1. Introduction
Natural gas is one of the most widely used
sources of energy in the world due to its
environmental issues and global warming
concerns. As such, the demand for natural gas,
as a primary energy, are increasing more and
more. It is predicted that the natural gas
demands increase at an average rate of 2.4
percent annually until 2030 in the world
(Najibi & Taghavi, 2011). Normally, the
locations of natural gas resources and end
points for various applications are far apart. As
a result, it must be transformed from the
deposit and production sites to the consumers
through the pipeline networks. Increasing the
natural gas demands and pipeline networks
requires focus on the overall efficiency of
transmission. Currently, it is low because much
of energy is wasted in the transmission
network; these energies are recoverable.
Consequently, the development of high
efficiency gas transmission network is a key
issue in order to satisfy the growing demand
from the various customers (Guo & Ghalambor,
2014).
In the transmission network, gas flows through
pipes and various devices such as regulators,
valves, and compressors. The gas pressure is
reduced mainly due to wall friction and also
heat transfer between the gas and the
surroundings. Hence, compressor stations
should be installed to boost the gas pressure
and keep it moving to the desired destinations.
Gas compression is usually performed using
the centrifugal or reciprocating compressors
driven by the gas turbines or the electrical
motors. Gas turbines are more common
because of several reasons. First of all, their
operating cost is lower than the electrical
motors. Secondly, installation cost of a new
interconnecting electric power transmission
line may be high. Finally, it may be difficult to
obtain the necessary regulatory approvals for
the electrical motors. However, it is estimated
that 3-5% of the transported gases are
consumed by the gas turbines as fuel (Borraz-Sánchez & Ríos-Mercado, 2005; Wu, Rios-Mercado, Boyd, & Scott, 2000) and the exhaust
gases from turbines with high temperature are
released into the environment which leads to
the serious environmental pollution(Dai, Wang,
& Gao, 2009).
After passing all compression stations, the gas
pressure must be reduced at distribution
points. Currently, most pressure reduction
stations use expansion valves, which may
waste lots of energy associated with high
pressure gas. Hence, energy recovery from
pressure compression and reduction stations
can be regarded as one of the focal points to
improve the overall efficiency of gas
transmission network.
Edgar et al. (Edgar, Himmelblau, & Bickel,
1978), Cobos-Zaleta et al. (Zaleta & Ríos-Mercado, 2002), Ríos-Mercado et al. (Ríos-Mercado, Kim, & Boyd, 2006), Kabirian et al.
(Kabirian & Hemmati, 2007) and Safarian et
al. (Safarian, Saboohi, & Kateb, 2013)
presented different models or various
procedures to optimize the gas transmission
network. The objective of their models was to
minimize the cost or energy consumption of
transmission system. The effect of pressure
reduction station and the function of energy
recovery from current gas transmission
network were not considered in their work.
In the current study, minimizing the energy
consumption and maximizing the flow rate
through the pipes were the basic issues of this
study and the main objective of this research
was to maximize energy recovery and excess
power by utilizing different technologies on the
compression and reduction stations. Evaluation
of various technologies such as turbo-expander,
organic rankine cycle, air bottoming cycle, and
steam air bottoming were conducted in the
literature (Cho, Cho, & Kim, 2008; Maddaloni
& Rowe, 2007); It is worth mention that, this
research study, investigating all these systems
together, presents a comprehensive model to
compare them. Following the most efficient
system can be identified.
In 2007, Maddaloni and Rowe investigated the
application of turbo expander in the gas
pressure reduction station to produce
electricity. The electricity could either be
routed back into the electric distribution grid or
used to produce small amounts of hydrogen.
They found that at their assumed peak
efficiencies, electricity could be extracted from
the pressure reduction system with 75%
exegetic efficiency, and hydrogen could be
produced with 45% energetic efficiency.
In the compression stations, one common
solution for increasing the performance of a gas
turbine is to combine it with a steam cycle or
with an organic rankine cycle. This can be used
either to generate electricity alone, or to co-generate both electric power and heat for
industrial and home purposes (Safarian &
Bararzadeh, 2012).
Due to lower vaporization heat, organic fluids
are preferred to water; this is critical especially
when the available power and the heat source
temperature are low. In this situation, these
fluids can better follow the heat source to be
cooled; this reduces the temperature
differences and therefore, irreversibilities at
the evaporator. Furthermore, the turbines for
organic cycles could provide higher efficiencies
at partial loads. In addition their complexity is
usually less due to the lower enthalpy drop of
the fluid(Drescher & Brüggemann, 2007;
Larjola, 1995).
