Present Value calculations indicates that the choice of the carbon steel condenser ensures a shorter period of time for the cash return (2 years) when compared to stainless steel (5-7 years).
Practical Usage of Recovered Energy
The recovery of waste/low grade heat can only be beneficial if there is a potential use for it, such as power generation, combustion air and boiler feedwater preheating, steam generation, furnace load preheating, absorption refrigeration systems, and space heating among others.
Organic Rankine Cycle (ORC)
A suitable option to make use of low grade heat for power generation is the Organic Rankine Cycle (ORC) [Lee 1988]. The heat source is used to evaporate an organic fluid, which usually has a much lower boiling temperature than water. Organic Rankine Cycles are mainly used in large power plants and in some renewable energy plants such as solar and geothermal to produce power [Galanis 2009]. One of the advantages of ORC over conventional Rankine Cycle is its better performance in recovering heat from low temperature sources [Yamamoto 2001]. A simple ORC model is presented in Figure 2.7.
Figure 2.: Organic Rankine Cycle model, adapted from Yamada [2012].
The cycle consists of an evaporator, expander, condenser and pump. Depending on the pressure level in the turbine inlet, they can be classified as supercritical, subcritical and superheated ORCs [Dai 2009]. In the state point 1, the fluid exit the condenser as saturated liquid and then is compressed in the pump to the required pressure, state point 2. The working fluid is completely evaporated until it becomes superheated vapour in the evaporator, state point 3. The vapour is expanded to the condensing pressure (state point 4) and finally, the fluid enters the condenser where heat is removed until it converts in saturated liquid, state point 1.
Researches have studied and evaluated different working fluids to obtain the best energy recovery performance in ORCs [Drescher 2007, Dai 2009, Galanis 2009]. For instance, Chen [2010] evaluated 35 working fluids and analysed the influence of their properties in the performance of the ORC and found that isentropic and dry fluids are preferred. On the other hand, it was reported that wet fluids require superheating. However, the report did not include any analysis regarding the nature of the heat source, and this is a relevant factor that has a direct influence in the thermal match between the cycle and the energy source.
Drescher [2007] suggested that for cases where heat is recovered from biomass power plants, the best suitable fluids are those belonging to the family of the alkybenzenes since they showed the highest efficiencies. Nevertheless, it is worth mentioning that issues regarding safety and environmental performance are very important factors that should be taken into account when selecting a proper working fluid. Regarding safety aspects, auto ignition and flammability are quite important and within the environmental aspects, it should be included the ozone depletion potential and the atmospheric lifetime [Galanis 2009].
Quoilin [2009] reported several examples of site in Europe running Organic Rankine Cycles using different low grade heat sources. When the heat source is a flue gas, temperatures related to the source are above 120°C and only when the heat source is renewable (geothermal, solar) temperatures are lower. This dissertation evaluates the Organic Rankine Cycle using flue gases as low grade heat source at temperatures below 120°C.
Absorption Refrigeration Systems
The absorption refrigeration system is a close cycle where cooling capacity is generated from a heat source; this is the major difference with vapour compression refrigeration systems which utilizes mechanical work to drive the cycle. The absorption cycle uses a refrigerant-absorbent mixture known as the solution pair which should be environmentally friendly, for instance, ammonia-water or Lithium bromide-water [Joudi 2001, Pratihar 2010]. A single-effect absorption refrigeration system consists of four heat exchangers, namely, absorber, solution heat exchanger, generator, condenser and evaporator. Figure 2.8 shows a schematic of the process.
Figure 2.: Absorption refrigeration cycle, single-effect.
Refrigerant in liquid phase (2) passes through the expansion valve (3) and pressure is reduced and partial evaporation takes occurs, then evaporation is completed in the evaporator by taking heat from the source that is being cooled down and thus, cooling capacity is produced. The coefficient of performance (COP) of single-effect systems varies between 0.5 and 0.7 [Nimpunyakampong 2011]. To drive the cycle, the refrigerant needs to be boiled (1) in the generator and therefore, heat is required to do so and it represents the major energy input of the cycle. Low grade/waste heat is a suitable source to be used in the generator (Qgenerator), this is one of the reasons that makes heat powered refrigeration systems a viable economic option [Colonna 2003].
