SAE TECHNICAL PAPER SERIES
2006-01-0040
2-step Variable Valve Actuation: System Optimization and Integration on an SI Engine M. Sellnau, T. Kunz, J. Sinnamon and J. Burkhard Delphi Corporation
Reprinted From: Variable Valve Actuation 2006 (SP-2007)
2006 SAE World Congress Detroit, Michigan April 3-6, 2006 400 Commonwealth Drive, Warrendale, PA 15096-0001 U.S.A. Tel: (724) 776-4841 Fax: (724) 776-5760 Web: www.sae.org
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2006-01-0040
2-step Variable Valve Actuation: System Optimization and Integration on an SI Engine M. Sellnau, T. Kunz, J. Sinnamon and J. Burkhard Delphi Corporation
Copyright © 2006 SAE International
improvement was measured. Further improvements in vehicle fuel economy are expected with refinement in transient control and calibration.
ABSTRACT 2-step variable valve actuation using early-intake valve closing is a strategy for high fuel economy on sparkignited gasoline engines. Two discrete valve-lift profiles are used with continuously variable cam phasing. 2-step VVA systems are attractive because of their low cost/benefit, relative simplicity, and ease-of-packaging on new and existing engines.
INTRODUCTION Compromises inherent with fixed valve events on sparkignited engines have prompted engine manufacturers to consider variable valve actuation (VVA). Continuouslyvariable valve lift systems vary the valve lift and timing continuously over the operating range and usually employ early-intake-valve-closing (EIVC) as the primary strategy to reduce pumping losses and improve fuel economy. 2-step valve lift systems, which control valve lift using a low lift cam and a high lift cam, have received considerable interest due to their relative simplicity and ease-of-application. 2-step VVA has been used to enhance performance [1-4], deactivate one valve on a 4valve-per-cylinder engine to produce swirl [5,6], and to deactivate half the cylinders of an engine [7-9] for improved fuel economy.
A 2-step VVA system was designed and integrated on a 4-valve-per-cylinder 4.2L line-6 engine. Simulation tools were used to develop valve lift profiles for high fuel economy and low NOx emissions. The intake lift profiles had equal lift for both valves and were designed for high airflow & residual capacity in order to minimize valvetrain switching during the EPA drive cycle. It was determined that an enhanced combustion system was needed to maximize fuel economy benefit with the selected valve lift profiles. A flow-efficient chamber mask was developed to increase in-cylinder tumble motion and combustion rates.
2-step VVA may also be used with EIVC [2,10,11] and is a preferred strategy for fuel economy with the goal of providing most of the fuel economy benefit of continuous VVA at significantly lower cost. This approach also has the potential to simultaneously increase low-speed torque, while applying a late-intake-valve-opening (LIVO) strategy during the cold start for reduced HC. The main technical issues surrounding this approach are: 1. Optimization of valve lift profiles 2. Determination of cam phaser authority and timing 3. Understanding the effects of combustion enhancement on fuel economy and how it can be achieved.
A 2-step valvetrain mechanism was developed that features hydraulically-actuated switchable rocker arms and hydraulic lash adjusters (Type II valvetrain). The rocker arm is a dual-roller, single-slider design for compact packaging and low friction. The engine management system was modified for control and calibration of 2-step VVA, and to realize the full fuel economy potential of the system. Dynamometer tests on a multicylinder engine indicated a 6.9 percent fuel economy benefit relative to the production engine with exhaust cam phasing alone on the EPA city cycle. Combustion enhancement significantly contributed to the overall fuel economy benefit. Vehicle tests showed less fuel economy improvement than steady state dynamometer tests due, in part, to cam phaser control limitations. For warmedup Phase 3 EPA tests (cycles 19-23), a 5.5% improvement was measured with 46% reduction in NOx. For the whole EPA city test including cold start, a 4.8%
The objectives of the current work were to apply 2-step VVA using an EIVC strategy on a 4.2L line-6 sparkignited engine and optimize the system for fuel economy. An additional objective was to investigate reduced cold-start hydrocarbon (HC) emissions using a 1
cam profile on fuel economy was estimated for the EPA drive cycle using a 4-point approximation. The four steady-state test points were determined based on a drive cycle analysis for the production vehicle and are shown with fuel weighting factors in Table 1. A fifth point was used for simulation of idle-neutral conditions (see Table 1).
late-intake-valve-opening (LIVO) strategy [10]. This paper presents one approach to system optimization using equal lift for both intake valves of 4-valve-percylinder engines. Throughout the course of this work, emphasis was on integration of the valvetrain system with the engine management system (EMS), and demonstration of the total system on a vehicle. An electro-hydraulicallyactuated 2-step valvetrain system with intake and exhaust cam phasers was developed to meet requirements for high-volume series production. Control algorithms for valvetrain switching and calibration were also developed to realize maximum potential of the system on a vehicle. This involved both engine dynamometer and vehicle testing, supported by steady state and transient engine simulation.
Engine Speed NMEP % of Total Drive (rev/min) (kPa) Cycle Fuel 650 150 9.2 1050 200 24.5 1550 350 43.6 1900 650 22.7 650 100 idle-neutral Table 1. Speed-load operating conditions for simulation
Simulation results for fuel economy and NSNOx are shown in Figures 2 and 3, respectively. For each valve profile, results reflect the intake and exhaust timing that produced the highest thermal efficiency and nearly lowest NSNOx. The baseline for this comparison is the production engine with exhaust cam phasing only and no EGR. The reference case for the production engine with dual-independent cam phasing is also shown.
PRELIMINARY MODELING & SIMULATION A model of the 4.2L line-6 engine was developed using GT Power [12]. Combustion was modeled using Wiebe functions assuming that combustion duration is a function of dilution but is independent of valve lift [10]. Wiebe [13] parameters were fit to cylinder-pressurebased combustion data obtained during initial baseline testing of the production engine. Initially, it was assumed that the expected combustion deterioration at low valve lifts would be recovered with combustion system enhancements.
Results show that low-mode profiles with maximum lift of 6mm (total duration of about 204 crank angle degrees (CAD)) provided about 5.1 percent improvement in fuel economy relative to the production baseline with exhaust cam phasing only. However, NSNOx emissions were compromised because this profile does not permit optimum dilution levels during high load portions of the drive cycle. Simulation results indicated that higher lifts produced NOx closer to the production baseline with a small fuel economy tradeoff.
Fuel Econ. Benefit (%)
Flow characteristics for the production cylinder head were measured on the steady flow bench and inputted into the engine model. Tumble index as a function of valve lift is presented in Figure 1 and shows the expected tumble decrease as valve lift is reduced.
Tumble Index
0.8 10 mm
0.6 0.4
7 mm 5 mm
0.2 0 0
2
4
6
8
10
6 5 4 3 2 Production Cams with DICP
1 0 4
12
Valve Lift (mm)
5
6
7
8
9
10
11
Valve Lift (mm) Figure 2. Effect of Valve Lift on Fuel Economy Estimated from 4-Point Drive Cycle Simulation.
Figure 1. Tumble index for production cylinder head as measured on steady flow bench. Simulations were performed for a range of intake valve lifts at various intake and exhaust cam timings using GT Power and a custom software tool [10]. A family of valve lift profiles was designed for this purpose (see later section on valvetrain system). The effect of the low lift 2
efficiency required some exhaust retard for increased valve overlap and increased mixture dilution within combustion stability limits (COV IMEP: ~3%). Thermal efficiency was improved about 6.5 percent for both lowmode profiles at this operating condition. Exhaust retard was also important to achieve low NSNOx. NOx at peak efficiency was comparable for the three lift profiles as shown in Figure 7.
NSNOx Increase (%)
100 75 50 Production Cams with DICP
25 0 -25
EVO=79 EVO=117
4
5
6
7
8
9
10
11
Net Therm Eff (%)
Figure 3. Effect of Valve Lift on NSNOX Estimated from 4-Point Drive Cycle Simulation.
ENGINE DYNAMOMETER TEST RESULTS – PRODUCTION CYLINDER HEAD Based on preliminary simulation results, it was decided to modify and test a production 4.2L L6 engine. Intake camshafts were fabricated with low mode cam profiles of 5mm and separately, 7mm, and an intake cam phaser was fitted on the engine. The production exhaust cam profile and exhaust cam phaser were retained. The valve lift profiles and their range of timing variation are shown in Figure 4.
