Design Considerations and Experimental Results of a 60 W
6
Design Considerations and Experimental Results of a 60 W Compressed-Air-to-Electric-Power System D. Krähenbühl 1 , C Zwyssig 1 , H. Hörler 2 and J. W. Kolar 1 1 Power Electronic Systems Laboratory, ETH Zurich, Switzerland 2 Aerothermochemistry and Combustion Systems Laboratory, ETH Zurich, Switzerland Abstract—In many process applications, where a pressure reduction is required the energy ends up being dissipated as heat. Examples are throttling valves of gas pipelines and automotive engines or turbo expanders as used in cryogenic plants. With a new pressure reduction system that produces electricity while expanding the gas, this lost energy can be recovered. To achieve a high power density this energy generation system requires an increased operating speed of the electrical machine and the turbomachinery. This paper proposes a miniature compressed- air-to-electric-power system, based on a single-stage axial impulse turbine with a rated rotational speed of 350 000 rpm and a rated electric power output of 60 W. A comprehensive description including turbine and permanent magnet (PM) generator is given and measurements like maximum electric output power of 124 W and maximum system efficiency of 24 % are presented. I. INTRODUCTION n pressure reduction devices, such as valves, conventional throttles or turbo expanders, the excess process energy is usually wasted as heat. However, this energy could be recovered by employing a system that removes the energy from pressurized gas flow and converts it into electrical energy. One example is the replacement of the conventional throttle in automotive applications where a turbine in combination with a generator can actively throttle the intake air and thereby produce electrical power [1]. Measurements at constant speed have shown that up to 700 W electric power could be produced (turbine Ø 40 mm) and an extrapolation with a 50 % downsized turbine predicts that even more electric power could be produced. While it is necessary to transport natural gas at high pressures, end-users require gas delivery at only a fraction of the main pipeline pressure. Therefore, energy can be recovered at pressure reduction stations if throttling valves are replaced by expanders driving electrical generators [2]. For power recovery, turbines are generally rated from 150 kW to 2.5 MW, however, the pressure reduction process is usually done in several stages, and an array of small turbine-generator modules could replace one large pressure reduction valve [3]. Also, the turbo expanders used today in cryogenic plants transfer the excess power (in the kW range) to a brake compressor where the energy is finally dissipated into cooling water. If a generator would be employed for the braking of the turbo expander, energy could be recovered, and therefore the efficiency of such plants could be increased [4]. Several of the abovementioned applications, e.g. in automobiles, need ultra-compact power generation systems. Fig. 1. Picture of the ultra compact (22 mm x 60 mm) air-to-power demonstrator. Power density in both turbomachinery and electrical machines increases with increasing rotational speed [5], [6]. Therefore, for highest power density, these systems are operating at speeds between 100 000 rpm and 1 Mrpm at power levels of up to several kilowatts. Besides higher power applications, micro-turbines with less than 100 W power output and very high speeds have been reported in literature. In [7], a modular system consisting of an off-the-shelf air turbine from a dental drill, a permanent-magnet (PM) generator and a rectifier has been realized, with a maximal power output of 1.11 W and a maximal speed of 200 000 rpm. Drawbacks of this system are the poor power density (0.02 W/cm 3 ) and the large inlet flow rate of 45 l/min at maximum power output. The power electronics of the system consists of a three-phase transformer a diode bridge rectifier and a 5 V linear regulator. In [8], a single-stage axial micro gas turbine coupled to a commercial electrical machine with a maximal electric power output of 16 W at 160 000 rpm and 1 bar supply pressure has been introduced for later use in a gas turbine system. The maximal torque and mechanical power generated from the turbine is 3.7 mNm, respectively 28 W. The generator is connected to a variable three phase resistive load to measure the electric output power. The total system has a maximum efficiency of 10.5 % at 100 000 rpm and achieves a power destiny of 1.6 W/cm 3 , excluding power and control electronics. In [9], a PM generator, capable of supplying 8 W of dc power to a resistive load at a rotational speed of 305 000 rpm is shown. The stator uses interleaved, electroplated copper windings on a magnetically soft substrate. The rotor consists of an 8-pole SmCo PM, back iron and a titanium sleeve, because of the high centrifugal forces. The machine was characterized using an air-driven spindle. To provide a dc voltage, the ac generator voltages were first stepped up using a three-phase transformer and then converted to dc using a I
Design Considerations and Experimental Results of a 60 W
Microsoft Word - krähenbühl_MESA08_FINAL.docD. Krähenbühl1, C
Zwyssig1, H. Hörler2 and J. W. Kolar1
1Power Electronic Systems Laboratory, ETH Zurich, Switzerland
2Aerothermochemistry and Combustion Systems Laboratory, ETH Zurich,
Switzerland
Abstract—In many process applications, where a pressure
reduction is required the energy ends up being dissipated as heat.
