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POTENTIAL OF ORGANIC RANKINE CYCLES (ORC) FOR WASTE HEAT RECOVERY ON AN ELECTRIC ARC FURNACE (EAF)

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The organic Rankine cycle (ORC) is a mature technology to convert low temperature waste heat to electricity. While several energy intensive industries could benefit from the integration of an ORC, their adoption rate is rather low. One important reason is that the prospective end-users find it difficult to recognize and realise the possible energy savings. In more recent years, the electric arc furnaces (EAF) are considered as a major candidate for waste heat recovery. Therefore, in this work, the integration of an ORC coupled to a 100 MWe EAF is investigated. The effect of working with averaged heat profiles, a steam buffer and optimized ORC architectures is investigated. The results show that it is crucial to take into account the heat profile variations for the typical batch process of an EAF. An optimized subcritical ORC (SCORC) can generate an electricity output of 752 kWe with a steam buffer working at 25 bar. However, the use of a steam buffer also impacts the heat transfer to the ORC. A reduction up to 61.5% in net power output is possible due to the additional isothermal plateau of the steam.
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POTENTIAL OF ORGANIC RANKINE CYCLES (ORC) FOR WASTE HEAT
RECOVERY ON AN ELECTRIC ARC FURNACE (EAF)
Steven Lecompte*a, Oyeniyi A. Oyewunmib, Christos N. Markidesb, Marija Lazovaa, Alihan Kayaa, Bernd Ameela,
Martijn van den Broeka, Michel De Paepea
*Author for correspondence
aDepartment of Flow, Heat and Combustion Mechanics, Ghent University, Sint-Pietersnieuwstraat 41, 9000 Gent,
Belgium
bClean Energy Processes (CEP) Laboratory, Department of Chemical Engineering, Imperial College London,
London SW7 2AZ, United Kingdom
E-mail: steven.lecompte@ugent.be
ABSTRACT
The organic Rankine cycle (ORC) is a mature technology
to convert low temperature waste heat to electricity. While
several energy intensive industries could benefit from the
integration of an ORC, their adoption rate is rather low. One
important reason is that the prospective end-users find it
difficult to recognize and realise the possible energy savings. In
more recent years, the electric arc furnaces (EAF) are
considered as a major candidate for waste heat recovery.
Therefore, in this work, the integration of an ORC coupled to a
100 MWe EAF is investigated. The effect of working with
averaged heat profiles, a steam buffer and optimized ORC
architectures is investigated. The results show that it is crucial
to take into account the heat profile variations for the typical
batch process of an EAF. An optimized subcritical ORC
(SCORC) can generate an electricity output of 752 kWe with a
steam buffer working at 25 bar. However, the use of a steam
buffer also impacts the heat transfer to the ORC. A reduction
up to 61.5% in net power output is possible due to the
additional isothermal plateau of the steam.
INTRODUCTION
Vast amounts of thermal energy from various process
industries (in the form of flue-gas exhausts, cooling streams,
etc.), are currently being wasted by disposal into the
environment. These streams are generally considered to be low-
to medium-grade temperature and as such cannot be efficiently
utilized for conversion into power or shaft work by traditional
heat engines such as the steam Rankine cycle. However, the
recovery and reuse of these waste-heat streams can
significantly improve the energy and economic efficiencies of a
lot of process plants across a broad range of industries. Thus,
the deployment of suitable heat engines capable of efficiently
recovering and converting the wasted heat to power has been
identified as one of the major pathways towards a high
efficiency, sustainable and low-carbon energy future [1-3].
Various heat engines have been proposed for the
valorisation of low-temperature heat sources. Prime examples
of these engines include the organic Rankine cycle [4], the
Kalina cycle [5], the Goswami cycle [6] and supercritical
carbon dioxide (s-CO2) cycles [7]. Other novel engine
configurations include various thermoacoustic and
thermofluidic heat engines [8-10].
The organic Rankine cycle (ORC) in particular is an
attractive proposition due to its similarity with the well-
established steam Rankine-cycle engine and the accompanying
wealth of operational and maintenance experience.