Lift (mm)
2-Step L= 5.0
4 3 2 1 0 270
Two Step L= 5.0 mm
33
Two Step L= 7.0 mm
32
Production Cam L= 10.4 mm
Exhaust Retard
310
450
540
350
Two Step L= 5.0 mm
-20 -30
Exhaust Retard
Two Step L= 7.0 mm
-40 360
330
-10
6 5
+ 6.5%
34
0
2-Step L= 7.0
7
Adv anc e
Figure 5. Measured Net Thermal Efficiency for 5mm, 7mm and Production Lift Profiles as a Function of Intake and Exhaust Timing at 1550 rpm, 350 NMEP.
Production Intake Cam
8
35
IVO (Crank Deg)
NPMEP (kPa)
9
EVO=107
Inta ke
31 290
11
Production Exhaust Cam
EVO=97 EVO=127
36
Valve Lift (mm)
10
EVO=87 EVO=122
Production Cam L= 10.4 mm
630
Crank Angle
-50 290
Figure 4. Valve Lift Profiles Used During Testing of Production 4.2L L6 Engine.
310
330
350
IVO (Crank Deg)
Part-load engine dynamometer results at 1550 rpm, 350 kPa net mean effective pressure (NMEP) are shown in Figures 5 through 7. This speed-load condition represents a region encountered heavily by the production vehicle during the EPA City test cycle. The figures show test results for the 5mm, 7mm, and production cam profiles as a function of intake and exhaust cam timing. Intake valve opening (IVO) and exhaust valve opening (EVO) are referenced at zero cam lift. As expected, operation using shorter duration, lower lift profiles reduced net pumping work (EIVC effect) and resulted in improved net thermal efficiency relative to the production lift profile. Achievement of best
Figure 6. Measured Net PMEP for 5mm, 7mm and Production Lift Profiles as a Function of Intake and Exhaust Timing at 1550 rpm, 350 NMEP
3
Lost Dilution Tolerance Exhaust Retard
Production Cam L= 10.4 mm
15
Burn Duration (Crank Deg)
NSNOx (g/kW-hr)
20 Two Step L= 7.0 mm
Two Step L= 5.0 mm
10 5
Intake Advance
0 290
310
330
350
60 COV > 3 %
50
5.0 mm
40 30
0 - 10% MBF
20
10 - 75% MBF
7.0 mm
L= 10.4 mm Production
10 0
IVO (Crank Deg)
5
10
15
20
25
30
35
Residual (% Mass) Figure 7. Measured NSNOx for 5mm, 7mm and Production Lift Profiles as a Function of Intake and Exhaust Timing at 1550 rpm, 350 NMEP.
Figure 8. Combustion Duration for 5mm, 7mm and Production Lift Profiles as a Function of Charge Dilution at 1550 rpm, 350 NMEP.
Combustion Characteristics Tumble Ratio at 90 BTDC
Combustion duration for 0-10 and 10-75 percent mass burn fractions are plotted as a function of charge dilution in Figure 8. Dilution was computed using simulation. As expected, a significant amount of combustion degradation was suffered with shorter-duration, lower-lift profiles. The 0-10 duration was increased by at least 5 CAD and up to 12 CAD depending on dilution, while the 10-75 main duration was increased at most 3 CAD. The main impact of combustion degradation was reduced dilution tolerance, which was reduced by 7 to 10 percent dilution. This is expected to significantly limit intake cam advance, the EIVC effect, and the associated fuel economy improvement attributed to 2-step VVA at combustion-limited operating conditions [10].
0.7 0.6 0.5 0.4
EVO=79 EVO=107
Production Cam L= 10.4 mm
EVO=127
0.3 0.2 0.1
Two Step Cam L= 7.0 mm Two Step Cam L= 5.0 mm
0.0 -0.1
290
300
310
320
330
340
IVO (Crank Deg)
Combustion degradation for EIVC may be caused by several factors including reduced in-cylinder tumble, increased time for tumble decay, and lower effective compression ratio (ECR), which produces lower mixture temperatures [14]. Figure 9 shows in-cylinder tumble at 90 CAD BTDC for the three lift profiles from GT Power simulation. Mixture motion is nearly absent for low lift profiles, so an enhanced combustion system, with higher tumble for low-lift profiles, would likely yield significant improvements in fuel economy.
Figure 9. Tumble Ratio for 5mm, 7mm and Production Lift Profiles with Production Cylinder Head as a Function of Intake and Exhaust Timing at 1550 rpm, 350 NMEP (Simulated). 4-Point EPA Fuel Economy Estimate Fuel economy for EPA drive cycle was estimated using the four speed-load points listed in Table 1. Results are presented in Tables 2 and 3. For the 5mm and 7mm profiles, fuel economy was improved by an estimated 4.5 percent and 4.8 percent, respectively. The baseline in this comparison was the production engine with exhaust cam phasing only.
4
Speed/Load Point 650/150 1050/200 1550/350 1900/650
% of Total Drive Cycle Fuel
Net Thermal Efficiency Production (%)
9.2 24.5 43.6 22.7 100.0
23.8 27.1 33.5 37.5
airflow rate [15,16,17]. Therefore chamber masks must be developed carefully to minimize the airflow penalty.
Net Thermal Efficiency % Drive 2-Step (%) Improvement Cycle Fuel 24.9 29.3 35.2 37.7
4.8 8.0 5.3 0.4
FTP Urban Cycle Improvement (%) =
An alternate approach for combustion enhancement is differential timings and/or lift of the two intake valves. By differentially controlling the lift of the two intake valves at any crank position, intake swirl can be created in low mode without an attendant reduction in airflow in high mode. However, when operating in low mode, the airflow and residual capacity of the engine would be significantly reduced for any total intake duration. Therefore this approach was not preferred for best fuel economy.
8.8 22.7 41.4 22.6 95.5
4.5
Table 2. 4-Point Fuel Economy Estimate for 5mm Lift Profile on EPA City Cycle.
Speed/Load Point 650/150 1050/200 1550/350 1900/650
% of Total Drive Cycle Fuel
Net Thermal Efficiency Production (%)
9.2 24.5 43.6 22.7 100.0
23.8 27.1 33.5 37.5
Net Thermal Efficiency % Drive 2-Step (%) Improvement Cycle Fuel 24.3 29.2 35.4 38.5
2.1 7.5 5.6 2.6
FTP Urban Cycle Improvement (%) =
Chamber Mask Development An experimental study was conducted on the flow bench to determine the effect of mask geometry on tumble and airflow. The parameters used to define chamber mask geometry are shown in Figure 10. Mask height, H, was varied from 2 to 6mm and mask clearance, C, was varied from 0.5 to 1.5mm for mask arc-lengths, T of 90 and 180 degrees.
9.0 22.8 41.3 22.1 95.2
4.8
Table 3. 4-Point Fuel Economy Estimate for 7mm Lift Profile on EPA City Cycle
Engine simulation was used to evaluate the tradeoff between increased mixture motion and reduced full-load performance. In Figure 11, tumble ratio at part load using the 7.0 mm cam profile is plotted against full load BMEP using the production high-lift profile. A wide range in estimated tumble ratio was observed with corresponding reduction in peak power. Results show that 90-degree masks are more efficient than 180-deg masks, and significantly improve the tumble-airflow tradeoff.
COMBUSTION SYSTEM ENHANCEMENT It has been shown that operation on short-duration, lowlift profiles degrades combustion and reduces combustion dilution limits. Consequently, enhanced combustion has the potential to increase the fuel economy benefit attributed to 2-step VVA, while reducing NOx emissions. Combustion enhancements will have greatest benefit at operating conditions that are combustion-limited. A condition is said to be combustion-limited when the dilution limit is reached at intake timings retarded from best efficiency timings for which pumping loss is minimized [10]. Since low-mode cam profiles with longer duration operate under combustion-limited conditions over a greater portion of the drive cycle, they will benefit to a greater degree from combustion enhancement.
Mask design IV was selected to moderately increase incylinder tumble motion. This 90-degree mask featured a height of 3mm and a clearance of 1mm. Tumble index as a function of valve lift for this mask and the production cylinder head is presented in Figure 12. As shown, chamber masks generally produce increased tumble at low valve lifts (for low mode operation), with tumble generally unchanged at high valve lifts. This is a good match to needed flow characteristics for 2-step VVA applications. Mask IV produced a tumble ratio of 1.4 with a predicted peak power reduction of 2.4 percent (simulation). Figure 13 shows a CAD model of Mask IV design.
