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Engine Control and Performance Enhancement with Variable Valve Train for Gas Engines
C. Christena,1 and E. Codana,2
a ABB Turbo Systems Ltd. Bruggerstrasse 71a, CH-5401, Baden, Switzerland
Abstract: With regard to NOx emissions and fuel efficiency, modern gas engines are starting to overtake their diesel counterparts in the corre-sponding power classes. However, in terms of power density and espe-cially load acceptance capability, the diesel engine is still a substantial step ahead. The present paper focuses on highly turbocharged premix gas engines with Miller timing. It shows by means of engine cycle simulations how the use of advanced engine power control based on variable valve actuation is able to improve the weak points of today’s gas engines, and even to attain a further increase in thermal efficiency. Further, it is dem-onstrated how ABB’s new valve train system VCM is able to fulfill the re-quirements of the postulated control concept. Engine test bed results have shown accurate and reliable operation as well as immediate response to control commands.
Key Words: Variable Valve Train, VCM, Gas Engine, Power Control
1 E-mail: [email protected], URL: http://www.abb.com/turbocharging 2 E-mail: [email protected], URL: http://www.abb.com/turbocharging
1 Introduction
The design goals of modern lean-burn gas engines focus on finding the best compromise on the trade-off between fuel efficiency and load accep-tance capability for a desired power density and a required emissions limit. Hence, corresponding development activities concentrate mainly on four key aspects:
· For ever higher power density and fuel efficiency, the general ten-dency to knocking combustion has to be reduced by means of leaner air/fuel mixtures, the Miller cycle and careful design of component cooling and possibly applying cooled EGR.
· Minimizing cylinder to cylinder and cycle to cycle variations based on well distributed cylinder filling and high-energy ignition systems.
· Minimizing of the safety margin to the knock boundary even under changing fuel quality and ambient conditions by use of variable igni-tion timing and knock sensors.
· Development of control concepts for fastest load acceptance without compromising fuel efficiency.
The development potential inherent to the application of high pressure turbocharging in combination with Miller timing has been presented in [3]. Based on simulation results with focus on premix gas engines, the present paper shows how the use of a variable valve actuation system further con-tributes to all of the four key points mentioned. As a consequence, it im-proves engine efficiency even in connection with improved engine load ac-ceptance properties. In Chapter 2, the state-of-the-art concept of gas en-gine power control is discussed. Chapter 3 presents an advanced control concept based on a variable valve train. Then, this concept is investigated by engine simulation. The corresponding simulation model is described in Chapter 4, the simulation results presented in Chapter 5. The require-ments on the variable valve actuation system for the realization of the concept presented are summarized in Chapter 6. Finally, Chapter 7 pre-sents how ABB’s new variable valve train VCM is able to cover the stated requirements.
2 Conventional Control Strategies of Lean-burn Gas Engines
Highly turbocharged gas engines with lean-burn combustion are able to fulfill the stringent requirements of today’s emission regulations, e.g. “½ TA-Luft”, at high power density and engine efficiency. But, as illustrated in Figure 1, in order to achieve the attractive performance properties de-scribed above, the engine has to be operated in a very narrow range of air/fuel ratio lV, limited mainly by tendencies to knocking, misfiring and an upper limit of NOx emissions. For this reason, an accurate and reliable
control of the necessary air/fuel ratio together with a control device for trapping the required amount of mixture in the cylinders is crucial.
0.00
0.10
0.20
0.30
0.40
0.50
0.60
0.70
0.80
0.90
1.00
0.200 0.700
Air/Fuel Ratio
Mea
n Ef
fect
ive
Pres
sure
THC
rich
NOx
NOx Limit
Mechanical Limit
Applicable Operating Window
lean
Knocking
Misfiring
Figure 1: Operating range limitations of lean-burn gas engines
2.1 Port Injected Gas Engines
The output of port injected gas engines is controlled by the amount of fuel per working cycle injected into the intake port of the cylinder head, while the air/fuel ratio lV is controlled by adjusting charge air pressure. The fol-lowing control strategies are in use on today’s gas engines:
· Compressor recirculation [8] · Exhaust waste gate · Variable turbine geometry [13], [11] · Throttle valve · Skip firing [10]
On port injected gas engines as well as on diesel engines, load control is based on fuel flow, resulting in comparably good transient engine behav-ior. For this reason, the more fuel efficient but slower air/fuel control strategies on the exhaust side of the engine, like variable turbine geome-try (VTG) or exhaust waste gate (EWG), are preferred on this type of en-gine.
2.2 Premix Gas Engines
In the case of premix gas engines, gaseous fuel is admitted to the charge air path ahead of the turbocharging unit and the desired air/fuel ratio is adjusted to attain engine operation without knocking and misfiring at the required level of NOx emissions. Engine power thus results from the amount of air/gas mixture trapped in the cylinder. On advanced engines,
the clearance volume is always scavenged, because residual gas has an adverse influence on the tendency for knocking combustion. Valve timing must be suitably tuned to avoid excessive loss of mixture which would im-pair engine efficiency and emissions of unburned hydrocarbons (UHC). The trapping efficiency is typically kept above 99%. Under this condition, the cylinder mass balance can be calculated with external parameters, e.g. the dependency of engine power on charge pressure pRec can be ex-pressed by using the delivery ratio lR and the total air/fuel ratio lV,tot as shown in Eq. 1. Hence, for a given engine operating point (constant lV,tot and TRec) and a fixed valve timing (constant lR) engine power is propor-tional to the engine intake receiver pressure pRec, Eq. 2.