Furthermore, there is the design option of employing a number
of organic working fluids, ranging from refrigerants to
hydrocarbons and siloxanes [11], including working fluid
mixtures [12,13], to optimize the heat transfer (and heat
recovery) from/to the waste heat source and heat sink. ORC
engines also feature quite a number of architectures such as the
transcritical cycles [4,14], trilateral cycles [15], partial
evaporation cycles [4] and the basic subcritical cycles, to better
suit the characteristics of the heat source/sink.
NOMENCLATURE
Abbreviations
PPTD
Pinch point temperature difference
ORC
Organic Rankine cycle
SCORC
Subcritical ORC
TCORC
Transcritical ORC
PEORC
Partial evaporating ORC
EAF
Electric arc furnace
ORC systems have been studied and deployed for a variety
of applications and in various energy intensive industries,
spanning scales from a few kW to tens of MW. More recently,
ORCs have been applied for waste-heat recovery from
automobile and marine prime movers such as internal
combustion engines and diesel engines [16]. They have also
been applied for power generation and energy efficiency on
offshore oil and gas processing platforms [17], sometimes in
combination with windfarms [18]. In addition, over the years,
the ORC has also seen applications in the petroleum refining
industries from multiple waste-heat sources [19,20], and in heat
recovery from (rotary) kilns in the cement and steel industries
[21,22]. In these systems, an intermediate oil or water loop is
employed to recover heat from the waste-heat stream before the
heat is subsequently passed on the working fluid in the organic
Rankine cycle.
In the steel industry in particular, being one of the highest
energy and emission intensive sectors, there are multiple
avenue for waste heat recovery with ORCs from the various
cooling water loops in ore smelting furnaces. The electric arc
furnaces (EAF) are considered as the major candidates for
waste heat recovery applications [22]. Three different layouts
can be conceived: heat exchangers can be placed directly
outside the furnace (2000-1600 °C), just after the post
combustion (200–900 °C) or they can recover heat by replacing
the dry cooler. Inlet gases into conditioning system after the dry
cooler have temperature values of 80140 °C.
Turboden implemented the first ORC-based heat recovery
plant on an electric arc furnace at Elbe-Stahlwerke Feralpi
located at Riesa, Germany [23]. The furnace capacity is 133 t/h
with a tap-to-tap time of 45 min. The new 3 MW electrical
output ORC unit exploits a portion of the saturated steam
produced and recovers heat from the exhaust gases. The heat
recovery system was started up on December 2013. Since the
EAF is a batch process with high variable heat flow, the
inclusion of a steam accumulator was crucial. Compared to
thermal oil loops, which work at higher temperatures (280-310
°C), the inclusion of a steam buffer at 26 bar (228-245 °C)
implies a reduction in ORC efficiency. Using thermal oil was
however ruled out due to safety reasons (i.e., flammability).
The heat recovery process is divided into two main sections. In
the first section, the flue gas at a temperature of 1600 °C is
cooled by evaporative cooling. Cooling water at boiling point is
fed from the steam drum to the cooling loop. This replaces the
old cold water cooled ducts. The second section replaces the
existing water quench tower (comparable to the dry cooler in
the case presented in section 2). A Waste Heat Steam Generator
(WHSG) with vertical tubes was installed. The WHSG includes
an evaporator, superheater and economizer. The payback time
depends on the use of the steam. Direct use of the steam gives a
pay-back time of approximately 2 to 3 years. While using the
steam for a power generating system has a pay-back time of
approximately 5 to 6 years.
THE ELECTRIC ARC FURNACE
Figure 1 Lay-out of the electric arc furnace.
Data from an operational 100 MWe electric arc furnace in
Belgium is taken for this case study. The layout of the plant is
shown in Figure 1. There are two main heat sources available:
the low temperature water cooling loops and the exhaust gas
after the wet duct. The temperature in the water cooling loops is
around 30 °C to 40 °C. For the flue gasses, the minimum,
maximum and average temperatures and mass flow rates are
reported in Table 1. A typical waste heat profile from the flue
gas is shown in Figure 2. It is clear that the heat available for
recuperation shows a high variation in time. For the case under
consideration there are two specific constraints. First, the flue
gas is available directly before the dry cooler. Secondly, the
temperature of the flue gas entering the filters should be
between 80 °C and 140 °C. In the remainder of this work, only
the flue gasses are considered for waste heat recovery due to
the very low temperature of the water cooling loops and the
high thermal capacity of the flue gas.