Enhanced combustion also enables leaner mixtures and more retarded combustion for cold start conditions for reduced HC emissions. This combines with the 2-step late-intake-valve-opening (LIVO) strategy presented in a later section of this paper. Enhanced combustion via increased mixture motion may also reduce knock propensity. For operation at high loads on the short-duration, low-mode profile, effective compression ratio is increased. This increases the propensity for combustion knock, thus potentially limiting low and mid-speed torque. The method chosen to enhance combustion was combustion chamber masking. Chamber masks are known to increase tumble motion with some tradeoff in 5
Figure 13. CAD Model Showing Chamber Mask IV Design. Valve Lift Profiles for Enhanced Combustion System Figure 10. Parameters for Definition of Chamber Mask Geometry.
Results from simulation and engine testing were used to select intake and exhaust lift profiles for the enhanced combustion system. Zero-dimensional simulation codes, such as GT Power, cannot predict the improvement of combustion burn rates based on flow bench measurements of tumble or swirl. The enhancement of combustion with increased mixture motion is highly application specific.
Tumble Ratio @90btdc
2.5 V-alt
2.0
VIII III
1.5
VII
90-Deg Masks II VI
Selected
For the low-mode intake cam, the 7mm profile was selected. This cam provided better EPA fuel economy in real testing than the shorter 5mm profile. This profile, when used with an equal-lift EIVC strategy, provided good airflow capacity and sufficient lift area to achieve high charge dilution. As shown in a following section of this paper, this enables the vehicle to drive the whole EPA3 city cycle without switching to the high lift profile. Longer duration low-mode profiles have the advantage of providing good medium-load efficiency and NOx but exhibit limited efficiency for light-load combustion-limited conditions unless appropriate combustion enhancement is provided [10].
IV 180-Deg Mask
1.0
Prod Cyl Hd & Cams
0.5 0.0 900
920 940 960 Full-Load BMEP (kPa) (5500 rpm)
Figure 11. Tumble Ratio vs. Full-Load Performance for Several Mask Designs (from Simulation).
Tumble Index
3
The exhaust cam profile was modified based on simulation. It was found that a shorter exhaust duration (relative to the production exhaust profile) enabled more advanced intake timings for low mode at any dilution level. This increased the EIVC effect and reduced net pumping losses for improved fuel economy, as shown in Figure 14. Improvements are expected for combustionlimited operating conditions but will compromise peak power if exhaust duration is excessively short. This establishes a tradeoff between fuel economy benefit and performance for 2-step EIVC. Since the primary goal in this work was to maximize the fuel economy benefit of 2Step VVA, a small full load performance penalty was deemed acceptable.
Mask IV Design
2
1 Prod Chamber 0 0
2
4
6
8
10
12
Valve Lift (mm)
Figure 12. Tumble Index for Mask IV as Function of Valve Lift as Measured on Steady Flow Bench.
6
NPMEP (kPa)
40
Intake Default
Exhaust Default
1050 rpm / 200 kPa NMEP
30 20 EVL = 9.23mm (Tot Dur=278) EVL = 11.2mm (Tot Dur=304) EVL = 11.7mm (Tot Dur=322)
10 0 90
100
110
120
130
140
150
Figure 16. Valve Lift Profiles for Enhanced Combustion System.
160
EVO (CAD atdc) Figure 14. Effect of Exhaust Lift Profile on Net PMEP as a function of Exhaust Timing at 1050 rpm, 200 NMEP.
Cylinder Head for Enhanced Combustion System Integration for Enhanced Combustion
Use of a shorter exhaust lift profile required an exhaust port with higher flow capacity. Revised exhaust port and exhaust manifold were developed on the steady flow bench. Flow improvements between 25 and 40 percent were measured at medium-to-high valve lifts with the production exhaust valves. Results are shown in Figure 15.
The enhanced combustion system, including combustion chamber mask, high-flow exhaust port, and Delphi’s 2step valvetrain system (presented in a later section of this paper), was integrated on a cylinder head. Figure 17 shows a Unigraphics solid model of the integrated system. Cylinder heads were fabricated for subsequent dynamometer and vehicle testing. Tri-lobe camshaft
1
Curtain Area Cd
Pressure Drop: 7 kPa
2-Step rocker arms (12)
High-flow Port
0.8 0.6
Exhaust Phaser OCV
0.4
2-Step OCV
Production Port
0.2 0 0
2
4
6
8
10
12
Valve Lift (mm) Exhaust Cam
Figure 15. Discharge Coefficient as a Function of Valve Lift for Revised Exhaust Port with Production Valve.
Intake Cam Phaser
The high lift intake cam profile was also revised to take advantage of the intake cam phaser that was added to the engine. Simulations to optimize full-load torque and peak power were performed for a family of high-mode intake and exhaust profiles.
Intake Phaser OCV
Figure 17. Unigraphics Solid Model of Integrated System.
ENGINE DYNAMOMETER TEST RESULTS: ENHANCED COMBUSTION SYSTEM
Based on these results, a 9.7mm exhaust profile and a 8.5mm high-mode intake profile were selected for the enhanced combustion system. Valve lift profiles are shown in Figure 16.
Part-load test results for the 7mm profile with the enhanced combustion system at 1550 rpm, 350 kPa NMEP are shown in Figures 18 through 23. Data are presented as a function of charge dilution computed from simulation. Test conditions that exceed 7
As intake timing was advanced, dilution increased, pumping work reduced, and the engine was unthrottled (EIVC effect). Exhaust retard further increased dilution. Net thermal efficiency and gross indicated efficiency increased substantially, while NOx emissions decreased. A comparison of the cylinder pressurevolume diagram for the optimized 2-step system and the baseline production engine is shown in Figure 24.
40
Best Effic
A ke Inta
5
-20
EVO=111
EVO=121
EVO=131
EVO=141
30
30
35
PMEP (kPa)
EVO=121 EVO=141
-40
5
10
15 20 25 Residual (% Mass)
30
35
Figure 20. PMEP for Engine with Enhanced Combustion System.
20
15
Ad
EVO=111
EVO=121
EVO=131
EVO=141
EVO=151
va nc
10
Best Effic Lowest NOx
BSNOx (g/kW-hr)
In ta ke
e
5
COV > 3% 0 5
15 20 25 Residual (% Mass)
25
-60
33 10
20
-50
EVO=151
5
15
EVO=111 EVO=131 EVO=151
-30
COV > 3%
Net Thermal Efficiency (%)
34
10
Figure 19. Gross Indicated Thermal Efficiency (GITE) for Engine with Enhanced Combustion System.
Best Effic
e
EVO=141
Residual (% Mass)
37
k ta In
EVO=121
EVO=131
37
Fuel economy for EPA drive cycle was estimated using the four speed-load points and results are presented in Table 4. Relative to the production engine with exhaust cam phasing, the fuel economy for the enhanced system was improved an estimated 6.9 percent. This was a 2.1 percent increase over the production engine with the same 7mm lift profile.
35
EVO=111 EVO=151
The 10-75 burn durations are shown in Figure 23, and were similar to the baseline production engine. This shows that the primary effect of combustion enhancement was to improve early flame development.
e nc a v Ad
nce
39
38
As shown in Figure 22, combustion enhancement for the revised system was significant at this engine speed and load. Combustion durations for the baseline production cylinder head (shown previously in Figure 8) are also shown for comparison. The 0-10 burn duration at low dilution levels decreased by about 10 CAD, and the dilution limit was increased to 32 percent. This compares to a dilution limit of 25 percent for the production combustion system using the same 7mm profile (see Figure 8). Thus, the dilution tolerance using the low-mode, short-duration cam was restored to that of the baseline production engine, overcoming the effects of reduced tumble, increased tumble decay time, and lower effective compression ratio. Dilution limits and the 0-10 burn duration depended heavily on exhaust timing.
36
dva
COV > 3%
41
GITE (%)
combustion dilute limits (COV IMEP >3%) are shown beyond the dashed straight line in each figure.
10
35
Figure 21. BSNOx Combustion System.
Figure 18. Net Thermal Efficiency for Engine with Enhanced Combustion System.