( ) ecRecRnmitotV
thuRecR
RTLHp
bmep××+×
×××=
1,lhl
Eq. 1
ecRpbmep µ Eq. 2
Thus, engine torque control is conventionally achieved by adjusting the intake receiver pressure pRec by one or more of the following turbocharg-ing system variabilities:
· Throttle valve [2], [6] · Compressor recirculation [6] · Exhaust waste gate [2] · Variable turbine geometry · Variable compressor pre-swirl
For a sudden increase in engine power, e.g. to respond to a sudden load step on an engine operated in stand-alone mode, it should be possible to immediately raise the receiver pressure pRec. Thus, in engine steady state, the turbocharging unit has to be able to provide a higher receiver pressure which corresponds to a higher engine power than actually needed. The surplus receiver pressure is throttled down or blown off to the level neces-sary for the actual engine load by the control measures listed above. In the present document, the achievable difference between the increased and the actual engine power will be named power control margin. The higher the control margin the better the engine load response capability. But, in the case of conventional engine control concepts (e.g. throttle and compressor bypass control) this goes along with increased cylinder back pressure and consequently reduced engine efficiency. Control concepts on the turbine side (EWG and VTG) result in comparably lower engine ex-haust pressure. As a consequence, engine pumping losses during the ex-haust stroke and thus engine efficiency are less affected. As a drawback of these more fuel-efficient concepts, engine transient phases cannot be managed as fast, since for increased receiver pressure the turbocharger rotor(s) first needs to be accelerated. The advantages of both compressor side and turbine side control concepts can be combined and even en-hanced with a more advanced power control concept introduced in the fol-lowing chapter.
3 Advanced Torque Control Concept for Premix Gas Engines via Variable Valve Train
3.1 Working Principle
As introduced in Subsection 2.2, the power of a premix gas engine can be controlled by adjusting the receiver pressure and thus the amount of trapped fuel gas, e.g. by means of a throttle valve between compressor outlet and intake receiver. For a given valve timing and air/fuel ratio, the amount of fuel gas entering the cylinder per cycle can be described by the delivery ratio lR, see Eq. 3. With a variable intake valve train, the amount of fuel gas being delivered to the cylinder can be set by a corresponding inlet valve timing and/or lift. As a consequence the delivery ratio lR can be varied at every load point. For example, starting at filling-optimized valve timing, an earlier closure of the inlet valves decreases the amount of trapped fuel gas and the engine power can be adjusted by choosing the appropriate inlet valve timing. Thus, Eq. 2 can be extended to Eq. 4.
ecR
ecRecR
h
gasair
hecR
gasairR p
RTV
mmVmm +
=×
+=
rl Eq. 3
ecRR pbmep lµ Eq. 4
The influence of variable inlet valve closure timing is illustrated in the dia-grams in Figure 2 and Figure 3 (early Miller timing). As the valve closure timing is advanced, the cylinder filling, and thus the delivery ratio lR, is reduced. In the following, the control of engine power by means of inlet valve variability will be denoted as VCM (Valve Control Management). VCM, ABB’s new variable valve train system, will be described in more de-tail in Chapter 7. As will be shown, its ability is not limited to the variabil-ity of valve closure timing.
Crank Angle [°]
Valv
e Li
ft [m
m]
0.4
0.5
0.6
0.7
0.8
0.9
1.0
480 500 520 540 560 580 600
IVC Timing [°CA]
Del
iver
y R
atio
lR [-
]
Figure 2: Variable valve closure timing
Figure 3: Delivery ratio lR as func-tion of inlet valve closure timing
The control concept based on variable valve actuation postulated above has been studied by means of engine cycle simulations. Further details regarding the applied investigation method are given in the following chapter.
4 Simulation Model
4.1 Simulation Software
Engine cycle simulations have been carried out using ABB’s in-house simulation software SiSy, [9].
4.2 Engine and Turbocharging System Details
A simulation model of a highly turbocharged premix gas engine has been used to study the benefits of VCM. The performance of the simulated en-gine corresponds to the benchmark best in its power class. With two-stage turbocharging it attains attractive equivalent turbocharger efficiencies of hTC,2-st,eq > 70%. Future engine development is expected to follow the two-stage turbocharging path with increased power density and applica-tion of the Miller cycle. The corresponding engine model topology is shown in Figure 4. The en-gine is equipped with an intercooled two-stage turbocharging system. The desired air/fuel ratio lV,tot is mixed ahead of the low pressure compressor stage. Simulations have been carried out in a zero-dimensional mode, considering only the filling and emptying of volumes interconnected by orifice-type flow devices.
Throttle
Compressor Recirculation
Exhaust Gas Receiver
Air/Gas Mixture Receiver
Fuel Gas
Ambient
Engine
High PressureTurbocharger
Low PressureTurbocharger
Throttle
Compressor Recirculation
Exhaust Gas Receiver
Air/Gas Mixture Receiver
Fuel Gas
Ambient
Engine
High PressureTurbocharger
Low PressureTurbocharger
Figure 4: Engine simulation model topology
4.3 Combustion
Heat release has been simulated according to a Wiebe model with con-stant parameters. Generally, the start of combustion has been kept con-stant throughout the investigation. However, whenever combustion timing has been changed in the simulation results shown, it is explicitly stated. Of course, real engine combustion is influenced by a range of parameters. The air/fuel ratio lV in particular has an impact on combustion duration and efficiency which is difficult to model. Nevertheless, we are confident that the simplified assumption of constant combustion allows the evalua-tion of the influence of turbocharging and control systems at least on a qualitative basis.
4.4 Knocking
Combustion knock depends on cylinder charge temperature and pressure mainly in the region of TDC before and until the end combustion. By ap-plying the Miller cycle, cylinder temperature can be substantially lowered and the risk of knock decreased. On the other hand, by eliminating the Miller effect by means of a variable inlet valve timing, e.g. for improving transient engine response, knock tendency is correspondingly increased. For a fair comparison of the potential of different engine control concepts, an estimation of knock tendency had to be considered. As a result, deto-nation has been estimated by a widely used phenomenological knock model described in [4] and [12]. According to the model used, spontaneous ignition of a gas takes place after a certain period of time tKnock. According to Eq. 5, this time is a func-tion of temperature, pressure and mixture composition. With a changing state of the cylinder charge over time, spontaneous ignition occurs when the knock integral shown in Eq. 6 exceeds the value of unity. The model applied is not expected to provide exact information about the occurrence of combustion knock, but serves as an instrument for fair comparisons of engine operating points with different Miller timings or compression ratios e. Furthermore, autoignition at in-cylinder hot spots can not be considered.