Table 1 Details of the flue gas heat profile.
Variable
Max.
Min.
Temperature [°C]
578
68
Mass flow rate [Nm³/h]
185026
0
Heat transfer rate [MW]
28
0
Figure 2 Available heat during the batch process (reference
temperature of 90 °C)
INTEGRATION OF THE ORGANIC RANKINE CYCLE
Organic Rankine cycles (ORC) offer the possibility to
generate electricity from low capacity and low temperature heat
sources. The choice for an ORC to convert heat to power is
influenced by the maturity, simplicity and cost-effectiveness of
the technology.
Figure 3 Component layout of the basic organic Rankine
cycle.
Conceptually, the ORC is based on the classic (steam)
Rankine cycle. The main difference is that instead of water, an
alternative working fluid is used. Due to the organic working
fluid, a low boiling point is attainable. This is beneficial for low
temperature heat recovery. Also, the volume ratio of turbine
outlet and inlet can be reduced compared to water. This allows
using smaller and hence cheaper expanders. A reduced specific
enthalpy drop permits single-stage turbines instead of the
costlier multi-stage machines. Further benefits include: low
0
20
40
60
80
0:00:00
0:07:20
0:14:40
0:22:00
0:29:20
0:36:40
0:44:00
0:51:20
0:58:40
1:06:00
1:13:20
1:20:40
1:28:00
1:35:20
1:42:40
1:50:00
1:57:20
Thermal heat [MW]
Time [hh:mm:ss]
2
Condenser
G
Generator
Evaporator
Turbine
Pump
1
3
6
5
7
8
4
maintenance, favourable operating pressures and autonomous
operation [24]. The benefits associated to ORCs have already
been extensively proven by installations in the past [25].
The principle of the basic ORC is explained with help of
Figure 3. The key components are the evaporator, expander,
condenser and the pump. First, the hot working fluid leaves the
turbine (1) and is condensed in the condenser. The heat from
the condensation process is transferred to a cooling loop (7-8)
which typically consists of water or air. Subsequently, the
condensed working fluid enters (2) the pump and is pressurized
(3). Then, the working fluid enters the evaporator and is heated
to a superheated state (4). The temperature of the heat carrier
(5-6) is thus gradually reduced. The superheated vapour enters
the turbine in which it is expanded to provide mechanical
power. Next, this cycle is again repeated. The minimum
temperature difference between two streams is called the pinch
point temperature difference. In the above cycle, there is both a
point in both the condenser and the evaporator. The basic cycle
introduced above is called the subcritical cycle (SCORC). This
type of cycle is the de facto standard in commercial ORC
systems. However there is ongoing research to further increase
the performance of ORCs by looking at alternative cycle
architectures [4].
Finally, the possibilities of integrating an ORC with a heat
carrier loop are discussed. Starting with the low temperature
heat from the cooling loops, the ORC working fluid could be
pre-heated. However due to the high mass flow rates and low
temperatures this would result in large and expensive heat
exchangers. The main reason the heat from the water cooling
loops cannot be used is that the condensation temperature of the
ORC is not low enough to valorise this heat. Therefore, only
the heat of the flue gasses is a potential heat source for the
ORC.
Table 2 Details of the flue gas heat profile.
Thermal oil loop
Pressurized hot
water loop
Saturated
steam loop
High temperatures
(< ~400 °C)
Low temperatures
(< ~200 °C)
Medium
temperatures
(< ~300 °C)
High reliability
Simple technical
design
Complex
technical design
Flammable
substances
Certified
personnel
necessary
The next challenge is the high variance in temperature and
mass flow rate of the heat source. ORC systems typically
operate down to 10% of the nominal load according to Siemens
and Turboden [26, 27] (Maxxtec gives a figure of 15%). This
means that without thermal buffering the ORC would need to
restart frequently between operations. Considering the time to
start up (up to 30 minutes [28]) this provides an unworkable
situation. Therefore integration of an intermediate thermal
circuit, which can act as buffer, is crucial for the application
under consideration. The possible heat carrier options are
compared in Table 2. Thermal oil loops are normally not
considered in the steel industry due to flammability concerns.
