Experimental testing of a low-temperature organic Rankine cycle [618350]

Experimental testing of a low-temperature organic Rankine cycle
(ORC) engine coupled with concentrating PV/thermal collectors:Laboratory and field tests
George Kosmadakisa,*, Arnaud Landelleb,c,d,1, Marija Lazovae,1, Dimitris Manolakosa,
Alihan Kayae, Henk Huisseunee, Christos-Spyridon Karavasa, Nicolas Tauveronb,
Remi Revellinc, Philippe Haberschillc, Michel De Paepee, George Papadakisa
aDepartment of Natural Resources and Agricultural Engineering, Agricultural University of Athens, Iera Odos Street 75, Athens 11855, Greece
bCEA, LITEN, DTBH, SBRT, LS2T, 17 rue des Martyrs, F-38054 Grenoble, France
cUniversit /C19e de Lyon, INSA Lyon, CETHIL UMR5008, F-69621 Villeurbane, France
dADEME, 20 Avenue du Gr /C19esill /C19e, 49004 Angers, France
eDepartment of Flow, Heat and Combustion Mechanics, Ghent University, Sint-Pietersnieuwstraat 41, Ghent 9000, Belgium
article info
Article history:
Received 27 April 2016
Received in revised form12 August 2016Accepted 17 October 2016Available online 28 October 2016
Keywords:
Organic Rankine cycleConcentrating PV/thermalLow-temperatureScroll expander
Evaporator
Diaphragm pumpabstract
A detailed experimental investigation of a small-scale low-temperature organic Rankine cycle (ORC) with
R-404A is presented. The tests are first conducted at laboratory conditions for detailed evaluation of the
main components at both design and off-design conditions, for variable heat input up to 48 kW thand hot
water temperature in the range of 65 e100/C14C. A scroll compressor in reverse operation is used as
expansion machine and a dedicated helical coil heat exchanger is installed, suitable for high-pressure andtemperature operation. The ORC pump is a diaphragm pump coupled with an induction motor. The
rotational speeds of both the expander and pump are regulated with frequency inverters, in order to have
the full control of the engine operation. The ORC has been then connected with concentrating PV/thermal collectors, which produce electricity and heat and provide it to the ORC. These field tests are also
presented with the overall focus on the performance of the whole ORC unit and its power contribution to
the solar field. The tests have revealed that such low-temperature ORC unit can have adequate ef ficiency
and that its coupling with a solar field is feasible, increasing the power production of the whole system.
©2016 Elsevier Ltd. All rights reserved.
1. Introduction
The organic Rankine cycle (ORC) technology is suitable for heat
recovery applications of temperature even lower than 100/C14C[1].A t
such conditions its ef ficiency is rather low, usually in the range of
4e6%, but still there are cases where it can be cost effective,
especially for waste heat recovery. The main advantage at this
temperature range is the simple and low-cost heat source circuit,
since even liquid water can be used at low-pressure, while the use
of thermal oil is avoided. Moreover, a simple ORC con figuration can
be considered with a single expansion machine and no internal
heat exchangers [2,3] .The challenge is even bigger in small-scale systems with power
production of few kW. In such cases although the design is rather
simple to reduce costs, a very careful selection of each component is
crucial, in order to keep an adequate performance. Moreover, low-
temperature operation (below 100/C14C) brings some additional re-
strictions, since a limited number of organic working fluids can be
used for such purposes. The most important components in ORC
engines are: 1. the ORC pump, for which low attention is given most
of the times, 2. the expansion machine with intensive research
effort for producing/adapting expanders suitable for a wide power
range (of positive displacement type or even turbines for larger
systems), and 3. the evaporator, for which new correlations need to
be developed for designing a heat exchanger suitable to operate in
ORC conditions. When operating at low temperature, the
condenser also becomes an important component, since thermal
efficiency becomes highly sensitive to the temperature of the heat
rejection medium.*Corresponding author.
E-mail address: gkosmad@aua.gr (G. Kosmadakis).
URL:http://www.renewables.aua.gr/
1Second and third author have equal contribution.
Contents lists available at ScienceDirect
Energy
journal homepage: www.elsevier.com/locate/energy
http://dx.doi.org/10.1016/j.energy.2016.10.047
0360-5442/ ©2016 Elsevier Ltd. All rights reserved.Energy 117 (2016) 222 e236

A list of existing experimental ORC units is shown in Table 1 . This
list only includes similar set-ups working with a heat source of
70e100/C14C or an expander with inlet temperature under 100/C14C,
and a heat input from 10 to 100 kW or an expander output power
lower than 10 kW. Most of them are dedicated to solar applications,
but very few are finally tested when coupled with the solar field.
Another common application in that range is the waste heat re-
covery (WHR). The organic fluids R134a, R245fa and R123 are the
most common ones used. It is not easy and straightforward to
compare bench performances, since there is no standard de finition
for power and ef ficiency.
Some main highlights can be extracted from Table 1 . ORC unitsof that scale require low fluidflow rate and high pressure lift.
Reciprocating pumps, such as diaphragm pumps, are widely used
because they can handle such conditions, by providing constant
flow rate regardless the pressure variation. Multistage centrifugal
pumps are also used for such ORC units, but there are very few
available references.
For small-scale systems with power production lower than
around 20 kW, scroll expanders have been widely used and showed
adequate performance and expansion ef ficiency [20]. The present
authors have also used the same expansion technology (both open-
drive and hermetic ones) and revealed the good performance at a
wide range of pressure ratios [1]. This brings con fidence that such
Table 1
List of experimental small-scale low-temperature ORC (sorted by date).