Examples are throttling valves of gas pipelines and automotive
engines or turbo expanders as used in cryogenic plants. With a new
pressure reduction system that produces electricity while expanding
the gas, this lost energy can be recovered. To achieve a high power
density this energy generation system requires an increased
operating speed of the electrical machine and the turbomachinery.
This paper proposes a miniature compressed- air-to-electric-power
system, based on a single-stage axial impulse turbine with a rated
rotational speed of 350 000 rpm and a rated electric power output
of 60 W. A comprehensive description including turbine and
permanent magnet (PM) generator is given and measurements like
maximum electric output power of 124 W and maximum system
efficiency of 24 % are presented.
I. INTRODUCTION n pressure reduction devices, such as valves,
conventional throttles or turbo expanders, the excess process
energy is
usually wasted as heat. However, this energy could be recovered by
employing a system that removes the energy from pressurized gas
flow and converts it into electrical energy. One example is the
replacement of the conventional throttle in automotive applications
where a turbine in combination with a generator can actively
throttle the intake air and thereby produce electrical power [1].
Measurements at constant speed have shown that up to 700 W electric
power could be produced (turbine Ø 40 mm) and an extrapolation with
a 50 % downsized turbine predicts that even more electric power
could be produced. While it is necessary to transport natural gas
at high pressures, end-users require gas delivery at only a
fraction of the main pipeline pressure. Therefore, energy can be
recovered at pressure reduction stations if throttling valves are
replaced by expanders driving electrical generators [2]. For power
recovery, turbines are generally rated from 150 kW to 2.5 MW,
however, the pressure reduction process is usually done in several
stages, and an array of small turbine-generator modules could
replace one large pressure reduction valve [3]. Also, the turbo
expanders used today in cryogenic plants transfer the excess power
(in the kW range) to a brake compressor where the energy is finally
dissipated into cooling water. If a generator would be employed for
the braking of the turbo expander, energy could be recovered, and
therefore the efficiency of such plants could be increased
[4].
Several of the abovementioned applications, e.g. in automobiles,
need ultra-compact power generation systems.
Fig. 1. Picture of the ultra compact (22 mm x 60 mm) air-to-power
demonstrator.
Power density in both turbomachinery and electrical machines
increases with increasing rotational speed [5], [6]. Therefore, for
highest power density, these systems are operating at speeds
between 100 000 rpm and 1 Mrpm at power levels of up to several
kilowatts. Besides higher power applications, micro-turbines with
less than 100 W power output and very high speeds have been
reported in literature. In [7], a modular system consisting of an
off-the-shelf air turbine from a dental drill, a permanent-magnet
(PM) generator and a rectifier has been realized, with a maximal
power output of 1.11 W and a maximal speed of 200 000 rpm.
Drawbacks of this system are the poor power density (0.02 W/cm3)
and the large inlet flow rate of 45 l/min at maximum power output.
The power electronics of the system consists of a three-phase
transformer a diode bridge rectifier and a 5 V linear
regulator.
