the effect of intake temperature in a turbocharged multi...
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LUND UNIVERSITY
PO Box 117221 00 Lund+46 46-222 00 00
The Effect of Intake Temperature in a Turbocharged Multi Cylinder Engine operating inHCCI mode
Johansson, Thomas; Johansson, Bengt; Tunestål, Per; Aulin, Hans
Published in:ICE 2009
2009
Link to publication
Citation for published version (APA):Johansson, T., Johansson, B., Tunestål, P., & Aulin, H. (2009). The Effect of Intake Temperature in aTurbocharged Multi Cylinder Engine operating in HCCI mode. In ICE 2009 (pp. 1-15). SAE.
Total number of authors:4
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2009-24-0060
The Effect of Intake Temperature in a Turbocharged MultiCylinder Engine operating in HCCI mode
Thomas Johansson, Bengt Johansson, Per Tunest alFaculty of engineering, Lund University
Hans AulinGM Powertrain AB, Sweden
ABSTRACT
The operating range in HCCI mode is limited by the ex-cessive pressure rise rate and therefore high combustioninduced noise. The HCCI range can be extended with tur-bocharging which enables increased dilution of the chargeand thus a reduction of combustion noise. When the en-gine is turbocharged the intake charge will have a hightemperature at increased boost pressure and can then beregulated in a cooling circuit. Limitations and benefits areexamed at 2250 rpm and 400 kPa indicated mean effec-tive pressure. It is shown that combustion stability, com-bustion noise and engine efficiency have to be balancedsince they have optimums at different intake temperaturesand combustion timings. The span for combustion timingswith high combustion stability is narrower at some intaketemperatures and the usage of external EGR can improvethe combustion stability. It is found that the standard de-viation of combustion timing is a useful tool for evaluatingcycle to cycle variations. One of the benefits with HCCIis the low pumping losses, but when load and boost pres-sure is increased there is an increase in pumping losseswhen using negative valve overlap. The pumping lossescan then be circumvented to some extent with a low intaketemperature or EGR, leading to more beneficial valve tim-ings at high load.
INTRODUCTION
The homogenous charge compression ignition (HCCI)combustion offers high thermal efficiency, low throttlinglosses and low emissions. The limited operating rangein HCCI mode makes it to an alternative to the spark ig-nited (SI) engine at low load to improve the efficiency. Onthe other hand the diesel engine has shown some favor-able properties when operated with fuels with long ignitiondelay like gasoline [1] leading to low soot and nitrogen ox-ides (NOx) levels. Here the HCCI mode or a mix of it canbe attractive to cover the whole operating range. In HCCImode the air-fuel charge is auto ignited by controlling thetemperature and pressure history inside the engine cylin-der to initiate and time the combustion phasing at the rightmoment.
When operating with a homogenous charge the right con-ditions for the combustion timing has to be set before in-take valve closing (IVC). With the introduction of direct in-jection (DI) of the fuel there is an increased control space[2, 3] and even a possibility to change the operating condi-tions after IVC, so the preferred charge composition doesnot have to, or will be homogenous [4].
1
To reach auto ignition in a four-stroke engine the in-cylinder temperature is normally increased by exhaust gasrecirculation (EGR), air heating or both as outlined in theeighties by Thring [5]. The compression ratio can also beused to control the HCCI combustion timing [6, 7]. Theconcept with negative valve overlap (NVO) with early ex-haust valve closing to trap a certain amount of hot residu-als is widely accepted to control combustion timing. Herethe variable valve timing (VVT) is accurately controlled ei-ther mechanically or electrically.
The concept of re-breathing [8] might be more mechani-cally complicated but has an advantage in reduced heatlosses since it can be operated without recompression.The combustion timing in HCCI can also be adjusted withthe intake temperature by a thermal management system[9, 10], but this need a heat recovery system.
The combination of intake heating and NVO [11] in a natu-rally aspirated (NA) engine showed that the intake temper-ature had a low impact on engine operating but increasedtemperature could be used to extend the low load limit. Byboosting the engine in HCCI mode to extend the possibleload range [12, 13, 14, 15, 16], the intake temperaturewill be elevated and can therefore be cooled down to asuitable level with an intercooler. The boost pressure canbe controlled directly by the intake valve timing [16] whichgives large freedom to operate the engine with optimumefficiency, emissions, combustion noise, cyclic variationsand control strategies. With NVO the intake temperaturedirectly control how much internal EGR is needed for agiven combustion timing and therefore the in-cylinder rel-ative burn gas fraction. The chemical and thermodynamicproperties of the EGR [17] will influence the combustion.In addition to the residual gas, external cooled EGR canbe used and it was found that it could improve combus-tion efficiency while there seem to be low influence on thecombustion duration [18] in HCCI mode.
The main operating parameter in HCCI mode is the com-bustion timing here represented by the crank angle of 50%heat released (CA50). The CA50 timing will effect the en-gine performance substantially and need to be controlledexactly to run the HCCI engine successfully. If enginespeed goes up or if load is increased the CA50 window forstable engine running will be narrowed, making the enginecontrol more challenging.
The scope of this paper is to investigate how combus-tion timing, intake temperature and EGR influence a tur-bocharged NVO HCCI engine performance. This is doneby dividing the result in small fractions. By doing this theunderstanding on how to operate a turbocharged HCCIengine is expanded. Since the application of HCCI has tomeet consumer demand, emission legislation and controlcapability the real operating range is always a balance be-tween these. The result is also applicable to some extentfor a NA HCCI engine.