Info. Heat source Cold source ORC fluid&conditions ORC components Performances
Reference Heater Cooler Fluid ( țLubricant) Pump ( țSubcooler) Exp. power
Target application Transfer fluid Transfer fluid Max eMin temperature Evap. eCondenser ( țIHE) hORC
Temp. ePower Temperature High eLow pressure Expander eOutput type hCarnot Cycle ePlant
[4] n/a Tap R134a țOil Diaphragm 3.3 kW (e)
n/a Water Water n/a en/a PHE ePHE n/a
70/C14Cen/a 5/C14C n/a en/a Scroll Hermetic eEl. AC n/a ea19%
[5] Gas Air Chiller N-Pentane Diaphragm 1.4 kW (e)
Solar Water Direct 81/C14Ce31/C14C PHE en/a 4.3% (e)
93/C14Ce34 kW 31/C14C 3.8 bar e1.1 bar Turbine Radial eEl.a14%ea17%
[6] Electric n/a R123 n/a 0.15 kW (m)
n/a Direct n/a 70/C14Cen/a n/a eShell-and-Tube 1.2% (m)
n/ae13 kW n/a 3.5 bar en/a Turbine Radial en/a n/a en/a
[7] Engine Diesel Tap R134a țOil Diaphragm țSubcooler n/a
Solar &WHR n/a Water n/a en/a PHE ePHE n/a
83/C14Ce80 kW 7/C14C n/a en/a Scroll Hermetic eEl. n/a ea21%
[8] Solar țGas n/a HFE7000; N-Pentane Diaphragm n/a
Solar &Gas n/a Water 67/C14Ce19/C14C PHE ePHE 7.6% (e)
70/C14Cen/a n/a 1.3 bar e1.2 bar Turbine Radial en/a 14% en/a
[9] Electric Sea R134a Diaphragm 2.05 kW (m)
Solar Water Water n/a en/a PHE ePHE 4% (m)
70/C14Ce100 kW 25/C14C 22b e9b Scroll Open-Drive eEl. n/a ea13%
[10] Solar n/a R134a țOil Diaphragm 1 kW (m)
Solar Water Water 76/C14Ce35/C14C PHE ePHE 1.5% (m)
76/C14Ce80 kW n/a n/a en/a Scroll Open-Drive eMech.a12%en/a
[11] Solar țElectric Cooling Tower R245fa Diaphragm 1.64 kW (m)
Solar Direct țWater Water 78/C14Ce14/C14C PHE en/aa5.8% (m)
n/ae38 kW 14/C14C 6.7 bar 1.9 bar Rolling-Piston eEl. DCa18%en/a
[12] Solar Cooling Tower R245fa n/aa1.2 kW (a)
Solar Glycol Water 90/C14Ce25/C14C Shell-and-Tube ePHEțIHE 9% (a)
90/C14Ce10 kW n/a 9.5 bar e1.4 bar Rotary Vane en/aa18%ea25%
[13] n/a Cooling Tower R123 Centrifugal Multistagea2.4 kW (a)
n/a Oil Water 97/C14Ce20/C14C PHE -PHE 7.1% (a)
n/aen/a n/a 6 bar e0.5 bar Turbine Radial eEl. DCa21%en/a
[14] Electric Cooling Tower R245fa țOil Plunger 1.38 kW (e)
WHR Water Water 86/C14Ce26/C14C PHE ePHE 7.8% (e)
92/C14Cea18 kW 25/C14C 8.5 bar e2.1 bar Scroll eEl. ACa17%ea18%
[15] Electric Tap R134a Plunger 3.7 kW (e)
Solar Oil Water 88/C14Ce26/C14C PHE ePHE 5.6% (en)
n/ae63 kW 10/C14C 33 bar e7 bar Scroll Hermetic eEl.a17%en/a
[16] Electric Tank R245fa Diaphragm 3.5 kW (e)
Solar Water Water n/a en/a PHE ePHE 7.2% (en)
95/C14C/C0110 kW 15/C14C n/a en/a Scroll Hermetic eEl. AC n/a ea22%
[17] Gas Tank R134a Diaphragm țSubcooler 5 kW (e)
WHR Water Water 88/C14Ce33/C14C PHE ePHE n/a
90/C14Cen/a 15/C14C 25 bar e9.5 bar Scroll Open-Drive eEl. ACa15%en/a
[18] Gas n/a R123 n/a 0.4 kW (a)
WHR Water Water 89/C14Cen/a Shell-and-Tube ePHE n/a
90/C14Cen/a n/a 6.1 bar e0.7 bar Rotary Vane eEl. DC n/a
[19] n/a n/a R245fa Rotative Vane 1.2 kW (a)
n/a Water Water 90/C14Cen/a PHE ePHE 9.3% (a)
95/C14Ce11 kW 20/C14C 10 bar e2.4 bar Scroll eEl. n/a ea20%
Expander power type: (e)lectric ¼Wel,exp ; (m)echanical ¼Wmech,exp ; (a)diabatic ¼m.dh exp.
hORC type: (en) electric net ¼(Wel,expeWel,pp)/Q; (e)lectric ¼Wel,exp /Q; (m)echanical ¼Wmech,exp /Q; (a)diabatic ¼m.dh exp/Q.
hCarnot: cycle ¼1eTout,cond /Tin,exp ; plant ¼1eTin,sink /Tin,hot source .
aCalculated by authors of this paper with data from the reference.G. Kosmadakis et al. / Energy 117 (2016) 222 e236 223

expander can be even used at a supercritical cycle, which is the next
step in this research. One positive aspect is that for low-
temperature applications, the pressure ratio is low and usually in
the range of 2 e4[2], enabling the scroll expander to operate with
good ef ficiency.