In [8], a single-stage axial micro gas turbine coupled to a
commercial electrical machine with a maximal electric power output
of 16 W at 160 000 rpm and 1 bar supply pressure has been
introduced for later use in a gas turbine system. The maximal
torque and mechanical power generated from the turbine is 3.7 mNm,
respectively 28 W. The generator is connected to a variable three
phase resistive load to measure the electric output power. The
total system has a maximum efficiency of 10.5 % at 100 000 rpm and
achieves a power destiny of 1.6 W/cm3, excluding power and control
electronics.
In [9], a PM generator, capable of supplying 8 W of dc power to a
resistive load at a rotational speed of 305 000 rpm is shown. The
stator uses interleaved, electroplated copper windings on a
magnetically soft substrate. The rotor consists of an 8-pole SmCo
PM, back iron and a titanium sleeve, because of the high
centrifugal forces. The machine was characterized using an
air-driven spindle. To provide a dc voltage, the ac generator
voltages were first stepped up using a three-phase transformer and
then converted to dc using a
I
three-phase Schottky diode bridge rectifier. The dimensions of the
device are chosen with reference to a future integration into a
micro turbine engine. This leads to a generator power density of 59
W/cm3 and to a generator efficiency of 28 %.
A planar generator with an diameter of 8 mm consisting of a
permanent magnet disc rotor cut out of bulk SmCo or NdFeB protected
by a titanium sleeve, and a silicon stator with electroplated
three-phase planar coils is presented in [10]. The generator is
driven by a planar turbine, etched in the other side of the rotor.
Due to the turbine construction, the speed is limited to 100 000
rpm with 5 bar compressed air supply. A maximum power output of
14.6 mW was measured at 58 000 rpm with three Y-connected 50
resistors. Using a turbine of a dental drill, the rotor reached a
maximum speed of 420 000 rpm. With this setup, the highest electric
power output of 5 W (three Y-connected 12 resistors) was reached at
380 000 rpm with an electric efficiency of 66 %.
In [11], a compressed-air-to-electric-power system with a
rotational speed of 490 000 rpm and a power output of 150 W is
presented. The system is based on the reversal of an existing
turbocompressor system which reaches a maximal pressure ratio of
1.6 at a maximal rotational speed of 550 000 rpm and an electric
power input of 150 W. It is driven by a low voltage power
electronics with 28 V dc input. With new and specially designed
nozzle guide vanes, the turbo compressor system can be reversed and
operated as a turbine system.
In this paper, a compressed-air-to-electric-power system with a
rated rotational speed of 350 000 rpm and a rated power output of
60 W is presented (Fig. 1). First, different turbines are described
and compared, and then a full description of the system is given.
This compressed-air-to-electric-power system comprises of a
single-stage axial impulse turbine (Laval turbine), a PM-generator
and the power and control electronics. Secondly, measurements like
mass flow, electric output power and efficiencies of the different
system parts are presented.
II. TURBINE SELECTION There are several options for the turbine
which are
compared concerning, e.g., size, efficiency, rotational speed, and
simplicity in manufacturing.
A. Axial Turbine 1) Single-stage axial impulse turbine (Laval
turbine)
In impulse turbines, the drop in pressure (expansion) of
pressurized air takes place only in the stationary nozzles and not
between the moving rotor blades (p2 ≈ p3). This is obtained by
making the blade passage of constant cross-sectional area. The
nozzle vanes produce a jet of air of high velocity and the blades
change the direction of the jet, thus producing a change in
momentum and a force that propels the blades. Advantages of an
impulse turbine are the small leakage losses because of the small
pressure gradient over the rotor blades, lower rotating speed
compared with the reaction turbine and a minimal axial thrust which
results in low friction losses in the bearings. A further advantage
of this turbine type is the simple construction and the possibility
to use a shrouding band,
which would lead to a higher efficiency. Disadvantages of an axial
turbine are the losses in the nozzles because of the high
acceleration of the pressurized air, and the blade losses because
the air is highly deflected. These main disadvantages lead to lower
efficiencies than for reaction and radial turbines.
