COMPONENT MATCHING FOR ULTRA-LOW GRADE THERMAL
ENERGY RECOVERY OPERATING IN ORGANIC RANKINE CYCLE
_______________
A Thesis
Presented to the
Faculty of
San Diego State University
_______________
In Partial Fulfillment
of the Requirements for the Degree
Master of Science
in
Mechanical Engineering
_______________
by
Falgunkumar J. Patel
Summer 2013
iii
Copyright © 2013
by
Falgunkumar J. Patel
All Rights Reserved
iv
DEDICATION
I will like to dedicate my thesis to my professor Dr. Asfaw Beyene, my loving parents
and my friends, who inspired me and helped me in every difficult situation of my life.
v
ABSTRACT OF THE THESIS
Component Matching for Ultra-Low Grade Thermal Energy
Recovery Operating in Organic Rankine Cycle
by
Falgunkumar J. Patel
Master of Science in Mechanical Engineering
San Diego State University, 2013
The aim of this work is to study the influence of component selection and sizing on
the thermodynamic performance of an Organic Rankine Cycle (ORC) designed for ultra-low
grade thermal energy recovery. The selected working fluid of the ORC is a refrigerant whose
properties are adapted for energy recovery in a low temperature regime, below 100C.
Following a mathematical model, validation of the results was conducted using experimental
test using commercial water heater as a heat source, with temperature of the hot water
ranging from 75 to 90 °C. The system was modeled, designed, assembled, and tested. The
limits of the ORC system are evaluated and the results were tabulated. It was found that
component efficiencies and sizing have significant impact on system efficiency of ORCs.
Expanders are typically expensive, whereas the scroll expander, used in this study, has
limited efficiency. Because of the Carnot limit, issues given less attention in traditional
Rankine cycle performance analysis, such as pipe diameters on the liquid and vapor sides,
become critical in ultra-low grade thermal energy recovery.
vi
TABLE OF CONTENTS
PAGE
ABSTRACT ...............................................................................................................................v
LIST OF TABLES ................................................................................................................... ix
LIST OF FIGURES ...................................................................................................................x
ACKNOWLEDGEMENTS .................................................................................................... xii
CHAPTER
1
INTRODUCTION .........................................................................................................1 1.1 Low Grade Thermal Energy Sources .................................................................1 1.2 ORC Design Aspects .........................................................................................3 1.3 Literature Survey ...............................................................................................4 1.3.1 Sustainable Thermal Energy Source for Heat Supply ..............................4 1.3.2 ORC Design ..............................................................................................5 1.3.3 Prime Movers for ORC .............................................................................6 1.3.4 Organic Working Fluid .............................................................................7 2
WORKING FLUID .......................................................................................................9 2.1 Chemical and Physical Properties ......................................................................9 2.2 Environmental Factors .....................................................................................10 2.3 Safety ...............................................................................................................11 2.4 Other Important Factors ...................................................................................12 2.5 R-245fa ............................................................................................................13 3
ORC THERMODYNAMICS ......................................................................................14 3.1 Background ......................................................................................................14 3.2 Energy Conversion In ORC .............................................................................17 3.2.1 Pump Phase (Stage 1-2) ..........................................................................17 3.2.2 Evaporation Phase (Stage 2-3) ................................................................18 3.2.3 Turbine Phase (Stage 3-4) .......................................................................18 3.2.4 Condenser Phase (Stage 4-1) ..................................................................18 3.3 Cycle Efficiency...............................................................................................19 vii
3.4 Performance of the System ..............................................................................19 3.4.1 Assumptions of the Model ......................................................................19 3.4.2 Performance ............................................................................................20 3.4.2.1 Thermodynamic State ....................................................................20 3.4.2.2 Results of the Model ......................................................................21 3.4.2.3 ORC Refrigerant Piping .................................................................22 3.5 Varied Initial Conditions of the Cycle .............................................................24 3.5.1 Varied T1 .................................................................................................24 3.5.2 Varied Expander Efficiency ....................................................................27 4
ORC TEST BENCH ....................................................................................................31 4.1 ConFigureuration of ORC................................................................................31 4.2 Scroll Expander ................................................................................................33 4.2.1 Operating Principle of Scroll Compressor ..............................................33 4.2.2 Leakages and Sealing ..............................................................................34 4.3.3 Specification of the Scroll Expander ......................................................35 4.3 Pump ................................................................................................................35 4.4 Condenser ........................................................................................................37 4.5 Evaporator ........................................................................................................39 4.6 Condenser Water Pump ...................................................................................40 4.7 Boiler and Hot Water Pump .............................................................................40 4.8 Electric Generator ............................................................................................43 4.9 Measurement Devices ......................................................................................44 4.9.1 Pressure Transducer ................................................................................44 4.9.2 Thermocouple .........................................................................................45 4.9.3 Coriolis Flowmeter .................................................................................46 4.9.4 Paddlewheel Flowmeter ..........................................................................46 4.10 Data Acquisition (DAQ) ................................................................................47 5
DESCRIPTION OF EXPERIMENT ...........................................................................50 5.1 Leak Detection Test .........................................................................................50 5.1.1 Description of Leak Detection Test 1 .....................................................51 5.1.2 Description of Leak Detection Test 2 .....................................................51 5.2 Refrigerant Charge in ORC .............................................................................52 viii
5.3 ORC Performance and Results ........................................................................54 5.3.1 First Set of Tests .....................................................................................54 5.3.2 Second Set of Tests .................................................................................55 5.4 Theoretical Cycle Efficiency at Varied Turbine Efficiency ............................56 5.5 Discussions ......................................................................................................59 6
CONCLUSION ............................................................................................................60 6.1 Summary ..........................................................................................................60 6.2 Recommendation .............................................................................................61 REFERENCES ........................................................................................................................62 ix
LIST OF TABLES
PAGE
Table 1.1. Sustainable Thermal Energy Source (Except Waste Heat).. ....................................2 Table 1.2 Waste Heat Sources.. .................................................................................................3 Table 2.1. Safety Classification for Common Refrigerants. ....................................................12 Table 3.1.Theoretical Input Data for ORC Simulation ............................................................24 Table 3.2. ORC Simulation Results at Varying T1 ..................................................................24 Table 3.3 Theoretical Input Data for ORC Simulation at Varied Expander Efficiency ..........27 Table 3.4 ORC Simulation at Varied Expander Efficiency .....................................................28 Table 4.1 Scroll Compressor Specifications ............................................................................36 Table 4.2 Rotary Vane Pump and Electric Motor Specifications ............................................38 Table 4.3 Thermodynamic Properties of Two Streaming Fluids in Condenser ......................39 Table 4.4 Thermodynamic Properties of Two Streaming Fluids in the Evaporator. ...............40 Table 4.5 Submersible Pump Specifications ...........................................................................41 Table 4.6 Boiler and Hot Water Circulation Pump Specifications ..........................................42 Table 4.7. Pressure Transducer Specification ..........................................................................45 Table 4.8. Thermocouple Specification ...................................................................................45 Table 4.9. Coriolis Flow Meter Specification ..........................................................................47 Table 4.10. Paddlewheel Flowmeter Specification..................................................................48 Table 5.1 Experimental Parameters for First Test ...................................................................54 Table 5.2 Experimental Results of First Test...........................................................................55 Table 5.3. Experimental Parameters for Second Test ..............................................................56 Table 5.4 Experimental Results for Second Test .....................................................................57 Table 5.5 Theoretical Calculation of Cycle Efficiency and Turbine Work Done ...................58
x
LIST OF FIGURES
PAGE
Figure 1.1. Schematic diagram of the basic ORC. .....................................................................4 Figure 2.1. Wet, isentropic and dry saturation vapor curve. ....................................................10 Figure 2.2. Refrigerant safety group classification per ANSI/ASHARE standards. ...............11 Figure 2.3. R-245fa ph diagram. ..............................................................................................13 Figure 3.1. T-s diagram of R245fa...........................................................................................15 Figure 3.2. p-v and p-h diagram for R245fa. ...........................................................................15 Figure 3.3. T-s and p-h diagram for ideal/real ORC cycle. .....................................................16 Figure 3.4. Energy flow diagram of ORC................................................................................17 Figure 3.5. ph diagram of refrigerant R245fa. .........................................................................21 Figure 3.6. Simulation of P1 varying T1. ................................................................................25 Figure 3.7. Simulation of T4 varying T1. ................................................................................25 Figure 3.8. Simulation of thermal efficiency varying T1. .......................................................26 Figure 3.9. Simulation of mass flow varying T1. ....................................................................26 Figure 3.10. Simulation of turbine work varying T1. ..............................................................27 Figure 3.11. Simulation of thermal efficiency varying expander efficiency. ..........................28 Figure 3.12. Simulation of mass flow at varied expander efficiencies. ...................................29 Figure 3.13. Simulation of turbine work at varied expander efficiencies. ...............................29 Figure 4.1. Organic rankine cycle test bench. ..........................................................................32 Figure 4.2. Operating principle of scroll compressor. .............................................................33 Figure 4.3. Fixed and orbiting scrolls. .....................................................................................35 Figure 4.4. Inlet/outlet of the scroll compressor. .....................................................................36 Figure 4.5. Magnetic centrifugal pump on left and powerstat variable autotransformer
on right. ........................................................................................................................37 Figure 4.6. graphs pressure – flow rate at different motor speed; the model
TMOT151. ...................................................................................................................38 Figure 4.7. Condenser used in ORC. .......................................................................................39 Figure 4.8. Evaporator for ORC. .............................................................................................40 Figure 4.9. Condenser water pump. .........................................................................................41 xi
Figure 4.10. Boiler and water pump for ORC. ........................................................................42 Figure 4.11. PMA alternator. ...................................................................................................43 Figure 4.12. DC540 voltage and ampere output chart. ............................................................43 Figure 4.13. Pressure transducer. .............................................................................................44 Figure 4.14. Thermocouple. .....................................................................................................45 Figure 4.15. Phase shift along the tube’s length. .....................................................................46 Figure 4.16. Coriolis flow meter. .............................................................................................47 Figure 4.17. Paddlewheel flowmeter. ......................................................................................48 Figure 4.18. DAQ hardware.....................................................................................................48 Figure 4.19 DAQ software. ......................................................................................................49 Figure 5.1. ORC experimental set up.......................................................................................50 Figure 5.2. High pressure nitrogen test for peripheral leak detection test. ..............................51 Figure 5.3. Vacuum pump and associated pressure gauges. ....................................................52 Figure 5.4. Vacuum test for peripheral leak detection. ............................................................53 Figure 5.5. Graphical representation of experimental results, T3 [˚C] vs ηt. ..........................56 Figure 5.6. Graphical representation of experimental results, T3 [˚C] vs ηt. ..........................57 Figure 5.7. Graphical representation of theoretically calculated cycle efficiency with
turbine efficiency plot, ηt vs η. ....................................................................................58 xii
ACKNOWLEDGEMENTS
This dissertation would not have been possible without the guidance and the help of
several individuals who in one way or another contributed and extended their valuable
assistance in the preparation and completion of this study.
First and foremost, my utmost gratitude to Dr. Asfaw Beyene, Mechanical advisor of
San Diego State University whose sincerity and encouragement I will never forget. Dr.
Beyene has been my inspiration as I hurdle all the obstacles in the completion this research
work
I wish to acknowledge the work of Enrico Munari, Mike Lester, Greg Morris, and
everyone that helped make the hardware functional.
1
CHAPTER 1
INTRODUCTION
This thesis demonstrates the use of Organic Rankine Cycle (ORC) as an alternative to
the traditional steam Rankine cycle for converting low grade thermal energy into mechanical
work. Traditional steam powered Rankine plants are inefficient in converting low grade
thermal energy below 100 ˚C into mechanical work. This is due to the requirement for the
steam to be in a superheated state before expansion. This requirement of higher temperatures
for superheating can be eliminated by use of organic working fluid in place of water.
Organic working fluid requires low evaporation energy compared to water, so the former
requires less thermal energy to evaporate the working fluid [1]. Therefore, the organic
working fluid evaporates at lower temperature and pressure. This will avoid the need for
higher temperature to superheat the fluid allowing power to be harnessed from low grade
thermal energy sources.
1.1 LOW GRADE THERMAL ENERGY SOURCES
The temperature of heat sources below 100˚C is called low grade heat because a
Rankine cycle using water cannot work efficiently at such low temperatures [2]. Statistical
data indicates that 50% or more of total heat generated by industry is low-grade waste heat
[2]. Generally, low grade thermal energy produced during industrial processes as well as
natural ones like such as geothermal heating, ocean heating, solar heating etc. is discarded in
the atmosphere as waste heat. This thermal pollution has become major environmental
concern [1]. A growing scarcity of non-renewable energy sources is driving development of
technology that can use low grade thermal energy to produce useful work. This effort can
reduce the burden on primary energy sources like energy such petroleum, coal and natural
gas etc.
Low grade thermal energy can be recovered either directly or indirectly. Direct
recovery of thermal energy is more affordable but not feasible due to contamination
problems. Indirect recovery method is more feasible for low grade thermal energy recovery.
2
Indirect recovery involves the use of two surfaces as a medium for heat transfer from the heat
source to a working fluid. Mechanical components can be utilized that convert low grade
thermal energy to useful work.
One approach that offers a solution for the problem of low grade thermal energy
recovery is to produce electricity using an ORC system. Due to the low operating
temperature of ORC systems, organic fluids are used as working fluids [1]. An ORC system
can be retro-fitted to an existing plant to increase its overall thermal efficiency and
decreasing negative environmental impacts [1].
There are many applications of heat engines that use an external heat source to heat
the working fluid through a heat exchanger. The working fluid expands in the closed system
and produces usable work. Such systems are generally referred to as “external combustion
engines” [3].
Various kinds of thermal energy sources and the temperature ranges at which they
function are detailed in Table 1.1 [4]. From a thermodynamic point of view the temperature
associated with a particular heat source is the key parameter for power production.
Table 1.1. Sustainable Thermal Energy Source (Except Waste
Heat)
Thermal source
Temperature, ˚C
Solar
80-1500
Geothermal
60-350
Ocean heat
25-30
Biomass/biofuel
200-1000
Nuclear
250-900
Moderator heat
60-85
Waste incineration
1000-1500
Source: Md. Ali Tarique. “Experimental Investigation of Scroll Based Organic
Rankine Systems.” Master’s thesis, University of Ontario Institute of Technology,
2011.
Table 1.2 [4] shows the waste heat sources that are mainly generated as secondary
heat from certain applications. The main source of the secondary heat or low grade heat is the
waste heat produced from the electric generators, power plants and industrial processes such
3
Table 1.2 Waste Heat Sources
Thermal source
Temperature
Thermal source
(˚C)
Steam power plant
230-250
Temperature
(˚C)
Alkalines and chlorines
170-220
production
Gas turbine power
600-800
plants
Industrial inorganic chemical
120-200
production
Heating furnace exhaust
175-230
Industrial organic chemicals
Automobile engine
400-700
Petroleum crudes and
exhaust
120-300
intermediate refining
150-200
195-230
Glass factory
400-500
Pulp mills
140-200
Cement production
200-300
Paper mills
140-200
Blast furnace
500-700
Paperboard mills
140-200
Iron foundries
425-650
Copper production
760-815
Drying and baking oven
93-230
Nitrogenous fertilizers
plants
Source: Md. Ali Tarique. “Experimental Investigation of Scroll Based Organic Rankine Systems.” Master’s
thesis, University of Ontario Institute of Technology, 2011.
as casting, copper and steel production etc. To increase the efficiency and economy of such
industrial processes, the secondary heat should be transformed to some useful work using an
ORC system.
This conversion of energy to some particular application includes heating, cooling,
regenerative, ventilation, mechanical work, electric power etc.
1.2 ORC DESIGN ASPECTS
Two main thermodynamic constrains to keep in mind when designing external
combustion engines concern the heat source and heat sink. The temperature differential and
the nature of the fluids determine the magnitude of heat transfer occurring between the heat
source and the working fluid and similarly applicable after work done i.e. between the
working fluid and the heat sink. Figure 1.1 shows the basic configuration of the ORC. The
4
Heat source
Figure 1.1. Schematic diagram of the basic ORC.
goal of the ORC is to convert thermal energy into mechanical energy. The ORC cycle can be
summarized in following steps:




An organic fluid in liquid state is pressurized using a refrigerant pump
This pressurized fluid is then vaporized by mean of a heat source using a heat
exchanger.
The resulting pressurized vapor is expanded through a turbine to produce
mechanical work.
The vapor is then condensed in condenser.
1.3 LITERATURE SURVEY
The following are the literature review carried out putting more focus on heat supply,
ORC design, prime movers for ORC system and ORC working fluid.
1.3.1 Sustainable Thermal Energy Source for Heat
Supply
In 1883, Frank W Ofeldt presented the first patent on an ORC for steam boats which
used vaporized naphtha instead of water as the working fluid [5]. In 1994, J. Larjola et al. [6]
published a paper on producing electricity from industrial waste using an ORC-design which
used a high speed oil free turbo generator-feed pump. Use of a high- speed turbogenerator
5
makes an ORC simple, hermetic and reduces maintenance cost. In 1976, Parimal Patel and
Edward Doyle [7] designed an ORC operating on heat produced from the exhaust of a diesel
truck engine which produced an additional 36 horsepower and improved fuel economy by
15% over a typical duty cycle. In 2007, H.D. Madhawa Hettiarachchi et al. [8] proposed the
optimum design criteria for an ORC using a low temperature geothermal heat source where
the ratio of total heat exchanger area to net power output is used as an objective function. In
1961, Harry Zvi Tabor et al. were the first to build the prototype of a small solar power unit
working on the ORC principle [9]. The unit was exhibited in Rome at a United Nations
conference on new Energy sources. In all the cases above the range of heat source
temperature is between standard 25˚C and 250˚C, which is the low temperature range.
1.3.2 ORC Design
In 2005, D. Manolakos et al. [10] presented the design outline of a low temperature
solar ORC. Thermal energy produced from the solar collectors is used as a heat source.
Using this thermal energy the working fluid HFC-134a is converted into super-heated vapor
and pumped to an expander to generate power. The saturated vapor from the expander is
directed to condenser for condensation. In 2006, Pedro J. Mago et al. [11] designed a
regenerative ORC to investigate luid selection for ORC performance. They performed
various experiments using R113, R245ca, R123 and isobutene as a separate working fluid in
order to determine best thermal efficiency based on the boiling temperature of the fluid [11].
In their experiment with the working fluids the thermal efficiency of the system remained
constant while increasing the turbine inlet temperature. Their experiment also revealed that
organic fluids operating in saturated condition give better efficiency and reduce
irreversibility. The fluid that shows the best thermal efficiency is R113 with a boiling
temperature of 47.59˚C (the highest of the four). The fluid with lowest thermal efficiency is
isobutene with the lowest boiling temperature of -11.61 ˚C [11]. In 2006, H.D. Madhawa
Hettiarachchi et al. [8] presented the design criteria for an ORC using a low temperature
geothermal heat source in the range of 70˚C-100˚C. They used the ratio of the total heat
transfer area to total net power as an objective optimization criterion. They performed their
experiment using ammonia, HCFC 123 and n-pentane. They concluded that the difference in
the thermal efficiency is closely driven by evaporation and condensation temperature and on
6
the T-S curve. According to their experimental results, ammonia was the most preferred fluid
[8]. In 2008, Iacopo Vaja and Agostino Gambarotta [12] described the bottoming of the
internal combustion engine with ORC, thus identifying a combined power unit to enhance the
overall efficiency of the system. In this experiment the heat needed to vaporize the organic
fluid is sourced entirely from exhaust gases. They achieved a 12.5% increase in efficiency for
the combined power unit over the engine with no bottoming [12].
1.3.3 Prime Movers for ORC
The prime mover is the most important component of the ORCs that decides the
efficiency of the system. The pressure ratio across the prime mover is quite high while
operating the organic working fluid [6]. For low power output the concept of positive
displacement offers more benefits compared to turbines [13]. In 1985, O. Badr, Probert, and
O'Callaghan [13] proposed that multi-vane expanders are the most appropriate prime mover
for running an ORC. To demonstrate this, they generated mathematical models and computer
codes to analyze expander performance based on varying operating conditions. In 1994, R.
Zanelli and D. Favrat [14] modified the standard 3 KW scroll compressor to be a hermetic
scroll expander. A 10 KW ORC system was built to investigate performance of a scroll
expander using HCF 134a as a working fluid. In their experiment, they produced overall
isentropic efficiency in the range of 63 to 65 %. In 2007, James A. Mathias et al. [15]
performed some tests on the Gerotor and Scroll Expander to determine how to produce
power from low grade heat. Both expanders tested had an efficiency of around 0.85. In 2010,
X.D. Wang et al. [16] constructed an experimental setup which used flat plate solar collectors
as heat source along with a rolling piston expander and R245fa as a working fluid. They
found that the expander ran stably in their experiment producing expansion power of about
1.73 KW, with an average isentropic efficiency of about 45.2% [16]. In 2009, S. Quoilin [17]
evaluated the efficiency of the scroll type expander using R123. They utilized 40 steady
performance points relating to temperature, mass flow rate, and condensate temperature, and
expansion ratio. In his experiment, isentropic efficiency of the expander ranged from 42% to
68% [17].
7
1.3.4 Organic Working Fluid
The selection of a working fluid is very critical in an ORC to achieve high thermal
efficiency. In 2004, Lars J. Brasz and William M. Bilbow [18] presented a thermodynamic
cycle analysis that determines the ranking of working fluids for ORC applications. The result
of his analysis revealed that the ranking of working fluids depends on critical temperature
and vapor saturation curves which control the de-superheating heat transfer irreversibility
[18]. In 2011, E. H. Wang et al. [19] presented a study regarding selection of the working
fluid selection of an ORC for engine waste heat recovery. They analyzed the performance of
different working fluids operating in a thermodynamic model built in MATLAB and
REFPROP. Their analysis revealed that in the feasible working region, the performance of
R11, R141b, R113 and R 123 is slightly better than R245fa, R245ca, R236ea, R114 and
butane. But when environmental and safety factors are considered, R245fa and R245ca are
the most suitable working fluids for an engine waste heat-recovery system [19]. Tzu-Chen
Hung [20] studied waste heat recovery of ORC using dry fluids. The working fluids under
examination were Benzene, Toluene, p-Xylene, R113 and R123. The analysis revealed that
in the feasible working region, p-Xylene shows the highest efficiency, while Benzene shows
the lowest. The analysis also revealed that irreversibility depends on the type of heat source.
In 2006, P. J. Mago, L. M. Chamra, and C. Somayaji [21] presented a performance analysis
of different working fluids applicable for ORCs. The working fluids under examination were
R134a, R113, R245ca, R245fa, R123, isobutane, and propane, with boiling points between 43 ˚C and 48 ˚C. According to the authors, ORCs using R113 show the best thermal
efficiency, whereas those using propane show the worst. Y. Chen et al. [22] provided a
comparative study of the carbon dioxide trans-critical power cycle compared to an ORC with
R123 as working fluid in waste heat recovery. The authors found that when utilizing the low
grade heat source with equal means of thermodynamic heat rejection, the carbon dioxide
trans-critical power cycle had a slightly higher power output than the ORC.
A review of ORC application has been carried out from the above literature survey,
with special focus on ORC prime movers and ORC working fluid.
The review of ORC Prime movers pointed out that expanders are one of the main
components of ORC that determines the cycle efficiency. The positive displacement type of
8
expanders is preferably used in small scale ORC applications. In this experiment, we used
positive displacement type of expander by modifying existing scroll compressor.
The review of ORC working fluids spots out the most widely used working fluids in
ORC application, i.e. R134a, R123, R245fa and isobutane. In this experiment, we used
R245fa as the working fluid taking into consideration an optimal operating temperature range
in terms of cycle efficiency as well as environmental benefits.
9
CHAPTER 2
WORKING FLUID
The selection of the working fluid for ORC is very important due to its low
temperature and the resulting limit imposed by Carnot efficiency. Due to low operating
temperature the heat transfer inefficiencies of the working fluid are inversely proportional to
the cycle efficiency. This efficiency mainly depends on the thermodynamic properties of the
fluid. However, thermodynamic characteristics are not the only criteria that influence the
working fluid selection. Criterias which are not related to thermodynamic properties include:
environmental factors, toxicity, flammability, cost, availability and material compatibility
[23].
2.1 CHEMICAL AND PHYSICAL PROPERTIES
The common property of all organic working fluids used in power plant technologies
is their low boiling temperature. A small amount of superheating at the evaporator’s exhaust
is always used to avoid quick condensation of working fluid in the expander. The slope of the
vapor saturation curve is another important factor to be considered [1]. The saturation curve
in T-s diagram can have a positive, negative or vertical slope. Figure 2.1 shows different
refrigerant’s saturation curve in T-s diagram.
The preferred characteristic of the low temperature Rankine cycle is the isentropic
saturation vapor curve i.e. R245fa in Figure 2.1 [17]. The saturation curve for the R245fa is
parallel to the line of expansion and fits very well in the curve. The saturation curve for R22
has a negative slope. Therefore, fluid gets saturated before the expansion and begins to
condense creating the prospect of expander breakdown. The saturation curve for isopentane
has a positive slope, indicating that fluid requires more cooling at the expander exhaust [17].
Therefore, vapor saturation curves for working fluids with positive and negative slope are not
desirable.
10
Figure 2.1. Wet, isentropic and dry saturation vapor curve. Source: Quoilin, Sylvain.
“Experimental Study and Modeling of a Low Temperature Rankine Cycle for Small
Scale Cogeneration.” Unpublished master’s thesis, University of Liège, 2009.
2.2 ENVIRONMENTAL FACTORS
ODP i.e. ozone depleting potential and GWP i.e. greenhouse warming potential are
two important parameters considered in environmental safety along with the ALT factor.
ODP is the ratio of ozone column destruction per unit mass of released gas [24]. CFC 11 is
the reference for all chemicals in determining ODP and has assigned ODP value of 1. All
other refrigerants including CFCs and HCFCs, have ODP values between 0.1 and 1. Modern
11
refrigerants such as HFCs and HFEs do not contain chlorine due to its negative impact on
environment and therefore, have ODPs of zero [25].
GWP is the factor that indicates how much heat greenhouse gases can trap over a
specific period of time. It compares the amount of heat trapped by a certain mass of gas in
relation to the amount of heat trapped by the same amount of carbon dioxide. So carbon
dioxide is the reference gas with a GWP value of 1. GWP is based on the radiative qualities
and decay rate of a chemical [26]. ALT is the length of time a gas will remain in the
atmosphere based on its decay rate.
2.3 SAFETY
The working fluid should be non-corrosive, non-toxic and non-flammable. The
ASHARE classification of good refrigerant is significantly out the danger zone. ASHARE
has defined refrigerant toxicity and flammability accordingly. Figure 2.2 shows refrigerant
safety group classification as per ANSI/ASHARE standards 34-1992 [27]. Table 2.1 [26]
gives the main characteristics of organic fluids from an environmental point of view.
Figure 2.2. Refrigerant safety group classification per
ANSI/ASHARE standards.