Concerning the evaporator, two types of heat exchangers (plate
and shell-and-tube heat exchangers) have been mainly investi-
gated for maximizing the net cycle ef ficiency of Organic Rankine
Cycles [21,22] . However, there is a lack of experimental data
regarding heat transfer in the evaporators, designed and suitable to
work in ORC conditions. The researchers focus most of the times on
optimization at system and component level, taking into consid-
eration all possible heat exchangers for such applications. In order
to predict the performance of the set-up at different operational
conditions, system and component models have been developed
[23]. Here, a helical coil design was selected as it could be produced
in a cost-effective way for this prototype unit (more compact than a
shell-and-tube heat exchanger) and easily integrated in the test set-
up. The performance of this component at subcritical state is re-
ported in this work.
All previous research activities are very important especially in
small-scale systems, in order to evaluate and compare the perfor-
mance of the ORC. Here, an experimental study is implemented,
testing a small-scale ORC with a net capacity of 3 kW. The tests are
first conducted in the laboratory with variable heat input and
temperature [24,25] for evaluating each key component and the
ORC as a whole. Then, the ORC has been coupled with concentrating
PV/thermal collectors. These collectors produce electricity from the
PV cells and heat, which is provided to the ORC. The field tests
during a winter and summer day are also presented and discussed.
2. The installed ORC and concentrating PV/T collectors2.1. The installed ORC engine at the laboratory
The developed ORC engine has been installed at the laboratory
for performance tests under controlled conditions. The heat input is
provided by an electric heater and its heat production can be
altered covering a large range of its capacity (from 25% of the total
heat capacity: 12 e48 kW
th) by operating different number of
electric resistances and switching on/off the heater, keeping con-
stant the hot water temperature. Different hot source temperatures
have been examined from 65/C14Cu pt o1 0 0/C14C with the heat transfer
fluid (HTF) being pressurized water at around 2.5 bar at maximum
temperature, and circulated with an inline centrifugal pump (Wilo
IPL 32/160) at constant speed of 2900 rpm.
A simpli fied design of the system installed at the laboratory is
depicted in Fig. 1 , together with the heating and cooling circuits.
Further details for the test-rig in the laboratory are provided in
Ref.[25].
An electric brake (heavy duty unit, manufactured by Bonitron),
is connected with the frequency inverter of the expander's induc-
tion motor, in order to control the test conditions and evaluate the
performance of this expansion machine [25]. The engine cooling
during the laboratory tests is accomplished with a cooling water
circuit, using a conventional shell-and-tube heat exchanger. Cold
water with a temperature of around 16/C14C is circulated and drawn
from a large water reservoir with capacity of 320 m3, rejecting the
heat of the ORC engine (no cold water temperature increase was
noticed during the whole testing period, keeping the heat rejection
conditions constant). The condensation temperature of the organic
fluid with this method is around 25/C14C(fluctuating according to the
engine load).
The ORC pump is a triplex diaphragm pump manufactured by
Hydra Cell (model G-10X), coupled with a 3 kW induction motor(Valiadis K132S, 6-pole, 86.4% nominal ef ficiency), driven by a 4 kW
frequency inverter (Siemens SED2-4/32B).
The organic fluid selected is R-404A after screening many po-
tential fluids using environmental (zero Ozone Depletion Potential
eODP), cost and ef ficiency criteria [25,26] . Although R-404a has a
moderate Global Warming Potential (GWP), there is already a
replacement fluid (R-407f) with similar properties but still high
cost. Moreover, one of the biggest challenges was the modi fication
of a commercial scroll compressor (Copeland ZP137KCE-TFD with
swept volume of 127.15 cm3/rev., maximum isentropic ef ficiency
75.2%, and built-in volume ratio of around 2.8 at compressor mode)
to operate as scroll expander (in reverse operation). A new casing
had to be made, while many internal parts have been re-designed
for better matching its operation as expander (such as the inlet
volume before the fluid enters the steady scroll) [27]. The two scroll
geometries have been kept the same (same built-in volume ratio as
the original compressor).
An evaporator of a helical coil type with capacity of 41 kW has
been developed for this application as well ( Fig. 2 ). The shell of the
helical coil heat exchanger is formed by two concentric cylinders in
which a metal coil tube is fitted. The flow paths are arranged in
counter flow with the hot water flowing downwards in the shell
and the working fluid R-404A circulating upwards in the coil. The
heat transfer between both fluids takes place across the coil wall.
This evaporator is a key component in the installation, because it isthe link between the ORC engine and the heat source. Such helical
coil heat exchanger is designed and built speci fically for such ORC
installation, suitable to operate at relatively high pressure and
temperature (capable for both sub- and supercritical working
conditions) [28].
The ORC engine was installed in the laboratory and it is depicted
inFig. 2 . All components have been mounted on the same structure
and an electric panel is included. Further details are provided in
Ref.[25].
2.2. The installed CPV/T collectors with ORC
The solar field has been installed at the AUA campus (in Athens,
Greece). The field has been prepared (cleared and leveled) and the
10 collectors have been installed on concrete foundation. Each
Fig. 1. ORC design (laboratory con figuration).G. Kosmadakis et al. / Energy 117 (2016) 222 e236 224

collector has electric capacity of 1 kWp, concentration ratio of
around 10, and heat production of 4.1 kW th. They have been
adapted to operate at temperature up to 95/C14C[24]. After finalizing
the laboratory tests, the ORC engine has been moved to the field
and connected with the piping circuit of the solar collectors. The
solar field and the small housing, where the ORC is placed with all
control and electric panels, are shown in Fig. 3 .
The condenser of the ORC has been replaced with an evaporative
condenser, for keeping the condenser pressure as low as possible,
due to the pump inlet pressure limit of 17 bar. This condenser is also
depicted in Fig. 3 . An air-chiller is included in this set-up for
dissipating the produced heat from the collectors in case the ORC is
not operating (appropriate by-pass pipes and valves have been
installed).