2) Single-stage axial reaction turbine In reaction turbines a part
of the expansion of compressed
air takes place between the rotor blades. For a reaction of 0.5 the
expansion takes place in equal shares in the stationary nozzles and
between the rotor blades. Drawbacks of a reaction turbine are the
additional friction losses in the bearings, because of additional
axial thrust, due to the pressure gradient over the rotor blades
and the more complex construction. Advantages are the better
efficiency and also the possibility of adding a shrouding
band.
B. Radial Turbine Inward-flow radial (IFR) turbine
Concerning the efficiency the radial turbine is the best choice,
but there are some major disadvantages: For low flow rates the
blade height of the turbine (< 0.5 mm) or the turbine diameter
get very small (< 1 cm). This leads to expensive turbines which
are difficult to manufacture. Also the curved geometry of the rotor
blades, the spiral casing and the radial outlet lead to a difficult
production. Due to the higher rotating speed the life time of the
bearing is reduced.
C. Further Systems Reciprocating engine
Theoretically, a reciprocating engine could be used instead of an
axial- or radial turbine. However the disadvantages like the piston
lining, lubrication and vibrations and the rather low speed (e.g.
big size of the generator) are so dominant that this approach is
not an option.
Due to simplicity, size and rotating speed a single-stage axial
impulse turbine has been chosen.
III. TURBINE AND NOZZLE GUIDE VANE DESIGN
A. Turbine Design Assuming adiabatic flow through the turbine,
the
corresponding ideal enthalpy temperature drop ΔT(1-3s) can be
calculated with
1 1 1
3 1 (1 3 ) 219.2 Ks sT T T −= − Δ =
(2) with p1 = 3 bar, p3 = 1 bar and T1 = 300 K. For the real
expansion the increase of entropy, e.g. losses, must be considered.
The isentropic efficiency ηis was assumed to be 30 %, which leads
to the actual expansion drop of
(1 3) (1 3 ) 324.2 K 275.8 Ks isT T Tη− −Δ = Δ = → =
(3) and to the theoretical mass flow of
(1 3 )
g2.5 s
= = Δ
(4)
where cp is the specific heat capacity and Pmech = 60 W. The
effective turbine inlet area A (and thereby the radii r1, r2 and
r3) can be calculated with the flow function, described in [12].
Furthermore, the velocity diagram at rotor entry and rotor outlet
can be calculated as shown in Fig. 2. The relatively large absolute
velocity c2 = 335 m/s is near the sonic speed.
B. Nozzle Guide Vanes Design Due to the fact that the selected
turbine type is an impulse
turbine, the pressure drop, e.g. the acceleration of the air, fully
takes place in the nozzle guide vanes. This means that the outer
inlet area decreases, e.g. the radius r1 reduces to r2, while r3
remains constant, and the nozzle guide vane output area equals the
turbine inlet area. The second important function of the guide
vanes is the deflection of the air stream to the correct angle,
such that the velocity diagram is consistent as shown in Fig.
2.
C. Leakage Losses All turbomachines suffer from losses associated
with the
leakage of some fluid around rotors and stators. Tip leakage is
driven by the pressure difference between the blade suction and
pressure side. Also the manufacturing tolerances cannot be
decreased proportional with the turbine scaling, therefore the
leakage losses become more dominant for small turbines. Due to the
fact that the rotor blades are only 0.5 mm high, the tip clearance
between rotor and casing must be as low as possible. The first
measurement was made with a tip clearance of 0.1 mm (dc/dh = 20 %),
which is clearly too much. For big turbomachinery the tip clearance
is in the range of 1 % to 2 % of the rotor height. Therefore, for
the further measurements tip clearance must be reduced to values
below 10 %.
