2-stage turbo charging on medium speed engines –
future supercharging on the new LERF-test facility
2-stufige Aufladung von mittelschnell-laufenden Dieselmotoren – eine zukünftige Aufladung auf dem
neue LERF-Motorprüfstand
Christer Wik, Heikki Salminen a
Klaus Hoyer b
Christoph Mathey, Stefan Vögeli c
Panagiotis Kyrtatos d
a Wärtsilä Finland Oy
Järvikatu 2-4, 65101 VAASA, FINLAND
b Paul Scherrer Institut
5232 Villigen PSI, SWITZERLAND
c ABB Turbo Systems Ltd
Bruggerstrasse 71a, 5401 BADEN, SWITZERLAND
d Aerothermochemistry and Combustion Systems Laboratory, Swiss Federal Institute of Technology (ETH)
Clausiusstrasse 33, 8092 ZÜRICH, SWITZERLAND
Abstract / Kurzfassung:
Points requiring further attention in application of extreme Miller cycles
together with 2-stage turbo charging have been related to low load operation and engine start-up due to the substantially cooler combustion chamber when applying a Miller cycle. Focus of the work reported in this paper
has been put on investigating different solutions for coping with variable
speed, low load operation as well as exploring the limits of the Miller cycle
utilising a 1D simulation code. Furthermore a new test facility for 2-stage
turbo charging system testing on medium-speed diesel engines, the Large
Engine Research Facility (LERF), which is a new research platform realized
within the Competence Center for Energy and Mobility (CCEM) at the Paul
Scherrer Institut (PSI), is presented.
Key Words: 1-D simulation; 2-stage turbocharging; Diesel engines
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1 Introduction
Application of a Miller cycle to the engine process is a key to combining
low emissions and high cycle efficiency. The so-called Miller effect allows
for a low combustion air temperature at a given end-of-compression pressure, promoting low levels of nitrogen oxides (NOx) formation, while allowing for a large expansion ratio at the same time, and thus, a high
thermal efficiency. However, application of the Miller process tends to diminish the mean effective pressure. This drawback can be compensated
by utilizing a higher boost pressure. When a very early intake valve closing point is employed, preserving the mean effective pressure requires the
boost pressure to be taken to a level unattainable by singe-stage turbo
charging. By application of two-stage turbo charging, the boost pressure
can be brought to the required level, with the added benefit of better
overall turbo charging system efficiency, afforded by inter-cooling between the compressor stages. The improved turbo charging efficiency is a
direct result of the thermodynamic advantage connected with cooling between the compressor stages. On the other hand, the effect of lower compression end temperature has the unwanted side effect of making the ignition of the fuel more difficult, especially under low load and starting
conditions.
Potential with application of extreme Miller cycles together with two-stage
turbo charging has as such been proven and reported in many different
sources [1, 2]. Points requiring further attention are, as mentioned above,
related to low load operation and engine start-up due to the substantially
cooler combustion chamber when applying a Miller cycle. Also the potential performance gain when going for even more extreme Miller cycles
than tested so far is still open. This is why major focus of the work reported in this paper has been put on investigating different solutions for
coping with low load operation respectively load response issues as well as
exploring the limits of the Miller cycle. Investigations have been done utilising a 1D simulation code and outcome of these will be reported in the
paper.
This paper will also focus on presentation of a new test facility for 2-stage
turbo charging system testing on medium-speed diesel engines, the Large
Engine Research Facility (LERF) which is a new research platform realized
within the Competence Center for Energy and Mobility (CCEM) at the Paul
Scherrer Institut (PSI).
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2 Large Engine Research Facility (LERF)
The Large Engine Research Facility (LERF) is a new research platform realized within the Competence Center for Energy and Mobility (CCEM) at the
Paul Scherrer Institut (PSI). See figure 1 for a view of the test facility.
One purpose of this facility is to test new turbo charging systems, in combination with flexible air and fuel delivery control means, and their impact
on engine performance and NOx emission. This should not compromise on
engine efficiency or increase of unburned hydrocarbons and particulate
matter. Beginning with the ground breaking for the new building in April
2008, the installation at PSI has been completed within 7 months and the
first start was realized in October 2008.