In 2009, Desai et al. (Desai & Bandyopadhyay,
2009) proposed a methodology for accurate
optimization of an ORC as a co-generation
process to generate shaft-work, with16
different organic fluids. In addition, they
investigated the benefits of integrating ORC
with the background process and reported on
the applicability of proposed methodology with
illustrative examples.
In 2010, Roy et al. (Roy, Mishra, & Misra,
2010) performed parametric optimization and
performance analysis of a waste heat recovery
system along with Organic Rankine Cycle to
generate the power. This analysis was
performed for R-12, R-123, and R-134a as
working fluids. They also found that R-123 had
the maximum work output and efficiency
among all the selected fluids.
Combining the gas turbine cycle with an air
bottoming cycle (ABC) is another method that
has been found to increase the performance of a
gas turbine (Korobitsyn, 1998). In1995,
Kambanis (Kambanis, 1995) and in 1996,
Bolland (Bolland, Forde, & Hande, 1996),
reported that by using the exhaust gas of a
simple gas turbine in the air bottoming cycle
the efficiency of the combined cycle improved
about 47% and 46.6% respectively.
In the recent study, Ghazikhani et al.
(Ghazikhani, Passandideh-Fard, & Mousavi,
2011) developed a model for the steam injection
in the gas turbine with air bottoming cycle.
They also found two new cycles with ABC.
These cycles were: the Evaporating Gas turbine
with Air Bottoming Cycle (EGT-ABC), and
Steam Injection Gas turbine with Air
Bottoming Cycle (STIG-ABC). Their findings
indicated that EGT-ABC had a lower
irreversibility and higher output compared to
the STIG-ABC.
In this article, the authors developed a model
for the technical analysis of transmission
network considering characteristics for
pressure compression and reduction stations
and the energy recovery technologies. The
developed model estimated net possible output
power, overall efficiency and system energy loss
to evaluate the performance of the gas
transport networks. It is should be noted that,
the model utilized equations of real gases for
estimation of enthalpy which produced accurate
results no longer needing Moulier graphs.
2. Methodology
2.1. Definition of Transportation Efficiency
Transportation efficiency can be regarded as a
function of the overall system design, the
efficiency of individual components, and the
way that the system is operated. Transportation
efficiency is defined as the amount of fuel
burned or electric power used per unit of the
throughput (i.e., British thermal unit (Btu) or
kW/Mcf). In addition to this general definition,
there are three other related measures.
1. Hydraulic efficiency is a measure of the loss
of energy (pressure drop) caused by the
friction of the flowing gas in the pipeline
facilities.
2. Thermal efficiency applied to a prime mover
(engine, turbine or motor), t measures as
fraction of the potential energy of an input
fuel or electric power which is converted
into useful energy; this energy can be used
to drive a compressor. The amount of energy
that is not converted into useful energy is
considered as “waste heat” in the exhaust.
3. Compressor efficiency measures how much
energy is expended in compressing the gas
compared in comparison with overall energy
used by the compressor. Inefficient
compressors heat the gas instead of raising
its pressure and thus have lower efficiency
values.
2.2. Proposed Model
The model was developed for the existing gas
pipelines networks from supply to demand
nodes. The demand nodes were major locations
of natural gas consuming in the study area.
The demand nodes were either consumption
regions in the study area or export terminals of
natural gas from the study area to the outside.
In contrast, supply nodes were resources
locations for natural gas processing in the
study area. These nodes included either
refineries, natural gas producing plants, or the
import terminals of natural gas from the
outside of the study area (Safarian et al., 2013).
In addition the model considered all network
units i.e. gas compressor, air compressor, gas
turbine, combustion chamber, expansion valves
and heater in the compressor stations, and
pressure reduction stations.
The model consisted of four sub-models that
are defined below:
1. Base scenario,
2. Reduction stations along with turbo-expander,
3. Compression stations affixed to ORC,
4. Compression stations affixed to ABC and
SI-ABC
The main thermodynamic assumptions that
were used in the present analysis are reported
in Table 1.Second Iran gas transmission
network was chosen as the case study in this
work. This network with 7 major compression
stations was one of the most important network
systems in Iran.