Two options thus remain: the pressurized hot water loop and
the saturated steam loop. Pressurized hot water loops are found
in waste incinerators [29] or the steel industry [30]. Steam
loops are also found frequently in industry. An additional
benefit of steam loops is the possibility to install steam
accumulators to efficiently buffer heat. As such, the steam loop
was selected for the further analysis.
METHODOLOGY
Cases under investigation
In this work five distinct cases are investigated. In the first
case (Case 1), the heat profile is simplified by assuming an
average value of the mass flow rate and the temperature. The
cycle is optimized (to maximize the net power output) without
any constraints on the evaporation pressure or the level of
superheating. The resulting cycle can thus be a subcritical ORC
(SCORC, i.e., a cycle with superheating), a transcritical ORC
(TCORC, i.e., a cycle with evaporation pressure and
temperature above the critical point) or a partial evaporation
ORC (PEORC, i.e., a cycle with the turbine inlet conditions
between saturated liquid and saturated vapour). In the next case
(Case 2), an additional constraint is added. Instead of working
with novel cycle architectures, the ORC is constraint to the
commercially available subcritical type. The working
parameters of the ORC are again optimized. Subsequently
(Case 3) an intermediate steam loop at 26 bar (the same value
as for the Elbe-Stahlwerke Feralpi plant [23]) is added which
provides the opportunity to buffer heat. In this way, the ORC
can operate at steady conditions and the risk of hotspots, like in
a directly heated evaporator, is reduced. Also, the ORC is again
optimized to work with the intermediate heat loop. In the fourth
case (Case 4), the actual heat source profiles are introduced
corresponding to a single batch process. Finally, in the last case
(Case 5), the intermediate steam pressure and the ORC
operating parameters are optimized.
Boundary conditions and models
From a thermodynamic viewpoint, the configuration
analysed with the assumptions from Table 3 is similar to that of
Elbe-Stahlwerke. Approximately the same working fluid
condensate temperature (T2) is attained. The minimum
evaporator pinch point temperature difference is varied in order
to keep the flue gas temperature above 90 °C. The results of the
simulations are summarized in Table 4. The working fluid used
is MDM. In scientific literature MDM is frequently reported as
good working fluid for the application considered in this case
study. Furthermore, it is known that Turboden and Maxxtec
operate several of their ORCs with MDM [31]. The modelling
and optimization approach has been extensively described in a
previous work by the authors [32]. Thermophysical data is
taken from CoolProp [33].
Table 3 Details of the flue gas heat profile.
Variable
Value
Working fluid
MDM
T7 [°C]
26
T8 [°C]
46
Minimum PPTD evaporator [°C] 5 (or higher to keep T6 >
90 °C)
Minimum PPTD condenser [°C]
5
Isentropic efficiency pump [-]
0.7
Isentropic efficiency turbine [-]
0.8
RESULTS AND DISCUSSION
An overview of the results for the five cases is presented in
Table 4. First (Case 1), simulations are performed on the
averaged values from Table 1. The temperature at the inlet (T5)
is 283 °C and the normalized mass flow rate is 37.7 kg/s. No
intermediate steam loop is considered. The T-s diagram of the
ORC corresponding with maximum power output is given in
Figure 4. The resulting net power output is 1132 kWe. The
optimal cycle would be a partial evaporation ORC (PEORC). In
this cycle, the working fluid is heated to a state between
saturated liquid and saturated vapour.
Figure 4 T-s diagram of Case 1 (PEORC)
However, PEORCs are commercially not available.
Therefore the constraint is added (Case 2) that the cycle type
should correspond to a subcritical ORC. The expander inlet
should thus at least attain the point of saturated vapour. The T-s
diagram of the optimized cycle under the imposed constraint is
given in Figure 5. The net power output is reduced to 989 kWe
which corresponds with a decrease of 12.63%. The higher net
power output from a PEORC is also confirmed in literature
[32]. In these temperature ranges, the expected increase would
be around 10% [32]. Care should be taken with these results as
the same pump and expander efficiencies are assumed for both
cycle architectures.
Figure 5 T-s diagram of Case 2 (SCORC)
Next (Case 3) an intermediate steam loop is introduced.