8
15 20 25 Residual (% Mass)
for
Engine
with
30
35
Enhanced
Speed/Load Point
Prod Cyl Head L = 7.0 mm
40 30
650/150 1050/200 1550/350 1900/650
COV > 3%
50
Best Effic
0-10 Burn Duration (CAD)
60
% of Total Drive Cycle Fuel 9.2 24.5 43.6 22.7 100.0
Net Thermal Net Thermal Efficiency Efficiency % Drive Production (%) 2-Step (%) Improvement Cycle Fuel 23.8 27.1 33.5 37.5
24.9 29.3 36.4 39.6
5.0 7.9 8.8 5.6
8.8 22.7 40.1 21.5 93.1
FTP Urban Cycle Improvement (%) =
Prod Cyl Head L = 10.4 mm
20 10
EVO=111
EVO=121
EVO=131
EVO=141
6.9
Table 4. 4-Point Fuel Economy Estimate for Enhanced Combustion System with 7mm Lift Profile on EPA City Cycle.
EVO=151
0 5
10
15 20 25 Residual (% Mass)
30
35
2-STEP VALVETRAIN SYSTEM DESCRIPTION A 2-step Valvetrain system was developed that features switchable rocker arms on all intake valves, dualindependent cam phasing, and solenoid-actuated oil control valves (OCV). Figure 25 shows a sectional view of the 2-step system. It is a Type II valvetrain system with hydraulic lash adjusters (HLA).
Figure 22. Combustion Duration for 0-10 Percent Mass Burn Fraction for Engine with Enhanced Combustion System.
EVO=111
EVO=121
EVO=131
EVO=141
Cam profile switching is accomplished by replacing the production roller finger followers (RFF) and camshaft with 2-step rocker arms (shown in Figure 26) and a trilobe camshaft with three-cam-lobes-per-intake-valve. The outer lobes actuate the valves in low lift mode and the center lobes actuate the valves in high lift mode.
COV > 3%
50
EVO=151
40 30 20
Best Effic
10-75 Burn Duration (CAD)
60
Prod Cyl Head L = 7.0 mm
10
2-Step Rocker Arm
Tri-lobe Cam Oil Control Valve
0 5
10
15 20 25 Residual (% Mass)
30
35
Figure 23. Combustion Duration for 10-75 Percent Mass Burn Fraction for Engine with Enhanced Combustion System.
Cyl Pressure (kPa)
150
100 IVC Optimized 2-Step
50
Oil Supply Gallery Production Engine
Figure 25. 2-step Valvetrain System.
0 0
0.2
0.4
0.6
0.8
Volume (L) Figure 24. Net Pumping Loop Comparison (simulation).
9
High Mode Slider Cam Drive Torque (N*m)
10.0
Low Mode Rollers (2)
8.0 6.0
High Mode Slider
4.0
2.0
Low Mode
Figure 26. 2-step Rocker Arm. 0.0 0
The rocker arm default operation is low lift mode. In this mode, oil pressure to the HLA is regulated by the OCV below the switch pressure. Valve motion is controlled by the outer cam lobes, which contact roller followers. The center slider swings between the walls of the outer arm and does not affect valve motion. An internal lostmotion spring is used to keep the high lift follower in contact with the high lift center cam lobe. Lubrication of the cam followers and rocker arm bearings is accomplished by oil splash in the cylinder head. The high lift slider has low contact force in this mode and splash lubrication is adequate.
1000
2000
3000
4000
5000
6000
7000
Engine Speed (rev/min)
Figure 27. Measured Cam Drive Torque for Low Mode and High Mode. The 2-step rocker arm has higher mass and inertia than the production RFF. This necessitates higher valve spring force to satisfy valvetrain dynamic requirements. The effect of increased spring force is included in the valvetrain friction measurements shown in Figure 27. The actuation control system for 2-step valvetrain was developed to minimize oil system losses. The OCV was designed with a variable orifice to restrict oil pressure to the HLA gallery in low mode with virtually no loss of oil flow or pressure. In high mode, the OCV opens fully to apply full engine oil pressure to actuate the rocker arm lock pins.
When high mode operation is commanded, the OCV is energized and full engine oil pressure is supplied to the HLA gallery. Full oil pressure passes through the HLA into the 2-step rocker arm to extend the lock pin into its high mode position. The lock pin engages the high lift slider preventing it from swinging through the outer rocker arm. This causes valve motion to be controlled by the center high-lift lobe. Lubrication is supplied to the slider interface by an oil hole in the rocker arm that is opened only during high mode operation. This provides an oil jet directly to the cam slider interface and conserves oil in low mode.
With the addition of hydraulically-actuated valvetrain devices, some engine applications may require use of larger oil pumps to provide adequate oil pressure and flow. With the approach used in this work, hydraulic system losses were minimized and test engines could be operated using the existing engine oil pumps. There was no need to increase oil pump capacity.
For low mode, the roller followers provide low valvetrain friction for low-to-medium engine speeds where the majority of engine operation occurs. For high mode and generally higher speeds, a single slider follower is used. During development, both single slider and split dual slider designs were tested. The single slider design was found to provide both the lowest friction and lowest rocker arm width.
This system can be relatively easily applied to both existing production engines and new engines with Type II valvetrains that use HLA and RFF. No major machining center changes are required. Valves, HLA, and camshafts remain in the same locations. However, an OCV must be added between the engine oil pump and the HLA oil gallery.
Figure 27 shows measured drive torque of the intake cam in low and high modes. The higher friction in high mode is due to the slider interface. This has no effect on EPA fuel economy, and has small effect on full-load power.
Valve Lift Profile Development Complexity of valve lift profile development increases with the additional degrees of freedom for 2-step VVA with dual independent cam phasing. Use of modeling techniques greatly reduced the cost and time of optimization and allowed much broader study than possible using engine dynamometer testing alone. Optimization was performed for the intake low mode profile, intake high mode profile, and the exhaust profile. Families of valve lift profiles were designed to evaluate different levels of duration and lift area, and were 10
categorized by limiting speed capability and cam lobe radius of curvature (ease of manufacture). This resulted in over 50 valve lift profile designs for analysis. The most aggressive cam profile designs were used for dynamometer and vehicle testing in order to establish the potential maximum benefit of the 2-step VVA system.
main stages: 1. After a 4,000 mile on-road break-in, the vehicle was tested in its baseline production configuration (ECP only) for EPA 3 emissions. 2. While the engine hardware remained unchanged, the production powertrain controller was replaced with a Delphi controller including torque-based software. The vehicle was calibrated for drivability and emissions and retested. 3. The 2-step VVA system with enhanced combustion system was applied. The torquebased-control software was modified for 2-step valvetrain control. The vehicle was recalibrated for fuel economy and emissions, and retested.
Intake cam profiles were designed with negative radius of curvature as low as 84 mm because this is within the capability of many cam grinders. These more aggressive profiles allowed more lift area for a given lift duration. Table 5 shows design characteristics for the valve lift profiles used in this development. Figure 16 shows the valve lift profiles, including default positions and phase ranges. Tota l Va lve Lift Dura tion
At 0.15 mm Va lve Lift Are a
Dura tion
Are a mm*cra nk°
mm
cra nk°
mm*cra nk°
cra nk°
Inta ke Low Mode
7
210
550.36
146.38
525.7
Inta ke High Mode
8.5
282
1110.92
230.48
1073.46
Exha us t
9.7
286
1175.7
220.66
1139.7
Ma x Ca m Conta ct S tre s s
Inta ke Low Mode
In stages 2 and 3, a separate aftermarket commercial transmission controller was utilized to provide full transmission shift control. The transmission controller was calibrated to mimic the shift schedule and torqueconverter-clutch lockup schedule of the production vehicle.
Engine Spe e d for Ca m Ne ga tive 25% Spring Ma rgin Ra dius of Curva ture
MP a
re v/min
mm
1015
4380
-84
Inta ke High Mode
725
7200
a ll pos itive
Exha us t
1170
7350
-285
Engine Management System for 2-step VVA The engine management system for control and calibration of the 2-step valvetrain system was developed to utilize the full potential of 2-step VVA on the vehicle. The new EMS was developed using Delphi’s torque-based-control software for dualindependent cam phasing. Critical areas of the software that were modified included cam phasing, spark control, speed-density calculations, and functionality for 2-step altitude control.
Table 5. Design Characteristics of Valve Lift Profiles Vane-type hydraulic cam phasers were used in this study. The exhaust phaser was a production design with 50 crank degrees of authority. The intake phaser was a prototype design with 70 crank degrees of authority.