In addition to the single-effect system, there are two more well-known configurations for absorption refrigeration systems; the double-effect and the multi-effect. The double-effect absorption system is similar to the single-effect shown in Figure 2.8 with the difference that includes an extra generator. Hence, the temperature of the refrigerant exiting the generator is not limited by its condensing pressure, as it is in the single-effect configuration, and therefore higher COPs are achieved in double-effect systems, between 1.0 and 1.5. Another important advantage of this configuration is related to its capacity to operate with heat sources at higher temperatures than single-effect systems [Nimpunyakampong 2011]. In multi-effect systems, the difference is that more than two generators are used and therefore the system can operate with heat sources at higher temperatures and achieve better COP. However, the challenge is focused in the equipment design due to the number of effects and the temperature resistance of typical construction materials.
Cooling Capacity for the Brayton Cycle
The performance of gas turbines is highly dependent on the temperature of the inlet air of the compressor, i.e., the ambient temperature. High temperatures of the air restrict the air mass entering the compressor and therefore reduce the output power of the gas turbine. It is reported that that a rise of 1°C in the ambient temperature results in 1% drop from the gas turbine rated capacity [Mohanty 1995] and its output power reduces around 0.5% to 0.9% [Popli 2012]. Consequently, it can be concluded that an improvement in the efficiency of the Brayton cycle can be achieved by cooling down the temperature of the intake air in the compressor. Figure 2.9 presents a schematic of the system.
Figure 2.: The Brayton Cycle - Cooling of intake air.
Popli [2012] evaluated, in terms of thermodynamic and economic feasibility, three options to reduce the inlet temperature of air considering United Arab Emirates ambient conditions; i) evaporative media coolers, ii) electrically driven mechanical vapour-compression chillers and iii) single-effect water-lithium bromide absorption chillers. The results showed that three single-effect absorption chillers (driven by steam), using 17 MW of waste heat from the gas turbine, could provide 12.3 MW of cooling to cool down the inlet air temperature to 10°C. On the other hand, for the same ambient conditions (55°C), evaporative coolers would only provide 2.3 MW of cooling capacity and mechanical vapour-compression chillers would require 2.7 MW of additional electric energy to provide the same amount of cooling as absorption chillers. The economic payback period for the absorption refrigeration system is between 1.3 and 3.4 years. It should be noted that the turbine exhaust was used to generate steam to drive the absorption chillers; the objective of this dissertation is to evaluate thermodynamically the absorption refrigeration system, using a flue gas stream directly as energy source into the generator.
Cooling Capacity for the Steam Rankine Cycle
Probably the most widely used method to produce power is the Rankine cycle where the main goal is to produce the maximum power at high thermal efficiency. One way to achieve this is by reducing the outlet pressure of the steam turbine (condenser pressure) and thus increasing the enthalpy drop. Consequently, the turbine output work is increased as well [Bekdemir 2002]. Usually, the condenser operates with water at ambient temperature as cooling medium; therefore, the minimum allowable pressure is set by this temperature. Absorption refrigeration could be used to produce cooling water at lower temperatures for the condenser, thus the outlet pressure of the steam turbine can be reduced to produce more power. Figure 2.10 shows the schematic of system.
Figure 2.: Steam Rankine Cycle - Cooling water at low temperature
Haseli [2008] investigated the optimum cooling water temperature in the steam condenser through an exergy analysis. It was reported that for a fixed steam mass flow (1 kg/s), when the condensing temperature increases from 46°C to 54°C, the optimum cooling water temperature increases from 16.78°C to 25.17°C and exergy destruction decreases from 172.5 kW to 164.6 kW. It can be observed that the lower the temperature of condensation of steam, the temperature of the cooling water should be lower as well. However, the report does not mention the consequence of reducing the condensing temperature in the power generation.
From the literature review it can be observed that research is being directed toward low grade heat recovery to enhance thermal efficiency within chemical and utilities processes. Due to the increase of fuel prices, these practices are becoming more common in the industrial sector. Flue gases are large sources of energy which are usually utilized to preheat combustion air and the feedwater in boilers; however, these gases are discharged to the atmosphere at relative high temperatures (120°C - 200°C). There is still scope for energy recovery to produce useful work through flue gases being discharged at temperatures lower than 120°C. Thus, this dissertation evaluates the power generation using a flue gas stream as heat source in the Organic Rankine Cycle. Additionally, it evaluates the use of the same source to produce cooling capacity through an absorption refrigeration system in order to enhance the power production in the Brayton Cycle and the Steam Rankine Cycle.