Toxicity
Class A: Toxicity has not been identified at concentration less than or equal to
400 ppm by volume [27].
Class B: There is evidence of toxicity at concentrations below 400 ppm by
volume [27].
12
Table 2.1. Safety Classification for Common Refrigerants
refrigerant
Class
ODP
GWP ALT Toxicity flammability
R-11
CFC
1
3600
45
A1
B11
R-113
CFC
0.9
5330
85
A1
1
R-22
HCFC
0.034
1710
12
A1
1
R-123
HCFC
0.012
53
1.3
B1
1
R-134a
HFC
0
1320
14
A1
1
R-245fa
HFC
0
1020
7.6
B1
1
R-717
Ammonia
0
<1
B2
2
R-601
n-pentane
0
~20
R-601a
Iso-pentane
0
~20
Source: “Global-warming potential.” Wikipedia. Last modified June 5, 2013.
http://en.wikipedia.org/wiki/Global-warming_potential.

Flammability
Class 1: No flame propagation in air at 21˚C and 101kPa [27].
Class 2: lower flammability limit greater than 0.1 kg/m3 at 21˚C and 1 atm,
heat of combustion less than 19,000 KJ/kg [27]
Class 3: Highly flammable as defined by LFL (Lower Flammable Limit) less
than or equal to 0.1 kg/m3 at 21˚C and 1 atm or heat of combustion greater
than or equal to 19,000 KJ/kg [27].
2.4 OTHER IMPORTANT FACTORS

Decomposition temperature
Unlike water, organic fluids usually suffer chemical decomposition at high
temperature. Refrigerant decomposes in the system or container when exposed
to high temperature, heat or electrical shock. For this reason the maximum
heat source temperature is limited by the chemical stability of the working
fluid [17].

Good availability and low cost
Temperature critical working fluids required for ORC application are
expensive. The availability of organic fluids in the market is also an important
factor.
Acceptable pressures
Working fluids with very high working pressures have a negative impact on
system reliability. Very high pressure creates high system resistance, requiring
use of expensive components.
Material compatibility
The organic fluid should be non-corrosive to the engineering equipment used
in experiments such as heat exchangers, pipes, pumps, expanders etc.


13
2.5 R-245FA
R-245fa is the working fluid used in this experiment. It is an HFC refrigerant which
replaces R-11 and R-123 as they have ODP of 1 and 0.012 respectively. The chemical
formula of R-245fa is CHF2CH2CF3 and its chemical name is 1,1,1,3,3-Pentafluoropropane.
The boiling temperature of R-245fa is 14.6˚C (58.28˚F) at 1 atm. The decomposition
temperature of R-245fa is more than 250˚C (482˚F) [28]. The main application of R245fa is
in centrifugal chillers but it is also used in ORC as working fluid due to its high heat transfer
coefficient and high critical temperature of 154˚C [28]. Therefore, R245Fa increases the
range of temperature for superheating. Figure 2.3 [24] shows the thermodynamic P-S
diagrams for R-245fa.
Figure 2.3. R-245fa ph diagram. Source: Quoilin, S. "An Introduction to
Thermodynamics Applied to ORCs." Master’s thesis, University of Leige, 2008.
14
CHAPTER 3
ORC THERMODYNAMICS
3.1 BACKGROUND
The Rankine cycle is an idealized thermodynamic cycle that converts thermal energy
into mechanical work. 90 percent of the total power produced in the world is created through
this Rankine cycle [29]. Unlike the traditional Rankine cycle, an ORC uses high molecular
mass organic working fluids like n-pentane, toluene, R245fa etc. [29]. The use of high
molecular mass fluids allows the low grade heat recovery from the heat sources like
industrial waste, solar panels, geothermal etc. The working principle of the ORC is the same
as that of the traditional Rankine cycle.
The thermodynamic state of a working fluid can be defined at any point of the
process if at least two thermodynamic properties at that point are known. Therefore, the
thermodynamic process can be represented in a two axis diagram, most commonly known as
a thermodynamic diagram. Various combinations of thermodynamic properties are used to
illustrate the process through the graph [24]. T-s, p-h, p-v and h-s are the most common ones.
Thermodynamic diagrams for R245fa are provided in Figure 3.1 [24] and Figure 3.2 [24].
In these diagrams several zones can be identified:




Liquid zone: zone at the left of the liquid line corresponds to liquid state
Two phase zone: It is the region in which liquid and vapors exist together. This
zone is bounded by liquid saturation line on left and vapor saturation line on right
[30].
Vapor zone: the zone at the right of the vapor line corresponds to vapor state.
Critical zone: it lies above the critical point where liquid and vapor cannot be
identified. This zone corresponds to very high pressure and temperature [24].
The ORC can also be studied in perspectives, i.e. ideal and real thermodynamic cycle
(Figure 3.3 [24]).
In the ideal thermodynamic cycle four processes can be distinguished:
1. Isobaric evaporation (1-4): there is no pressure drop across the heat exchanger.
Evaporation can be divided into three zones : preheating (1-2), evaporation (2-3)
and superheating (3-4) [24].
15
Figure 3.1. T-s diagram of R245fa. Source: Quoilin, S. "An Introduction to
Thermodynamics Applied to ORCs." Master’s thesis, University of Leige,
2008.
Figure 3.2. p-v and p-h diagram for R245fa. Source: Quoilin, S. "An Introduction to
Thermodynamics Applied to ORCs." Master’s thesis, University of Leige, 2008.
16
Figure 3.3. T-s and p-h diagram for ideal/real ORC cycle. Source: Quoilin, S. "An
Introduction to Thermodynamics Applied to ORCs." Master’s thesis, University of
Leige, 2008.
2. Isentropic expansion (4-5): In isentropic process, there is no change in entropy i.e.
adiabatic process. An isentropic process is an idealization of the expansion
process in which there is no heat transfer between system and surrounding i.e. no
heat loss. Entropy is directly related to heat transfer i.e. dQ = TdS (dQ = heat
transfer, T = temperature, dS = change in entropy), so if dQ = 0 then dS = 0 [31].
3. Isobaric condensation (5-8): the condensation process is achieved without
pressure change. It is subdivided into de-superheating (5-6), condensation (6-7)
and sub cooling (7-8) [24].
4. Isentropic pumping: the compression in the pump is isentropic. In other words,
the process is reversible and there is no change in entropy. Since dS = dT = 0, the
isentropic pumping process cannot be represented in the thermodynamic cycle.
In a real cycle there is irreversibility, which is the primary reason for decreasing
efficiency. The following are the main causes of irreversibility in the process:



During expansion: During the expansion process only part of the energy that is
recovered from the pressure difference across the expander is converted into
useful work. The rest is converted into heat and lost. This loss will reduce the
expander efficiency and ultimately system cycle efficiency [24].
Heat exchanger: The serpentine path in the heat exchanger ensures a better heat
transfer rate but the noticeable degree of pressure loss in heat exchanger reduces
the pressure related power recovery [24].
Pump and fluid friction: The moving parts involved in the pumping process create
electromechanical losses and internal leakage at various stages in the cycle,
leading to heat transformation and ultimately, system irreversibility. Fluid friction
causes a pressure drop in the pump, heat exchangers and piping which also leads
to heat loss and consequent system irreversibility [24].
17
3.2 ENERGY CONVERSION IN ORC
Figure 3.4 [32] depicts the basic flow diagram of the ORC. The working fluid
operates in the closed loop. The heat source enters the system through the evaporator where
the working fluid is heated to saturation or to its super heating point. After isobaric heat
addition which occurs between stage 2 and stage 3, high pressure vapor expands in the
turbine (3-4). The stream of the expanded fluid is passed directly through condenser where it
condenses to temperature T1 and pressure P1. After leaving condenser the working fluid
enters the pump via an expansion tank. In the pump the pressure is raised to P2 and the fluid
is returned to the evaporator.
Evaporator
Turbine
T3, P3
Wt
Stage
Qin
T2,
P2Increase
Stage2
Stage 4
Flammability
Stage 1
Pump
T4, P4
Wp
Qout
T1, P1
Condenser
Figure 3.4. Energy flow diagram of ORC. Source: Yamamoto, Takahisa,
Tomohiko Furuhata, Norio Arai, and Koichi Mori,. “Design and Testing of
the Organic Rankine Cycle.” Energy 26, no. 3 (2001): 239–251.
3.2.1 Pump Phase (Stage 1-2)
The refrigerant pump is the driving mechanism of the ORC system. The organic fluid
which leaves the condenser at reduced pressure P1 is pumped to the evaporator and regains
pressure P2. The pump work Wp is calculated [32]:
.
Where,
denote the density of the working fluid
18
denotes the adiabatic efficiency of the circulation pump.
The specific enthalpy of the working fluid at the pump outlet is given by :
Where
is the specific enthalpy of the working fluid at the pump inlet.
3.2.2 Evaporation Phase (Stage 2-3)
The evaporator heats the high pressure working fluid existing the pump. As the heat
loss in the evaporator is neglected, the amount of heat added to working fluid is equal to the
amount of heat extracted from the heat source; as is given by:
is the evaporator heat input
Where,
is specific enthalpy at the evaporator inlet
is specific enthalpy at the evaporator outlet
3.2.3 Turbine Phase (Stage 3-4)
The turbine converts the potential energy of the highly pressurized gases into kinetic
energy for power production. The stream of high pressure vapor expands in the turbine case,
transfer its energy to rotating shaft of the turbine. Following an assumed adiabatic expansion
process, the vapor returning to low pressure
and low temperature
.
Stipulating the turbine efficiency , turbine output is given by:
.
Where,
is the specific enthalpy at the turbine outlet.
3.2.4 Condenser Phase (Stage 4-1)
Vapor from the turbine at low pressure undergoes isobaric phase change into
saturated liquid form by transferring heat to environment or to the coolant. The pressure of
the working fluid in the condenser is considered the same as the lower pressure of the ORC
. The condenser load output,
fluid, can be given by the equation:
which defines the rate of latent heat transfer by working
19
3.3 CYCLE EFFICIENCY
Cycle efficiency is the ratio of net useful work done to the heat input during the
evaporation phase. The efficiency of the ORC is much less than the commercial Rankine
cycle due to its lower temperature range.
.
Work done per unit mass of fluid in turbine
Work done per unit mass of fluid by pump
.
The heat provided per unit mass of fluid by the hot water in the evaporator:
.
The efficiency is defined as:
=
(3.1)
It should be noted that the above expression is only valid when expansion and
compression is adiabatic. But in a real situation, there is minimum temperature difference
between the two streaming fluids in the evaporator.
The amount of heat transferred in heat exchanger is given as:
is overall heat transfer coefficient
is heat transfer surface area in m2
is logarithmic mean temperature difference [33]
∆
ln ∆
∆
∆
∆ = temperature difference between two streams at point A
∆
= temperature difference between two streams at point B
3.4 PERFORMANCE OF THE SYSTEM
The following assumptions were made to derive the theoretical performance of the
system.
3.4.1 Assumptions of the Model


The power output ( ) = 3KW
The temperature at outlet of condenser T1 = 28 ˚C(82.4 ˚F)
20





The temperature at inlet of scroll expander T3 = 90 ˚C (194 ˚F)
The overall efficiency of the scroll expander
= 60% [34]
The efficiency of the pump is fixed at 90%
The pressure drop at condenser is 5% i.e. P1 = 0.95*P4
The pressure drop at evaporator is 5% i.e. P3 = 0.95*P2
3.4.2 Performance
Thermodynamic states and ORC system performance is carried out based on the
assumptions.
3.4.2.1 THERMODYNAMIC STATE
Figure 3.5 shows the thermodynamic cycle of refrigerant R245fa
State 1: Saturated liquid
T1 = 28 ˚C (82.4 ˚F)
P1 = 1.7 bar (25 Psi)
State 2: sub cooled liquid
P2 = 10.5 bar (152 psi)
T2 is determined from h2 and P2
State 3: saturated vapor
T3 = 90 ˚C (194 ˚F)
P3 = 0.95* P2 = 10 bars (145 psi)
h3 is determined from P3 and T3.
State 4: superheated vapor
P4 = P1 / 0.95 = 1.79 bar (26 psi)
T4 is determined from h4 and P4
(3.2)
21
Figure 3.5. ph diagram of refrigerant R245fa.
3.4.2.2 RESULTS OF THE MODEL
The following results provide the theoretical size of the components required in the
experiment.




Pump
= 120 W
= 1.6 bar
= 10.5 bar
=0.16 kg/s
Evaporator
Qevap = 27 kW
Tinlet = 30 ˚C
Toutlet = 90 ˚ C
Expander
P3/P4 = 5.6
Tinlet = 90 ˚C
Toutlet = 55 ˚ C
Condenser
22

Qcond = 35 kW
Tinlet = 55 ˚C
Toutlet = 28 ˚ C
Thermal efficiency
= 7.9 %
3.4.2.3 ORC REFRIGERANT PIPING
The design of refrigerant piping involves following criteria [35]:





Ensure proper refrigerant feed to the evaporator.
Provide practical refrigerant line sizes without excessive pressure drop.
Maintaining a clean and dry system.
Condenser drain line 100 fpm or less [35].
Liquid line 125 to 450 fpm [35].
The pipe size selection for most of the refrigerant is done using special purpose
pressure-drop chart provided by manufacturer. However, some refrigerants do not have such
charts. We use Darcy-Weisbach to estimate pipe diameter in such cases [36]:
In S.I. units: ∆
(3.3)
Where,
∆
is pressure drop (Pascal) Assumed
is friction factor (dimensionless) Calculated from Eq. 12
is Length of pipe (meter) Measured
is diameter of pipe (meter) Unknown?
is velocity of fluid (m/s) Velocity is measured in term of Diameter of pipe
from Equation (3.6)
is density of fluid (kg/m3) Provided by manufacturer.
is the function of the Reynold's number and roughness of the pipe [36]. Since flow is
turbulant, we use Blasius equation to calculate the friction factor [37].
Where,
.
.
(3.4)
is Reynolds number (dimensionless).
(3.5)
Length of the pipe
= 6 fts =1.83 meters,
Velocity of the fluid
Where,
is Volume flow rate given by
/
(3.6)
23
3.4.2.3.1 Diameter of the Pipe on the
Liquid Side of the ORC System
Assuming 5 % of the pressure drop across the pipe for pressure P1.
∆ = 5000 Pascals
= 1400 kg/m3 at 15 ̊ C for R245fa.
= 0.16 kg/s from 4.4.2.2 Results of the model.
0.00043 kg/m-s
Substitute all the above values in Equation (3.5) and Equation (3.6), Reynolds
number and Velocity is given by
&
.
(3.7)
Substitute value of Reynolds number from Equation (3.7) into Equation (3.4) friction
factor is given by
0.067
.
(3.8)
Substitute the above known values and Equation (3.7) and (3.8) in Equation (3.3), the
diameter of the pipe for the liquid side of the ORC system is determined.
Therefore,
.
0.18X10
0.0089 meters = 0.35"
3.4.2.3.2 Diameter of the Pipe on the
Vapor Side of the ORC System
Assuming 5 % pressure drop across the pipe for pressure P1.
∆ = 5000 Pascals
= 57 kg/m3 at 90 ̊ C for R245fa.
= 0.16 kg/s from 4.4.2.2 Results of the model.
0.0000127 kg/m-s at 90 ̊ C
Substitute all the above values in Equation (3.5) and Equation (3.6), Reynolds
number and Velocity is given by
&
. (3.9)
Substitute value of Reynolds number from Equation (3.9) into Equation (3.4) friction
factor is given by
0.067
.
(3.10)
24
Substitute the above known values and Equation (3.9) & (3.10) in Equation (3.3), the
diameter of pipe for the vapor side of the ORC system is determined.
.
Therefore,
0.186X10
0.014 meters = 0.55"
3.5 VARIED INITIAL CONDITIONS OF THE CYCLE
Results are calculated from varying temperature T1 and varied expander efficiency to
optimize the ORC system theoretically.
3.5.1 Varied T1
This section shows the performance of the cycle at different T1 (lowest temperature of
the cycle). Table 3.1 provides the assumed input data for ORC simulation.
Table 3.1.Theoretical Input Data for ORC Simulation
P1
T1
P2
P3
T3
0 ˚C – 30 ˚C 0.5 – 1.8 bar 10.5 bar 10 bar 90 ˚C
90%
60%
Output data: Table 3.2 provides ORC simulation results at varying T1.
Table 3.2. ORC Simulation Results at Varying T1
T1[˚C] P1[bar] T4[˚C]
Thermal
Mass
efficiency( )[%]
flow[kg/s]
[kJ/kg] Qin[kW] Qout[kW]
0
0.5
38
11.6
0.09
32.2
25
20.17
10
0.8
44
10.4
0.11
27.5
28
25
20
1.2
50
9.2
0.13
23.3
31.2
28.3
30
1.8
56
7.85
0.16
18.7
36.8
34
A Figure 3.6 shows the graphical representation of pressure P1 varying T1. The graph
shows that at lower temperature, the pressure inside the system decreases below atmospheric
pressure which makes it difficult to operate refrigerant pump and ultimately increasing the
prospect of system failure.
A Figure 3.7 shows the graphical representation of T4 varying T1. The graph shows
that the temperature T4 is directly proportional to the temperature T1.
25
Figure 3.6. Simulation of P1 varying T1.
Figure 3.7. Simulation of T4 varying T1.
A Figure 3.8 shows the graphical representation of thermal efficiency varying T1. The
graph shows that at is thermal efficiency inversely proportional to the temperature T1.
A Figure 3.9 shows the graphical representation of the mass flow rate varying T1. The
graph shows that mass flow rate is directly proportional to the temperature T1.
A Figure 3.10 shows the graphical representation of the turbine work varying T1. The
graph shows that turbine work is inversely proportional to the temperature T1.
26
Figure 3.8. Simulation of thermal efficiency varying T1.
Figure 3.9. Simulation of mass flow varying T1.
27
Figure 3.10. Simulation of turbine work varying T1.
Figures 3.6 – 3.10 show that lower T1 is always better to increase the efficiency of the
system. Figure 3.6 shows the simulation of pressure P1 at varied temperature T1. The region
shown in Figure 3.6 reveals the pressure P1 below atmospheric pressure. This will increase
the chances of atmospheric leaking into the system. Moreover, the cooling fluid used in the
experiment is tap water, which makes it more difficult to achieve the required temperature in
that region. Therefore, for theoretical calculations, the temperature T1 is taken as 28 ˚C, i.e.
around ambient temperature of the water and air.
3.5.2 Varied Expander Efficiency
The expander is one of main component of an ORC system that determines the
efficiency of the cycle. The expander efficiency is varied from 30% to 90% to gage system
performance. The following Table 3.3 shows the input data for the system simulation at
different expander efficiencies.
Table 3.3 Theoretical Input Data for ORC Simulation at Varied Expander
Efficiency
P1
T1
P2
P3
T3
P4
T4
p
t
1.7 bar 28 ˚C 10.5 bar 10 bar 90 ˚C 1.79 bar 55 ˚C 90% 30% - 90%
Output data: Table 3.4 provides ORC simulation results at varied expander efficiency.
28
Table 3.4 ORC Simulation at Varied Expander Efficiency
Expander efficiency( t)
Thermal efficiency( ) Mass flow( )
Wt [kJ/kg]
[%]
[%]
[kg/s]
30
3.8
0.31
9.58
40
5.17
0.23
12.77
50
6.54
0.19
15.96
60
7.91
0.16
19.15
70
9.28
0.13
22.34
80
10.65
0.12
25.54
90
12.02
0.1
28.73
Figure 3.11 shows the graphical representation of the thermal efficiency varying
expander efficiency. The graph shows that the thermal efficiency of the system is directly
proportional to the expander efficiency.
Figure 3.11. Simulation of thermal efficiency varying expander
efficiency.
Figure 3.12 shows the graphical representation of the mass flow rate varying
expander efficiency. The graph shows that mass flow rate is inversely proportional to the
expander efficiency.
29
Figure 3.12. Simulation of mass flow at varied expander efficiencies.
Figure 3.13 shows the graphical representation of the turbine work varying expander
efficiency. The graph shows that turbine work is directly proportional to the expander
efficiency.
Figure 3.13. Simulation of turbine work at varied expander
efficiencies.
Figures 3.11, 3.12 and 3.13 show the improvement in the performance of the ORC
system with increased expander efficiency. Figure 3.11 shows that if a scroll expander of
30
60% is replaced with a turbine of higher efficiency, total system efficiency can improve from
8% to 11%.
31
CHAPTER 4
ORC TEST BENCH
To facilitate experiments on the ORC system , a test bench was set up at engineering
laboratory EL101 at San Diego State University. The main goal is to prove the efficiency of
the ORC with a variety of different components. Figure 4.1 shows the schemetic view of the
experimental test bench.
4.1 CONFIGUREURATION OF ORC
The working fluid used in the ORC is R245fa. The heat source is consists of a heat
exchanger sourced with hot water from a boiler at a temperature ranging from 65-90 ˚C. The
expander employed is an oilfree scroll compressor configured to run as an expander. The
expander is linked with asynchronous machine by means of belt and pulley coupling. Laser
speed gun is used to determine the rotational speed of the expander. The heat sink i.e.
condenser is sourced with cold tape water at around 16 ˚C. Temperature transducers and
paddle wheel flow meters were used on both the inlet and the outlet of heat source and heat
sink to determine the water temperature and mass flow rate.
The working fluid is pumped in liquid state at high pressure through the heat source
via a coriollis mass flow meter. High pressure vapor created due to the heat exchange process
in the heat source is expanded in the expander transfers energy to the expander's rotating
shaft. This shaft is connected to asynchronous machine via a belt and pulley mechanism
producing electric power. After the expansion process the working fluid passes through the
heat sink i.e. condenser. An optional regenerator can also be inserted between the expander
and condenser to preheat the working fluid before it enters heat source. After heat rejection in
condenser the working fluid is passes through the expansion tank where it gets more time for
phase transformation before it is pumped to the evaporator. Figure 4.1 shows the
configuration of the ORC used in this experiment along with different measuring devices at
different points. The thermodynamic properties evident at different points helps to prove the
feasibility of the ORC as a low grade heat recovery system.
32
Hot water circulation pump
Evaporator
Hot water
From boiler
T,P
T,P
T
∆P
Expander
Transmission
Belt
Pump
W
Expansion
Tank
T
T,P
T,P
Electric
generator
Cold
Water
Submersible pump for cold
water circulation
Pressure transducer and thermocouple
T, P
Coriolis flow meter
Rotary flow meter
W
Electric power measurement
Figure 4.1. Organic rankine cycle test bench.
33
4.2 SCROLL EXPANDER
For this experiment we used the scroll compressor running in reverse to operate as an
expander. A scroll compressor working as an expander is of the positive displacement type.
The review of ORC Prime movers from section 1.3.3 pointed out that positive displacement
type of expanders are preferably used in small scale ORC.
4.2.1 Operating Principle of Scroll Compressor
A scroll compressor uses two interleaving vanes, one fixed and the other mobile to
pump compress or pressurize fluids. The vane geometry may be involute, archimedean spiral
or hybrid curves [38]. The mobile scroll orbits eccentrically without rotating, creating
compressed fluid pockets betweeen two vanes. For this experiment we used compressor that
produces compressed pockets by offsetting the center of rotation of two vanes and co rotating
them in synchronous motion [39]. Figure 4.2 [17] (a) shows the fluid pocket created at the
circumfurance of the vanes. Figure 4.2 (b) the inside vane is fixed and as the outer spiral
orbits, the fluid pocket is pushed towards the center. The pocket size decreases as it moves to
the center as shown in Figure 4.2 (c). Simultaneously the size of the fluid pocket decreases
and the pressure of the fluid increases as shown in Figure 4.2 (d). The compressed fluid is
finally discharged through the port located toward the center of the spirals. In the expander
the fluid flows from center to periphery [17].
Figure 4.2. Operating principle of scroll compressor. Source: Quoilin, Sylvain.
“Experimental Study and Modeling of a Low Temperature Rankine Cycle for Small
Scale Cogeneration.” Unpublished master’s thesis, University of Liège, 2009.
34
4.2.2 Leakages and Sealing
The performance of the expander is directly influenced by internal leakages and
mehanical loses. The fluid pockets created between two vanes tries to find a state of
equilibrium. If the gas on one side of the vane is at higher pressure than the gas on the other
side, then the higher pressure gas will try to seek a path toward lower pressure gas, resulting
in leakage. Generally, there are two types of leakage paths, i.e. radial and axial. Axial
leakage occurs between the tip of the vane walls where the pocket is trying to find its way
toward the low pressure region, while radial leakages occur between the vane flanks [17].
Scroll expanders come with or without lubrication depending upon the particular
application. Lubrication reduces the friction between vanes and reduces the leakage area.
However, adapting a scroll compressor to function in reverse raises the following problems:


When the vanes rotate the vane in the opposite direction, the oil pump for
lubrication might not work properly, if connected directly to the compressor shaft.
The solution is a separate oil circuit to be designed and added to the system.
Compability of the working fluid with the lubricating oil is not guaranted if the
compressor is not designed for organic fluids [17].
To overcome the above stated problems an oil free scroll expander was used for this
experiment. Two types of sealing are embedded in the expander to reduce leakage:


Radial sealing: The vane flanks of the expander never touches each other.
Tolerance of the vanes is so precise that even a thin film of oil can avoid leakage.
But in this oil free system, a circular peripheral sealant is added to the inner wall
of the fixed vane. The sealant isolates the inside of the scroll from the outside,
thus avoiding leakage to the outside [17].
Axial sealing: Usually scroll expanders are designed with machine grooved vane
tips and lubricating oil creates a film around the tip groove touching baseplate of
the opposite scroll, making system leak free. But in an oil free scroll expander, tip
grooves are nettled with sealant and which prevents radial leakage [17].
Figure 4.3 is a view of two scroll vanes. The orbitng scroll is on the right. The seals
are visible at the tip of each scroll. In expander mode the working fluid enters greater than
atmospheric pressure. So, any leakage to the outside of the scroll expander induces a loss of
mechanical work. To prevent or reduce this loss an extra layer of sealant is applied radially
and axially. The contact effort between two vanes is increased by means of an allen screw
and sealant adhesive is applied between the two scrolls. The drawback of this technique is an
increase in friction between the two parts.
35
Figure 4.3. Fixed and orbiting scrolls.
4.3.3 Specification of the Scroll Expander
Figure 4.4 shows the inlet and outlet of the scroll compressor adapted for this test
bench. Due to budget constrains we used most of the component those are readily available
off shelf. The scroll compressor used in this experiment is taken from the old ORC
experimental set up.Table 4.1 provides the specification of the scroll compressor used in this
experiment.
4.3 PUMP
A magnetically driven rotary vane pump is used in the experiment, so the driving
torque is transmitted magnetically instead of mechanically. Also an electric motor with high
rotational speed is used to gurantee performance. Powerstat variable autotransformer was
used in this experiment to controls the rotation speed of the electric motor Figure 4.5 shows
the Rotary Vane pump and Powerstat variable autotransformer used in this experiment.
The performance of the pump (pressure, flow rate) depends on the motor speed.
According to the theoretical calculations done in chapter 3, the maximum pressure required
to operate ORC cycle is 10.5 bar at 0.16 Kg/S of mass flow rate. Keeping those theoretical
36
Outlet
Inlet
Figure 4.4. Inlet/outlet of the scroll compressor.
Table 4.1 Scroll Compressor Specifications
Brand
Sanden, installed on honda civic
Model
TRSA09 (85.7 cm3/rev)
Refrigerant
HFC134a
Lubricating
SP-10
High pressure side
197 kPa
Mass
3.8 kg
Type
Engine belt driven compressor
37
.
Figure 4.5. Magnetic centrifugal pump on
left and powerstat variable
autotransformer on right.
calculations in mind we used the manufacturer's (Fluid O Tech) provided graphs of pressure
against flow rate (Figure 4.6) at different motor speed to determine the right pump and motor
speed for our application.
Table 4.2 shows specifications of the rotary vane pump and electric motor.
4.4 CONDENSER
Figure 4.7 shows the condenser used in the ORC system. The condenser used is a
plate type heat exchanger in which a stack of corrugated plates are brazed together. Cold
water is the cooling fluid used in the condenser. Wp series of GEA heat exchanger is used.
Due to budget constrains we used most of the component those are readily available
off shelf. The condenser used in this experiment is taken from the old ORC experimental set
up.
38
Figure 4.6. graphs pressure – flow rate at different motor speed; the model TMOT151.
Table 4.2 Rotary Vane Pump and Electric Motor Specifications
Model
Fluid O Tech TMOT151
Type
magnet drive rotary vane
Housing
Brass
inlet/outlet ports
3/8"
motor speed
3600 rpm
motor power
0.5 kW
max temperature
60°C
max system pressure
18 bar
39
Figure 4.7. Condenser used in ORC.
Table 4.3 shows some of the thermodynamic properties of fluids passing through the
heat exchanger.
Table 4.3 Thermodynamic Properties of Two Streaming Fluids in Condenser
Cold side (cold water)
Hot side (hot refrigerant)
Flow rate (gpm)
2
15
Inlet temperature (˚ C)
16
49
Outlet temperature (˚C)
34
30
Pressure drop (PSI)
3.6
4.5 EVAPORATOR
Figure 4.8 shows the evaporator used in the ORC system. It is a plate typ heat
exchanger. In this case the SL 70 series sondex brazed heat exchanger is used. The heat
source is hot water from the boiler at around 90 ˚C. Due to budget constrains we used most
of the component those are readily available off shelf. The evaporator used in this experiment
is taken from the old ORC experimental set up.
Table 4.4 shows some of the thermodynamic properties of fluids passing through the
evaporator.
40
Figure 4.8. Evaporator for ORC.
Table 4.4 Thermodynamic Properties of Two Streaming Fluids in the Evaporator.
Cold side (refrigerant)
Flow rate (gpm)
2
Hot side (hot water)
16
Inlet temperature (˚ F)
86
217
Outlet temperature (˚ F)
194
110
Pressure drop (PSI)
3.72
4.6 CONDENSER WATER PUMP
Figure 4.9 shows the submersible water pump used to circulate cold water from the
storage tank to the condenser cold side. The Table 4.5 shows the condenser water pump
specifications. Due to budget constrains we used most of the component those are readily
available off shelf. The condenser water pump used in this experiment is taken from the old
ORC experimental set up.
4.7 BOILER AND HOT WATER PUMP
A high temperature circlation pump, between the evaporator ad boiler is used to
circulate hot water through the heat exchanger. Hot water from the boiler is used as
circulating fluid. A domestic boiler serves as a heat source for the ORC system. Hot water at
41
Figure 4.9. Condenser water pump.
Table 4.5 Submersible Pump Specifications
Type
Submersible pump
Power
0.25 kW
Volts
115 V
Amps
8.5 A
Frequency
60 Hz
Flow rate
14.12 gpm
Fluid
Water
approximately 95 is ˚C circulated through the heat exchanger and back to the boiler.
Temperature transducers and paddle wheel flow meteres were used on both the inlet and
outlet of heat source i.e. evaporator to determine the water temperature and mass flow rate
enabling measurement of temperature differential and heat transfer. According to the
theoretical calculations done in chapter 3, the maximum expander inlet temperature required
to operate ORC cycle is 90˚C. Keeping this theoretical calculation in mind we used the
manufacturer's (RHEEM) recommended boiler. Due to budget constrains we used most of
the component those are readily available off shelf. The hot water circulation pump used in
this experiment is taken from the old ORC experimental set up.
Table 4.6 shows the boiler and hot water pump specifications. Figure 4.10 shows the
boiler and hot water pump used in the ORC.
42
Table 4.6 Boiler and Hot Water Circulation Pump Specifications
Boiler
Capacity
50 gallons
Power source
240 VAC 50/60 Hz 9 KW 3PH
Current
24 amps
Thermal efficiency
98 %
Hot water pump
Power
1/8 KW
Max water temperature
110 ˚ C
Max pressure
125 Psi
Speed
3250 rpm
Figure 4.10. Boiler and water pump for ORC.
43
4.8 ELECTRIC GENERATOR
A belt driven permanent magnet alternator generates power. Permanent magnet
alternators (PMA) are a better option than traditional electric generators. PMA produce
power at lower RPMs. We used the manufacturer's (Windblue) provided voltage Figure 4.11
pictures the alternator used. Figure 4.12 shows the voltage and ampere output chart of DC540
PMA used for this experiment.
Figure 4.11. PMA alternator.
Figure 4.12. DC540 voltage and ampere output chart.
44
4.9 MEASUREMENT DEVICES
Measurement devices helped to prove the feasibility of ORC system. The following
describes about the measurement devices used in the experiment.
4.9.1 Pressure Transducer
The PX 309 series of pressure transducer from omega is used for pressure
measurement. A Micro machined silicon diaphragm is installed in the device with a piezoresistive strain gauge diffused into it. Silicon is fused with the diaphragm. The resistor has an
approximate value of 3.5 K- Ohm. Pressure induced strain changes the resistance across the
piezo resistive strain guage, increasing resistance across the radial resistors and decreasing it
across transverse resistors. The resistors are connected to the wheatstone bridge. This strain
change across the wheatstone bridge produces output signals in milliamperes which are
directly proportional to the pressure [40]. The output signal is connected to the LABVIEW in
the computer through a data aquisition system (DAQ) which calculates the pressure value.
Figure 4.13 shows the pressure transducer used in the ORC and Table 4.7 shows the
specification of the pressure transducer.
Figure 4.13. Pressure transducer.
45
Table 4.7. Pressure Transducer Specification
Excitation
9 to 30 Vdc (< 10 mA)
Output
4 to 20 mA
Accuracy
0.25 %
Pressure range
0 -300 psi
4.9.2 Thermocouple
The junction of dissimilar metals produces an electrical potential which is propotional
to the temperature with some fixed predictable or repeatable relationship depending on the
kind of alloys used. Extension wires are used as an intermediate connection between the
measuring point and measuring device i.e. DAQ. Figure 4.14 shows the thermocouple used
in the experiment. Table 4.8 shows it specification.
Figure 4.14. Thermocouple.
Table 4.8. Thermocouple Specification
Model
TC-K-NPT-U-72
Type
K (chromal alumed)
Measurement range
-40 ˚C to 1000 ˚C
Sensitivity
41 µV/ ˚C
46
4.9.3 Coriolis Flowmeter
A coriolis flowmeter is used to measure the refrigerant flow rate. It works on the
principle of coriolis force and consists of a vibrating flowtube and electrical transmitter. The
electrical transmitter maintains vibration of the flowtube and also measures the amount of
vibration. The flow tubes are vibrated at resonant frequency just like the fork. The fluid flow
through flow tubes is oscillated by the flowtube vibration which induces coriolis force along
the tube’s length. This force is translated into the phase shift shown in Figure 4.15 [40]. The
phase shift i.e. sinusodial wave generated by vibration of the flowtube and fluid acceleration
is detected by two electromagnetic sensors located at the tube’s leg, as shown in Figure 4.15.
The processing fluid’s temperature and density is also determined through the digital analog
because almost all the variables are functions of the resonant frequency. Figure 4.16 shows
the coriolis flow meter used in the experiment. Table 4.9 shows the flowmeter’s
specifications [6].
Figure 4.15. Phase shift along the tube’s length. Source: De Jonge,
T., T. Patten, A. Rivetti and L. Serio. “Development of a Mass
Flow Meter Based on the Coriolis Acceleration for Liquid,
Supercritical and Superfluid Helium.” Paper presented at the 19th
International Cryogenic Engineering Conference (ICEC 19),
Grenoble, France, July 2002.
4.9.4 Paddlewheel Flowmeter
The paddlewheel flowmeter has a rotating paddle which generates a signal according
to the velocity of the fluid flowing through the pipe cross section. The average velocity of the
moving fluid is read as mass flow rate. A Flowmeter is placed ahead the evaporator and
47
Figure 4.16. Coriolis flow meter.
Table 4.9. Coriolis Flow Meter Specification
Model
TCM-28K-AE-SGSS-CSDS
Flow range
Upto 600 kg/hr
Tube material
316 stainless steel
Pressure rating
200 bar
Accuracy
0.10 % and zero stability
Excitation
9-30 Vdc
Maximum Temperature
125 ˚C
condenser water side. The flowmeter also includes a temperature sensor which measures 0 ˚C
to 100 ˚C. Figure 4.17 shows the paddlewheel flowmeter used in the experiment. Table 4.10
shows the specifications of the paddlewheel flowmeter used in ORC.
4.10 DATA ACQUISITION (DAQ)
Data acquisition is the process of converting the signals into digital values. Sensors
guages physical system parameters and provide information to the DAQ hardware. The DAQ
hardware feeds the computer with through its DAQ software. DAQ software converts the
signals into digital values which reflect accurate physical parameter for this experiment.
LABVIEW DAQ software used in the experiment. Figure 4.18 shows the DAQ hardware
48
Figure 4.17. Paddlewheel flowmeter.
Table 4.10. Paddlewheel Flowmeter Specification
Type
Paddlewheel
Accuracy
2 FS
Repeatablility
1 FS
Power
12 to 18 Vdc @ 50 mA
Output
4 to 20 mA
Temperature
0 ˚C to 100 ˚C
Max viscosity
5 cps
Figure 4.18. DAQ hardware.
connecting the sensor output signals to the computer. Figure 4.19 shows the LABVIEW
software used for the digital calibration.
49
Figure 4.19 DAQ software.
50
CHAPTER 5
DESCRIPTION OF EXPERIMENT
This chapter describes various tests performed in the laboratory. For each test, a
summary of results and a detailed analysis are provided. Figure 5.1 show the experimental set
up of the ORC system.
Figure 5.1. ORC experimental set up.
5.1 LEAK DETECTION TEST
Two different kind of leak detection test were performed on ORC system to
materialize the working of the cycle.
51
5.1.1 Description of Leak Detection Test 1
All ORC components were set up and adapted as shown in Figure. 4.1 (ORC test
bench of chapter 4). A series of leak detection tests were performed to determine the
peripheral leaks in the cycle.
Starting procedure:





The entire measuring apparatus i.e. DAQ system was checked and activated for
the test.
High pressure Nitrogen at 12 bar was introduced to ORC to produce the
maximum operating pressure of ORC i.e. 10.2 bar.
Pressurized ORC test best was kept for about an hour.
If a drop in pressure is detected, there is a leak somewhere in the system.
Soap solution (bubble soap method) was employed to detect peripheral leaks in
the system.
Results:

A pressure drop was detected in the ORC system through DAQ system. The
location of the leak was identified with the soap solution. The leak was fixed after
removing nitrogen from the system. Thereafter, the leak detection test was
repeated until a perfect leak free system was achieved. Figure 5.2 show the graph
of Pressure Vs Time for the final leak free ORC system.
Figure 5.2. High pressure nitrogen test for peripheral
leak detection test.
5.1.2 Description of Leak Detection Test 2
Similar approach to leak detection was applied in test #2. Instead of pressurizing the
ORC test bench, a vacuum was created in the circuit. Figure 5.3 shows the vacuum pump
52
Figure 5.3. Vacuum pump and associated
pressure gauges.
used in the experiment. This test was performed only if the ORC system retained its pressure
in leak detection test 1.
Starting procedure:




The entire measuring device i.e. DAQ system were checked and turned on during
test.
A vacuum was created in the cycle with vacuum pump.
The vacuumed ORC test best was kept still for about an hour.
If a change in pressure is detected there is a leak in the system. If a leak was
detected leak detection test 1 was repeated until the problem was resolved.
Results:

No change in pressure change was detected in the ORC system. So, the ORC test
bench was ready to perform experiments using refrigerant R245fa. Figure 5.3
shows the graph of Pressure Vs Time for the leak free ORC system.
5.2 REFRIGERANT CHARGE IN ORC
As the circuit is completely emptied after the vacuum test, a evaluation of
the refrigerant charge is needed. The volume of each element has to be calculated. The
regions of the cycle where the fluid is in vapor state are not taken into account, the density of
the vapor being negligible compared to that of the liquid.
53
Figure 5.4. Vacuum test for peripheral leak detection.
The volumes taken into account are the following:
1. Volume of the pump
Vpump= 0.011 Cubic ft. = 0.000311485 cubic mt.
2. Volume of the piping.
The tubes concerned are the tubes between the outlet of the condenser and the
inlet of the evaporator. The volume includes that of the expansion tank. The total
length of the tubes is approximately 85 inches and their internal diameter 0.5
inches.
Vtubes + Vexpansion tank= 0.018 + 0.32 Cubic ft. = 0.338 cubic ft. = 0.009 cubic
mt.
3. Volume of the evaporator:
The total volume of the heat exchangers is 0.42 cubic ft. The working fluid only
occupies half of that volume, the other half being filled with hot water. An
arbitrary assumption is that half of the evaporator volume is filled with liquid to
ensures there is enough refrigerant to fill the circuit. Adjustment of the refrigerant
charge for the desired level is done by adding or removing refrigerant during the
test.
The volume is thus expressed by :
Vevaporator = 0.21 cubic ft.= 0.005946538 cubic mt.
4. Volume of the condenser:
The approach is the same as for the evaporator. In this case as well, it is also
assumed that 1/2 of the condenser is filled with liquid. The total volume of
condenser is 0.29 cubic ft.
Vcondenser= 1/2 * 0.29 = 0.145 cubic ft. = 0.0041 cubic mt.
The amount of refrigerant that needs to be introduced in the circuit is given by:
Mass of R245fa = liquid density of R245fa X Total volume
= 1018 X 0.017
= 17 Kg
54
The refrigerant charge to be introduced in the circuit is thus evaluated to 18 kg, in
order to take into account the refrigerant mass in vapor state and the possible
refrigerant losses during transfer.
5.3 ORC PERFORMANCE AND RESULTS
The following are the experiment performed to determine the expander efficiency
used in the ORC cycle.
5.3.1 First Set of Tests
All the ORC components are set up as shown in Figure 4.1. The refrigerant is charged
inside the ORC system using the refrigerant recovery method in reverse direction. The main
aim of this test is to calculate expander efficiency at 0.80 kg/s mass flow rate of hot water
circulating through the evaporator.
Starting procedure:





The entire measuring device i.e. DAQ system were checked and activated during
test.
Turn on the water heater and hot water circulation pump. Fix the mass flow rate at
0.8 kg/s.
When water temperature reaches 92 ˚C, turn on condensing water before
operating the ORC system
Turn on the refrigerant pump at full load (300W).
Measure the rpm of the expander connected to asynchronous machine.
Parameters:
Table 5.1 describes the conditions under which the experiment was performed:
Primary results: Table 5.2 describes the results associated.
Figure 5.5 provides the graphical representation of experimental results, T3 [˚C] vs
,
From Figure 5.5 shows clearly that increasing the heat source temperature increases expander
efficiency. The average experimental value of expander efficiency is about 2.145 %.
Table 5.1 Experimental Parameters for First Test
Parameters
Value
Hot water source THWin
92 ˚C- 75 ˚C range
Hot water flow rate
0.8 kg/S
Cold water source TCin
16 ˚C
Cold water flow rate
0.55 kg/S
55
Table 5.2 Experimental Results of First Test
T3
P3
P4
[˚C]
[bar]
85
8.3
65
3.0
17
2.6
20
8.4
85
8.5
67
2.65
17
2.28
20
83
8.5
68
2.5
17
2.2
82
8.35
66
2.5
17
80
8
64
2.4
79
7.7
63
77
7.3
75
Expander
efficiency
(%)
0.07
8.6
0.07
61
2.5
19
8.6
0.07
59
2.4
2.1
19
8.4
0.07
57
2.3
31
2
24
8.1
0.07
55
2.39
2.3
27
1.9
26
7.8
0.07
52
1.98
61
2.21
29
1.8
26
7.4
0.07
47
1.9
7.021
59
2.11
29
1.8
25
7.09
0.07
46
1.89
73
6.7
57
2.03
28
1.7
25
6.76
0.07
45
1.86
71
6.3
56
1.95
28
1.7
25
6.43
0.07
45
1.83
T1
P1
T2
P2
Refrigerant
flow rate
(kg/s)
Turbine
work
(Power)
(watts)
60
T4
[˚C] [bar] [˚C] [bar] [˚C] [bar]
2.4
5.3.2 Second Set of Tests
All the ORC components are set up as shown in Figure 4.1. The refrigerant is charged
inside the ORC system using the refrigerant recovery method in reverse direction. The water
heater is turned on. When the water heater temperature reaches 92 ˚C, before running the
ORC system the condenser water is turned on. The purpose of this test is to calculate
expander efficiency at 0.80 kg/s mass flow rate of hot water circulating through evaporator.
Starting procedure:





The entire measuring device i.e. DAQ system was checked and activated for the
test.
Turn on the water heater and hot water circulation pump. Fix the mass flow rate at
0.6 kg/s.
When water temperature reaches 92 ˚C, turn on condenser water before operating
the ORC system
Turn on the refrigerant pump at full load (150W).
Measure the rpm of the expander that is connected to the asynchronous machine.
Parameters:
Table 5.3 describes the condition in which the experiment is performed:
Primary results: Table 5.4 describes the results associated.
56
Figure 5.5. Graphical representation of experimental
results, T3 [˚C] vs .
Table 5.3. Experimental Parameters for Second Test
Parameters
Value
Hot water source THWin
92 ˚C- 75 ˚C range
Hot water flow rate
0.6 kg/S
Cold water source TCin
16 ˚C
Cold water flow rate
0.55 kg/S
Figure 5.6 provides the graphical representation of experimental results, T3 [˚C] vs
. From Figure 5.6 it is clear that increasing the heat source temperature improves expander
efficiency. The average experimental value of expander efficiency was about 2.042 %.
5.4 THEORETICAL CYCLE EFFICIENCY AT VARIED
TURBINE EFFICIENCY
At various turbine efficiencies ( ), from 2 % to 90 %, theoretical work done by a
turbine is calculated using Equation (3.2) from chapter 3. Accordingly, ORC efficiency is
calculated using Equation (3.1) from chapter 3. Table 5.5 reveals the theoretical value of
cycle efficiency and turbine work done at varied turbine efficiencies.
Figure 5.7 shows the graphical representation of cycle efficiency at varied turbine
efficiencies,
vs
using results from Table 5.5. Assuming 10% cycle efficiency as the
payoff for the ORC operation with low grade thermal energy recovery, the minimum turbine
57
Table 5.4 Experimental Results for Second Test
84
8.1
[˚C]
64
3.0
17
2.6
[˚C]
20
8.2
0.05
Turbine
work
(Power)
(watts)
60
84
8.3
66
2.65
17
2.28
20
8.4
0.05
61
2.29
81
8.2
67
2.5
17
2.2
19
8.3
0.05
60
2.21
81
8.1
66
2.5
17
2.1
19
8.14
0.05
58
2.1
80
7.8
64
2.4
31
2
24
7.9
0.05
55
2.1
79
7.5
63
2.3
27
1.9
26
7.6
0.05
54
1.98
77
7.1
61
2.21
29
1.8
26
7.2
0.05
51
1.9
75
6.91
59
2.11
29
1.8
25
7.09
0.05
45
1.86
72
6.5
56
2.03
28
1.7
25
6.65
0.05
45
1.85
71
6.1
56
1.95
28
1.7
25
6.23
0.05
45
1.83
T3
P3
T4
P4
T1
P1
T2
[bar] [˚C] [bar]
[˚C] [bar]
[bar]
Refrigerant
flow rate
(kg/s)
P2
Figure 5.6. Graphical representation of experimental
results, T3 [˚C] vs .
Expander
efficiency
(%)
2.3
58
Table 5.5 Theoretical Calculation of Cycle Efficiency
and Turbine Work Done
Turbine
efficiency
2%
Turbine work
done t
60
Cycle
efficiency
-2.2 %
10 %
245
-0.5 %
20 %
490
1.8 %
30 %
735
4.1 %
40 %
980
6.4 %
50 %
1225
8.8 %
60 %
1470
9.2 %
70 %
1715
11.8 %
80 %
1960
14.8 %
90 %
2205
17.5 %
Figure 5.7. Graphical representation of theoretically calculated cycle efficiency with
turbine efficiency plot, vs .
59
efficiency required is 60%. The graph shows that if the usual scroll expander efficiency of
2% is improved using a turbine with higher efficiency, overall efficiency of the ORC system
can increase to as high as 17 %.
5.5 DISCUSSIONS
From Figure 5.5 and 5.6 it is clear that the current conventional expander used in this
experiment has an efficiency of 2 %. From the above experimental results and theoretical
evaluation, it is clear that the ORC component that shows the greatest improvement is the
expander. The scroll compressor was modified to work as an expander and the following
might be account for its low efficiency:




Seals installed in the modified expander may have been damaged by friction
which could cause internal leakage.
Expansion of the working fluid might have occured before entering the expander
vanes. This would reduce the amount of recoverable pressure for the turbine.
Friction associated with the pipes, heat exchangers, curves etc. might reduce the
refrigerant flow rate, ultimately lowering the turbine work.
Improvement to the expander could be achieved with increased pressure.
However, pressure cannot be increased given the limitations of heat source.
Substituting a heat source that can provide consistent high temperature or
providing a redundancy system for the boiler might solve heat source problem.
60
CHAPTER 6
CONCLUSION
In this study, an ORC system was developed using a compact scroll expander for low
grade thermal energy recovery. HFC-245fa was employed as the working fluid.
The following observations were made:
1. The measured efficiency of the scroll expander was 2.4%, 2.3%, and 1.83% for
the temperature T3 = 85°C, 80°C, and 70°C, respectively. The maximum
expander power was 60W in the case of T3 = 85°C. A larger pressure difference
between the expander inlet and outlet resulted in larger expander power.
2. The measured thermal efficiencies of the cycle with 2.4 % expander efficiency
and 60 W expander power was -2.2 %. This negative efficiency shows that the
ORC component responsible for most of the improved efficiency is the expander.
3. Assuming variable turbine efficiency ranging from 2 % to 90 %, a calculation
of work done by turbine and ORC thermal efficiency is shown in Table 5.5.
Assuming a 10 % ORC thermal efficiency is the target efficiency of ORC system,
a minimum of 60 % expander efficiency is necessary.
The Figure 5.7 in chapter 5 is the graph plot between cycle efficiency and turbine
efficiency that explains how turbine efficiency affects overall cycle efficiency.
6.1 SUMMARY
The use of R245fa as the working fluid offers the potential for recovery of low grade
thermal energy from a variety of heat sources. The working fluid is advantageous as it has
high thermal capacity to absorb heat and is environmentally friendly. So, R245fa can provide
mechanical power at low risk and low operating maintenance.
The experimental results show that R245Fa provides the ORC functionability with off
the shelf components. The system achieved only negative efficiency without optimization of
ORC components. The theoretical analysis shows that if the current scroll expander of 2%
efficiency is replaced with a higher efficiency turbine unit, the efficiency of the ORC system
can reach to 17%. Such efficiency with low cost components, can make the system cost
competitive with solar power PV systems both residential and industrial applications.
61
6.2 RECOMMENDATION
The ORC test bench can be improved to get more accurate results and better control
over the parameters which will provide higher turbine work and thus, higher cycle efficiency.
The following are some of the recommendations applicable to the ORC system.
Improvement in Test bench:




Changing current heat source (water heater) to stable, higher heat source.
The maximum temperature of current heat source is approximately 92 ˚C. But
during ORC operation, it was noticed that the temperature drops very quickly. It
takes lot of time to return it to 92 ˚C. Thus, the ORC system never achieved a
steady state condition. This can be improved by providing a constant and higher
heat source with high heat recovery technology. This would result in steady state
operation of ORC system.
Installation of direct transmission from expander to asynchronous machine .
This would preclude the uncertainty of belt transmission and increase expander
shaft power output.
Improvement of expander.
This is the most important component that requires major improvement. As seen
from the results in chapter 5, the conventional expander used in experiment has an
average efficiency of about 2 %. Optimization of the expander has the highest
potential for boosting total efficiency of the ORC. The best option would be the
use of a hermetic expander. The following are the major advantages associated in
use of a hermetic expander:
- No external leakage over the long term, preventing refrigerant loses.
- Reduction in heat, due to increase in temperature differential across the
expander.
The lubrication needed for a scroll expander is advantageous, as it reduces
leakage and friction loses. However, lubrication raises several problems such as
lowering pump efficiency and lowering the heat transfer rate of the condenser and
expander due to sticking in heat exchanger plates etc.
Optimize pipe diameter
Optimize the pipe diameter to reduce the isentropic work done during the
pumping phase. The diameter of the pipe shall correspond to the current flow rate
that is achieved through the experiments.
62
REFERENCES
[1]
Sebastien Declaye. "Design, Optimization and Modeling of an Organic Rankine Cycle
for Waste Heat Recovery." Master’s thesis, University of Leige, 2009.
[2]
T. C. Hung, T. Y. Shai and S. K. Wang. “A Review of Organic Rankine Cycle for the
Recovery of Low Grade Waste Heat.” Energy 22, no. 7 (1997): 661-667.
[3]
"External Combustion Engine.” Merriam Webster, accessed November 7, 2012,
http://www.merriam-webster.com/dictionary/external%20combustion.
[4]
Md. Ali Tarique. “Experimental Investigation of Scroll Based Organic Rankine
Systems.” Master’s thesis, University of Ontario Institute of Technology, 2011.
[5]
Towne, Paul. “The Naphtha Engine.” Last modified October 1991.
http://gasengine.farmcollector.com/Gas-Engines/The-Naphtha-Engine.aspx.
[6]
J. Larjola. “Electricity from industrial waste heat using high-speed organic Rankine
cycle (ORC).” International Journal of Production Economics 41, no. 1 (1995): 227–
235.
[7]
Patel, Parimal and Edward Doyle. Compounding the Truck Diesel Engine with an
Organic Rankine-Cycle System. SAE International, No. 760343. Last modified
February 1, 1976. http://papers.sae.org/760343/.
[8]
Madhawa Hettiarachchi, H. D., Mihajlo Golubovic, William M. Worek, and Yasuyuki
Ikegami. “Optimum Design Criteria for an Organic Rankine Cycle Using LowTemperature Geothermal Heat Sources.” Energy 32, no. 9, (2007): 1698–1706.
[9]
"Harry Zvi Tabor.” Wikipedia. Last modified April 25, 2013.
http://en.wikipedia.org/wiki/Harry_Zvi_Tabor.
[10] Manolakos, D., G. Papadakis, Essam Sh. Mohamed, S. Kyritsis and K. Bouzianas.
“Design of an Autonomous Low-Temperature Solar Rankine Cycle System For
Reverse Osmosis Desalination.” Desalination 183, no.1–3 (2005): 73–80.
[11] Mago, Pedro J., Louay M. Chamra, Kalyan Srinivasan, and Chandramohan Somayaji.
“An Examination of Regenerative Organic Rankine Cycles Using Dry Fluids.” Applied
Thermal Engineering 28, no. 8–9 (2008): 998–1007.
[12] Vaja, Iacopo, and Agostino Gambarotta. “Internal Combustion Engine (ICE)
Bottoming with Organic Rankine Cycles (ORCs).” Energy 35, no. 2 (2010): 1084–
1093.
[13] Badr, O., S. D. Probert and P. O'Callaghan. "Performances of Multi-Vane Expanders."
Applied Energy 20, no. 3 (1985): 207–234.
[14] Zanelli, R., and D. Favrat.“Experimental Investigation of Hermetic Scroll ExpanderGenerator.” Paper presented at the International Compressor Engineering Conference,
West Lafayette, IN, July1994.
63
[15] Mathias, James A., Jon R. Johnston Jr., Jiming Cao, Douglas K. Priedeman, and
Richard N. Christensen. “Experimental Testing of Gerotor and Scroll Expanders Used
in, and Energetic and Exergetic Modeling of, an Organic Rankine Cycle.” Journal of
Energy Resources Technology 131, no. 1 (2009): 9.
[16] Wang, X. D., L. Zhao, J. L. Wang, W. Z. Zhang, X. Z. Zhao and W. Wu. “Performance
Evaluation of a Low-Temperature Solar Rankine Cycle System Utilizing R245fa.”
Solar Energy 84, no. 3 (2010): 353–364.
[17] Quoilin, Sylvain. “Experimental Study and Modeling of a Low Temperature Rankine
Cycle for Small Scale Cogeneration.” Unpublished master’s thesis, University of Liège,
2009.
[18] Brasz, Lars J., and William M. Bilbow. "Ranking of Working Fluids for Organic
Rankine Cycle Applications." Paper presented at the International Refrigeration and Air
conditioning conference, West Lafayette, IN, July 2009.
[19] Wang, E. H., H.G. Zhang, B. Y. Fan, M. G. Ouyang, Y. Zhao, and Q. H. Muc. “Study
of Working Fluid Selection of Organic Rankine Cycle for Engine Waste Heat
Recovery.” Energy 36, no. 5 (2011): 3406–3418.
[20] Hung, Tzu-Chen. “Waste Heat Recovery of Organic Rankine Cycle Using Dry Fluids.”
Energy Conversion and Management 42, no. 5 (2001): 539–553.
[21] Mago, P. J., L. M. Chamra, and C. Somayaji. “Performance Analysis of Different
Working fluids for Use in Organic Rankine Cycles.” Journal of Power and Energy 221,
no. 3(2007): 255-263.
[22] Chen, Y., P. Lundqvist, A. Johansson, and P. Platell. “A Comparative Study of the
Carbon Dioxide Transcritical Power Cycle Compared with an Organic Rankine Cycle
with R123 as Working Fluid in Waste Heat recovery.” Applied Thermal Engineering
26, no. 17–18 (2006): 2142–2147.
[23] “Refrigerant Safety.” U.S. Environmental Protection Agency. Last modified August 19,
2010. http://www.epa.gov/ozone/snap/refrigerants/safety.html.
[24] Quoilin, S. "An Introduction to Thermodynamics Applied to ORCs." Master’s thesis,
University of Leige, 2008.
[25] “Overview of Greenhouse Gases.” U.S. Environmental Protection Agency. Last
modified June 21, 2013. http://www.epa.gov/climatechange/ghgemissions/gases.html.
[26] “Global-warming potential.” Wikipedia. Last modified June 5, 2013.
http://en.wikipedia.org/wiki/Global-warming_potential.
[27] “Ashrae 15: Safety Code for Mechanical Refrigeration.” American Society of Heating,
Refrigerating and Air-Conditioning Engineers, Inc.
https://law.resource.org/pub/us/cfr/ibr/002/ashrae.15.1994.pdf.
[28] “Safety Data Sheets.” BOC. Last modified 2011.
http://www.bocsds.com/uk/sds/special/R245fa.pdf.
[29] “Rankine cycle.” Wikipedia. Last modified June 1 2013.
http://en.wikipedia.org/wiki/Rankine_cycle.
64
[30] Spakovszky, Z. S. “8.1 Behavior of Two-Phase Systems.” Last modified September 3,
2007. http://mit.edu/16.unified/www/FALL/thermodynamics/notes/node61.html.
[31] Jahono. “General Engineering.” Answerbag. Last modified April 13, 2010.
http://www.answerbag.com/q_view/1998740.
[32] Yamamoto, Takahisa, Tomohiko Furuhata, Norio Arai, and Koichi Mori,. “Design and
Testing of the Organic Rankine Cycle.” Energy 26, no. 3 (2001): 239–251.
[33] “Log Mean Temperature Difference.” Wikipedia. Last modified February 27 2013.
http://en.wikipedia.org/wiki/Log_mean_temperature_difference.
[34] Lemort, Vincent, Sylvain Quoilin, Cristian Cuevas, and Jean Lebrun. “Testing and
Modeling a Scroll Expander Integrated into an Organic Rankine Cycle.” Applied
Thermal Engineering 29, no. 14–15 (2009): 3094–3102.
[35] Holder, Roger D. “Refrigeration Piping & Charging Residential AirConditioning.”
Accessed March 24, 2013.
http://www.refrigtech.com/Knowledge_Center/Knowledge_Piping_Charging.pdf.
[36] Stoecker, Wilbert F. “Refrigerant Piping.” In Industrial Refrigeration Handbook. New
York: McGraw-Hill, 1998.
[37] Jukka Kiijarvi. Darcy Friction Factor Formulae in Turbulent Pipe Flow.
Kirkkonummi, Finland: Lunowa Fluid Mechanics No.110727, 2011.
[38] Arai, Nobukatsu, Hirokatu Kousokabe, Eiji Sato, and Kenji Tojo. Scroll compressor
with means for end plate bias and cooled gas return to sealed compressor spaces. U. S.
Patent 4216661, filed December 8, 1978, and issued August 12, 1980.
[39] “Scroll compressor.” Wikipedia. Last modified April 24, 2013.
http://en.wikipedia.org/wiki/Scroll_compressor.
[40] De Jonge, T., T. Patten, A. Rivetti and L. Serio. “Development of a Mass Flow Meter
Based on the Coriolis Acceleration for Liquid, Supercritical and Superfluid Helium.”
Paper presented at the 19th International Cryogenic Engineering Conference (ICEC 19),
Grenoble, France, July 2002.