In order to precisely control the air-fuel mixture in the cylinder during switch transitions, and avoid missing switch events, an event-based switching strategy was used. The timing of switch commands to the 2-step OCV was accomplished by scheduling cylinder-specific reference events and interrupts. A timing diagram for event-based switching is shown in Figure 28.
Valvetrain geometry and mechanical limitations are important to consider for application of 2-step valvetrain systems. Intake cam advance and exhaust cam retard can produce valve-to-piston or valve-to-valve interference depending on lift profiles. Some engines may require provisions to improve valve-to-piston clearance to attain enough intake cam advance for maximum fuel economy benefit.
Command sent here (floating interrupt)
Hardware switches here
Window begin
Target Switch Window end Tooth
Calculations performed here Minimum switch lift Valve Lift #1
VEHICLE DEVELOPMENT AND TESTING
Fuel boundary #3 Fuel boundary #2
VVL switch boundary #3
The system was integrated on two development vehicles (2002 Chevrolet Trailblazer) and calibrated for drivability, fuel economy, and emissions performance. The primary goal was to demonstrate the fuel economy potential of 2step VVA for warmed-up operating conditions.
Switch Teeth Trigger Tooth
Valve Lift #3
Valve Lift #5
The baseline production vehicle employed an integrated engine-transmission powertrain controller that restricted access to the controller software and algorithms. For this reason, vehicle development took place in three
BDC#6
TDC#2
BDC#1
Figure 28. Switching 11
TDC#3
BDC#2
TDC#4
BDC#3
TDC#5
BDC#4
TDC#6
BDC#5
TDC#1
Timing Diagram for Event-Based Cam
Calibration of the cam phasers for 2-step VVA is dependent on the state of the intake valve lift, as shown in Figure 29. Intake and exhaust phaser timing are controlled independently. The base cam timing is determined for the default lift state, which is low lift. Phasing compensation is used only in the operating region where both valve lift states are possible. The application of the phasing compensation is based on a global transition variable that is enabled after all cylinders have switched.
Always De-active; phasing calibrated for de-active
Can be either; phasing calibrated for de-active
Always active: phasing calibrated for active
optimum. The model-based calibration yielded fuel economy generally within 1% of the dynamometerdetermined “optimum”. Agreement between SBC prediction and dynamometer measurements was poor only for the 1200 rpm / 200 kPa GIMEP point, which was a very light-load condition. This “best-efficiency calibration” for steady state conditions was subsequently slightly modified on the vehicle for drivability and emissions performance.
Engine Speed (rpm) Delta calibrated for the active state
+
VCPC Base Phasing
1200
1600
VVLC Phasing Delta
Figure 29. Phaser Compensation Spark timing logic is enhanced for 2-step VVA by adding a spark compensation term for the additional lift state. A calibrated model delivers optimum spark timing for the range of operating conditions, cam phaser positions, and lift states.
GIMEP (kPa)
Error (SBC-to-Dyno) BSFC
200
-18.6% *
300
-1.10%
400
-1.40%
600
-1.00%
200
-1.30%
400
-1.80%
600
-1.00%
2000
400
0.00%
2800
600 400
-0.40% -1.20%
Average (excepting 1200/200)
-1.02%
* Combustion performance at this condition was inadequately predicted by the SBC model
Table 6. Speed/load points used for SBC Validation.
Airflow calculations are based on a Pneumatic State Estimator (PSE). Volumetric efficiency compensation for 2-step VVA is accomplished by creating an additional VE model for the second lift state. Transition functions are used between lift states. For low barometric pressure and high torque conditions, altitude compensation provides earlier transition to high lift.
Vehicle Test Results Figures 30 through 34 show vehicle test results for the three main stages of vehicle development. The focus of this initial effort was demonstration of the fuel economy potential of 2-step VVA for warmed-up operating conditions. Cold start calibration was developed to a preliminary level only. All vehicle results reported are based on replicate tests for the EPA drive cycle performed at one vehicle test site.
Vehicle Calibration for Warm Operating Conditions Vehicle development was assisted by use of simulationbased engine calibration (SBC) techniques [18]. SBC was used to create initial cam phasing maps, spark advance tables, and 2-step switching schedules for the 2-step valvetrain system. Through application of SBC tools and techniques, initial vehicle operation was quickly achieved at a high level of initial performance.
As shown in Figure 30, fuel economy for the entire EPA test (23 cycles with cold start) was improved 12.4 percent relative to the original production baseline (exhaust cam phasing only). The vehicle was able to complete the entire EPA drive cycle with the engine operating only using the low lift cam. Relative to the same production vehicle, except modified with Delphi EMS, fuel economy for the entire EPA test was improved 7.5 percent. Fuel economy for Phase 3 (cycles 19-23), which excludes the cold start portion, was improved 5.5 percent.
Subsequent validation of model accuracy and basic soundness of SBC was performed on the engine dynamometer for a range of speed/load points as shown in Table 6. At each speed/load point, both intake and exhaust cam phasing were varied to determine optimum brake specific fuel consumption, and this was compared to SBC-determined phasing. The SBC phasing calibration proved reasonably close to the measured
12
Production Production + Delphi EMS
+ 5.5%
20
+ 7.5%
+ 12.4%
Tailpipe NOx (wt. mg/mi)
Fuel Economy (mpg)
Production Production + Delphi EMS Delphi EMS + 2 Step & Dual Phasing + Enhanced Combustion
2 Step & Dual Phasing + Delphi EMS + Enhanced Combustion
25
15 10 5 0
Phase 1
Phase 2
Phase 3
Total
Figure 30. EPA City Fuel Economy for 2-Step Development Vehicle and 2002 Production Baselines.
Tailpipe NMHC (wt. mg/mi)
2-step 2002
Fuel Economy (mpg)
+ 4.8%
15 10 5 0
Phase 3
10 5 0
0.9 0.8 0.7 0.6 0.5
- 88 %
0.4 0.3 0.2 0.1 0.0
Carbon monoxide emissions (shown in Figure 34) appear to have degraded, but this was a deliberate artifact of the hot restart strategy used during Cycle 19 of the EPA 3 test. Under these conditions, NOx was a much bigger concern than CO, so the engine was calibrated to trade off CO emissions for improved NOx performance. Overall CO remained well below the development target limit.
Production 2005 = Exh. Cam Phasing 2-step 2002 = Dual Phasing, 2-step Intake, Delphi EMS, & Enhanced Combustion
Phase 2
15
Figure 33 shows the reduction in HC observed following the replacement of the production controller and subsequent installation of the 2-step VVA system.
Production 2005
Phase 1
- 46 %
Figure 33. Non-Methane HC Tailpipe Emissions Measured on Phase 3 of EPA City Cycle for 2-Step Development Vehicle and 2002 Production Vehicle.
During the development period, hot emissions were monitored for the Phase 3 EPA Test (excluding cold start portion). Figures 32, 33, and 34 show NOx, nonmethane hydrocarbon (NMHC), and CO emissions, respectively. NOx emissions were reduced about 46 percent relative to the production vehicle. This resulted from both increased charge dilution afforded by dual cam phasing, and increased dilution tolerance provided by combustion system enhancements.
+ 5.5%
20
Production Production + Delphi EMS Delphi EMS + 2 Step & Dual Phasing + Enhanced Combustion
As shown in Figure 31, the 2-step development vehicle demonstrated 4.8 percent improved fuel economy (entire FTP test) relative to the production 2005 vehicle. Fuel economy for Phase 3 (excluding cold start portion) was improved 5.5 percent. Overall, these results are reasonably consistent with, though somewhat less than expected based on steady-state engine dynamometer and simulation results. Further improvement may be achieved with continued development of controls and calibration.
20
25
Figure 32. NOx Tailpipe Emissions Measured on Phase 3 of EPA City Cycle for 2-Step Development Vehicle and 2002 Production Vehicle.
As 2002 was early in the production lifetime of the tested engine, the manufacturer was likely to have made continued improvements to the engine hardware, control software, and calibration. For this reason, it was decided to also test a 2005 production vehicle. As expected, the fuel economy of the 2005 vehicle was improved over the 2002 production vehicle.