Methodology and Modelling
Overview
Methodology
In order to achieve the main aims of this dissertation, which are the evaluations of the potential of low grade heat sources to produce power in the Organic Rankine Cycle (ORC) and to enhance the power generation in the Brayton Cycle and the steam Rankine Cycle, appropriate models have to be constructed to study the cycles.
The characterisation of the heat source in terms of the quantity of heat available is relative simple when temperatures, mass flows and heat capacities are known, however, the evaluation of the heat source regarding its quality is more complex; therefore a model is required to obtain the relevant data. Likewise, a model of the Organic Rankine Cycle is necessary to examine its performance in power generation regarding the heat source. The major parameters of interest in the ORC are the power generation and the thermal efficiency; therefore, heat loads in the turbine, the evaporator and pump are essential outputs of the model.
On the other hand, cooling capacity is required to improve the power generation in the Brayton Cycle and also in the steam Rankine Cycle. Thus, apart from the models that simulate the Brayton Cycle and the steam Rankine Cycle, an absorption refrigeration system model is necessary to assess the cooling capacity generation through the use of low grade heat. The outputs of interest of this model are the heat input and the cooling capacity generation of the system. The models have to be stand-alone and also be capable of being integrated with other models. Thus, it is required a modelling software with the capacity to model and simulate complex processes.
Aspen HYSYS
All models were constructed and simulated in Aspen HYSYS process simulation software [Aspen Technology, Inc. 2006]. These models include material, heat and work streams representing the heat source, the Organic Rankine Cycle, absorption refrigeration system, the Brayton Cycle and steam Rankine Cycle. Regarding the advantages offered by the software, in addition to its simple user interface, HYSYS offers a vast database of chemical properties for many components, in particular for those used in this work.
Modelling of Low Grade Heat Sources
Flue Gas Stream
The selected heat source is a flue gas stream. It can be found in the literature a wide range of mass flow rates (1 kg/s - 300 kg/s) for flue gas streams exiting boilers and gas turbines of different capacities. In this work the mass flow rate was fixed in 5.6 kg/s (20000 kg/h), unless indicated otherwise. The composition of the flue gas depends on the fuel used; Table 3.1 presents some values regarding flue gas compositions reported in the literature.
Table 3.: Composition of flue gases.
Molar fraction
Component
[Alves 2004]
[Sass 2005]
[Chen 2012]
CO2
0.029
0.087
0.121
O2
0.105
0.017
0.032
N2
0.610
0.721
0.603
H2O
0.257
0.174
0.244
The selected composition for this work (Table 3.1) is obtained through an arithmetic average of the reported compositions shown in Table 3.1. Concerning what temperature should be considered for the flue gas, generally researchers do not use the same temperature to describe a low grade heat sources, nevertheless, many of them agree that this temperature should be below 250°C. For instance, Galanis [2009] considers 150°C, U.S. DOE [2008] 250°C and C. Wang [2012] 123°C. In this work the temperature of the flu gas stream, 180°C, is calculated through an arithmetic average of six reported temperatures that describe low grade heat sources [U.S. DOE 2008, Galanis 2009, Saidur 2010, Wang 2012, Chen 2012]. The dew point temperature of the stream is calculated by the simulation.
Enthalpy and Exergy Analysis of the Flue Gas
To describe the quality and quantity of the heat source, a simple fuel gas stream is simulated in HYSYS according to the characteristics presented in section 3.1.1, using the Peng-Robinson equation of state as the fluid package. The stream is simulated at different temperatures keeping the pressure constant at 100 kPa and thus enthalpy and entropy data is extracted.
Enthalpy Change
The enthalpy change is calculated using the ambient conditions as reference (Section 4.1), thus the enthalpy change is given by equation 3.1:
(3.1)
Where:
Exergy
The specific exergy of the source is calculated with the extracted data of entropy, considering ambient conditions as reference, thus its value is given by equation 3.2:
(3.2)
Where:
Organic Rankine Cycle
The model for the organic Rankine Cycle is provided by the Centre for Process Integration (CEAS) of the University of Manchester. The cycle utilizes 1,1,1,2-Tetrafluoroethane (R-134a) as working fluid, which is reported to be a suitable option when the heat source is at low temperature and it can be applied in a supercritical Rankine Cycle [Velez 2012, Chen 2010]. In terms of environmental impact, R-134a has zero Ozone Depletion Potential and its Global Warming Potential is 1300 [Roy 2011]. The Peng-Robinson equation of state is the fluid package used in the simulations and the objective of the model is to evaluate the performance (specific power generation and thermal efficiency of the cycle) using a low grade heat source as energy input. The schematic diagram of the simulation is shown in Figure 3.1.