Figure 1: LERF test facility outside and inside view
The test engine is a Wärtsilä 6L20 four-stroke, 6-cylinder common rail diesel engine having a rated power of 1080 kW at 1000 rpm nominal speed.
The diesel engine and the variable speed brake system are mounted on a
common base frame which itself rests on spring damper elements on a
large concrete foundation. The brake system includes an induction engine
(ABB AMA 450L6L) and a frequency/voltage converter which feeds the
generated electric power via a 16 kV transformer to the electrical grid.
This brake system allows running the engine at variable speed but still
feeding the generated power synchronously to the power grid. The engine
cooling is implemented using a shell and tube heat exchanger that transfers the waste heat to water from the Aare River.
To comply with the Swiss regulations for the conservation of air quality, a
urea based SCR catalyst was installed to minimize the NOx emissions of
the exhaust gases. The catalytic converter system has an independent
controller using an engine load signal for feed forward control and fine3/14
tuning the urea injection rate using as feedback the actual NO concentration downstream of the catalyst. Figure 2 shows the SCR installation together with the urea injection point and its long mixing pipe for even urea
distribution before the catalyst.
Figure 2: SCR installation and mixing section for urea injection
2.1
Instrumentation of the LERF test facility
The installed measurements can be grouped into three categories.
2.1.1
Engine efficiency for the fixed load cases
In the first group we gather all system variables that allow estimating the
engine efficiency for the fixed load cases. These variables are acquired via
the PUMA analogue interface and are averaged over a period of two minutes.
For measuring the shaft power a torque flange (Kistler 0325DF) is installed. Fuel consumption is measured using two screw displacement flow
meters (Kral OMC) in the supply and return line, respectively. The air consumption is calculated measuring the pressure difference over a venturi
nozzle mounted upstream of the compressor inlet. In addition, the heat
flows entering and leaving the system boundaries carried by cooling water
and exhaust flows are measured. All temperature differences are measured using PT100 thermo-elements. The cooling water flow rates are metered using two flow meters mounted in the LT and HT return pipes. Exhaust gas flow rate is estimated from the mass balance and exhaust gas
composition is used to apply the correct heat capacities for the estimation
of the exhaust gas enthalpy.
2.1.2
Transient cylinder and gas exchange pressures
The second group of measurements addresses the transient cylinder and
gas exchange pressures during the working and gas exchange cycles.
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The exact timing signal is supplied by a crank angle encoder mounted at
the free end and being aligned with the TDC measured from two piston
head positions of cylinder #6, which was chosen for the measurements. A
timing belt with 2:1 drive ratio connected to a 180° disk with inductive
pickup allows identifying the combustion stroke TDC position.
The pressure sensors used for the cylinder indication are supplied by Kistler Instrumente AG. A piezoelectric (p/e) pressure sensor mounted with
a water cooled adapter measures the transient cylinder pressure. The induced charge signal is fed to a charge amplifier providing a linearly scaled
voltage signal to be acquired by the transient recorder. For the cylinder
reference pressure a piezoresistive (p/r) pressure transducer is used
mounted with a water cooled adapter that has a pneumatic switch valve,
opening only when the cylinder pressure is below the applied control pressure. This signal is used as the absolute pressure reference for the p/e
sensor. Cylinder head sensor locations for pressure indication measurements can be seen in figure 3.
Figure 3: Cylinder head sensor locations for pressure measurements
Additionally to the in-cylinder pressures, the transient pressures upstream
of the intake and downstream of the exhaust valves are measured using
two p/r sensors mounted on the cylinder head through water cooled
switching adapters opening when the control pressure is applied. The p/r
transducer signals are fed to the data acquisition system via a preamplifier also giving linearly scaled voltage signals.
2.1.3
Exhaust gas analysis
The third group of measurements encompasses the detailed exhaust gas
analysis using three independent measurement systems.
The most powerful analyzer used is a Fourier-Transform Infrared Spectrometer (FTIR) which can simultaneously quantify a large variety of
emission species. In our case the instrument provides measurements of
NO, NO2, NOX, N2O, CO, CO2, H2O, NH3, SO2, COS, AHC, C2H2, C2H4,
C2H6, C3H6, C4H6, NC8, HCHO, and HCD with an acquisition rate of 1 Hz.