RTX
XKnock epX
32
1 ××= -t Eq. 5
11 dtIonautoigniti
IVC
t
t Knockk ò=
t
Eq. 6
4.5 Valve Train Power Consumption
In the presented simulation results, no additional power loss due to valve train variability has been considered. These results are thus independent
of any valve train concept featuring the same valve lift profiles as used in the present investigation. For a discussion of VCM power consumption re-fer to Chapter 7.2.
5 Results
5.1 Increased Engine Efficiency
5.1.1 Improved Gas Exchange Work
The improvement of engine efficiency, achievable by switching from a conventional engine control concept like throttle valve or compressor by-pass control to a more advanced IVC control with VCM is best explained best using a simple example. Assume a throttle controlled engine is oper-ated at a given load point which requires a receiver pressure of pRec,1 = 5 bar and a throttle valve pressure drop of DpThr = 500 mbar at an assumed delivery ratio of lR,1 = 0.6. When changing from throttle control to VCM, receiver pressure is increased to pRec,2 = 5.5 bar while the deliv-ery ratio has to be reduced by about the same percentage (10%) to roughly lR,2 = 0.54, see Eq. 4. Since on the exhaust side of the engine almost no changes are perceived (roughly same pTI at same load and tur-bocharger specification), the theoretical gas exchange work is improved by pRec,2 – pRec,1 = 0.5 bar, see Figure 5. Although the reduction of the de-livery ratio lR also causes more Miller losses (refer to Figure 6, where Miller loss is defined as the difference of the compression work resulting from moving the piston from BDC to Vh,pRec at receiver pressure level and the isentropic compression work from BDC pressure until receiver pres-sure level), the effective gas exchange work is improved, thus resulting in an engine efficiency gain, while the control margin is maintained.
V / Vh [-]
p /
bmep
[-]
Throttle / Bypass ControlVCM
0.1
0.2
0.3
0.0 0.2 0.4 0.6 0.8 1.0 1.2
V / Vh [-]
p /
bmep
[-]
Exhaust Stroke Flow Loss
Intake Stroke Flow LosspRec
pTI
Miller Loss
pac
BDCTDC Vh,pRec
Norm. Cylinder Pressure
Figure 5: Schematic indicator diagram of conventional throt-tle/bypass control and VCM
Figure 6: Illustration of theoreti-cal and effective gas exchange work, flow and Miller losses
With the presented engine model in the foregoing Chapter, simulations have been carried out at nominal engine power. While controlling engine power by a throttle valve, inlet valve closure has been continuously ad-vanced. As a consequence the delivery ratio lR is continuously reduced. Due to the reduced cylinder filling, the throttle valve starts to open and receiver pressure increases in order to achieve the same trapped mass in the cylinders and thus same engine power. Due to increasing engine pres-sure drop, the trapping efficiency would be reduced. In order to keep the latter on a constant level, valve overlap has been symmetrically reduced. As the throttle opens and receiver pressure increases, the brake specific fuel consumption is reduced by Dbsfc = 4 g/kWh, as can be seen in Figure 7. Based on a partial loss analysis, Figure 8 explains the reason for this efficiency gain. Starting from the high pressure cycle efficiency which in-creases slightly with increasing Miller effect lR (red line), the theoretical gas exchange work pRec – pTI increases efficiency up to Dh = 4% when re-ceiver pressure is increased by opening the throttle valve (1). Unfortu-nately, a large part of this efficiency gain is offset by the increasing Miller loss (2) which reduces the maximum engine efficiency gain to about Dh = 1% (in this case, flow (3) and friction (4) losses are almost inde-pendent of inlet valve timing and do not cause any further change in effi-ciency gain).
0.00
0.10
0.20
0.30
0.40
0.80 0.85 0.90 0.95 1.00 1.05
norm. Delivery Ratio lR / lR,ref [-]
Dp T
hr /
p Rec
,100
% [-
]
-6
-4
-2
0
2
Dbsf
c [g
/kW
h]
Throttle Valve Pressure DropSpec. Fuel ConsumptionReference: Throttle Control
-1
0
1
2
3
4
5
6
7
8
0.7 0.8 0.9 1 1.1 1.2 1.3
norm. Delivery Ratio lR / lR,ref [-]
Dh
[%]
1) Theoretical Gas 1) Exchange Work
2) Miller Loss
3) Flow Loss
4) Friction Loss
Indicated Efficiency High-Pressure Cycle
Engine Thermal Efficiency
1
2
34
Figure 7: Decrease of fuel con-sumption at nominal engine power by changing from throttle to VCM control
Figure 8: Partial efficiency loss analysis at nominal engine power
In the indicator diagram of Figure 9, the change in effective gas exchange work between throttle control and VCM is shown.