2.3. Base Scenario
In this scenario, gas transmission network is
considered without energy recovery and it
includes simple pressure compression and
reduction stations. In this sub-model, the aim
is to minimize the energy used in the gas
compressors. This can be written as following
(O'Neill, Williard, Wilkins, & Pike, 1979):
Where α is the unitary energy price ($/kW),
ηtherm is the compressor thermic efficiency and calculated by (Menon, 2005):
Wj is compressor required power that is
Where T1, P1 and Z1 are input temperature,
pressure and compressibility factor andT2, P2
and Z2 are output ones, respectively. ηa stands
for the compressor adiabatic efficiency, Q is gas
flow rate through the pipeline (MMSCMD) and
γ is the ratio of specific heats which assumed to
be constant.
According to Weymouth equation (equation 3),
based on the input conditions, designed
maximum and minimum pressure and type of
pipe material, the maximum flow rate through
the pipeline must be calculated. Considering
this, the demand should always be lower than
maximum flow rate to satisfy the restrictions
for flow capacity.
Also, at each end point, the demand must be
guaranteed at a minimal pressure. On the
other hand, the gas transmission company
cannot take gas at a pressure higher than
predefined value. Mathematically (O'Neill et
al., 1979):
Fig. 1 shows a schematic of gas compression
station. The inlet air enters the compressor at
state1. Considering an isentropic efficiency of
ηcomp for the compressor and a constant
pressure ratio of rc that can be calculated as:
Because of high pressure of the inlet natural
gas, the reduction valve must be used. At the
next step, compressed air combusts with
medium pressure fuel. Consequently, exhaust
gases exit from combustion chamber and enters
the turbine. At final state, turbine output work
is obtained which depends on exhaust
temperature (TIT) and mass flow. Turbine
output power could be estimated from equation
2 but, the outcome of this equation was not
accurate. To have accurate value of outcome,
equation of states for real gases was used in
our model as follows (Abbott, Smith, & Van
Ness, 2001)
Every component has unique value of ; for
example this is -74920 (kJ/kgmol) for methane.
Second part of equation 9, can be simplified to:
Table 2, shows all constants for calculation of
methane specific heat. Last part of equation 9,
is called residual enthalpy which appears when
natural gas is assumed to be real not ideal, and
also gas pressure is more than atmospheric.
Where Z, Tr and w are gas compressibility
factor, reduced temperature and acentric
factor, respectively.
In the same way, the power for other units in
compression station is shown below:
The other parameter which is different in real
and ideal gasses is entropy. Regarding this,
entropy for real gas can be written as:
is the standard entropy of ideal gas that is
constant for each component and it is 183.48
(kJ/kgmol.K).for methane.
Passing all compression stations, near demand
nodes, the gas enters to gas pressure reduction
stations. The process that currently performs in
reduction stations is shown in Fig. 2. At the
first stage gas enters to the heater to make up
reducing temperature during expansion process
and then it passes under the constant enthalpy
process in Joule-Thomson valves.
The model estimates dissipation rate of
pressure energy in the expansion valves,
required amount of heat and total loss of
energy in the reduction stations. The result of
this part provides us a general overview about
the amount of energy which is being wasted in
gas pressure reduction stations. This can be
written:
2.4. Affix of turbo-expander to
reduction stations
In this part, the authors considered using
turbo-expander instead of J-T valves. For this
purpose, they develop model to calculate output
power and required heat of turbo-expander
system. Fig. 3 shows a simple schematic of an
improved station.
Equations 9-15, 19 and 20 are the base
equations in this part. Inlet conditions of heat
exchanger and outlet conditions from turbo
expander are key inputs. Considering an
isentropic efficiency of ηexp for the expander and
pressure drop of 1.46(%) during heater.
The most important operational challenge in
the turbo-expander is hydrate formation due to
the slight amount of water in the gas. Two
factors which are intensified hydrate formation
are the low temperature and high pressure. So
a proper temperature for outlet heat exchanger
should be found out considering this
phenomenon. At the first step, the expander is
considered as isentropic process and then using
trial and error for T2 and equation of expander
efficiency, the model estimates the correct
answer for T2.
The output power and required heat can be
obtained as follows: Having considered the heat
exchanger pressure drop, turbo-expander
efficiency, fuel mass flow rate and generator
and gearbox efficiency, we can have:
The thermodynamic properties of natural gas
are used in the model and the procedure of
model solution is given in flow chart1 in the
Appendix.
2.4.1. Validation
To validate the first sub-model, the results of
the model were compared to those of the
experiments performed by Pozivil (Poživil,
2004) (Fig 4) and also to the simulation results
(Fig 5 and 6). For the output power and heat
duty against temperature inlet expander and
inlet flow of natural gas, a good agreement was
observed. The discrepancy between the two
results is less than 6%.