The main design of the combined system is shown in Fig. 4 .A
back-up electric heater is also mounted, in order to operate and test
the ORC even if there is not adequate solar irradiation and heat
production from the collectors.2.3. Test data and processing
The location of the measuremen t instruments is depicted at
the three circuits in Fig. 1 (hot water circuit, ORC engine, cold
water circuit), in order to measure the key properties and eval-
uate the performance of this engine at laboratory conditions.
These instruments are mainly temperature and pressure sensors,
in order to calculate the thermodynamic state of the organic fluid
and hot/cold water at each location. The temperature sensors are
of Pt100 type (accuracy up to ±0.2/C14C), while the pressure
transmitters are manufactured by Keller (type 21Y with mea-surement error up to 1% of the full scale). With the above un-
certainties, the thermodynamic properties are calculated with an
accuracy of around 1.2%. Flow meters are not used, since steady-
state conditions are examined, once the engine has reached a
balanced operation at each case. The power production is
measured directly from the expander frequency inverter. The
heat input is calculated from the ORC side, since the organic fluid
pump is of diaphragm type and has a linear correlation of flow
rate with speed with a constant parameter of 0.0205 (L/min)/rpm,
which provides a very reliable calculation of the volume flow rate
(accuracy estimated at 2%). The mass flow rate is then calculated
with the measured temperature and pressure of the fluid at the
pump outlet using EES/REFPROP database for R-404A [29].T h e
pump shaft speed and electric consumption are measured from
the inverter.
The accuracy of the calculated parameters is given by the
following expression [29]:f
t;tot¼ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi
P
i/C18
vt
vXi/C192
f2
t;is
.I nTable 2 is given
the relative measurement error for each main parameter (mean
value), showing that it is low and would not in fluence the relative
differences of the results.
The model of pump power balance proposed in Ref. [30] pro-
vides a very detailed understanding of losses and can suggest so-
lutions for further improvements. This model is visualized in Fig. 5 ,
including the motor and a variable speed drive (VSD).
The model parameters C 2,C3,C4&C5are provided by pump
and motor speci fications. The Parameter C 1is estimated from
experiment, minimizing the error-objective function F¼
Pð_Wel;meas/C0 _Wel;estiȚ2, with electric power consumption given
by: _Wel¼C1ț _WmechțC2:_W2
mechțC3:_U2and hydraulic power
by: _Wmech¼C4:_UțC5:_Whyd.
Fig. 2. ORC installed in the laboratory.
Fig. 3. Solar field and ORC in the housing at the AUA campus.G. Kosmadakis et al. / Energy 117 (2016) 222 e236 225

3. Results and discussion
First, the laboratory tests are presented, leading to the validation
of each main component and the whole ORC unit. Then, field tests
are presented during winter and summer days.
3.1. Laboratory tests
Most of the tests in the laboratory have been conducted at
subcritical operation. The hot water temperature varied from 65/C14C
up to 100/C14C. The operating conditions that have been examined
concern the variation of the ORC pump and expander speeds, which
in turn affect the cycle and fluid properties. The ORC pump fre-
quency is altered from 10 Hz up to 50 Hz (192 e960 rpm) and theexpander frequency is regulated from 10 Hz up to 45 Hz
(580e2610 rpm). By regulating the pump speed, the heat input is
varied up to almost 55 kW th, being almost linear, as presented with
detail in Ref. [25]. In this way the heat input to the ORC can be
controlled effectively, while at the same time the flow rate of the
organic fluid is adjusted. The regulation of the organic fluid pump
speed has an important effect on high pressure as well, especially
when the expander speed is kept constant, as it will be shown later.
3.1.1. ORC pump
The HTF temperature does not have any effect on the pump
performance and therefore this parameter is not included in the
pump analysis. The main parameters considered here are pressure
difference and pump speed. The ORC pump has been investigated
over a large range of pressure differences (in the range of
5.8e17.3 bar) and speeds (in the range of 384 e864 rpm). The global
efficiency of the pump is given by Eq. (1), while the hydraulic power
by Eq. (2).
hg;pp¼ _Whyd ;pp._Wel;pp (1)
_Whyd ;pp¼_V:/C0Ppp;out/C0Ppp;in/C1(2)
Fig. 4. Combined system (CPV/T with ORC) installed at the AUA campus.
Table 2
Accuracy of calculated parameters.
Parameter Relative error (%) Range (Lab) Range (Field)
Heat input to ORC 2.62 12 e48 kW th25e40 kW th
Expander power production 2.62 0.5 e3 kWe 0.5 e2k W e
Pressure ratio 1.40 1.7 e2.6 1.4 e2
Expansion ef ficiency 2.66 20 e85% 65 e75%
Thermal ef ficiency 3.71 0 e4.2% 1 e4%
Fig. 5. Energetic chain of the pumping system with motor and VSD.G. Kosmadakis et al. / Energy 117 (2016) 222 e236 226

The model of pump power balance has been then applied. The
deviation between the measured electricity consumption ð_Wel;measȚ
and estimated one ð_Wel;estiȚis less than 10%, as shown in Fig. 6 ,
while all parameters are provided in Table 3 .
The maximum pressure difference imposed by the ORC at
design conditions is 35 bar. According to pump manufacturer data,
the maximum power required by the pump is 1.486 kW (35 bar,
960 rpm). Therefore, the motor rated power is two times the power
requested for ORC maximum conditions. This oversize has been
implemented, in order to provide the pump with adequate torque
during starting, but has a negative impact on the ef ficiency, since
the motor always operates below its nominal power by at least 50%.
The pump global ef ficiency as a function of hydraulic power is
presented in Fig. 7 . The maximum pump global ef ficiency reaches32%, for measured hydraulic power of 0.475 kW.