0.1
0.2
0.3
0.0 0.2 0.4 0.6 0.8 1.0 1.2
V / Vh [-]
p /
bmep
[-]
Throttle Control
VCM
Figure 9: Indicator diagram with throttle control and VCM
5.1.2 Skip Firing
Skip firing (also known as cylinder cut-off) is being used on port injected gas engines for part load performance improvement, [10]. At constant speed engine operation without lV control, part load air/fuel mixture be-comes very lean, causing the combustion to become less and less com-plete with an accompanying increase in UHC emissions. If the engine is throttled to a reasonable air/fuel ratio, fuel consumption is high because of a deterioration in gas exchange work. As a solution, gas admission is turned off on some cylinders and their firing is skipped (no ignition, no combustion). In order to maintain the same engine power, the same amount of fuel gas is distributed among fewer cylinders which conse-quently run on a richer mixture. UHC emissions are kept low at improved gas exchange work. In the case of premix gas engines, the cylinder firing cannot be simply omitted by skipping ignition since the fuel gas is already admitted ahead of the turbocharger compressor(s). Evidently, if the fuel gas would pass the cylinders without combustion, UHC emission would be dramatically high, giving rise to the risk of explosion in the exhaust manifold. The only way to manage skip firing on premix engines is to keep the inlet valves closed during a whole working cycle, which can be enabled by VCM. If the necessary engine power can be provided by fewer cylinders, the mean brake effective pressure of the remaining cylinders can be raised accord-ing to Figure 10. For example, at 50% engine load and only half the cylin-ders firing (50% skipped cylinders), the cylinders fired would run at nomi-nal brake mean effective pressure. In order to increase the power of the fired cylinders at a certain engine operating point, cylinder filling has to be increased (at constant air/fuel ratio the amount of trapped fuel gas needs to be increased, Eq. 4). As a consequence, the Miller timing has to be re-
duced which eventually also leads to a reduction of the Miller losses in the gas exchange work. Since the engine is operated at the same power and air/fuel ratio, the turbocharger turbine does not notice the changed engine operating mode (assuming skip firing is being applied to an engine with a large number of cylinders, e.g. 8 to 12 cylinders per bank). The receiver pressure does not change either, because the effective engine flow area and the compressor power remain the same. The theoretical gas exchange work pRec – pTI is unchanged while the Miller losses are reduced and con-sequently the effective gas exchange work (see estimated gas exchange work in Figure 11, wherein points A, B and C show the optimal percentage of fired cylinders) and consequently engine efficiency are both improved. Because of knocking and mechanical constraints, the number of non-firing cylinders is limited to a certain extent.
0.0
0.2
0.4
0.6
0.8
1.0
0.0 0.2 0.4 0.6 0.8 1.0
Pe / Pe,100% [-]
bmep
/ bm
ep10
0% [-
] 100%80%60% 90%70%50%Firing Cylinders
A
B
C
0%
1%
2%
3%
4%
5%
40% 50% 60% 70% 80% 90% 100%
Fired Cylinders [%]
(Dp G
E-D
p GE,
ref)
/ bm
ep10
0% [-
]
Engine Load 75%
50%25%
C
A
B
Figure 10: Diagram with bmep per firing cylinder as function of engine load
Figure 11: Estimation of im-provement of gas exchange work as the number of fired cylinders is reduced
Simulation results regarding skip firing have been generated by simply carrying out the engine system calculations with a reduced number of cyl-inders, e.g. 7 instead of 10. It was assumed that the non-firing cylinders have a negligible indicated mean effective pressure. Friction mean effec-tive pressure arising from the non-fired cylinders is distributed among the fired cylinders. As shown in Figure 12, simulations predict a reduction in fuel consumption of more than 5 g/kWh in the whole region from idle up to 80% engine load compared to the reference engine with throttle control. Below 25% engine load the gain in engine efficiency is even more pronounced. The benefit in fuel consumption arises mainly from the improved gas exchange work shown in Figure 13 as the Miller losses are reduced at roughly con-stant receiver pressure.
-15
-10
-5
0
0 0.2 0.4 0.6 0.8 1
Pe / Pe,100% [-]
Dbsf
c 42.
7 [g/
kWh]
VCM
Skip 4 out of 10
Throttle Control
Skip 2 Skip 1Skip 3
-1.5
-1.0
-0.5
0.0
0.0 0.2 0.4 0.6 0.8 1.0
Pe / Pe,100% [-]
Gas
Ex.
Wor
k D
p GE
[bar
]
Throttle ControlVCMVCM, skipping 2 out of 10VCM, skipping 4 out of 10
Figure 12: Decrease of specific fuel consumption when changing from throttle control to VCM and skip firing
Figure 13: Gas exchange losses can be reduced by distributing the required power among a reduced number of cylinders
Cylinder cut-off is assumed to be beneficial for part load engine perform-ance. If the cylinder is completely throttled by advancing the IVC timing, process temperatures near engine idle become very low and combustion quality questionable. By making use of the VCM skip firing capability, the Miller effect can be reduced to a reasonable level – see Figure 14. Further, the cylinder brake mean effective pressure is lifted to a higher level which helps to reduce part load UHC emissions – see Figure 1. If skip firing is not applicable on a certain engine design, at low engine power throttling should be realized by non-Miller valve timing at reduced lift according to Figure 15.
450
500
550
600
650
10 20 30 40 50 60 70
pTDC [bar]
T TD
C [°
C]
Throttle ControlVCMSkipping 2 out of 10Skipping 4 out of 10
Engine Idle to 25% Power
Crank Angle [°]
Valv
e Li
ft [m
m]
Figure 14: Skip firing enables better conditions for combustion at low engine load
Figure 15: Variable inlet valve lift for enabling throttling with lower in-cylinder temperature reduc-tion
5.1.3 Cylinder Balancing
In order to obtain the maximum efficiency from an engine, every single cylinder needs to perform as close as possible to the knock limit with op-timum combustion timing. For many reasons, e.g. manufacturing toler-ances, pressure fluctuations on the intake and exhaust sides or compo-nent wear, etc., the trapped amount of fresh gas/air mixture will never be the same for all cylinders. With a variable valve train, cylinder filling can be balanced adaptively. However, the minimum allowable safety margin with regard to knock depends to a great extent on the cycle-to-cycle variation of cylinder filling. For this reason, in addition to a reliable high energy ignition system, highly accurate inlet valve closure timing is re-quired.