Figure 3.: Simple Organic Rankine Cycle, HYSYS modelling.
The physical properties of R-134a are presented in Table 3.2.
Table 3.: Physical properties of R-134a.
Properties
Value
Molecular Weight
102
Normal Boiling point (°C)
-26
Ideal Liquid Density (kg/m3)
1242
Critical Temperature (°C)
101
Critical Pressure (kPa)
4055
The simulation includes the following devices and parameters:
Evaporator: Simulated using a shell and tube heat exchanger without pressure drop and utilizing the Exchanger Design (Weighted) model due to the change of phase involved in the process. R-134a passes through the tubes and the flue gas through the shell. The temperature and mass flow rate of the heat source are fixed and thus the simulation calculates the mass flow rate of R-134 that can be evaporated at the highest possible pressure, which depends on the temperature of heat source. Qin (kW) is the energy delivered by the flue gas to the R-134a fluid and it is calculated by the software.
ORC-Turbine: Simulated using an expander with 75% of adiabatic efficiency. This value is in the range of the efficiencies for the organic turbine reported by Kang [2012], and it is also the turbine efficiency used by Quoilin [2009] in his optimization of the ORC. The power generated, Worc-turbine (kW), depends on the amount of R-134a evaporated in the boiler. The outlet pressure was kept constant and is set by temperature of the condenser. Typical values of this temperature are in the range of 25°C - 60°C [Quoilin 2009].
Condenser: A cooler, with no pressure drop, was used to simulate the condenser. The temperature of the condenser is usually fixed by the accessible cooling water [performance analysis] and therefore in this simulation was set at 30°C.
Pump: The power consumption is represented by Wpump (kW). Regarding the efficiency of the pump, Quoilin [2009] reported an efficiency of 80% for the pump of the simulated ORC, likewise, Yamada [2012] reported an efficiency of 60%. Thus, the adiabatic efficiency in this simulation is set at 75%.
It is reported that the thermal efficiency for a simple Organic Rankine Cycle (R-134a), similar as shown in Figure 3, increases with the turbine inlet pressure [Roy 2010, Roy 2011] for pressures below the critical point of R-134a. In addition, Yamada [2012] reported that the thermal efficiency for the same type of cycle (Figure 3.0) increases with the inlet temperature of the turbine for pressures above the critical pressure (supercritical ORC). From the latter report, it can also be observed that for turbine inlet temperatures below 120°C, the subcritical ORC attains better thermal efficiencies (10%) than the supercritical ORC (8%), where the pressure for the subcritical condition is set at 94% of the critical pressure. For turbine inlet temperatures from 120°C and above, supercritical ORC achieves better thermal efficiencies, approximately 12%.
Consequently with the literature review, the pressure in the pump outlet is set at the highest level possible, according to with the restrictions imposed by the heat source, such as heat transfer and the temperature cross condition in the evaporator. Based on the parameters used by Yamada [2012], for turbine inlet temperatures above 120°C, the pressure for the supercritical condition is set in a 20% higher than the critical pressure, namely, 4866 kPa.
The organic Rankine Cycle thermal efficiency is:
(3.3)
Steam Rankine Cycle
Two models for the steam Rankine Cycle are developed to evaluate different parameters and behaviours. The first model is used to study the impact of the heat source temperature on the specific power production and cycle thermal efficiency. In addition, this model is compared with the Organic Rankine Cycle. The second model is built with the aim of assessing the effect of the condenser temperature on the power production increase and its thermal efficiency.
Steam Rankine Cycle - The Heat Source Temperature
This model is developed using the ASME steam fluid package contained in HYSYS and its objective is to evaluate the impact of the heat source temperature on the specific power production and the thermal efficiency of the cycle. In addition, the model is compared with the Organic Rankine Cycle regarding the temperature of the heat source. As stated in section 3.1, the flue gas was simulated using the Peng-Robinson fluid package. The schematic diagram of the simulation is presented in Figure 3.2.