2.4.2. Sensitivity Analysis
A sensitivity analysis was conducted in order to
better understand the effect of key-parameters
of process performance. In this analysis, the
effect of an additional percentage of parameters
P1, T1 and flow rate on the output power and
heat duty were investigated using the model.
The expander outlet pressure and temperature
were 1825 kpa and 18ºC, respectively.
Fig. 7 and 8 show the details of the sensitivity
analysis. Two figures are plotted for the
capacities of 37, 139.73 and 371 MSCMH.
As seen in Fig. 7, although output power and
required heat had a direct relationship with
inlet pressure and flow when inlet pressure
was close to outlet pressure, the expansion
system efficiency will be greater due to the low
difference between two graphs.
It is evident in Fig. 8 that expander power and
heat duty were more sensitive to inlet
temperature. Following this, with using boiler
in maximum load, more power would be
obtained without the extra cost. In addition,
there was a specific temperature for each inlet
pressure where output power and required heat
were equal; on the other hand expansion
system efficiency is 100%.
2.5. Affix of ORC to Compression
Stations
2.5.1. Thermodynamic Analysis of ORC
The ORC system consisted of an evaporator,
turbine, condenser and pump. It could be
classified into two groups according to the level
of turbine inlet pressure, including supercritical
ORCs and sub-critical ORCs (Safarian &
Aramoun, 2015). In the present study, the sub-critical ORCs were investigated.
As is shown in Fig. 9, the working fluid left the
condenser as saturated liquid (point1). Then, it
was compressed by the liquid pump to the sub-critical pressure (point2). The working fluid
was heated in the evaporator until it became
superheated vapor (point3). In this research,
heating process to working fluid was considered
to be indirect. In other words, an inductor fluid
such as oil over took heat transfer to working
fluid. The reason of this was to increase the
security and management level of process. The
superheated vapor flowed into the turbine and
expanded to the condensing pressure (point4),
and then, the low pressure vapor led to the
condenser and condensed by air. The condensed
working fluid flew in to the receiver and was
pumped back to the evaporator, and a new
cycle began.
In the mentioned cycle, if the temperature T4
was considerably higher than the temperature
T1, it might be useful to implement an internal
heat exchanger (IHE) into the cycle as shown in
Fig.10.This heat exchanger is also depicted in
Figs.9 by the additional state points 4a and 2a.
The turbine exhausts flowed in to the internal
heat exchanger and cool in the heat exchanger
in the process (4–4a) by transferring heat to
the compressed liquid that was heated in the
process (2–2a)(Vaja & Gambarotta, 2010).
Each process in the ORC can be described as
follows:
Process 2 to 3: This was the heat absorption
process in the evaporator. The pressure drop
due to evaporation was considered. The amount
of heat transferred from the waste heat to the
working fluid is (Wei, Lu, Lu, & Gu, 2007):
If the internal heat exchanger is added, the
amount of heat transfer can be calculated by:
Process 3 to 4: This was a non-isentropic
expansion process in the turbine. Ideally, this
was an isentropic process 3–4s. However, the
efficiency of the energy transformation in the
turbine never reached 100%, and the state of
the working fluid at the turbine outlet is
indicated by state point 4. The isentropic
efficiency of the turbine can be expressed as:
Process 4 to 1: This was a constant pressure
exothermic process in the condenser.
Process 1 to 2: This was a non-isentropic
compression process in the liquid pump. The
isentropic efficiency of the pump can be
expressed as:
The thermal efficiency of the ORC is defined on
the basis of the first law of thermodynamics as
the ratio of the net power output to the added
heat.
The procedure of model solution is given in
Flow chart 2 in the Appendix.
2.5.2. Validation
To validate the model, the results of the model
were compared to those of the experiments
performed by Dai et al. (Dai et al., 2009). Fig.
11 shows a comparison of model results with
these experiments for 3 different working
fluids. Fig. 11 shows net power output against
turbine inlet temperature where a good
agreement is observed. The discrepancy
between the two results was less than 5%.
Although as the turbine inlet temperature
increased, the net output power for ammonia
and water increased correspondingly, for the
butane, an increase in turbine inlet
temperature led to a reduction in net output
power. For instance when TIT varied from 90
to 135 (C) the efficiency of ORC-butane
decreased by 7.9 % averagely, although it was
increased by 7.2 % and 2.1 % for ammonia and
water, respectively. Consequently, for high TIT
ammonia was better choice among these three
working fluids.