D. Reynolds Number The Reynolds number characterizes the flow in
the turbine
(laminar / turbulent) and is defined as
Re hcd v
= (5)
where c denominates the speed of the airflow, dh the height of the
air flow channel (rotor height) and v the kinematic viscosity. As a
consequence of miniaturization, the Reynolds number will decrease,
and therefore the flow will become more laminar. In the used axial
turbine the Reynolds number is in the range of 11 000, with is
still in the area of turbulent air flow.
IV. ELECTRICAL CONSIDERATIONS The rotor of the PM generator
consists of a diametrically magnetized cylindrical SmCo or NdFeB
permanent magnet encased in a retaining titanium sleeve ensuring
sufficiently low mechanical stresses on the magnet. The
eccentricity is minimized by shrink fitting the sleeve on the
permanent magnet and grinding the rotor. Additionally, the fully
assembled rotor is balanced. The stator magnetic field rotates with
high frequency (5.8 kHz), it is therefore necessary to minimize the
losses in the stator core by using amorphous iron. In order to
minimize the eddy current losses in the three-phase air-gap copper
winding, the winding is realized with litz-wire. The motor
has
Fig. 2. Drawing of the single-stage axial impulse turbine (Laval
turbine) with rm = 4.5 mm and the corresponding nozzle guide vane
and velocity diagrams at rotor inlet and outlet for the rated
rotational speed. c2 and c3 are the absolute velocities at the
rotor inlet and outlet. v2r and v3r are the relative velocities at
the rotor inlet and outlet. u is the rotor speed at radius
rm.
TABLE I: COMPUTED THERMODYNAMIC DATA Parameter Description Value
Unit T1 Inlet air temperature 300 K p1 inlet pressure 3 - 8 bar p3
outlet pressure 1 bar ηn nozzle guide vane efficiency 90 % ηis
isentropic efficiency 40 % rm median radius 4.5 mm dh rotor blade
height 0.5 mm nr rated speed 350 000 rpm u rated stator speed at rm
165 m/s
TABLE II: MEASURED GENERATOR DATA Parameter Description Value Unit
nmax maximum speed 500 000 rpm ΨPM magnet flux linkage 0.32 mVs
Uind back EMF at rated speed 11.7 V LS stator inductance 2.1 μH RS
stator resistance 0.12
a peak phase-to-phase voltage of 20.2 V at 350 000 rpm (with the
NdFeB permanent magnet). The generator design has been optimized,
considering the total losses [13] (not included in the optimisation
process are the bearing losses), therefore, in the rated operating
point of the turbine the generator efficiency is 93 %. A detailed
description of a similar motor/generator has been presented in
[14]. In Table II the measured electrical data of the PM generator
is summarized, and in Fig. 3 the computed efficiency of the
generator in the entire power-speed plane is shown.
The first measurements were made with varying a resistive
three-phase load. In a second step, a bi-directional power
electronics consisting of an active three-phase rectifier and an
additional boost converter will be used. The power electronics,
analyzed in [15], shows an efficiency of 95 % at rated power.
The output of the system is controlled to 24 V dc, allowing for
direct connection to applications/loads in contrary to the variable
dc output voltage in [7] and the variable three-phase ac voltages
in [8].
V. SYSTEM INTEGRATION In Fig. 4 a solid model of the
compressed-air-to-electric-power system is shown. In the following
section the system integration, e.g. the air flow, the rotor
dynamics and the power density is described.
A. Air Flow The compressed air enters through a common
pneumatic
connector that can be screwed into the system on the right hand
side. The pressurized air then gets diverted into eight channels
that are arranged symmetrically in the generator casing and the
ball bearing shields (indicated with arrows in Fig. 4). This leads
to higher effort in the construction of the casing, but the
generator and the ball bearings can be cooled and as a positive
side effect, the inlet air gets heated up which leads to higher
outlet air temperature and therefore less problems with dew point
and icing. Calculations show that the temperature rise due to waste
heat from the generator is in the range of 5 K. The pressurized
inlet air then reaches the nozzle guide vanes. In the first part of
the nozzle guide vanes the area is decreased to the effective
turbine area and thereby the pressurized air is accelerated and the
pressure drops to almost outlet pressure. In the second part the
accelerated air is diverged to the right angle α2. The air then
passes through the turbine and leaves the system on the left hand
side to the environment. A detailed view of the air flow in the
nozzle guide vanes and the turbine can be seen in Fig. 1 and Fig.