Figure 3.: Steam Rankine Cycle, heat source temperature impact, HYSYS modelling
The simulation includes the following devices and parameters:
Boiler: Simulated using a shell and tube heat exchanger without pressure drop and utilizing the Exchanger Design (Weighted) model due to the change of phase involved in the process. Water passes through the tubes and the heat source, flue gas, through the shell. The temperature and mass flow rate of the heat source are fixed and therefore the simulation calculates the amount of steam that can be produced at the highest possible pressure, considering a superheated temperature of 10°C above the saturation temperature. However, when higher pressure levels were reached, this value was increased. The minimum approach temperature assumed was 10°C. Qin (kW) is the energy delivered by the flue gas to the water/steam and it is given by the software.
Steam Turbine: Simulated using an expander with 75% of adiabatic efficiency. This value is within the efficiency range of steam turbines reported by Bahadori [2010], additionally, is the same efficiency assumed for the turbine in the ORC, allowing a sensible comparison between both cycles. Wturbine (kW) is the power generated and depends on the steam produced in the boiler. The outlet pressure was kept constant at 10 kPa.
Condenser: A cooler, with no pressure drop, was used to simulate the condenser.
Pump: The adiabatic efficiency was set at 75%. This value might seem low, however, it was selected to maintain the setting parameters of both cycles (organic and steam) as similar as possible. Wpump (kW) represents the power consumption.
The cycle thermal efficiency is:
(3.4)
Steam Rankine Cycle - The Condenser Temperature
The objective of the modelling is to evaluate the temperature level of the cooling medium used in the condenser; therefore, a simple steam Rankine Cycle is simulated using the ASME Steam fluid package to produce 36000 kg/h of superheated steam at 9000 kPa and 400°C. It must be noted that this cycle is not driven by the low grade heat source. A schematic diagram of the simulated cycle is shown in Figure 3.3.
Figure 3.: Simple steam Rankine Cycle, HYSYS modelling.
The base model includes:
Boiler: A heater without pressure drop was used to simulate a boiler that produces 36000 kg/h of superheated steam at 9000 kPa and 400°C. Qin (kW) is the energy required to produce the steam and as stated before, it is not provided by the low grade heat source.
Steam Turbine: An expander with 75% of adiabatic efficiency was selected to simulate the steam turbine. The arguments to select this value for the efficiency were presented in section 3.4.1. The initial outlet pressure is set in 100 kPa and Wturbine (kW) is the power generated.
Condenser: This was simulated with a shell and tube heat exchanger using the Exchanger Design (Weighted) model due to the change of phase of the steam. Water at 20°C and 100 kPa was used as cooling medium passing through the tubes, without pressure drop, and steam passes through the shell with negligible pressure drop. The cooling water conditions just mentioned were selected according to the ambient conditions assumed in this work (Section 4.1). The outlet temperature of the cooling water was fixed and therefore the simulation calculates mass flow required, i.e., the cooling capacity. For practical reasons, a minimum approach temperature of 10°C was selected for the condenser.
Pump: With a 75% of adiabatic efficiency and a constant pressure difference of 8900 kPa. Wpump (kW) is the power required by the pump.
The cycle thermal efficiency is calculated as it follows:
(3.5)
The outlet pressure of the turbine is reduced in order to evaluate the power generation and cycle thermal efficiency. The pressure reduction leads to a temperature reduction as well, thus cooling water is required at lower temperatures than ambient conditions and therefore a refrigeration system is needed. The production of low temperature cooling water is evaluated with the absorption refrigeration system driven by low grade heat.
The Brayton Cycle
The model of the Brayton Cycle was developed by Saddiq 2012. The fluid package used to simulate the cycle was the Peng-Robinson Twu (PR-Twu) equation of state, and in this work a simple model is utilized in order to assess the impact of the temperature of the inlet air of the compressor. A schematic diagram of the Brayton cycle is shown in Figure 3.4.
Figure 3.: Brayton cycle, HYSYS modelling.
The devices and conditions included in this model are:
Compressor: The polytropic efficiency of the compressor is set in 86% and the operating mode is centrifugal. Wcompressor (kW) is the power required to compress air at ambient conditions (100 kPa) to 1755 kPa.
Natural Gas: Is the fuel selected and it has a molar composition of 73.1% of methane, 23.8% of ethane, 2.8% of nitrogen and 0.3 of Hydrogen sulphide at 15°C and 1755 kPa.