25
30
Total
Figure 31. EPA City Fuel Economy for 2-Step Development Vehicle and 2005 Production Vehicle. 13
12
Exhaust
80
+ 121 %
40 30
Low-Lift, 2-Step Intake, Retarded for Cold Start
6 4
BDC
60
8
TDC
70
50
High-Lift Intake
10
Lift (mm)
Tailpipe CO (wt. mg/mi)
Production Production + Delphi EMS Delphi EMS + 2 Step & Dual Phasing + Enhanced Combustion
2
20 0 270
10
360
0
450
540
630
Crank Angle
Figure 34. CO Tailpipe Emissions Measured on Phase 3 of EPA City Cycle for 2-Step Development Vehicle and 2002 Production Vehicle.
Figure 35. Cam Profiles Used for Cold Start Tests.
Intake Valve Flow Velocity (m/s)
COLD START INVESTIGATION Previous researchers have reported large 50 to 60 percent cold start HC reductions enabled by VVA [15, 19-21]. In the current work, engine experiments were performed to investigate the potential for cold start, engine-out HC reduction using 2-step VVA. The test engine was a 4-valve, 0.6 L, port-fuel injected, singlecylinder research engine fitted with a fully flexible electro-hydraulic valve actuation system. Instrumentation included a Cambustion Fast-FID HC Analyzer and an ETAS Lambda Sensor (fast-response heated wide-range O2 sensor) to measure cycleresolved HC and gas-phase equivalence ratio, respectively. The fuel injector controller permitted cycleby-cycle scheduling of injection pulse width.
Low -Lift, 2-Step Intake, Retarded for Cold Start IVO = 390
300 200 100 0 -100 -200
High-Lift Intake IVO = 340
-300 -400 -500 TDC
BDC
-600 270
360
450
540
630
Crank Angle
Figure 36. Intake Valve Flow Velocity (from Simulation) The cold-start VVA strategy that was investigated is illustrated in Figure 35. A low-lift, short-duration cam profile that is normally advanced during warm operation to implement an EIVC fuel economy strategy, may be retarded to achieve a late-intake-valve-opening strategy for the cold start. An intake cam phaser with increased position authority would be required. As shown in Figure 36, LIVO and low valve lift produce high intake flow velocities well into mid-stroke for enhanced fuel-air mixing, while the short lift duration results in valve closing near BDC for a high effective compression ratio (ECR) and higher mixture temperatures. These features were expected to enable leaner cold operation without misfire or partial–burn cycles. Note that if the high lift production profile were similarly retarded to achieve high intake velocity, excessively low ECR would result and startability would be seriously degraded.
Test Procedure Throughout the tests, room temperature (20 C) coolant and air was supplied to the engine. The test was initiated by first motoring at 1200 RPM with valves actuated using a 2-stroke mode (intake valves open on each downward stroke, exhaust valves open on each upward stroke) to avoid compression heating and keep the combustion chamber walls at near coolant temperature. The fully flexible, electro-hydraulic VVA system was programmed to provide 2-stroke operation. The intake throttle was preset to give 250 kPa NMEP during fired operation, and spark timing was retarded to place crank angle at 50% mass burn at about 25 degrees ATDC. It was assumed that retarded combustion was desirable for HC post-oxidation and faster catalyst light-off. When steady motoring conditions were achieved, the combustion test was initiated by simultaneously enabling fuel injection and normal 4-stroke valve actuation. Data was recorded for about 15 seconds, or about 150 cycles. Injection was performed on the closed valves during the compression stroke. Injector pulse width was varied cycle-by-cycle according to a preset schedule entered 14
into the engine controller. The goal was to schedule injection pulse width such that the first 10 cycles burned at approximately 0.9 equivalence ratio, then transitioning to stoichiometric. In cases where this target value resulted in misfires, fueling level was increased until stable combustion was obtained. Several test runs were required to determine this fueling schedule. The iteration process was aided by using a two-component fuel puddle model [22]. The fueling schedule for the next run was calculated after fitting the puddle model to lambda sensor results from the previous run.
combustion due to either wall wetting (to the crankcase) or exhaustion of liquid fuel. This is consistent with expected improved mixture preparation for the low lift profile.
It was desired to compute the ratio of injected fuel mass to in-cylinder vapor-phase fuel mass. The injector was calibrated against the fuel flow meter so that fuel mass could be calculated for each cycle. Measured airflow rate, along with the lambda sensor data was used to calculate in-cylinder fuel vapor mass for each cycle.
5
10
Rich
15
20
Cycle Number
Figure 38. Measured Gas-Phase Equivalence Ratio. Measurements of cumulative HC emissions, shown in Figure 40, are significantly reduced for the low lift profile. A 35 percent reduction from baseline was observed by 4 seconds after start. Overall, these preliminary results suggest that high intake velocities and high ECR resulting from the low-lift, short-duration profile improve mixture preparation and combustion quality. Additional cold start development on a vehicle, including investigation of the effects of injection timing, spark retard, idle RPM, and catalyst light-off may yield larger HC reductions. TRANSIENT SIMULATION AND CONTROL Proper transient control of engines with VVA is necessary to achieve the full fuel economy and NOx potential expected from steady state dynamometer testing. Optimum transient performance requires that phaser motion be coordinated with engine airflow dynamics such that the engine’s charge dilution is maintained near the combustion dilute limit under all part-load operating conditions. Previous research has indicated that fuel economy loss from ideal levels may be significant [23]. As part of the current work on 2-step VVA, an investigation was initiated to understand the cause and magnitude of these losses.
45
Injection Duration (msec)
Low-Lift, 2-Step Intake, Retarded for Cold Start IVO=390, IVC=534
0
Engine behavior changed rapidly during the first 10 cycles, followed by a slow drift toward steady state. In both cases, the first cycle required a relatively long injection pulse width to achieve stable first cycle combustion. Subsequently, the differences between the two cases were mainly confined to the next 5 to 7 cycles. As shown in Figure 37, the low-lift, shortduration profile required significantly lower fueling to achieve mixture requirements. For the high-lift baseline case, richer mixtures were required to maintain stable combustion.
40 L o w -L ift, 2 -S te p In ta k e , R e t a r d e d f o r C o ld S t a r t IV O = 3 9 0 , IV C = 5 3 4
25 20
H ig h - L if t In t a k e IV O = 3 5 5 , IV C = 6 0 5
15 10 5 0 0
1.0
0.6
Figures 37 through 40 show experimental test results for two cases. A 5mm valve lift profile retarded to an IVO of 390 CAD was used for the 2-step case. The baseline case was a high lift profile retarded to an IVO of 355 CAD. The baseline was representative of modern 4valve engines with dual-independent cam phasing.
30
1.2
0.8
Test Results
35
High-Lift Intake IVO=355, IVC=605
1.4
Lean
Equivalence Ratio
1.6
5
10
15
20
C y c le N u m b e r
Figure 37. Injection Duration The ratio of “injected-to-vapor-phase” fuel mass is plotted in Figure 39. The low-lift, short-duration profile shows significantly lower values for this ratio for the first 3 cycles, indicating that less liquid fuel escaped 15
Mass Ratio: Fuel Inj / Fuel Vapor
Figure 41 shows the system architecture for transient co-simulation. Preliminary vehicle tests were performed using prescribed transient throttle maneuvers to validate the predictive capability of this approach.
10 H ig h -L ift In ta k e IV O = 3 5 5 , IV C = 6 0 5
8
6
Typical Phaser Response for Tip-In Transient
L o w -L ift, 2 -S te p In ta k e , R e ta r d e d fo r C o ld S ta r t IV O = 3 9 0 , IV C = 5 3 4
4
Figures 42 through 46 illustrate typical simulation results. A throttle pedal increase from 0 to 7 percent in 0.2 seconds was specified to simulate a light acceleration from a standing start – a maneuver that is common on the FTP test for this vehicle. During this transient, engine gross IMEP increased from 230 kPa at idle to about 400 kPa after which quasi-steady vehicle acceleration was achieved. In these simulations, the desired steady-state phaser positions were commanded as a function of engine speed and desired IMEP (torquebased control) without any phaser rate limiting. Two cases were compared. The first case was for “normal” oil temperature of 90 C, and the second was for high oil temperature of 120 C.
2
0 0
5
10
15
20
C y c le N u m b e r
Figure 39. Ratio of “Injected-to-Vapor-Phase” Fuel Mass.