5.2 Engine Load Acceptance
In island power generation mode, where only a few power generation de-vices cover prevailing power demand (in the simplest case there is only one engine providing power) engine speed depends on the difference be-tween engine torque and load torque. For example, if the generator load is suddenly increased, engine crankshaft speed slows down, because the generator torque is higher than the one provided by the engine. Actuated by the engine speed governor, engine torque temporarily overcompen-sates generator torque demand in order to accelerate the engine crank-shaft back to the desired generator speed. The maximum frequency drop and its recovery time need to be within certain limits depending on the requirements of the power consumers in the grid. Genset engines are categorized according to ISO-8528 in terms of recovery time and fre-quency drop. In the case of today’s premix gas engines, torque can only be increased in accordance with the achievable increase in receiver pressure (Eq. 4), by opening the throttle valve (or closing the compressor bypass or the ex-haust waste gate, respectively). Thus, the more stringent the require-ments for engine load acceptance, the higher the necessary control mar-gin, e.g. higher throttle valve pressure drop. At engine steady state, the higher control margin results in reduced engine thermal efficiency. At high Miller effects, control margins at engine part load are revealed to be small, as can be shown in Figure 16 which shows the normalized throt-tle valve pressure drop for a variation of delivery ratio in the range of lR = 0.6-0.9. For each case, the turbocharger with ideal turbine and com-pressor characteristics has been adjusted in order to provide a constant ratio of compressor boost pressure and the required receiver pressure at full engine load. Based on Eq. 1, the relationship of Eq. 7 can be derived, which gives a rough estimation of how much engine torque can be imme-diately increased, starting from different engine load points (Figure 17). This example shows that compared to a non-Miller engine, load steps
achievable with high Miller effects are substantially reduced at engine part load.
( )1,1,
2,1,
1,1,
2,2,
1
2
threc
thThrrec
threc
threc
ppp
pp
bmepbmep
hh
hh ×D+
=» Eq. 7
0.0
0.1
0.2
0.3
0.0 0.2 0.4 0.6
Pe / Pe,100% [-]
Dp T
hrot
tle /
p Rec
,100
% [-
]
Delivery Ratio 0.9Delivery Ratio 0.8Delivery Ratio 0.7Delivery Ratio 0.6
0.0
0.1
0.2
0.3
0.0 0.2 0.4 0.6
Pe / Pe,100% [-]
DP e
/ P e
,100
% [-
]
Delivery Ratio 0.9Delivery Ratio 0.8Delivery Ratio 0.7Delivery Ratio 0.6
Figure 16: Normalized throttle valve pressure loss
Figure 17: Estimation of achiev-able engine power steps
In the case of variable inlet valve actuation, cylinder filling can be varied in a wide range as illustrated in Figure 18. After a load step, torque can be increased by changing from a low steady state delivery ratio to the maxi-mum one – limited by engine knocking – resulting in maximum cylinder power. The possible increase in engine torque can be estimated by the relationship shown in Eq. 8.
1,1,
2,2,
1
2
thR
thR
bmepbmep
hlhl
» Eq. 8
Figure 19 shows the benefit of variable inlet valve timing on engine torque rise capability. Since the torque steps shown are only based on rough es-timations without consideration of system transients, they shall be inter-preted as a qualitative comparison of the two control strategies. It can be seen that engine load steps are especially challenging at engine part load and that engine power control with variable valve train promises that achievable load steps can be substantially increased in the critical range.
0.0
0.2
0.4
0.6
0.8
1.0
0.0 0.2 0.4 0.6 0.8 1.0Pe / Pe,100% [-]
Del
iver
y R
atio
lR [-
] Decrease of Miller Effect
VCMConventional Control
0.2
0.0
0.2
0.4
0.6
0.8
0.0 0.2 0.4 0.6 0.8 1.0Pe / Pe,100% [-]
DP e
/ P e
,100
% [-
]
VCM
Conventional Control
Knocking Limit
Figure 18: Increased delivery ra-tio lR by reduction of Miller timing
Figure 19: Estimation for achiev-able power step
As expected from the estimates shown in Figure 19, engine simulations confirm a substantial improvement in transient engine performance. Figure 20 shows engine speed during engine load imposition from idle to full load in four equal steps. The difference of the two control concepts is obvious for the first two load steps where the engine speed drop and re-covery time are massively reduced with VCM, while the knock reference value has not been exceeded in both cases. If a certain limit of engine speed drop and recovery time is set as a con-straint, e.g. according to ISO-8528 performance class G2, the maximum achievable load steps can be evaluated. In Figure 21, the maximum load steps starting from a) engine idle and b) 16.5% engine power (range of lowest throttle control margin) are compared. From both starting points, VCM allows steps of roughly 25% engine load, while throttle control only allows steps of 10% or even less.
0.5
0.6
0.7
0.8
0.9
1.0
1.1
1.2
Time [s]
n / n
100%
[-]
VCM
Throttle Control10 s
0.00
0.05
0.10
0.15
0.20
0.25
0.30
0.35
0.40
Step from Idle Step from 16.5% Load
Load
Ste
p [-]
Throttle ControlVCM
Figure 20: Load imposition from idle to full load in four equal steps
Figure 21: Maximum load steps for given limits of speed drop dfdyn = 10% and recovery time tf,in = 5 s
In the simulation results shown above no enrichment of the admitted air/fuel mixture has been considered. While the additional potential for
VCM is expected to be small (limited by knocking even at part load be-cause of the reduced Miller effect during transients) the throttle control version would benefit from the reduced air/fuel ratio. Although variable valve actuation systems also have some potential for improved load rejection behavior, e.g. increased gas exchange pumping losses, for fast load rejection a feed-forward controlled compressor bypass valve will be necessary in order to avoid turbocharger surging.