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For comparison to values obtained using the FTIR, measurements are
done in parallel of NO and NO2 concentrations using a Chemiluminescence
detector (CLD) and the total HC using a Flame Ionization Detector (FID).
2.2
LERF test facility for 2-stage turbo charging tests
Conversion of the LERF test facilities towards a functioning 2-stage turbo
charging (TC) system setup was started in June 2009 and start up of the
engine is planned for September 2009. The test system setup will be
same as reported in reference [2] and could be viewed in figure 4.
Figure 4: Planned 2-stage turbo charging system setup for the LERF test
facility
The 2-stage TC system provided by ABB Turbo Systems is being designed
for a maximum charge air pressure of 10 bar a. As a first step there will
be implementation of a variable inlet valve closure (VIC) system which
uses a controllable retarder allowing control of the intake valve timing.
This variability is especially important for engine start-up and partial
loads, where the valve timing should stay unchanged with respect to the
original engine setup, since a strong Miller ratio, i.e. early inlet valve closure, will result in too low in-cylinder temperatures and incomplete combustion.
As a next step there will be an upgrade of the valve timing system to entail fully variable intake and exhaust valves as to also employ internal exhaust gas recirculation and other means to further push the envelope for
reduced emissions.
3 1D-simulations
Simulations have been performed at ABB Turbo Systems Ltd, Swiss Federal Institute of Technology (ETH) Zürich, and Wärtsilä Finland Oy, of
which the first two partners have performed all transient simulations.
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3.1
Used model and general definitions
During the development process, different levels of Miller cycle application
were referred to by the intake valve closing point (IVC), in terms of crankshaft degrees. However, a more general way of defining the Miller effect
level is by calculating the portion of the stroke available for cylinder filling
defined as follows:
m* = sM/s
Eq. 1
This parameter may be called the Miller ratio. Here sM is the amount of
stroke that the piston has moved down from the top dead centre when
intake valves are closed. Assuming that the Miller effect would be more
dependent on volumetric rather than time ratios, it is more descriptive of
the Miller effect than cam timing angle degrees, as it is not dependent on
the geometry (or the existence) of the crank gear. Thus, when m* = 1,
there would be no Miller effect, and the Miller ratio would tend towards
zero with increasing amounts of Miller effect application. In practical
terms, a Miller ratio of around 0.5 might be considered rather drastic, as
then half of the effective piston stroke would not be utilized for air induction. As a side note, if the flow in the engine manifolds would be mainly
dependent on flow dynamics, camshaft timing degrees may well be a better indicator of the extent of the Miller effect than the Miller ratio m*,
however, this does not seem likely in the case of medium speed engines.
The term compression ratio originally was coined for traditional engines
with the idea that the dead centres of the piston would demarcate the operating phases of the cycle. When referring to a Miller cycle engine, in fact
it would be more appropriate to refer to the expansion ratio instead of the
compression ratio. In general, the expansion ratio of the engine can be
considered equal to the geometric compression ratio, whereas the compression ratio will be dependent on the Miller ratio m* utilized, as follows:
εC = m* (ε - 1) + 1
Eq. 2
The IVC closing points, Miller ratios and corresponding compression ratios
used in the simulations are shown in Table 1.
Due to the nearly adiabatic and isentropic expansion-compression process
that takes place within the engine cylinder, compression temperature in all
cases will be mainly a function of the effective compression ratio if the receiver air temperature is not greatly varied. Looking at the effective compression ratios in Table 2, it becomes apparent that with low values of
Miller ratio, end-of-compression temperature may become a problem
when diesel-type combustion is to be initiated. This should come as no
surprise, as the very purpose of the Miller effect is to bring the cylinder
charge temperature down, and was originally used to prevent uncontrolled
spontaneous ignition from happening in gas engines.