2.
B. Rotor Dynamics In order to run the system in between two
critical speeds,
the bending modes of the rotor and turbine assembly are determined
with finite element simulations. The spring constant of the bearing
system is taken into account, which shifts the bending modes to
lower frequencies. The length of the shaft is adjusted such that
rated speed (350 000 rpm, 5833 Hz) falls between the second (285
420 rpm) and the third (723 900 rpm) bending mode (Fig. 5). In Fig.
6 a picture of the axial impulse turbine and rotor with assembled
high speed bearings is shown.
C. Power Density The integrated system has a volume of 22.8 cm3 (d
= 2.2 cm
l = 6 cm) and the electronics interface has a volume of 60.8 cm3 (l
= 4.5 cm, b = 4.5 cm, h = 3 cm). This leads to a maximal generator
and turbine power density of 4.4 W/cm3 and of 0.8 W/cm3 including
the power and control electronics. Due to the fact that the
electronics were made for motor applications and not for compact
generator applications, a specific redesign and integration will
increase the overall power density of the system. Integrating the
power electronics into the turbine-generator system will avoid an
additional heat sink if it is thermally attached to the generator
casing and thereby cooled by the air flow.
Fig. 3. Computed efficiency of the generator (including bearing
losses).
Fig. 4. Solid model of the ultra compact (22 mm x 60 mm)
air-to-power demonstrator.
a)
b)
c)
Fig. 5. Bending modes of the rotor. First bending mode at 198 840
rpm, 3314 Hz (a), second bending mode at 285 420 rpm, 4757 Hz (b)
and third bending mode at 723 900 rpm, 12065 Hz (c).
titanium sleeve
axial turbine
Fig. 6. Impulse turbine and rotor with assembled high speed
bearings.
0 100 200 300 400 500 0
5
10
15
1
2
3
4
5
20
40
60
80
100
120
140
6 bar
Fig. 7. Measured losses of the high speed motor over speed. The
total power losses include the bearing losses, windage losses and
core losses.
Fig. 8. Measured and calculated mass flow as a function of the
inlet pressure.
Fig. 9. Electrical power generated by the turbine and generator
system as a function of speed and supply pressure.
0 200 400 600 0
1
2
3
4
5
6
5 bar 6 bar
5
10
15
20
25
30
)
p = 2 bar p = 3 bar p = 4 bar p = 5 bar p = 6 bar
0 200 400 600 0
5
10
15
20
25
)
p = 2 bar p = 3 bar p = 4 bar p = 5 bar p = 6 bar
Fig. 10. Torque generated by the turbine as a function of speed and
supply pressure.
Fig. 11. Turbine efficiency as a function of speed and inlet
pressure.
Fig. 12. System efficiency as a function of speed and inlet
pressure.
VI. MEASUREMENTS An experimental test bench is built in order to
verify
theoretical considerations and the compressed-air-to-power system
concept. It includes a mass flow sensor and several thermocouples
and pressure sensors and the three-phase variable resistive load.
Measurements of mass flow, electric output power and efficiencies
of the different system parts will be presented in the following.
The system has been tested up to an inlet pressure of 6 bar and a
maximal outlet electric power of 124 W. The
turbine-generator-system has been built with a rotor tip clearance
of 0.1 mm and without a shrouding band.
A. Measurement Setup For the measurement, the operating point could
be changed
by varying the resistive three-phase load and the supply pressure.