5.3 Engine Operation on a Propeller Curve
In the foregoing section it has been shown how the control margin of en-gines operated at constant speed suffers from strong Miller effects at en-gine part load. This section will show that, as engine speed is reduced for a given engine power, e.g. fixed pitch propeller (FPP) operation mode, en-gine operation with conventional control devices becomes even more chal-lenging. Eq. 9 recalls the fact that for a premix gas engine at constant air/fuel ratio lV, engine torque mainly depends on receiver pressure and delivery ratio lR. At constant speed engine operation, torque increases in linear progres-sion with engine power, Eq. 10. However in FFP mode, since engine speed is reduced, at a given engine power the brake mean effective pressure is higher than the corresponding one at constant speed engine operation, Eq. 11. The stated equations of bmep as function of engine power are plotted in the diagram of Figure 22.
ecRR pbmep lµ Eq. 9
Constant speed engine operation mode:
%100%100,%100%100, bmepbmep
nbmepnbmep
PP
Eng
Eng
e
e == Eq. 10
FPP engine operation mode:
32
%100,%100
23
%100%100,%100%100,÷÷ø
öççè
æ=Û÷÷
ø
öççè
æ==
e
e
Eng
Eng
e
e
PP
bmepbmep
bmepbmep
nbmepnbmep
PP
Eq. 11
The ratio of brake mean effective pressure resulting from constant speed and FPP engine operation:
GENecR
FPPecR
e
e
GEN
FPPGENFPP p
pP
Pbmepbmep
,
,3
1
%100,/ =÷
÷ø
öççè
æ==
-
a
Eq. 12
According to Eq. 9 and Eq. 12, the latter being derived from Eq. 10 and Eq. 11, the necessary receiver pressure for a given engine power is in-creased by the factor aFPP,GEN which increases significantly towards engine
part load, e.g. aFPP,GEN(50%) = 1.26, aFPP,GEN(25%) = 1.59. This ratio is also plotted in Figure 22 as red solid line. Figure 23 shows the normalized pressures before and after the throttle valve, taken from the constant speed example given in the foregoing section; the resulting throttle valve pressure drop is shown in Figure 25 as a solid black line. Since engine torque at a certain air/fuel ratio and constant valve timing is proportional to the receiver pressure pRec, the receiver pressure required for FPP opera-tion increases by the factor aFPP,GEN. On the other hand, the pressure be-fore the throttle valve will remain on a similar level, because the turbine receives the same mass flow at constant power and air/fuel ratio; only fuel consumption will be somewhat lower in FPP mode. The FPP mode throttle valve pressure drop resulting from the pressure before throttle and the theoretically required FPP receiver pressure (pRec,GEN multiplied by aFPP,GEN), is also shown in Figure 25 as solid red line. The FPP control mar-gin drops to zero at about 50% rated power and below. Hence, the engine cannot be operated below 50% engine power. If the cylinder filling is in-creased by a variable inlet valve train (eliminating the Miller effect) the necessary FPP receiver pressure can be reduced (Eq. 9); see the red dashed line in Figure 23. The resulting increase in throttle valve pressure drop is shown in Figure 25. It clearly shows that VCM enables operation of the engine on a propeller curve. In this example the delivery ratio is in-creased according to Figure 24. Of course, the issue regarding insufficient charge pressure can be somewhat alleviated by lowering the premix air/fuel ratio and/or retarding ignition, but both measures are at the ex-pense of increased fuel consumption and are only possible within a limited range. Indeed, in any case combustion will be retarded by a certain de-gree to avoid engine knocking.
0.0
0.2
0.4
0.6
0.8
1.0
0.0 0.2 0.4 0.6 0.8 1.0Pe / Pe,100% [-]
bmep
/ be
mp 1
00%
[-]
1.0
1.2
1.4
1.6
1.8
2.0
bmep
FPP
/ bm
epG
EN [-
]
Generator ModeFPP Modebmep-Ratio
0.00
0.25
0.50
0.75
1.00
1.25
1.50
0.0 0.2 0.4 0.6 0.8 1.0Pe / Pe,100% [bar]
p b,T
hr ,
p Rec
/ p R
ec,1
00%
[-]
Const. Speed Pressure before ThrottleConst. Speed Rec. PressureFPP Rec. PressureFPP Rec. Pressure VCM
Figure 22: Comparison of bmep in FPP and constant speed engine operating mode as a function of engine power
Figure 23: Required receiver pressure for generator and FPP engine operation
1.0
1.2
1.4
1.6
1.8
2.0
0.0 0.2 0.4 0.6 0.8 1.0Pe / Pe,100% [-]
l R,V
CM
/ l R
,ref [
-]
0.0
0.1
0.2
0.3
0.4
0.0 0.2 0.4 0.6 0.8 1.0Pe / Pe,100% [bar]
Dp T
hr /
p Rec
,100
% [-
]
Constant Speed, Throttle ControlFPP, Throttle ControlFPP, VCM
VCM
Figure 24: Example for variable cylinder filing for FPP engine op-eration at high nominal Miller timing
Figure 25: Throttle valve pres-sure drop according to Figure 23
For engine simulations in FPP operating mode (see results in diagrams of Figure 26), the air/fuel ratio has been kept at a constant value of lV = 2.0 (simulation cases VCM and Throttle A). As expected from the estimation calculation above, because of the increased part load mean effective pres-sure compared to constant speed operating mode, the necessary receiver pressure is increased and the throttle valve control margin disappears in the region below 50% engine load. In order to decrease the required re-ceiver pressure at engine part load, the premixed air/fuel ratio lV,tot is en-riched towards part load on two different gradients (simulation cases Throttle B and C). In order to achieve a minimal throttle control margin (first diagram of Figure 26) a reduction in the air/fuel ratio substantially below lV = 1.5 is necessary. The corresponding increase in in-cylinder temperature leads to higher heat losses which are reflected in a noticeable increase in fuel consumption. A further increase in control margin could be achieved by retarding combustion, but this would further decrease engine efficiency and increase turbine inlet temperatures to an even higher level. In Figure 27, the necessary variation of inlet valve closure timing is plot-ted against engine power. At engine part load, due to the reduced Miller effect, the start of combustion has to be retarded in order to maintain the knock margin.