7/14
Table 1: IVC point, Miller ratio, and effective compression ratios used
IVC (BBDC)
m*
εC (ε = 16)
0
1,00
16,0
33
0,94
15,1
73
0,71
11,7
83
0,63
10,4
100
0,48
8,2
It was deemed necessary to study the engine response to various levels of
Miller effect by simulating the engine system with the aid of onedimensional engine simulation software. Suitable process parameter selection for the extreme Miller cycle engine, part load operation including special measures to facilitate starting and to improve part load characteristics
were assessed to. Figure 5 shows the so-called project map of the GTPower W6L20 engine model.
Figure 5: GT-Power engine model
While this kind of model is advanced, it can never fully represent the reality. In order to confirm the simulation results, and to certify that the desired performance indeed is attained on real engines, the favourable configurations found with the aid of simulation need to be proven by practical
tests. Accuracy of the engine model was validated against earlier test engine results [2], using realistic turbo maps in the model.
Figure 6 shows the measured and simulated brake specific fuel consumption (SFOC) and temperature before high-pressure (HP) turbine which
both fit towards measurement results. As can be seen in figure 6, the GTPower model was found to produce a very good match at high loads. Due
to turbocharger maps and measurement accuracies as well as non performed model adjustments at low loads, the match is slightly worse there.
8/14
700
Measured
1,20
Simulated
1,15
t 5HP (°C)
1,25
600
500
Temp b. turbine
Normalised SFOC
1,30
1,10
1,05
1,00
400
300
0,95
0,90
200
20
30
40
50
60
70
80
90
100 110
20
30
40
Load (%)
50
60 70
Load (%)
80
90
100 110
Figure 6: Simulated results vs. measurements on Lab 6L20 / 2-stage TC,
MDO, 1000 rpm constant, IVC-83
3.2
Steady state simulations results
Steady state simulations have been performed for optimising the Miller
ratio, scavenging period, exhaust valve opening timing, and valve lift of
the engine.
It is typical of internal combustion engines that there are strong interactions between the different parameters that can be varied and therefore it
does not suffice to try to optimize just one variable while keeping the others constant. For this reason, a DoE (Design of Experiments) analysis was
applied to the variables chosen for the parametric study. Locations of
measured and generated data points are shown on the NOx emissions vs.
SFOC plane in figure 7. The line of points located closest to the origin
would represent the Pareto front of optimal solutions, each point representing a unique combination of the pertinent variables.
Figure 7: Measured and generated data points in the NOx vs. SFOC plane
9/14
The model is actually pointing the way that the Pareto front will be found
outside of the studied range altogether. The Pareto front analysis suffered
from a lack of data points to properly represent the influence of variation
of parameters, and there was probably not a complete enough set of variables included in the study. However, it is in the nature of the DoE approach that scarce data might be studied first and then, with the aid of
the DoE methods, the true optima may be approached and approximated
in an iterative process, quicker and closer than by other means.
Variation of the Miller effect by point of intake valve closure was studied
by a series of simulation runs. Simulations were made up to an intake
valve closing point of 100°CA before bottom dead centre (BBDC) as seen
in table 1. It was found that the specific fuel consumption of the engine
will decline when a high level of Miller effect is utilized, and the optimum
point for intake valve closure would be close to 83°CA BBDC. This would
correspond to a Miller ratio of m* = 0.63. At Miller ratios below this, the
pumping work increases considerably leading to higher SFOC. The needed
air receiver pressure increases exponentially with lower Miller ratio. See
figure 8.
3,0
1
0,99
2,5
0,97
0,96
2,0
0,95
p3
0,94
Normalised p 3
Normalised SFOC
0,98
1,5
SFOC
0,93
0,92
1,0
-30
-40
-50
-60
-70
-80
-90
-100
-110
IVC
Figure 8: Simulations results with varying Miller degree
The other parameter showing a big influence on fuel consumption and
other engine performance data was the scavenging period. With high
Miller degrees there is a clear optimum towards shorter scavenging period
since the pressure drop over the cylinder during the scavenging period
tends to get larger while the charge air density tends to get higher, as
well. Thus, if equal scavenging area integral is provided in both cases, the
scavenge air ratio of the extreme Miller cycle engine will be larger than
the baseline engine’s. The effect of valve overlap was studied by varying
the point of exhaust valve closure (EVC) only and the results could be
seen in figure 9. Choice of optimum scavenging period is always a compromise of lowest fuel consumption and acceptable thermal load (exhaust
10/14
gas temperature before turbine). As a conclusion, about 15…20°CA
shorter scavenging period vs. the reference would be the optimum choice.