Most of the losses are located in the driving part of the pumping
system (inverter and motor) and represent 60 e80% of the electric
consumption. The static losses have been calculated to 883 W, being
as analogous to the value of 240 W obtained in Ref. [30] in a similar
configuration, considering that the motor nominal power is two
times higher and the inverter is not included in the motor (separate
part). The induction motor operates between 5 and 20% of its nom-
inal power (equal to 3 kW, which has been deliberately oversized to
have adequate power during supercritical operation) with ef ficiency
much lower than its nominal one. The global pump ef ficiency could
be higher in case of smaller motor sizing [31], but allows a smooth
starting and adequate power during transient operation (important
when the ORC is supplied with solar thermal energy).
Fig. 6. Estimated vs. measured pump electric consumption.
Table 3
Parameters of the pump power balance model.
Equation Parameter Unit Value
Motor &Variable Speed Drive (VSD) _Wel¼C1ț _WmechțC2:_W2
mechțC3:_U2 C1 W 883
C2 W/C013.67/C210/C05
C3 W/rpm21.54/C210/C04
Pump _Wmech¼C4:_UțC5:_WhydC4 W/rpm 0.178
C5 e 1.174
Fig. 7. Pump global ef ficiency as a function of hydraulic pumping power.G. Kosmadakis et al. / Energy 117 (2016) 222 e236 227

3.1.2. Evaporator
The heat exchanger was designed for a heat transfer capacity of
41 kW, inlet temperature of the heat transfer fluid of 95/C14C and
mass flow rates of the HTF and the organic fluid of 2.5 kg/s and
0.25 kg/s, respectively. Using the LMTD method, the design has
been concluded. More details on the design procedure and the
geometry can be found in Ref. [28].
The heat transfer is calculated from the measured data, using
Eq.(3).
Q¼ _mORC/C0hevap;out/C0hevap;in/C1(3)
where Qis the heat transferred, _mORCis the organic fluid mass flow
rate, hevap,in and hevap,out are the enthalpies at the inlet and at the
outlet of the evaporator.
Fig. 8 shows the heat transfer as function of the mass flow rate of
the working fluid for different heat transfer fluid inlet tempera-
tures. The heat transfer increases by increasing the mass flow rate
in the coil, while keeping the mass flow rate constant at the annular
(shell) side. A maximum value of the heat transfer of 55 kW this
achieved at higher flow rates. The highest mass flow rate of
_mORC¼0.34 kg/s was reached at relatively low heating fluid tem-
perature of 65/C14C.
A comparison of the nominal designed value with the mea-
surements was also possible. At the nominal mass flow rate of the
organic fluid _mORC¼0.25 kg/s and HTF temperature of 95/C14C a heat
transfer rate of 46 kW thwas achieved, which is higher than the
designed one. The main reason for this difference is that the
evaporator was designed for supercritical operating conditions
using supercritical heat transfer correlations. Details about the heat
transfer correlations that have been used for designing the heat
exchanger are included in Ref. [28], while the development of new
more accurate correlations is currently under investigation.
The pinch point temperature difference between the heating
and organic fluid in the heat exchanger was then calculated. Since
the temperature was measured only at the inlet and outlet of the
heat exchanger, the thermal match pro file was determined from a
simple model developed in EES environment [29]. This thermal
match for the three inlet measured temperatures for HTF temper-ature of 65
/C14C, 80/C14C and 95/C14C is examined next.
Both the mass flow rate of the organic fluid and the inlet tem-
perature at hot and cold side have a great in fluence on the pinchpoint temperature difference. For all measurements the mass flow
rate of the heat transfer fluid and the organic fluid R-404A was kept
constant at 2.7 kg/s and 0.3 kg/s respectively. It is clear from Fig. 8
that an improved thermal match is obtained at higher mass flow
rate of the organic fluid, approaching the design condition.
The inlet temperature of the organic fluid depends on its mass
flow rate and the inlet temperature of the heating fluid. Hence, for
inlet HTF temperature of 65/C14C the organic fluid's inlet temperature
is 26/C14C. A minimum temperature difference of 13/C14C is reached
before the evaporation starts along the coil because the tempera-
ture difference of both fluids at the inlet of the heat exchanger is
lower compared to the inlet temperatures of the HTF of 80/C14C and
95/C14C, as presented in Fig. 9 .
At higher temperatures at the hot side, for example at 95/C14C, the
temperature difference between both fluids and before the evap-
oration in the coil occurs is up to 36/C14C. This indicates that the
exergy losses at such operating conditions are higher, due to the
higher average temperature difference between the HTF and R-
404A.
By analysing the results from the measurements the pinch point
temperature difference at the outlet of the evaporator is below10
/C14C, which is a lower value than the designed one. Hence, an
improved thermal match at the outlet of the heat exchanger and
pinch point temperature difference of only 2/C14C is realized for all
measurements due to the high superheating of the organic fluid.
3.1.3. Scroll expander
The expansion machine is directly coupled with a three-phase
asynchronous motor/generator (capacity of around 10 kW) inside
the hermetic casing. This motor operates up to around one third ofits nominal power, avoiding overheating, due to the absence of
cooling (in compressor mode it is cooled by the organic fluid
eusually a refrigerant eitself, which enters at low temperature).
The variation of both the pump and expander speed has a strong
effect on electricity production and expansion ef ficiency. The po-
wer production is depicted in Fig. 10 for different pump speeds and
HTF temperature as a function of pressure ratio. It should be noted
that for constant pump speed, the pressure ratio is adjusted
through the variation of expander speed, covering the whole
expander speed range for high pump speed and a lower one (from
10 up to 25 e30 Hz) for low pump speed.