Combustor: This was modelled with a conversion reactor, without pressure drop, where the reaction was: CH4 (methane) + 2·O2 (oxygen) ïƒ CO2 (carbon dioxide) + 2·H2O (water)
Gas Turbine: An expander was used with a polytropic efficiency of 84.3%; this value was proposed by Saddiq 2012. Wgas-turbine (kW) is the power produced.
The cycle thermal efficiency is obtained by:
(3.6)
Absorption Refrigeration System
The HYSYS model for the single-effect absorption refrigeration system used in this work was elaborated by Nimpunyakampong [2011]. This model was simulated using the sour Peng-Robinson fluid package contained in the HYSYS database. The objective of this model is the evaluation regarding the cooling capacity generation of the absorption refrigeration system, using ammonia-water as refrigerant and absorbent respectively. The schematic of the system is shown in Figure 3.5.
Figure 3.: Absorption Refrigeration System (Ammonia/Water)
The description of the system and its devices is as it follows [Nimpunyakampong, 2011]:
The Absorber: Was modelled using a refluxed absorber where the two input streams are fixed. The absorber operates at 354 kPa and there is not pressure drop within the 15 stages, in addition, total condenser is assumed the distillation outlet stream is null.
Pump: The outlet pressure of the pump is set in 1167 kPa.
Heat Exchanger (SHE): This was simulated with a shell and tube heat exchanger using the Exchanger Design (Weighted) model to address any change of phase, where the cold stream passes through the tubes and the hot steams through the shell.
Generator (Desorber): A distillation column without pressure drop was used to model the generator. The mole fraction of the light key component in the bottom, ammonia, is set in 0.3355 and the mole fraction of the heavy key component in the distillate, water, is set in 0.00. The external reflux ratio is 0.25. An extra heater (Temp. Adjust 2) is required to heat up the distillate at 100°C, finally, the sum of Qg1, Qg2 and Qg2 (Figure 3.5) is heat input in the generator.
Condenser: It was modelled with a cooler without pressure drop. The outlet temperature of the condenser, which is saturated liquid of the refrigerant ammonia, is set in 30°C.
Evaporator: The objective of the evaporator is to receive heat from a low temperature source to evaporate the ammonia at constant temperature, 0°C. Hence, a heater was selected without pressure drop. Is in this equipment where the cooling capacity is generated.
Expansion Valves: These devices are used to reduce the pressure (flow expansion). Valve 1 expands the ammonia from 1167 kPa to 510 kPa, and Valve 2 expands the weak solution from 1167 kPa to 354.4 kPa.
The coefficient of performance (COP) of the system is obtained by:
(3.7)
Where,
Qe (kW) is the heat required in the evaporator and Qg (kW) in the heat input in the generator.
Finally, as stated in section 1.2, the economic analysis is not included in the objectives of this dissertation and therefore there is not a methodology used in this regarding.
Results and Discussion
The results obtained in the simulations and the calculations made to evaluate the potential for extra power production, using the flue gas stream described in section 3.1 as heat source, are presented in this section in the following order:
Selected ambient conditions.
The results of the flue gas stream simulation; enthalpy and exergy analysis.
The Organic Rankine Cycle simulation and its comparison with steam Rankine Cycle. Evaluation of the performance for both cycles regarding the temperature of the flue gas stream.
The results of the Brayton cycle model and its dependency of the net power production regarding the temperature of the atmospheric air.
The steam Rankine Cycle and its relation with the condenser cooling capacity (temperature and pressure).
The power production in the Organic Rankine Cycle using the flue gas stream as energy source, and the absorption refrigeration system simulation and its capacity to generate cooling, using the same energy source.
Comparison between the power production in the Organic Rankine Cycle and the extra power production in the Brayton Cycle and the steam Rankine Cycle.
Ambient conditions
Taken into account the fact that this work is evaluating the use of a low grade heat source to produce cooling capacity and generate extra power in a Brayton Cycle and also in a steam Rankine Cycle, the conditions selected were:
Atmospheric temperature: 35°C
Atmospheric pressure: 100 kPa
Cooling water temperature: 20°C
The relative humidity of air was considered negligible since it was not incorporated in simulation model delivered by Saddiq 2012, and therefore it was not included in the characteristics of the air intake in the Brayton cycle.