Reference
t5
NOx
Normalised parameters
SFOC
-10
-5
0
5
10
15
20
25
30
35
40
Scavenging period
Figure 9: Simulations results with varying scavenging period at IVC -83
Changes of the other parameters, exhaust valve opening timing and valve
lift, showed only a small influence on fuel consumption and NOx emissions
and are not in more detail presented here. The only bigger effect seen was
an increase in thermal load with decreased valve lifts.
3.3
Transient simulation results
Different configurations as listed in table 2 have been performed in order
to screen the possibilities and influences with changes in certain parameters on load acceptance behaviour for the W6L20 engine equipped with a
2-stage turbocharging system. Furthermore, different technologies as 3pulse exhaust gas system, air injection into the air receiver, and Powertake-in (PTI) system for improved load acceptance at IVC -83 have been
simulated. Boundary conditions used have been:
• constant speed operation
• 1080 kW used for 100% load
• load steps: 0-33-66-100% in 0.5 s ramps per step
• smoke limiter set to minimum λC=1/1.2/1.5 for full load range
• when VIC on: VIC on (100°CA delayed IVC) for first load step, then
back to early IVC
A summary of all simulation results with different Miller degrees is seen in
table 2. It can be noted that the VIC system improves performance radically and with VIC on at IVC -73, we reach almost same load acceptance
behaviour as for the reference case.
11/14
A VIC system would also aid steady-state part load operation as the air
amount inside the cylinder would increase considerably.
Smoke limiter setting is greatly influencing the results as well since this
allows for only a certain drop in air-to-fuel ratio and the lower it is allowed
to drop, the faster the reaction will be. Unfortunately also transient smoke
emissions are greatly influenced and the lower it is allowed to drop, the
higher smoke emissions we see.
Table 2: Transient simulations done and most important results, red values do not fulfil class requirements (max. 10% speed drop, 5 s rec. time)
TC system
Load step
IVC
Smoke
VIC
Speed
Recovery
(BBDC) limiter (on/off) drop (%) time (s)
1-stage (ref)
0-33%
33
1.0
off
2.5
2.4
2-stage
0-33%
83
1.0
off
25
15.1
2-stage
0-33%
83
1.0
on
3.8
3.0
2-stage
33-66%
83
1.0
off
8.4
6.3
2-stage
66-100%
83
1.0
off
2.6
1.7
2-stage
0-33%
73
1.0
on
2.5
2.2
2-stage
0-33%
73
1.2
on
5
4.9
2-stage
0-33%
73
1.5
on
16.4
14.8
Reaching acceptable loading criteria for 0-33% load at IVC -83 would be
possible with utilisation of a VIC system, to delay the inlet valve closure
timing with 80…100°CA, and with fine tuning of the smoke limiter and system inertias. But for the load step 33-66% as well as for two-step loading
(0-50-100%) it would not be possible without additional means. Applying
delayed inlet valve closure timing up to 50% load would help the situation
considerably.
Resulting load acceptance when applying additional technologies at IVC 83 in a 2-stage turbocharged W6L20 engine can be seen in figure 10. For
air injection into the receiver, air with 10 bar and at 25°C temperature
was used whilst the PTI required an electrical input of 20 kW to accomplish three-step loading. The area of the on/off-valve for air injection respectively the amount of air injected was adjusted to still have some margin to the surge limit on both turbochargers.
Changing to a 3-pulse exhaust system is not having any impact on the
results for the critical load step 33-66% even though the impact is rather
big for the first load step (0-33%).
One potential for improved load acceptance is the implementation of air
injection into the receiver or a PTI device either on the low-pressure (LP)
or high-pressure (HP) charger. With these additional means, three-step
loading is easily achievable and the PTI device would even accomplish 2step (0-50-100%) loading.