Cummulative Fast-FID HC
7 .E + 0 5
L o w -L ift, 2 -S te p In ta k e , R e ta r d e d fo r C o ld S ta r t IV O = 3 9 0 , IV C = 5 3 4
6 .E + 0 5 5 .E + 0 5
Figures 42, 43 and 44 show exhaust and intake phaser responses, and the oil pressure at the inlet to the phaser oil control valves. For Toil=90C, the exhaust phaser response was quite fast with a response lag of about 0.2 second. The oscillation in the response is due to cam torque pulsations (six cycles per engine cycle). For Toil=120C, however, a large lag in phaser response was seen, and the reason for this is apparent in the oil pressure plot. At engine idle, the oil pressure was only slightly above the minimum level required to move the phaser. As soon as the phaser begins moving, its oil consumption causes pressure to drop below this minimum. The phaser moves slowly in an intermittent start-stop fashion until the engine speed rises to a value that supplies sufficient oil pressure. When oil pressure is inadequate, the amplitude of oscillation in phaser position is increased.
H ig h -L ift In ta k e IV O = 3 5 5 , IV C = 6 0 5
4 .E + 0 5 3 .E + 0 5
2 5 % R e d u c tio n at 10 Seconds
2 .E + 0 5 1 .E + 0 5
3 5 % R e d u c tio n at 4 Seconds
0 .E + 0 0 0
20
40
60
80
100
120
140
C y c le N u m b e r
Figure 40. Cumulative HC Emissions Model Description Detailed models of the engine system were created and coupled in co-simulation. Matlab/Simulink provided the central control of information passed between the submodels. The engine and VVA system were modeled using GT-Power [12]. The lubrication system, cam phasers, oil control valves, and valvetrain cam switching mechanism were modeled with AMESim [24]. Engine and transmission controller code were written in Simulink/Stateflow to provide the essential features of the Delphi torque-based EMS system.
Exhaust Phaser (AMESim) VVA Mechanism (AMESim, Simulink)
Lubrication System (AMESim)
Information Flow (Simulink)
140
E V O [c ra n k d e g
Intake Phaser (AMESim)
Intake phaser response (Figure 43) was significantly slower at Toil=120C, but was not as drastically affected as for the exhaust phaser.
Controller (Simulink, Stateflow) Engine (GT-Power)
T o il = 9 0 C
130
D e s ire d 120
110
T o il = 1 2 0 C 100
Vehicle (Simulink)
1 .0
1 .5
2 .0
2 .5
T im e [s ]
Figure 42. Exhaust Phaser Position Response
Figure 41. System Architecture for Transient Simulation 16
3 .0
If phaser response is too fast, then valve overlap consistent with near-unthrottled, part-load operation will be obtained, while manifold pressure and engine speed are still well below steady-state values. The result is a dilution overshoot. Dilution overshoots will necessitate steady-state phaser calibration with less valve overlap, lower dilution, and less than optimum thermal efficiency. Significant compromises in fuel economy and NOx emissions can be expected during the relatively steady portions of the drive cycle.
3 50
IVO [crank deg]
3 40 3 30
T o il = 1 2 0 C 3 20
D e s ire d
3 10
T o il = 9 0 C
3 00 2 90 2 .0
2 .5
3 .0
Residual Fraction [%]
1 .5
T im e [s ]
Figure 43. Intake Phaser Position Response 350 300
Oil Pressure [kPa]
T o il = 9 0 C
50 1 .0
T o il = 1 2 0 C
40
D es ired 30
20
T o il = 1 20 C 250
T o il = 9 0 C 10
200
1.0
1.5
2 .0
2.5
3.0
T im e [s ]
150 100
Figure 45. Dilution Response
50
M in im u m p r e s s u r e r e q u ir e d to m o v e p h a s e r .
The response of Gross IMEP (essentially torque) is shown in Figure 46. IMEP decreases in correspondence with peaks in dilution as air is displaced by residual gas. This is followed by a sudden, rapid IMEP increase as dilution drops toward the steady-state value. This demonstrates that inadequate dilution control can result in undesirable engine torque response.
0 1.0
1 .5
2 .0
2 .5
3 .0
T im e [s ]
Figure 44. Oil Pressure Response The response of internal residuals for the engine is shown in Figure 45. The desired level represents a moderately aggressive phaser calibration. A modern engine with good combustion performance will generally tolerate somewhere between 30 and 35% total dilution. The 27% shown here for the desired dilution provides some margin for control error but at the expense of some fuel economy. If the phasers could be precisely controlled and properly coordinated with the engine’s airflow response, then a best-fuel-economy calibration could be used without inducing combustion-related drivability and emissions problems.
450
D e s ir e d
GIMEP [kPa]
400 350 300
T o il = 1 2 0 C 250 200
T o il = 9 0 C
150
Overshoot of dilution is seen for both oil temperatures. The slower case with Toil=120C shows dilution remaining within the allowable margin. However, with phasers capable of response rates as high as those simulated here with normal oil temperature, excessive dilution and engine misfire would occur. Cylinder pressure combustion data measured on the test vehicle showed evidence of excessive dilution excursions during fast throttle tip-ins.
1 .0
1 .5
2 .0
2 .5
3 .0
T im e [ s ]
Figure 46. Gross IMEP Response Effect of Phaser Rate Limiting Based on simulation of various phaser control strategies, it was shown that dilution overshoot could be avoided by rate-limiting the phaser commands. Figures 47 through 17
49 show dilution, thermal efficiency and cumulative NOx responses for two different rate-limiting strategies during a tip-in-tip-out transient. (The tip-in is the same throttle maneuver used to generate Figures 42 to 46.) For efficiency and NOx, the response for parked phasers is shown for comparison.
1. Indirect Benefit – Avoiding dilution excursions during transients that compromise drivability enables phaser calibration for best economy and lower NOx. 2. Direct Benefit - Maintaining desired dilution during transients improves economy and NOx during those transients.
For the “best single-phase-rate” strategy, the rate was limited at the same value for both tip-in and tip-out. The tip-in dilution spike is suppressed by slowing the phaser, but a tip-out spike is then produced. This rate limit was chosen to produce moderate tip-in and tip-out spikes of about equal magnitude. This is a large improvement over the non-rate-limited case (Figure 45 at Toil=90). This strategy shows the best efficiency and NOx response, but some of this advantage would be lost on a driving cycle test because a somewhat sub-optimal steady-state phaser calibration would still be required due to the dilution overshoots.
B est S in g le P h as e R a te
Brake Therm Effic [%]
30
25 S low T ip -In , F ast T ip -O u t
20
P a rk ed P h as ers
15
10
For the “slow tip-in, fast tip-out” strategy, the rates were specified independently to suppress all dilution overshoots so that an optimum steady-state calibration can be implemented. However, slowing the tip-in rate to this degree causes a prolonged dilution deficit, with corresponding efficiency and NOx penalties, as shown in Figures 48 and 49. Further simulations showed that prolonged dilution deficits persist at throttle rates that are common during the FTP, so a significant fuel economy and NOx penalty may be expected.
1
4
5
6
7
8
Figure 48. Transient Thermal Efficiency for Two Different Phaser Rate-Limiting Strategies 0 .1 4
P a rk e d P h a s e rs
Cummulative N0x [g]
Residual [%]
3
T im e [s ]
30 25 20
2
0 .1 2
S lo w T ip -In , F a s t T ip -O u t
0 .1 0 0 .0 8 0 .0 6 0 .0 4
B e s t S in g le P h a s e R a te
0 .0 2
D esired
0 .0 0
15
B est Single Phase R ate
10
1
3
4
4
5
6
7
8
Figure 49. Cumulative NOx Emissions for Transient.
5 2
3
T im e [s ]
Slow Tip-In, Fast Tip-O ut
1
2
5
6
7
SUMMARY AND CONCLUSIONS
Tim e [s]
Engineering teams within Delphi collaborated to develop and integrate a 2-step VVA system using an earlyintake-valve-closing strategy. A 2-step valvetrain and the engine management system were developed. Extensive engine dynamometer and vehicle testing was completed to evaluate fuel economy and emissions potential of the system. The following conclusions may be made.
Figures 47. Transient Dilution Response for Two Different Phaser Rate-Limiting Strategies Model-Based Phaser Control More sophisticated model-based phaser control is required to obtain dilution control for optimized torque response, fuel economy and emissions.
1. For application of 2-step on the production engine, dynamometer tests indicated a fuel economy benefit of 4.5 and 4.8 percent for 5mm and 7mm lift profiles, respectively. The baseline was the production engine with exhaust cam phasing only.
Two significant sources of fuel economy and NOx benefits are expected:
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2. Combustion analysis indicated that combustion characteristics for the low-mode profiles were significantly degraded from levels for the production intake lift profile as tested on the production cylinder head.