Steam at 26 bar and 230 °C enters the ORC. These are the same
values as for the Elbe-Stahlwerke Feralpi case. The steam is
condensed and subcooled to 90 °C. The minimum pinch point
temperature difference between the flue gas and the
intermediate steam loops is fixed at 5 °C. The net power output
in this case is further reduced to 380.5 kWe. In contrast, the
thermal efficiency of the ORC only shows a minor change. The
large reduction in net power output is attributed to the
mismatch between the steam loop and the flue gas. As such, the
flue gas exit temperature also rises above 90 °C. The results in
a T-s diagram are shown in Figure 6. Note that with higher inlet
temperatures the pinch point can shift to the inlet of the steam
loop and thus more heat can be transferred.
Figure 6 T-s diagram of Case 3 (subcritical ORC with
steam loop)
Therefore the analysis was redone (Case 4) with time
varying inputs of the real heat source profiles. It is assumed that
a sufficiently large steam accumulator is present, such that the
steam pressure of the intermediate loop can be assumed to be
constant in time. The resulting T-s diagram is shown in Figure
7. The extremes of the flue gas heat profile can easily be
identified. The temperature distribution plot furthermore shows
that much of the time the temperature of the waste heat is larger
s [kJ/(kgK)]
-0.6 -0.4 -0.2 0 0.2 0.4
T [°C]
0
50
100
150
200
250
300
2
3
6
5
4
1
78
s [kJ/(kgK)]
-0.6 -0.4 -0.2 0 0.2 0.4 0.6
T [°C]
0
50
100
150
200
250
300
4
1
2
3
6
5
7
8
s [kJ/(kgK)]
-0.6 -0.4 -0.2 0 0.2 0.4 0.6
T [°C]
0
50
100
150
200
250
300
1
4
5
6
32
78
Flue gas profile
aver. temperature
than the average value. The total time averaged thermal input is
now 5034 kW. The resulting time averaged net power output is
685.6 kWe.
Figure 7 T-s diagram of Case 4 (subcritical ORC with
steam loop and varying heat input).
Table 4 Simulation results of the five different cases.
Variable
Value
Case 1: PEORC
Thermal power in [kW]
7470
Net power output [kWe]
1132
Thermal efficiency ORC [%]
13.35
Mass flow rate cooling water [kg/s]
84.25
Case 2: SCORC
Thermal power in [kW]
7475
Net power output [kWe]
989
Thermal efficiency ORC [%]
13.35
Mass flow rate cooling water [kg/s]
86.2
Case 3: SCORC + steam loop
Thermal power in [kW]
2794
Net power output [kWe]
380
Thermal power out [kW]
0
Thermal efficiency ORC [%]
13.62
Mass flow rate cooling water [kg/s]
50.1
Case 4: SCORC + steam loop + variable heat
Averaged thermal power in [kW]
5034
Averaged net power output [kWe]
685.6
Averaged thermal efficiency ORC [%]
13.62
Averaged mass flow rate cooling water [kg/s]
50.1
Case 5: SCORC + steam loop + variable heat + optimized
Averaged thermal power in [kW]
5429
Averaged net power output [kWe]
752
Averaged thermal efficiency ORC [%]
13.85
Averaged mass flow rate cooling water [kg/s]
53.9
In the last step (Case 5), both the pressure of the
intermediate steam loop and the ORC operating parameters are
optimized. The results of the optimization are presented in
Table 4. The time averaged net power output is increased by
9.7% to 752.4 kWe in comparison with Case 4 where the steam
pressure is not optimized. There is now a slight superheating of
roughly 6 °C to attain the maximum net power output.
Furthermore, the optimized steam pressure (25 bar) is close to
the values of the Elbe-Stahlwerke case (26 bar). In Figure 8, the
instantaneous steam flow rate generated from the waste heat
stream is shown. It is obvious that there are very large
variations during a single batch process. The steam flow rate
varies from 0 kg/s to 10.79 kg/s. This graph can be used to size
the thermal capacity of the steam buffer to allow a constant
time averaged steam flow rate to the ORC of 2.11 kg/s.
Table 5 Optimal operating parameters for Case 5.
Variable
Value
Pressure steam loop[kPa]
2,504
Evaporation pressure ORC [kPa]
194.8
Superheat ORC [°C]
5.6
Figure 8 Instantaneous steam flow rate to buffer vessel for
a single batch process in Case 5.