The maximum power production is equal to 3.2 kW, and is
observed for moderate pump speed (pump frequency equal to
35 Hz) and pressure ratio (equal to 1.95) for HTF temperature of
95
/C14C. But even for the moderate HTF temperature, the ORC engine
shows high power production for a pressure ratio of around 2, due
to the increased expansion ef ficiency at such conditions.
The maximum expansion ef ficiency is observed at a moderate
expander speed/pressure ratio, due to the low electrical ef ficiency
of the asynchronous generator at low speeds. In Fig. 11 is shown the
expansion ef ficiency as a function of pressure ratio for the three
different HTF temperatures. The expansion ef ficiency is calculated
with Eq. (4)and includes all possible losses (electrical, friction, heat
transfer, etc.) and provides a reliable evaluation parameter of all
types of positive displacement expansion machines.
nex¼Wmeas
_mORC/C0hexp ;in/C0hexp ;out ;is/C1 (4)
where W meas is the measured electricity production, _mORCis the
organic fluid mass flow rate, h exp,in is the enthalpy of the organic
fluid at the expander inlet, and h exp,out is the enthalpy of the organic
fluid at the expander outlet for isentropic expansion.
The maximum expansion ef ficiency is equal to 85% and is
observed for a pressure ratio of 2 (for pump frequency of 35 Hz),
Fig. 8. Heat transfer as function of the mass flow rate of the organic fluid for different
HTF temperature.G. Kosmadakis et al. / Energy 117 (2016) 222 e236 228

which is highly relevant to the built-in volume ratio of the original
compressor [32]. This optimum pressure ratio value is lower than in
most of the ORC units (common pressure ratio values for maximum
expansion ef ficiency are around 3 e4[33]), since the original
compressor is intended for air-conditioning applications with R-410a, where the pressure ratios are not very high (in the range of
2.5e3). For lower pump speeds and pressure ratios, the expansion
efficiency is decreased. Moreover, as HTF temperature decreases,
the maximum expander ef ficiency is observed for even lower
pressure ratio (reaching even 1.65 for the 65
/C14C case with an ef fi-
ciency of up to 60%). Finally, the expansion ef ficiency is, in general,
higher for increased HTF temperatures, since at such conditions the
volume ratio increases and approaches the built-in volume ratio of
the expander, limiting under-expansion. A small improvement of
volumetric ef ficiency can be also observed, which also aids the
expansion ef ficiency increase.
The peak expansion ef ficiency is high and fully justi fies the ef-
forts for replacing the casing and optimizing some expander pa-
rameters, although for most of the operating conditions this
efficiency is in the range of 45 e70% and similar to other reported
values for scroll expanders of similar scale and temperature
[9,14,20,34] . Also, the selection of the ZP series of expansion ma-
chine with low built-in volume ratio (about 2.8) seems to be ideal
for the speci fic application [35,36] , since the pressure ratio is
restricted to values around 2 for the R-404A fluid.
3.1.4. ORC engine eenergy analysis
The thermal ef ficiency as a function of the pressure ratio is
depicted in Fig. 12 . The thermal ef ficiency is expressed as the net
power output (power produced minus the ORC pumping work)
divided by the heat input [37]. A maximum of 4.2% thermal ef fi-
ciency is reached for a pressure ratio of 1.95. At constant pressure
ratio, ORC thermal ef ficiency increases with the expander speed,
but a minor improvement is noticed above 30 Hz. At constantexpander speed, the thermal ef ficiency increases with the pressure
ratio, which is directly related to the expander ef ficiency shown
previously, with an optimum pressure ratio of around 2. However,
such pressure ratio is only achieved for low expander speed. For
high expander speed, the pressure ratio is limited by the pumpmaximum flow rate. Thermal ef ficiency of 5 e6% could be expected
for higher expander speeds, due to the decreased electrical losses in
the expander induction generator.
3.1.5. ORC engine eexergy analysis
Exergetic ef ficiency is recommended for evaluation of low-
temperature power systems and comparison of their performance
for different heat source temperature [38,39] . Exergetic ef ficiency
expression depends on the heat source type and the application
[40]. Solar applications can be considered as a sealed source type,
and the exergy supplied to the ORC is determined by Eq. (5), with T
0
the sink water inlet temperature in Kelvin.
DEsup¼ _mHTF/C2hHTF ;in/C0hHTF ;out/C0T0/C0sHTF ;in/C0sHTF ;out/C1/C3(5)
The ORC exergetic ef ficiency is then de fined as the net power
output divided by the exergy supplied by the heat source:
Wnet=DEsup. The exergetic ef ficiency as a function of relative pres-
sure for constant HTF temperature of 95/C14C is depicted in Fig.13 . The
maximum exergetic ef ficiency increases with pump frequency and
the optimum pressure differs with pump frequency. However, a
reduced cycle ef ficiency at higher pressure is induced by the low
expander speed (as a result of low expander ef ficiency). Consid-
ering only data with expander frequency above 25 Hz, the exergetic
efficiency constantly increases with the relative pressure.
For each heat source temperature (65, 80, 95 and 100/C14C), the
maximum exergetic ef ficiency points are presented and linked with
the pressure-temperature map shown in Fig. 14 , providing the
exergetic ef ficiency value. This map can be used to conclude to the
Fig. 9. Pinch point temperature difference for HTF temperature of 65, 80 and 95/C14C.G. Kosmadakis et al. / Energy 117 (2016) 222 e236 229

optimum pressure of this speci fic set-up for different HTF tem-
perature level according to the calculated exergetic ef ficiency. It can
be observed that the optimum pressure is increasing with
maximum temperature, but the optimum exergetic ef ficiency
seems to be limited below 20%.