2.5.3. Sensitivity Analysis
A sensitivity analysis was conducted for n-pentane as working fluid to better understand
the effect of parameters on the process
performance. Fig.12 shows the effect of gas
turbine outlet temperature on output power of
n-pentane turbine and ORCs efficiency. The
output work was increased by increasing flu
gas temperature, because more amount of n-pentane could be evaporated by this way. But
the ORCs efficiency was decreased due to the
increase in Q in equation 34.
As seen in Fig.13 the obtainable work was
grown by increment of flu gas rate, because of
increase in heat absorption process in the
evaporator. The net output power was
enhanced too. In addition, Fig.13 shows a
changeless trend for ORCs efficiency because
the ratio of increase of net output power and
required heat were kept constant.
The effect of n-pentane turbine inlet pressure
on net output power and ORCs efficiency at
constant inlet temperature (490C) is displayed
in Fig.14. The net output power and ORCs
efficiency were augmented by increasing of
inlet pressure.
2.6. Affix of ABC and SI - ABC to
Compression Stations
An Air Bottoming Cycle system consists of an
air compressor, regenerator and turbine. Fig.15
shows such a combined cycle in which the
exhaust of an existing, topping gas turbine was
sent to a gas-air heat exchanger which heated
the air in the secondary gas turbine cycle.
ABC was proposed in the late 1980s as an
alternative for the conventional steam
bottoming cycle. Nowadays, this cycle was
considered as a compact and simple bottoming
cycle in the various applications: as an
upgrading option for simple-cycle gas turbines
in the offshore industry, a hot-air co-generation
plant, and a heat recovery installation at high-temperature furnaces.
Fig.16 also shows Steam Injection Gas Turbine
with Air Bottoming Cycle (SI-ABC). The
topping exhaust gases had high temperature
after passing through the regenerator. Thermal
energy of these gases could be used for
evaporating of water. The steam was then
mixed with ABC compressor discharged air in a
mixer (Ghazikhani et al., 2011). The
evaporating process was performed in the
HRSG. The temperature of the exhaust gas
was decreased in about 120 C by generating
steam. The amount of the injected steam per
unit fuel flow is 5-6 (kg/kg-fuel) (Nishada,
Takagi, & Kinoshita, 2005)
The model developed in this study included the
calculation of three cycles: simple gas turbine,
ABC, and SI-ABC. Each process in the ABC
can be described as follows:
Process 1 to 2: This was the heat absorption
process in the regenerator. The pressure drop
due to regenerator was considered. Enthalpy of
flu gas was calculated by equations 9-15.
Equation 35 is utilized for estimation of air
outlet temperature from regenerator.
Process 3 to 4: The inlet air entered the
compressor at state 3. The compressor inlet
power was calculated as:
Air outlet temperature from regenerator was
the function of outlet enthalpy which can be
estimated by energy balance equation around
regenerator.
Process 5 to 6: This was a non-isentropic
expansion process in the turbine. The
isentropic efficiency of the turbine can be
expressed as:
The power generated by the turbine can be
given by:
SI-ABC process was similar to ABC; the
differences between the ABC gas turbine and
SI-ABC were the HRSG and a mixer which
provided steam for the bottoming cycle. The
procedure is given in Flowchart 3 and 4 in the
Appendix for ABC and SI-ABC.
2.6.1. Validation
To validate the model, the results of the model
for ABC and SI-ABC were compared with those
of the experiments performed by Ghazikhani et
al. (Ghazikhani et al., 2011). Fig.17 displays
variations of compression station overall
efficiency with ABC or SI-ABC against TIT. SI-ABC system had more output power than ABC
at the same bottoming cycle pressure ratio and
TIT. This was due to more heat recovery in the
regenerator in the SI-ABC cycles. It could
produce exhaust with a lower temperature and
more inlet mass to bottoming turbine. In
addition, form Fig.17a good agreement between
the model results and experimental data can be
observed. The discrepancy between the two
results was less than 4%.
2.6.2. Sensitivity Analysis
The effect of key parameters of process on its
performance was evaluated by a sensitivity
analysis. In Fig.18, the thermal efficiency of
the ABCs and SI-ABCs varied against
bottoming pressure ratio. The thermal
efficiencies of steam injection system were
higher than ABCs. Fig.19 shows, the efficiency
reduces as the ambient temperature is
increased. In addition, in the ABC, the reduction
rate of efficiency with ambient temperature was
steeper. In ABCs the effectiveness of the heat
recovery in the bottom-cycle was also decreased
by increasing ambient temperature due to a
smaller difference between the two stream
temperatures. Fig.19 also shows the superiority
of the SI-ABC to have the highest efficiency in
different ambient temperatures.