CONCLUSIONS
In this work, a comprehensive analysis on the integration
of an organic Rankine cycle (ORC) coupled to an electric arc
furnace (EAF) was provided. As the EAF is a batch process
with large time variation in available thermal capacity and
temperature, it was concluded that buffering of the heat is a
necessity. The use of a steam loop is identified as a
straightforward solution for the buffering need.
The subsequent analysis was subdivided in five different
cases. The following main conclusions could be drawn from the
results. Firstly, the partial evaporation cycle (PEORC) provides
performance benefits in line with previous research in
literature. The PEORC shows approximately a 10% better net
electricity output compared to the SCORC. Secondly, the use of
a steam buffer greatly reduces the heat transfer to the ORC due
to the additional isothermal plateau of the steam. The ORC
electric power output is decreased with up to 61.5%. In
addition, the use of time averaged input values is not sufficient
to accurately simulate ORC/EAF systems as this gives biased
results. Finally, the optimal pressure of the steam buffer is 25
bar which closely resembles the 26 bar found in the Elbe-
Stahlwerke Feralpi case.
s [kJ/(kgK)]
-0.6 -0.4 -0.2 00.2 0.4 0.6
T [°C]
0
100
200
300
400
500
600
bins T [°C]
0
100
200
300
400
500
600
Occurence [%]
0
20
40
4
2
6
3
5
1
6
8
7
Flue gas profile
min. temperature
Flue gas profile
max. temperature
Time [seconds]
0 1000 2000 3000 4000 5000 6000 7000
Steam mass flow rate [kg/s]
0
2
4
6
8
10
12
ACKNOWLEDGMENTS
The results presented in this paper were obtained within
the frame of the IWT SBO- 110006 project The Next
Generation organic Rankine cycles (www.orcnext.be), funded
by the Institute for the Promotion and Innovation by Science
and Technology in Flanders. This financial support is gratefully
acknowledged. The case study was provided ENGIE Electrabel
in the framework of the ‘Value of waste heat’ project. This
work was also supported by the UK Engineering and Physical
Sciences Research Council (EPSRC) [grant number
EP/P004709/1]. Data supporting this publication can be
obtained on request from steven.lecompte@ugent.be.
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... A schematic representation of a simple ORC is presented in The cycle parameters used in the calculations are summarized in Table 4.1. Based on the previous studies [105,106], for the operating conditions used in this study, the temperature difference between the turbine inlet temperature (TIT) and the heat source temperature (the primary heat exchanger inlet temperature) is in the range of 5 to 25°C. ...
Thesis
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Emissions of greenhouse gases (GHGs) are one of the main problems of using fossil energy sources such as coal and natural gas. Thus, decarbonization of the energy sector, i.e., the increased use of the lower carbon intensity and carbon-free energy sources and technologies, such as solar and wind, is needed to meet increasing energy demand and reduce emissions. Although the Sun’s energy falling on Earth has the capacity of being the largest source of electricity by the mid-century, efficiency, intermittency, and cost are the challenges for the solar and other renewable energy technologies. Concentrating solar power tower (CSP-T) technology is one of the technologies for converting solar energy to electric energy (electricity). In a CSP-T plant, the large number of mirrors or lenses is used to concentrate incoming solar energy to a small area. Thus, solar energy is first converted to the thermal energy in the solar receiver, and then to the mechanical and electrical energy in a plant power block. This study focuses on the CSP-T technology integrated with power cycles such as Rankine and Brayton. Selection of the power cycles, working fluids, and heat rejection systems was analyzed in this research study with the objective to improve conversion (mirror-to-electric generator) efficiency and reduce the cost. The selection of the best working fluid for a power cycle is traditionally performed by conducting a large number of parametric calculations over a range of cycle operating parameters for a number of candidate working fluids. A novel and systematic multi-step method was developed in this study for the selection of the best working fluid(s) for the commonly considered power cycles and specified (selected) cycle operating conditions (maximum temperature and pressure, and heat rejection temperature). The best working fluid gives the highest thermal efficiency, or the highest net power output of the power cycle. The power cycle modeling and design analysis were performed by employing EBSILON Professional Version 11 (EPV-11) software. Thermo-physical and environmental properties of the working fluids, and construction and operating cost of the CSP-T plant were the main criteria considered in the working fluid and power cycle selection procedure. To alleviate the negative effects of dry cooling in arid areas, where most CSP-T plans are (and will be) built, a direct air-cooled cooling tower (CT) with the “cold energy” thermal storage system (CE-TES), and a closed-loop hybrid-cooled CT are proposed. The effects of both cooling methods on cycle performance (net power output and electricity generation) were analyzed and compared.