Flue Gas Stream
As stated in section 3.1.1, the available low grade source has a temperature of 180°C with a mass flow of 5.6 kg/s (20000 kg/h). The composition of the stream is presented in Table 4.1.
Table 4.1: Composition of the flue gas stream.
Table 4.: Composition of the flue gas stream.
Component
Molar fraction
CO2
0.079
O2
0.051
N2
0.645
H2O
0.225
Enthalpy and Exergy Analysis
The dew point temperature calculated by HYSYS is 62.5°C and the simulations delivered the mass enthalpy and entropy of the stream at different temperatures. The enthalpy change along with the exergy of the flue gas is calculated through equations 3.1 and 3.2, using the ambient conditions as reference state. The results are presented in Figure 4.1.
Figure 4.: Enthalpy and Exergy of the flue gas stream.
Figure 4.1: Enthalpy and Exergy of the flue gas stream.
As expected, the graph in Figure 5.1 shows that at higher temperatures there are more enthalpy and exergy available in the heat source. In terms of heat quantity, which is described by the enthalpy, it can be noticed that is much bigger that the quality of the source, which is represented by the exergy. The exergy denotes the capability of the flue gas stream to produce useful work [ref.] and from equation 3.2 it is observed that exergy is a fraction of the enthalpy content. The breakpoint of the trend line of both the enthalpy and exergy is generated as consequence of the dew point.
The exergy/enthalpy ratio of the flue gas stream is presented in Figure 4.2.
Figure 4.: Exergy/Enthalpy ratio.
Figure 4.2: Exergy/Enthalpy ratio.
The exergy/enthalpy ratio is a useful measure that represents the maximum percentage of the enthalpy content that can be recoverable as shaft work. For instance at 160°C the ratio is 0.09, which means that a maximum 9% of the enthalpy available at this temperature can be transformed in power.
Regarding power generation through the well-established Rankine Cycles, it is known that the temperature level of the heat source has a relevant influence in the thermal efficiency in both cycles; steam and organic. Thus, the next section presents the results in terms of the thermal efficiency achieved in the simulations of the Organic Rankine Cycle and the steam Rankine Cycle, and its dependency on the temperature of the flue gas stream.
The heat source temperature and its influence in the thermal efficiency in the Organic Rankine Cycle and the steam Rankine Cycle.
This section presents the results obtained in the simulations for a fixed mass flow of the flue gas stream, 20000 kg/h. The variable parameter was the inlet temperature of heat source in the evaporator (Figures 3.0 and 3.1).
Organic Rankine Cycle Simulations (ORC)
As presented in section 3.2, this cycle was simulated in supercritical conditions to achieve high efficiencies; however, these conditions are constrained by the temperature of the heat source. This means that if the temperature of flue gas is not high enough, the cycle must be simulated in subcritical conditions. It has been assumed that cooling water is available at 20°C (section 4.1) and thus the outlet temperature of the condenser (Figure 3.) was set in 30°C. The results are shown in Table 4.2.
Table 4.2: ORC simulation results
Table 4.: ORC simulation results
Flue Gas
R-134a
ORC Turbine
Flue Gas
Pump
Tin (°C)
Tin (°C)
Pout pump (kPa)
T sat (°C)
Tout (°C)
Mass flow kg/h
Power (kJ/kg)
Qin (kJ/h)
W (kJ/h)
80
31
1800
63
70
1159
13
230353
1340
100
31
2100
69
90
3335
17
721902
4978
120
34
4866
101
110
7442
19
1215493
34214
140
34
4866
101
120
8492
25
1711218
39040
150
34
4866
101
130
8829
29
1959905
40591
160
34
4866
101
150
8707
34
2209154
40030
180
34
4866
101
170
9617
38
2709367
44212
220
34
4866
101
200
11559
43
3716839
53140
240
34
4866
101
220
12153
46
4224182
55870
When the temperature of the source is below 120°C, in this case 80°C and 100°C, the cycle operates in subcritical conditions and consequently the efficiencies are the lowest. These results are similar to the ones reported by Yamada [2012], where it can be observed that the efficiencies for the supercritical cycle are higher than those for subcritical cycles, for turbine inlet temperatures above 120°C and pressure above the critical. The maximum thermal efficiency reached by the cycle is around 12% when the turbine inlet temperature is 160°C. Table 4.2 shows that at constant pressure, superheating the fluid above 160°C does not improve the thermal efficiency