3. Results and Discussion
The proposed model was implemented for
second Iranian gas pipeline network. Table 3
shows the properties of this network. In this
network, the demand levels were different in
different quarters of year. So it must be
considered as an important variable which
should be satisfied. In this regards, the model
was evaluated for different demand levels.
Optimization of the required power for the gas
compressors in various demands was the first
step in the minimizing of the total power
consumption. Then the model used the energy
recovery technologies within the pressure
compression and reduction stations to estimate
the improved overall efficiency and net output
power. Table 4 displays the optimization
results for simple network in the case study.
The demand and the pressure ratio for all gas
turbines was considered 50 MMSCMD and 14,
respectively.
Having a clear comparison between the
mentioned systems, the variation of gas
transmission network overall efficiency against
the demand, is shown in Fig.20. The comparison
included six systems:
1. Simple pipeline
2. Simple pipeline with the turbo-expander in
reduction stations
3. Simple pipeline with ABCs in compression
stations and turbo-expander in reduction
stations
4. Simple pipeline with SI-ABCs in compression
stations and turbo-expander in reduction
stations
5. Simple pipeline with ORC-butane in
compression stations and turbo-expander in
reduction stations
6. Simple pipeline with ORC-pentane in
compression stations and turbo-expander in
reduction stations
Fig.20 shows that the overall efficiency of the
equipped transmission network by ORC-pentane and turbo-expander was higher than
other systems at the same flow rate. These
technologies increase the network overall
efficiency 13-28 % in span of 50- 90 MMSCMD.
As seen in the Fig.20, the overall efficiency of
SI-ABCs was higher than ORC-butane at lower
65 MMSCMD. Although the effect of turbo
expander on the overall efficiency was low at
low flow rates, it had great impact on the
efficiency at cold season when the demand level
is high.
Fig.21 displays the increase of the overall
efficiency with bottoming pressure ratio (rb) in
the range of 5-15 for all systems. The flow rate
was considered 70 MMSCMD. As seen in
Fig.21, the overall efficiency of the ORC-pentane was more sensitive to bottoming
pressure ratio than the other systems.
4. Conclusions
In this study, the effect of three technologies on
the overall efficiency of gas transmission
network was investigated. Based on an energy
analysis a computer program was developed to
survey improving the performance of gas
transmission system. The examined technologies
were organic rankine cycle, air bottoming cycle,
and steam injection air bottoming cycle which
were used in the pressure compression station
and turbo-expander which was utilized within
the pressure reduction stations.
The main conclusions were:
1) The turbo-expander outlet powers and
required heat had a direct relationship with
expander inlet pressure, temperature and flow-rate. In addition, when the inlet pressure
ranged from 450 to 750 pisa and gas flow was
maximum i.e. 866 MSCMH, the efficiency of
the expander system was82-44%, respectively.
2) The SI-ABC was found to have maximum
output power at the same bottoming cycle
pressure ratio and turbine inlet temperature
(TIT) in comparison with ABS. This was due to
more heat recovery in the regenerator in the
SI-ABC cycle that resulted a lower exhaust
temperature; and more inlet mass to bottoming
turbine causes a higher output work. Moreover,
the results displayed that the overall efficiency
was decreased as the ambient temperature had
been increased for ABC and SI-ABC.
3) In this study, an organic rankine cycle using
working fluid such as ammonia, butane, and
water was analyzed and the results were
compared together. When turbine inlet
temperature varied from 90 to 135 (C) the
efficiency of ORCs with butane as working fluid
decreased by 7.9 % averagely, although it
increased by 7.2 % and 2.1 % for ammonia and
water, respectively.
4) The ORC-pentane was more sensitive to
variation of bottoming pressure ratio.
The model was tested for Iran second gas
transmission network. The results showed that
the highest efficiency is obtained from
implementation of ORC with n-pentane as
working fluid and turbo-expander in pressure
compression and reduction stations,
respectively. The demand or gas flow rate
through pipeline was the most effective
parameter that needed to satisfy. So, variation
of the transmission system overall efficiency
was investigated based on the flow rate.
When the study network was equipped by
ORC-pentane and turbo-expander, the overall
efficiency grew by 22 % averagely, in the span
of 50 to 90 MMSCMD.