... Organic Rankine cycle (ORC) is a Rankine cycle which utilizes an organic working fluid, waste heat, and operates at relatively low temperature, thus having thermal efficiency of around 10e20% [7,8], i.e., significantly lower compared to the steam Rankine cycle. Organic fluids, such as Butane, RC318, R22, R123, Iso-butane are the choices of working fluids for an ORC [9,10]. ...
Article
Thermodynamic analysis and optimization of the power block of concentrated solar power (CSP) plants were performed in this study. Single and combined power cycles such as regenerative steam Rankine cycle with reheat (RSRC), organic Rankine cycle (ORC), combined Rankine/ORC cycle, regenerative Brayton cycle (RBC), regenerative Brayton cycle with recompression (RBCR), and combined Brayton/ORC cycle were compared. Thermodynamic performance of the power cycles was evaluated by performing parametric calculations over a range of operating conditions (maximum temperature, minimum temperature, maximum pressure). Selection of the best power cycle(s) is the main focus of this study. Performance maps which present performance information on the best power cycles in a graphical and straightforward manner were constructed. Results show that for the maximum cycle temperatures lower than 300 °C, the ORC has the highest thermal efficiency. For the medium maximum cycle temperatures (between 300 °C and 650 °C), the combined Rankine/ORC and RBCR are the best choices. For the maximum cycle temperature higher than 650 °C, depending on the maximum pressure, the combined Brayton/ORC cycle and RBCR give the highest thermal efficiency. Also, for low and medium maximum temperatures, RSRC produces the highest specific net work output, followed by the combined Helium Brayton/ORC cycle.
... A simple ORC using waste heat sources and different working fluids has been the focus of many papers and studies [15][16][17][18][19]. Saleh et al. [20] compared 31 different working fluids and their effect on the thermal efficiency for an ORC with different work cycle configurations. ...
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The analysis and optimization of an organic Rankine cycle (ORC) used as a bottoming cycle in the Brayton/ORC and steam Rankine/ORC combined cycle configurations is the main focus of this study. The results show that CO 2 and air are the best working fluids for the topping (Brayton) cycle. Depending on the exhaust temperature of the topping cycle, Iso-butane, R11 and ethanol are the preferred working fluids for the bottoming (ORC) cycle, resulting in the highest efficiency of the combined cycle. Results of the techno-economic study show that combined Brayton/ORC cycle has significantly lower total capital investment and levelized cost of electricity (LCOE) compared to the regenerative Brayton cycle. An analysis of a combined steam Rankine/ORC cycle was performed to determine the increase in power output that would be achieved by adding a bottoming ORC to the utility-scale steam Rankine cycle, and determine the effect of ambient conditions (heat sink temperature) on power increase. For the selected power plant location, the large difference between the winter and summer temperatures has a considerable effect on the ORC power output, which varies by more than 60% from winter to summer.
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In the present paper, we consider the employment of working-fluid mixtures in organic Rankine cycle (ORC) systems with respect to thermodynamic and heat-transfer performance, component sizing and capital costs. The selected working-fluid mixtures promise reduced exergy losses due to their non-isothermal phase-change behaviour, and thus improved cycle efficiencies and power outputs over their respective pure-fluid components. A multi-objective cost-power optimization of a specific low-temperature ORC system (operating with geothermal water at 98 °C) reveals that the use of working-fluid-mixtures does indeed show a thermodynamic improvement over the pure-fluids. At the same time, heat transfer and cost analyses, however, suggest that it also requires larger evaporators, condensers and expanders; thus, the resulting ORC systems are also associated with higher costs. In particular, 50% n-pentane + 50% n-hexane and 60% R-245fa + 40% R-227ea mixtures lead to the thermodynamically optimal cycles, whereas pure n-pentane and pure R-245fa have lower plant costs, both estimated as having ∼14% lower costs per unit power output compared to the thermodynamically optimal mixtures. These conclusions highlight the importance of using system cost minimization as a design objective for ORC plants.