Exergy analysis can be useful for each component of the ORC,
concluding to an exergy flow diagram. Fig. 15 shows such diagram
of: (a) the optimum subcritical point, and (b) transcritical point for
similar supplied exergy of 9.1 kW (heat input: 43 kW that 95/C14C).
The supplied exergy is used as a base index 100 for this figure.
Most of the exergy destruction occurs in the two heat ex-
changers (evaporator and condenser). The exergy loss in the
evaporator occurs because of the high average temperature differ-
ence between the HTF and the organic fluid. In transcritical oper-
ation, the pressure is increased and the mean temperature
difference is reduced. Therefore, exergy destruction in the evapo-
rator is 30% lower for transcritical operation (see Annex ). In the
condenser, there is no valorization of the heat rejected. Even if
superheating at the expander outlet is lower for transcritical case,
condenser exergy destruction is 40% higher because the conden-
sation pressure is 2.5 bar higher. Lower condensation pressure
could be achieved with a better design and control of thecondensation process. The use of an internal heat exchanger (IHE)
for recovering the de-superheating power could save 10 e30% of the
heat input and reduce exergy destruction both in the condenser
and evaporator.
Exergy destruction in the expander is much higher in the
transcritical case, mainly because the expander frequency is very
low/C015 Hz in average ein order to reach supercritical operation.
Such operating conditions resulted in low expander ef ficiency as
well, mainly due to the low electrical ef ficiency of the expander's
generator.
At supercritical conditions the pump exergy destruction slightly
increases by 15%, while the provided hydraulic power is 60% higher
than in the subcritical case. This is because of the high static losses
in the pump power balance presented in Section 3.1.1.
3.2. Combined solar field with ORC engine
Next, the ORC engine has been moved to the field and connected
with the CPV/T solar collectors. Tests have been conducted during
winter and summer days, in order to examine and evaluate the
combined system at different weather conditions. The main test
results for both winter and summer days are presented next.Fig. 10. Power production as a function of the pressure ratio for variable pump speed (HTF temperatures of 65, 80 and 95/C14C).G. Kosmadakis et al. / Energy 117 (2016) 222 e236 230

Fig. 11. Expansion ef ficiency as a function of the pressure ratio for variable pump speed (HTF temperatures of 65, 80 and 95/C14C).
Fig. 12. ORC Thermal ef ficiency function of pressure ratio with expander iso-speed fitting.G. Kosmadakis et al. / Energy 117 (2016) 222 e236 231

3.2.1. Winter tests
Thefirst performance tests of the combined system concern the
operation for almost steady state conditions during a sunny winter
day. The HTF pump speed is decreased (pump frequency of 30 Hz),
in order to reach higher temperatures, due to the moderate solar
irradiation.
InFig. 16 is shown the heat produced by the collectors, the heat
provided to the ORC, and the HTF temperature. Until around 12:15
the system load was increasing, due to the sunshine and the heat
produced was up to 40 kW (heat produced by collectors andabsorbed by ORC is almost equal with minor differences). The heat
provided to the ORC then continues to increase due to the high
thermal inertia of the evaporator, while the heat produced by the
collectors is decreased, due to the high HTF temperatures and
higher thermal losses from the collectors. Nevertheless, during the
whole test duration the operation is more or less stable with few
fluctuations, which makes it possible to evaluate the combined
system.
The HTF temperature is increased slightly over 70/C14C and the
HTF temperature difference (inlet/outlet difference) is around 5 K.
Fig. 13. ORC exergetic ef ficiency function of relative pressure for constant HTF temperature of 95/C14C.
Fig. 14. Efficiency in the expander inlet relative pressure and temperature map.G. Kosmadakis et al. / Energy 117 (2016) 222 e236 232

For the whole test duration the HTF temperature varies from 60/C14C
up to 72/C14C.
As the HTF temperature increases, so does the organic fluid
temperature and pressure. The pressure at the expander inlet in-
creases during this test period (see left Fig. 17 ), whereas the
expander pressure ratio, decreases, due to the increased coolingload of the condenser (the condensation pressure is increased). In
the same figure the performance of the ORC engine (expander and
cycle thermal ef ficiency) at such almost stable conditions is also
presented, together with the electric power production from the
ORC and the PV cells.
The ef ficiency of the expander reaches even 74%, while it holds
high values for the whole test duration (in the range of 65 e70%).
Similar conclusions have also been drawn during the laboratory
tests, as presented in the previous sections. The high expander ef-
ficiency brings an increased thermal ef ficiency as well, since themaximum reached is almost 4.6%. During the end of the testing
period, the thermal ef ficiency drops, due to the increased
condenser pressure.
Moreover, since solar irradiation was almost constant during this
period, the PV production is almost constant as well and equal to
around 5 kW. The ORC engine production is also stable and around
1.9 kW, while during the end of these tests it starts to decrease, as
already mentioned. The net power production of the ORC is lower
due to the pump consumption, which is almost 1 kW. The ORC en-
gine net production is around 20% of the PV production, which in-
creases the productivity of the combined system at such conditions.
3.2.2. Summer tests
The combined system has been also tested during different
summer days. In this section the test results during one represen-
tative summer day are presented.
The heat produced by the collectors, the heat provided to the
ORC, and the HTF temperature values are shown in Fig. 18 . There is
high fluctuation of the heat produced and absorbed by the ORC,
since the system operates at unsteady conditions.
The heat provided to the ORC is mostly in the range of
20e40 kW
th, while the HTF temperature reaches even 80/C14C. After
12:30 it is reduced due to few clouds that did not allow it to further
increase heat production and temperature.
The expander inlet pressure and pressure ratio are shown in
Fig. 19 . The pressure level is now much higher than during the
winter day, due to the higher condenser pressure, which shifts the
whole cycle to higher pressure. This brings a lower pressure ratio(around 1.6), which is not bene ficial for the expander, as also shown
inFig. 19 (right).