0.0
0.1
0.2
0.3
0 0.2 0.4 0.6 0.8 1 1.2Pe / Pe,100
dpTh
/ Pre
c [-]
VCM
Throttle Control A
Throttle Control B
Throttle Control C
1.0
1.5
2.0
2.5
0 0.2 0.4 0.6 0.8 1 1.2Pe / Pe,100 [-]
Air/
Fuel
Rat
io l
V [-]
VCM
Throttle Control A
Throttle Control B
Throttle Control C
400
500
600
700
800
0 0.2 0.4 0.6 0.8 1 1.2Pe / Pe,100 [-]
Turb
ine
Inle
t Tem
p. T
TI [°
C] VCM
Throttle Control A
Throttle Control B
Throttle Control C
0
5
10
15
20
25
30
0 0.2 0.4 0.6 0.8 1 1.2Pe / Pe,100 [-]
Dbs
fc [g
/kW
h]
VCM
Throttle Control A
Throttle Control B
Throttle Control C
Figure 26: FPP engine simulation results
-20
-10
0
10
20
30
40
0 0.2 0.4 0.6 0.8 1 1.2Pe / Pe,100 [-]
IVC
- IV
Cre
f [°C
A]
0
2
4
6
8
10
12
Df S
OC [°
CA
]
Shift of IVC timing
Shift of combustion start
Figure 27: Valve timing and start of combustion
6 Requirements on the Variable Valve Train
Based on the foregoing chapters, the requirements on the valve train sys-tem for enabling power control of premix gas engines by the use of the gas exchange organs are summarized. Beside the basic valve train system requirements, e.g. reliability, durability, ease of maintenance, etc., the main requirements on valve actuation variability are listed in Table 6-1.
Table 6-1: Requirements for valve train system for power control of premix gas engines
Engine Ability Valve Train Requirement
Cylinder power control via valve timing Step-less and cycle-to-cycle variable ad-justment of inlet valve closure timing
Cylinder power control by valve timing at part load
Step-less and cycle-to-cycle variable ad-justment of inlet valve lift, allowing for re-duced cylinder filling at reduced Miller ef-fect
Slight cycle to cycle deviation of cylinder filling
Slight variation in inlet valve closure timing
Control of cylinder scavenging Variable inlet valve opening or variable exhaust valve closure timing
Skip firing Cam lift without valve lift
Cylinder balancing Cylinder individual valve control
7 VCM: ABB’s New Valve Control Management
In cooperation with Schaeffler Technologies Gmbh & Co. KG, a well-known specialist for automotive engine components, ABB began developing a variable valve train system in 2009. It is based on Schaeffler’s UniAir/MultiAir® technology ( [1], [5]) and meets the requirements of 4-stroke engines with power outputs above 400 kW. The resulting product, VCM (Valve Control Management), enables variations in the timing and lift of the inlet and/or exhaust valves by means of electro-hydraulic valve actuation. As will be shown in the following sections, the VCM valve train system fulfills all the requirements listed in Table 6-1.
7.1 VCM Working Principle
The working principle of VCM, illustrated in the sketch of Figure 28, is ex-plained in the following paragraphs. Likewise, the achievable valve lift pro-files are schematically displayed in Figure 29. a) Full Lift Mode The rotating camshaft (1) drives a hydraulic pump (2), thus increasing the oil pressure in the high pressure chamber (3). As the pressure on the valve actuator unit (4) is increased, the valve is opened against the valve spring, gas and friction forces. During the descent of the cam profile, the valve springs return their stored energy to the camshaft.
b) Late Valve Opening / Reduced Valve Lift, f1-Mode If the control device (5) is open during actuation of the pump unit, the valve does not open. The increased oil pressure does not cause a move-ment of the valve, moving the pressure accumulator piston (6) instead. Thus, valve opening can be retarded (or even omitted). In addition to re-tarded opening, valve lift is reduced because the cam is no longer able to provide the necessary oil quantity for full valve lift. During the descent of the cam profile the valve springs return the energy stored to the cam-shaft. c) Early Valve Closure, f2-Mode If the valve control device is opened, the oil pressure changes according to the pressure accumulator spring characteristic. If the control device is opened during an open valve period, the valve closes as a result of the sudden drop in oil pressure. The pressure accumulator is charged. During the descent of the cam profile, the pressure accumulator returns the en-ergy stored to the camshaft and the high pressure oil chamber is refilled with the correct amount of oil. In every case, during the valves’ closing phase, the hydraulic brake (4) acts as a damping element for a smooth touch-down of the valves on their seats.
12
3
4 5
67
1122
33
44 55
66677
1) Cam
2) Pump unit
3) High pressure chamber
4) Actuator / brake unit
5) Control device (solenoid valve)
6) Pressure accumulator
7) Oil supply
Figure 28: VCM working principle sketch
Crank Angle
Valv
e Li
ft
a)
b) c)
f1
f2
Figure 29: VCM working modes: a) full valve lift, b) late valve opening (f1-mode), c) early valve closure (f2-mode)
7.2 Engine Test Bed Experience
A prototype module of the variable valve train system VCM, shown in Figure 30, has been tested on a small medium speed diesel engine on a test bed with the aim of checking its functionality and mechanical integ-rity. Such a VCM module mounted on the fired test engine is shown in Figure 31.
Figure 30: VCM module proto-type with variable inlet and ex-haust valves
Figure 31: VCM module mounted on the diesel test engine
On-engine testing of the VCM system demonstrated its maturity for indus-trial application. Valve timings can be changed in line with engine load ac-cording to predefined and freely editable timing maps for exhaust and inlet valve closure timings during steady and transient engine operation. Measurement results with regard to variable Miller timing and valve over-lap show a substantial potential for reduction of specific fuel consumption, NOx and soot emissions within the whole engine load profile in FPP and
constant speed operation, compared to an engine with a conventional camshaft. The positive impact of VCM on engine transient behavior was also demonstrated. A selection of measured valve lift curves on the fired test engine is shown in Figure 32. Endurance tests on the fired engine (> 300 running hours) and on the mechanical test bed (> 1000 running hours) have proven the reliable operation with high accuracy in valve timing events (±1.5 °CA) even at changing lubricating oil temperatures, even at varying lubrication oil temperatures.