ACKNOWLEDGMENTS The authors would like to thank Paul Armour, Dominic Borraccia, Jon Darrow, Hermes Fernandez, Tom Fischer, Ryan Fogarty, Cynthia Greene, Nick Hendriksma, Dan Kabasin, Gary Lippert, Natalie Payne, Jeffrey Pfeiffer, Jeff Rohe, and Jim Waters for their expertise and contributions provided in the course of this work. Also acknowledged is the Delphi engine test group including Bob Emens, Mike Helms, Charles Hernandez, Ron Kunst, Chuck McMahon, Ray Parker, and Chris Yates.
3. An enhanced combustion system was developed for 2-step EIVC using a chamber mask for increased incylinder tumble. Valve lift profiles were modified to maximize fuel economy potential. Dynamometer tests showed substantial improvement of burn characteristics and extension of dilute combustion limits. Fuel economy was improved 6.9 percent relative to the production engine with exhaust cam phasing only.
Delphi modified GM production engines for this study. The authors acknowledge support from GM Powertrain in the course of this work. Expertise from the following individuals is also acknowledged; Jim Bruso (EDS), Carl McQuillen (McQuillen Racing Engines), Butch Elkins (Diamond Racing), and the team at Soligen, Inc.
4. A 2-step valvetrain mechanism was developed that features hydraulically-actuated switchable rocker arms and hydraulic lash adjusters. The rocker arm is a dual-roller, single-slider design for compact packaging and low friction. 5. The engine management system was modified for control and calibration of 2-step VVA. New software and control algorithms were embedded using torque-based control to realize the full fuel economy potential from 2-step VVA.
REFERENCES 1. Hosaka, T., “Development of the Variable Valve Timing and Lift (VTEC) Engine for the Honda NSX”, SAE Paper 910008, 1991. 2. Brüstle, C., and Schwarzenthal, D., “VarioCam Plus – A Highlight of the Porsche 911 Turbo Engine,” SAE Paper 2001-01-0245, 2001. 3. Shibano, K., Kamimaru, S., Yamada, T., Watanabe, K., “A Newly Developed Variable Valving Mechanism with Low-Mechanical Friction,” SAE 920451, 1992. 4. Shikida, T., et al., “Development of the High Speed 2ZZ-GE Engine,” SAE Paper 2000-010671, 2000. 5. Horie, K., “The Development of a High Fuel Economy and High Performance Four-Valve Lean Burn Engine,” SAE Paper 920455, 1992. 6. Fukuo, K., et al., “Honda 3.0L, New V6 Engine,” SAE 970916, 1997. 7. Albertson, W., et al., “Displacement on Demand for Improved Fuel Economy without Compromising Performance in GM’s High Value Engines,” Powertrain International, Volume 7, Number 1, pg 25-40, 2004. 8. Fortnagel, M., Doll, G., Kollmann, K., and Weining, H., “Four Made of Eight – The New 4.3L and 5.0L V8 Engines”, MTZ Special Edition, July, 1999, pp 58-63. 9. Falkowski, A., et al., “Design and Development of the DaimlerChrysler 5.7l HEMI MultiDisplacement Cylinder Deactivation System,” SAE Paper 2004-01-2106, 2004. 10. Sellnau, M., and Rask, E., “2-step Variable Valve Actuation for Fuel Economy, Emissions, and Performance,” SAE Paper 2003-01-0029, 2003.
6. Vehicle tests were performed for the EPA City cycle. For Phase 3 EPA tests (cycles 19-23), a 5.5 percent improvement was measured with a 46 percent reduction in NOx. For the whole EPA test including cold start, a 4.8% improvement was measured. Vehicle fuel economy improvement was less than steady state dynamometer tests, in part, due to cam phaser control limitations. Further improvements in vehicle fuel economy are expected with refinement in transient control and calibration. 7. Preliminary cold start tests were performed on a single-cylinder engine using a 2-step late-intakevalve opening (LIVO) strategy with a shorter duration intake lift profile. HC emissions were reduced 35 percent by 4 seconds after start, relative to a baseline representing dual-independent cam phasing. 8. Transient engine simulations were performed to understand interactions between valvetrain components and the combustion system for different system temperatures and phaser response rates. It was shown that dilution excursions during transients will compromise drivability and limit cam phaser calibration. Model based dilution control techniques have potential to maintain desired dilution during transients for improved fuel economy and NOx.
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Delphi Energy and Chassis Systems 5500 W. Henrietta Rd. Henrietta, NY 14586
11. Kreuter, P., et al., “Variable Valve Actuation – Switchable and Continuously Variable Valve Lifts,” SAE Paper 2003-01-0026, 2003. 12. GT Power Software, Version 6.1, Gamma Technologies, Inc, Westmont, Illinois, 2005. 13. Vibe, I.I., “Brennverlauf und Kreisprozeb von Verbrennungs-Motoren,” VEB Technik, Berlin, 1970. 14. Nagumo, S., and Hara, S., “Study of Fuel Economy Improvement through Control of Intake Valve Closing Timing: Cause of Combustion Deterioration and Improvement,” JSAE Review 16 (1995)13-19, 1995. 15. Heuser, P., Kreuter, P., Reinicke-Murmann, J., Erz, R., and Peter, U., “The META VVH System – The Advantages of Continuously Mechanical Variable Valve Timing,” SAE Paper 1999-010329,1999. 16. Burkhard, J. et al., “Benefits of a Mechanical Variable Valve Actuation System,” Aachen Kolloquium, Fajrzeig-und Motorentechnik, Aachen, Germany, 2000. 17. Stein, R.A., Chou, T., Lyjak, J., “The Combustion System of the Ford 5.4L 3-Valve Engine,” 2003 Global Power Conference, Dearborn, Michigan. 18. Rask, E., and Sellnau, M., “Simulation-Based Engine Calibration: Tools, Techniques, and Applications,” SAE Paper 2004-01-1264, 2004. 19. Nakeyama, Y., et. Al., “Reduction of HC Emissions from VTEC Engine During Cold-Start Conditions,” SAE Paper No. 940482, 1994. 20. Pietch, I., “Einflub verringerter Einlabventilhube auf die Gemischbildung in Ottomoteren,” Essen 2000 Conference, 2000. 21. Pischinger, M., et al., “Benefits of the Electromechanical Valvetrain in Vehicle Operation”, SAE Paper 2000-01-1223, 2000. 22. Kirwan, J., Jorgenson, S., Matekunas, F., Chang, C.-F., “Engine Fuel Injection Control Method with Fuel Puddle Modeling,” U.S. Patent 6176222, January 23, 2001. 23. Jacquelin, F., et al., “Cam Phaser Actuation Rate Performance Impact on Fuel Consumption and NOx Emissions Over the FTP-75 Drive Cycle,” SAE Paper 2003-01-0023, 2003. 24. AMESim Software, Version 4.2, Imagine S.A., Roanne-France, 2004.
NOMENCLATURE AND UNITS ATDC BDC BMEP CAD ECP ECR EGR EIVC EMS EVO FTP GIMEP HC HLA ICP IMEP IVO L LIVO NMEP NPMEP NOx OCV PMEP PSE RFF RPM SBC TDC VVA
CONTACT For additional information, Mark Sellnau
[email protected] Delphi Research Labs 51786 Shelby Parkway Shelby Twp., MI 48315 Tim Kunz
[email protected] 20
After Top Dead Center Bottom Dead Center Brake Mean Effective Pres (kPa) Crank Angle Degrees Exhaust Cam Phasing Effective Compression Ratio Exhaust Gas Recirculation Early Intake Valve Closing Engine Management System Exhaust Valve Opening @ L=0 (CAD ATDC firing) Federal Test Procedure Gross Indicated Mean Effective Pressure Hydrocarbon Emissions (g/kW-hr) Hydraulic Lash Adjuster Intake Cam Phaser Indicated Mean Effective Pres (kPa) Intake Valve Opening @ L=0 (CAD ATDC firing) Lift (mm) Late Intake Valve Opening Net Mean Effective Pressure (kPa) Net Pump Mean Effect. Pressure (kPa) Oxides of Nitrogen (g/kW-hr) Oil Control Valve Pumping Mean Effective Pres(kPa) Pneumatic State Estimator Roller Finger Follower Revolutions per Minute Simulation Based Engine Calibration Top Dead Center Variable Valve Actuation