The ef ficiency of the expander reaches 65% and then it is
reduced to even 30%, leading to low thermal ef ficiency in the range
of 1e3.5%. An encouraging aspect is that even during transient
operation the ORC engine can maintain an adequate performance,
showing that it is suitable for such conditions. Moreover, PV pro-
duction is much lower than in the winter day, due to the high
ambient temperature, which also has a negative effect on ORC
production (due to high condenser pressure), resulting to low po-
wer production and much lower than in the winter day.
Fig. 15. Exergy flowecomparison between optimum subcritical and transcritical cycle.
Fig. 16. Heat produced by the collectors and absorbed by the ORC and HTF tempera-
ture during the winter day.G. Kosmadakis et al. / Energy 117 (2016) 222 e236 233

4. Conclusions
The detailed experimental results of the ORC engine testing at
the laboratory have been presented, revealing its performance
capability and evaluating its main components (pump, evaporator,
and expander). The tests concern a large range of HTF temperature
with variable heat input. Various parameters have been examined,
mainly when regulating the expander and pump speed, showing
the heat-to-power conversion ef ficiency of such engine. A detailed
energy and exergy analysis is presented, focusing on the main
components.
The pump global ef ficiency is low, as expected, while the
expander ef ficiency can reach high values at a small range of
operating conditions. Also, the evaporator outperforms in terms of
the heat transferred and the pinch point temperature difference at
the outlet of the heat exchanger, showing that more accurate cor-
relations need to be developed, for even more detailed designs and
use of less material (leading to lower costs). The most important
conclusions from the laboratory tests is that such ORC engine with
capacity of just 3 kW can reach an adequate thermal ef ficiency,
when operating at very low temperature.Fig. 17. Expander inlet pressure and pressure ratio of ORC engine (left), and expansion ef ficiency of the scroll expander, thermal ef ficiency of the ORC engine and power production
(right) during the winter day.
Fig. 18. Heat produced by the collectors and absorbed by the ORC and HTF tempera-
ture during the summer day.
Fig. 19. Expander inlet pressure and pressure ratio of ORC engine (left), and expansion ef ficiency of the scroll expander, thermal ef ficiency of the ORC engine and power production
(right) during the summer day.G. Kosmadakis et al. / Energy 117 (2016) 222 e236 234

The ORC engine has been then coupled with concentrating PV/T
collectors, which provide the heat produced to the ORC. Various
tests have been conducted during both winter and summer day, to
demonstrate the performance of the combined system. The ORC
engine could operate with good performance even at unsteady
conditions, increasing the productivity of such hybrid system dur-
ing both winter and summer day.
In the present study experimental test data are presented and
evaluated, and the advantages and the potential of such tech-
nology could be identi fied. Also, some first proof is provided,
whether such combined system can increase power production,
by utilizing the available low-temperature heat. Nevertheless, a
detailed techno-economic study is required, in order to identify
the advantages of the presented system in comparison with other
ORC units coupled with solar thermal collectors and operating
under similar conditions (mostly relevant to HTF temperature),
since for the same collectors' area the system produces more
electricity (from both the PV cells and the ORC), but its cost is also
higher.
Acknowledgements
The research leading to these results has received funding from
the European Union's Seventh Framework Programme managed by
REA-Research Executive Agency, http://ec.europa.eu/research/rea
([FP7/2007-2013] [FP7/2007-2011]) under grant agreement n/C14
315049 [CPV/RANKINE], FP7-SME-2012. The AUA and UGENTresearch teams would also like to thank their partners for theirwork within this project.
A part of this work was supported by the French Environment
and Energy Management Agency (ADEME), the French Alternative
Energies and Atomic Energy Commission (CEA) and the KIC
InnoEnergy.
A part of the results presented in this paper have been ob-
tained within the frame of the IWT SBO-110006 project The Next
Generation Organic Rankine Cycles ( www.orcnext.be ), funded by
the Institute for the Promotion and Innovation by Science and
Technology in Flanders. This financial support is gratefully
acknowledged.Annex
Exergy destruction rate in the evaporator is de fined as the
exergy destruction divided by the exergy supplied. There is a strong
correlation with the evaporation pressure and the maximum
temperature (see Fig. A-1 ). When pressure is increasing, exergy
destruction rate decreases, because the mean temperature differ-
ence between HTF and working fluid is decreased. But when the
heat source fluid temperature increases, and so does the maximum
cycle temperature, the destruction rate in the evaporator is also
increased. This is because the mean HTF temperature increases
faster than the working fluid mean temperature.
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Nomenclature
A:total heat transfer area (m2)
C1:pump parameter: VSD static losses (W)
C2:pump parameter: Motor part-load losses (W/C01)
C3:pump parameter: Motor part-speed losses (W/rpm2)
C4:pump parameter: Pump rotational friction (W/rpm)
C5:pump parameter: Pumping ef ficiency ( /C0)
E:exergy (kW)
h:speci fic enthalpy (J/kg)
m:mass flow rate (kg/s)
P:Pressure (bar)
Q:heat transfer (kW th)
s:speci fic entropy (J/kg K)
T:Temperature (/C14C)
V:volume flow rate (m3/h)
W:Power (kW)
h:efficiency
U:rotational speed (rpm)
Subscript
cd:condenser
el:electric (power)
esti: estimated (model)
evap: evaporator
exp: expander
hyd: hydraulic (power)
is:isentropic
in:inlet
meas: measured
mech: mechanical (power)
0:heat sink
orc:organic fluid
out: outlet
pp:pump
sup: supply
th:thermalG. Kosmadakis et al. / Energy 117 (2016) 222 e236 236