0
5
10
15
20
25
90 180 270 360 450 540 630
Crank Angle [°]
Valv
e Li
ft [m
m]
Figure 32: Post-processed valve lift measurements for various IVC and EVC timings
The industrial application of many highly variable valve trains was ham-pered by their increased power consumption. Corresponding measure-ments of VCM camshaft power consumption on a fired engine are ongoing. However, tests already performed on a non-fired cylinder head assembly indicate the power consumption to be much lower than the values quoted for other valve train systems that provide a similar degree of freedom in variability. In addition, VCM allows for steeper valve closure gradients since its hydraulic brake ensures a soft landing for the valves. This results in reduced gas exchange flow losses, which partially compensate the in-creased power demand of VCM.
8 Summary and Outlook
The introduction of two-stage turbocharging (Power2) has opened consid-erable development potential for gas engines in terms of power density and efficiency [3]. In the present document, it has been investigated by simulation how variable valve actuation can contribute to further en-hancements in the already attractive attributes of modern premix gas en-gines. In the investigated case studies, simulation results showed sub-stantial improvements in engine efficiency and engine load acceptance. It is demonstrated how skip firing can be enabled on premix gas engines and how engine performance can be improved with regard to fuel consumption and UHC emission, according to engine cycle simulations.
ABB’s variable valve system VCM fulfills the requirements of engine power control based on a variable valve train system while experience with VCM on a fired engine has proven its readiness for industrial application. Further studies will focus on the benefit of variable valve train systems for the dual-fuel engines which are expected to be an interesting option in achieving compliance with the IMO Tier III, the third stage of emission limitations from the International Maritime Organization.
Acknowledgements
The authors would like to thank Messrs. M. Haas, M. Berger and their team of Schaeffler Technologies GmbH & Co. KG, Herzogenaurach, Ger-many, for their valuable contribution in the design and development of the VCM system presented here.
9 Nomenclature
BDC/TDC Piston bottom/top dead center
bmep Brake mean effective pressure [bar]
FPP Fixed pitch propeller
Hu Lower heating value [J/kg]
IK Knock integral [-]
IVO /IVC Inlet valve opening / closure timing [°CA]
Lmin Stoichiometric air/fuel ratio [kg/kg]
m Mass [kg]
p Pressure [bar]
Pe Engine power [kW]
R Specific gas constant [J/kg/K]
T Temperature [K]
tt,in Frequency recovery time requirement for increasing load [s]
UHC Unburned Hydrocarbons
VCM Valve Control Management
Vh Cylinder displacement volume [m3]
VTG Variable turbine geometry
VVT Variable valve train
X1,2,3 Knock model constants
dfdyn Transient frequency deviation from rated frequency [%]
f Crank Angle [°]
f1 / f2 Late valve opening / Early valve closing mode (VCM specific)
hth Engine brake thermal efficiency [-]
hTC,2-st,eq Equivalent 2-stage turbocharger efficiency [-]
ll Charging efficiency [-]
lR Delivery ratio [-]
lV, lV,tot Trapped air/fuel ratio, total air/fuel ratio [-]
p*C Compressor compression ratio (total to total pressure) [-]
r Density [kg/m3]
tKnock Time until autoignition [ms]
Subscripts
ac Start of compression (BDC)
GE Gas exchange
Rec State of intake receiver
Thr Throttle valve
TI Turbine inlet
10 References
[1] Bernard, L. et al., Elektrohydraulische Ventilsteuerung mit dem Mul-tiAir-Verfahren, MTZ 12/2009
[2] Boewing, R, Plohberger, D., Thermodynamic Optimisation of three Gas Engine, Families for Higher Efficiency, Paper No. 126, CIMAC World Congress 2010, Bergen
[3] Codan, E., Vögeli, S., Mathey, C., Hochdruckaufladung bei Gasmoto-ren, 13. Aufladetechnische Konferenz 2008, Dresden
[4] Dimitrov D. et al., Eine Methode zur Vorausberechnung des Klopf-verhaltens von Gasmotoren, 4. Dessauer Gasmotoren-Konferenz, 2005
[5] Haas, M., Rauch, M., Elektrohydraulischer Vollvariabler Ventiltrieb, MTZ 03/2010
[6] M. Haidn, J. Klausner, J. Lang, Dr. Ch. Trapp, Zweistufige Hoch-druck-Turboaufladung für Gasmotoren mit hohem Wirkungsgrad, 15. Aufladetechnische Konferenz 2010, Dresden
[7] Mathey, C., Variable Valve Timing – A necessity for future large die-sel and gas engines, Paper No. 298, CIMAC Congress 2010, Bergen
[8] Nerheim, L., Nordrik, R, Bergen gas engine development, Paper No. 71, CIMAC World Congress 2004, Kobe
[9] Ninkovic, D., et al., The ABB Turbo Systems Program Suite for the Design and Optimization of Turbocharging Systems, Proc. Interna-tional Congress Motor Vehicles & Motors MVM 2006, Kragujevac, Serbia
[10] Nylund, I., Status and potentials of the gas engines, Paper No. 163, CIMAC World Congress 2010, Bergen
[11] Schmuttermair, H., et al., Optimierung von Gasmotoren durch Ein-satz variabler Turbinengeometrien, 13. Aufladetechnische Konferenz 2008, Dresden
[12] Soylu, S., Prediction of knock limited operating conditions of a natu-ral gas engine, Elsevier Science Ltd., 2003
[13] Humerfelt, T., Johannessen, E., Vaktskjold, E., Skarbö, L. , Devel-opment of the Rolls-Royce C26:33, Marine Gas Engine Series, Paper No. 54, CIMAC World Congress 2010, Bergen