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Engine Speed
1040
1020
ISO G2 Frequency Band
1000
n [rpm]
980
960
SPEX
VIC-Switching
3-Pulse
PTI-LP
940
PTI-HP
920
AIR
1st 3-Pulse
900
0
10
20
30
40
50
60
70
80
90
t [sec]
Boost Pressure
8
7
pRec [bar]
6
5
VIC-Switching
4
SPEX
3-Pulse
3
PTI-LP
2
PTI-HP
AIR
1
1st 3-Pulse
0
0
10
20
30
40
50
60
70
80
90
t [sec]
Configuration
0% - 33% (VIC)
33% - 66%
66% - 100%
SPEX
Speed
Reco
drop
time
[%]
[s]
3,8%
3,0
8,4%
6,3
2,6%
1,7
3-Pulse
Speed
Reco
drop
time
[%]
[s]
2,7%
2,0
8,4%
6,0
2,7%
1,7
PTI LP-TC (20kW)
Speed
Reco
drop
time
[%]
[s]
3,8%
3,0
7,4%
4,6
2,6%
1,7
PTI HP-TC (20kW)
Speed
Reco
drop
time
[%]
[s]
3,8%
3,0
7,1%
4,6
2,6%
1,7
AIR
Speed
Reco
drop
time
[%]
[s]
3,8%
3,0
4,9%
2,6
2,6%
1,7
Figure 10: Engine speed and boost pressure variations as well as speed
drops and recovery times when applying special technologies for improved
load acceptance behaviour at IVC -83 and 0-33-66-100% load
4 Summary and outlook
Investigations of different solutions for coping with variable speed low
load operation as well as exploration of limits with the Miller cycle have
been done with a 1-D simulation tool.
The simulation studies point the way to the future engine process enabling
very low emissions, with little or no fuel consumption penalty. A high degree of Miller effect utilization (very early closing of the intake valves) enables combustion in a low-temperature environment, with accompanying
low NOx emissions. Nevertheless, at very high Miller degrees, the specific
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fuel consumption of the engine starts to suffer. Therefore, the optimum
point of inlet valve closure is deemed to be within the range already studied i.e. a Miller ratio, m*, of about 0.63 seems to be the optimum for low
fuel consumption and concurrent low NOx emissions. This would correspond to an inlet valve closing timing of 83°CA BBDC and effective compression ratio of 10.4:1 with a geometrical compression ratio of 16:1. Together with this optimum Miller ratio, about 15…20°CA shorter scavenging
period would be needed for lowest fuel consumption and acceptable engine thermal load.
As regarding solutions for coping with variable speed low load operation,
investigations have been performed revealing that there is a high potential
by applying a variable inlet valve closure (VIC) device, power-take-in device, or air injection into the receiver. As a next step in the transient
simulations, turbo maps with more detail in the low pressure ratio range
will be inserted together with more realistic fuel injection control, in an
effort to try to reproduce low load running with better accuracy.
However, even with the best of the current state-of-the-art simulation
models, it is not possible to fully reproduce real engine operation with
every detail of the process. In particular, the uncertainties of NOx prediction make it essential that the suggested solutions be tested in practice.
At the new Large Engine Research Facility, an engine incorporating the
most promising design features will therefore be put under tests to assess
and verify in real world the ideas and results presented here. This new
test facility for 2-stage turbo charging system testing on medium-speed
diesel engines at the Paul Scherrer Institut (PSI), will form a unique and
optimum setup for further pushing the envelope towards reduced emissions when starting up in September 2009.
Acknowledgements
Work reported in this paper belongs to the Hercules β research program of
the European Commission Framework Program FP7. Partners involved in
the work are both the industrial partners ABB Turbo Systems Ltd, Wärtsilä
Finland Oy, and Kistler Instrumente AG as well as the academic partners
Paul Scherrer Institut, and Swiss Federal Institute of Technology (ETH)
Zürich.
References
[1]
E. Codan, C. Mathey, ‘Emissions – A new Challenge for Turbocharging’, paper no. 245, 25th CIMAC Congress, Vienna 2007.
[2]
C. Wik, B. Hallbäck, ‘Utilisation of 2-stage turbo charging as an
emission reduction means on a Wärtsilä 4-stroke medium-speed diesel engine’, paper no. 101, 25th CIMAC Congress, Vienna 2007.
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