Additionally to input and output pressure, the input and output
temperature of the air flow has been measured. As expected, the
efficiency could not be calculated depending on the temperature
drop, because the turbine is not sufficiently isolated from the
thermal losses of generator and the ball bearings. At the first
possible temperature measurement point after the turbine the air is
already heated up. For better verification the mass flow has been
measured in the low and
high pressure side. The pressurized air to electric power
efficiency has been calculated using
, (1 3 )
η −
= ⋅Δ ⋅
(6)
with ΔT(1-3s) from (1). The isentropic efficiency of the axial
turbine can now be calculated with the measured efficiency of the
generator shown in Fig. 7 (the copper losses must be added
separately, described in VI.B).
B. Initial Measurements Since the rotor bending modes, and the
interference fit of
rotor sleeve and the permanent magnet have been designed for a
maximal speed of 500 000 rpm, the generator could first be tested
as a motor up to a speed of 500 000 rpm. Thereby, the losses at the
rated speed of 350 000 rpm (6.2 W) and at 500 000 rpm (13 W) were
measured (Fig. 7). For measuring the bearing, windage and core
losses a deceleration test was used. This method is based on the
fact that in open loop operation (no electrical drive or break) the
rotational energy is used up by the losses, decreasing the
rotational speed accordingly. The gradient of this deceleration is
a measure for the losses. The dynamical equation for the rotor
is
Loss Loss
ω = − = −
(7)
where ω is the angular frequency, J the calculated rotor inertia
(2·10-8 kgm2), TLoss the total friction torque and PLoss the total
losses. Not included in the deceleration test are the copper losses
depending on the phase currents. The copper losses can easily be
added if the phase currents are measured during generator
operation.
C. Mass flow Measurements In Fig. 8 the measured and calculated
mass flow through the
turbine is plotted. It can be seen, that the measured mass flow can
be very well calculated with the flow function described in [12].
Also, the predicted dependence of mass flow (and therefore input
power) and supply pressure in (4) can be verified; the mass flow
does not depend on speed or load. A maximal mass flow of 4.7 g/s at
6 bar inlet pressure was achieved.
D. Speed Measurements Fig. 9 and Fig. 10 show the electric output
power and torque
as a function of speed and supply pressure. The maximal electric
power output is around 124 W at 370 000 rpm and the maximal
measured torque is 5 mNm at 240 000 rpm.
An increase of the resistive load causes a decrease of the torque
and therefore an increasing speed at a constant supply pressure.
The turbine generator system has been tested up to 455 000 rpm and
6 bar supply pressure.
E. Efficiency Measurements Fig. 11 and Fig. 12 show the turbine and
the system
efficiency as a function of speed and supply pressure. The maximum
turbine efficiency is about 28 %, while the maximum system
efficiency is 24 %. This can be compared to [8], where the system
efficiency is 10.5 %. The maximum turbine efficiency is not as high
as assumed in III.A. This is mainly due to the large tip clearance
and the absence of a shrouding band.
VII. CONCLUSION This paper shows the design and measurement results
of a
compressed-air-to-electric power system. The described system has
been optimized concerning power density (4.4 W/cm3) and system
efficiency (ηSystem = 24 %); the computed and measured values are
significantly higher compared to similar systems described in
literature so far. Also the generator efficiency (87 %) is
significantly higher compared to [8] (58 %) and [9] (28 %),
respectively. The better efficiencies can be achieved by system
integration, generator optimization and careful design of the
turbine. Due to the miniaturization, the isentropic efficiency
cannot be predicted analytically and has been verified
experimentally.
Measurements show that the system has a maximum power output of 124
W at 370 000 rpm and a maximum efficiency of 24 % at 350 000
rpm.
In a next step the tip clearance will be further reduced and the
use of a shrouding band will be investigated. Future steps are the
integration of the power electronics and a valve into the system
and the implementation of a digital control in order to provide a
constant dc output voltage for variable loads.
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