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By employing the SAFT-VR Mie equation of state, molecular-based models are developed from which the thermodynamic properties of pure (i.e., single-component) organic fluids and their mixtures are calculated. This approach can enable the selection of optimal working fluids in organic Rankine cycle (ORC) applications, even in cases for which experimental data relating to mixture properties are not available. After developing models for perfluoroalkane (n-C4F10+n-C10F22) mixtures, and validating these against available experimental data, SAFT-VR Mie is shown to predict accurately both the single-phase and saturation properties of these fluids. In particular, second-derivative properties (e.g., specific heat capacities), which are less reliably calculated by cubic equations of state (EoS), are accurately described using SAFT-VR Mie, thereby enabling an accurate prediction of important working-fluid properties such as the specific entropy. The property data are then used in thermodynamic cycle analyses for the evaluation of ORC performance and cost. The approach is applied to a specific case study in which a sub-critical, non-regenerative ORC system recovers and converts waste heat from a refinery flue-gas stream with fixed, predefined conditions. Results are compared with those obtained when employing analogue alkane mixtures (n-C4H10+n-C10H22) for which sufficient thermodynamic property data exist. When unlimited quantities of cooling water are utilized, pure perfluorobutane (and pure butane) cycles exhibit higher power outputs and higher thermal efficiencies compared to mixtures with perfluorodecane (or decane), respectively. The effect of the composition of a working-fluid mixture in the aforementioned performance indicators is non-trivial. Only at low evaporator pressures (<10bar) do the investigated mixtures perform better than the pure fluids. A basic cost analysis reveals that systems with pure perfluorobutane (and butane) fluids are associated with relatively high total costs, but are nevertheless more cost effective per unit power output than the fluid mixtures (due to the higher generated power). When the quantity of cooling water is constrained by the application, overall performance deteriorates, and mixtures emerge as the optimal working fluids.
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The ORC (organic Rankine cycle) is an established technology for converting low temperature heat to electricity. Knowing that most of the commercially available ORCs are of the subcritical type, there is potential for improvement by implementing new cycle architectures. The cycles under consideration are: the SCORC (subcritical ORC), the TCORC (transcritical ORC) and the PEORC (partial evaporation ORC). Care is taken to develop an optimization strategy considering various boundary conditions. The analysis and comparison is based on an exergy approach. Initially 67 possible working fluids are investigated. In successive stages design constraints are added. First, only environmentally friendly working fluids are retained. Next, the turbine outlet is constrained to a superheated state. Finally, the heat carrier exit temperature is restricted and addition of a recuperator is considered. Regression models with low computational cost are provided to quickly evaluate each design implications. The results indicate that the PEORC clearly outperforms the TCORC by up to 25.6% in second law efficiency, while the TCORC outperforms the SCORC by up to 10.8%. For high waste heat carrier inlet temperatures the performance gain becomes small. Additionally, a high performing environmentally friendly working fluid for the TCORC is missing at low heat carrier temperatures (100 C).
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Organic Rankine cycles (ORCs) are an established technology to convert waste heat to electricity. Although several commercial implementations exist, there is still considerable potential for thermo-economic optimization. As such, a novel framework for designing optimized ORC systems is proposed based on a multi-objective optimization scheme in combination with financial appraisal in a post-processing step. The suggested methodology provides the flexibility to quickly assess several economic scenarios and this without the need of knowing the complex design procedure. This novel way of optimizing and interpreting results is applied to a waste heat recovery case. Both the transcritical ORC and subcritical ORC are investigated and compared using the suggested optimization strategy.
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Power systems based on Organic Rankine Cycle (ORC) technology have been recognized as one of the most promising solutions in converting low- and medium-temperature heat into electricity. In this paper, experimental results of the utilization of charge air heat by means of ORC are presented. The experimental setup consists of a 1.6 MWe diesel engine and an ORC process utilizing charge air heat in which the turbine-generator has been replaced with an expansion valve. Thus, no mechanical or electrical power was extracted from the system and the primary focus of the experiments was to study the performance of the evaporator acting as a charge air cooler. The studied working fluids were R245fa and isopentane. The test runs were carried out at full engine load and at engine part loads. In addition, transient tests were carried out. Based on the measured values the ORC utilizing charge air heat was evaluated to be capable to increase the power output of the test engine by 2%.
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