EXPERIMENTAL SET-UP
ENGINE SYSTEM The test engine is an in-line fourcylinder gasoline engine with a total displacement of 2.2l.The cylinder head is a 4-valve design with a pent-roofcombustion chamber. There are some small squish ar-eas on the intake and exhaust sides. The piston has araised piston dome with a small bowl in the center. Theintake channel is side-drafted with a low tumble design.The engine has DI with a slightly canted, centrally placedinjector of the Spray-Guided type. The eight-hole solenoidfuel injector has a cone angle of 60o. The spark plug is lo-cated close to the fuel injector and has an extended tip.The spark is always operated as a safety measurementagainst misfires and it can also improve the combustionstability at late combustion timings. To achieve HCCI com-bustion the engine is operated with NVO with low lift andshort duration camshafts designed for a NA HCCI engine.In Figure 1 the minimum and maximum valve timings areplotted.
The VVT is controlled by hydraulic actuators at thecamshafts giving a separate 50 crank angle degree (CAD)adjustment on both intake and exhaust valve timings. Theengine is turbocharged by a fixed geometry turbine. Theexhaust manifold is of pulse type with short individual run-ners straight to the turbine inlet. On the intake side thereis a water cooled intercooler. The heated aluminum intakemanifold has short intake runners and a small volume.The cooling system has an electric driven water pump andthe coolant temperature is adjusted by the control system.Engine specifications are listed in Table 1.
Valve timings
Lift
[m
m]
TDC BDC TDC BDC TDC
1
2
3
4
5Exhaust max.Exhaust min.Intake min.Intake max.
Figure 1: Valve timing range and setup
Table 1: Engine specificationsNumber of cylinders 4Displacement 2198 cm3
Bore x Stroke 86 mm x 94.6 mmCompression Ratio 11.75:1Valve duration 125 CADValve lift 4.1 mmNVO range 74–174 CADTurbocharger B&W KP31Fuel Supply DI, up to 20 MPaFuel Type Gasoline, 95 RON
2
Figure 2: Exhaust manifold
Figure 3: Intake manifold
MEASUREMENT AND CONTROL SYSTEM The con-trol system is a combined data acquisition and enginecontrol unit from dSPACE. The signals from a fully instru-mented engine set-up are collected and analyzed. Forcombustion feedback there are individual cylinder pres-sure sensors. The control system has an in-cycle re-solved heat release calculation where main parameterslike CA50, peak cylinder pressure (PCP) and peak pres-sure derivative (dP/CAD) etc. are used for cylinder indi-vidual control in closed loop. Emission analysis and sootmeasurement are collected on an external PC and thedata is sent to the control unit.
To increase the operating range it is important that all thecylinders operate identically, meaning that they have thesame CA50 position, indicated mean effective pressure(IMEPnet) and peak pressure rise rate. This means thatthe control system has to perform cylinder balancing (CB).Injection of the fuel in NVO can give a reformation of thefuel with an increase in charge temperature and reductionin effective octane number leading to advanced combus-tion timing. The combustion timing will be more advancedif the fuel is injected in the beginning of the NVO than inthe end. The control system sets the CA50 position anduses the cylinder with the most advanced CA50 positionfor the exhaust valve closing (EVC) set point. The othercylinders are then advanced by adjusting the injection tim-ing and/or fuel amount in the first injection, until all cylin-ders have the same CA50 position. Since there will besome differences in injection timing when using CA50 bal-
ancing there will also be some differences in efficiency be-tween the cylinders. This results in different load, IMEPnet
in the cylinders. To maximize the load range there has tobe CB of IMEPnet as well. This is done at the main injec-tion event. Obviously this affect the CA50 balancing sinceincreasing the fuel amount will raise the combustion heat,leading to earlier CA50 for that cylinder. Therefore the CBhas to be constantly monitored and adjusted by the con-trol system. Table 2 shows the measurement and controlequipment.
Table 2: Measurement and control system.ECU and data logging dSPACE Rapid ProCylinder pressure sensor Kistler 6043AspCharge amplifier Kistler 5011BEmission analyzer Horiba 9100 MEXASoot analyzer AVL 415 S
TEMPERATURE AND EGR CONTROL SYSTEM Theintake temperature is controlled by an electrical throttle inthe bypass routing to the water-cooled intercooler circuit,see Figure 4. When the throttle is fully open the main airflow goes through the bypass routing due to less flow dis-charge compared to the intercooler circuit. By closing thethrottle the airflow is forced through the intercooler circuit.
The long route EGR system takes exhaust gas after theturbine and route it to the compressor inlet. To get a pos-itive displacement of exhaust gas there is a throttle in theexhaust system to increase back pressure. The compres-sor inlet can be used as an ejector to enhance the EGRflow, so the needed back pressure is quite small. The ex-ternal EGR routing is cooled by the engine cooling circuitand is controlled by an EGR valve near the turbocharger,see Figure 4.
Figure 4: Engine layout
The presented results are based on average results fromall four cylinders during 1000 cycles The deviation in en-gine load between measured test points is ± 3 kPa andin engine speed ± 2 rpm. Engine coolant temperature is∼ 80
◦ C out from the engine.
3
RESULTS AND DISCUSSION
The combustion phasing in HCCI mode is vital for sta-ble engine running and efficiency. In the NVO conceptthe EVC timing is the major control variable for the com-bustion phasing since it controls the amount of trappedhot residual gas. The intake valve timing can also beused to control the combustion timing but the influenceis weaker and not as predictive as the EVC timing whichmakes it less suitable for cycle to cycle combustion con-trol. It is normal to operate with almost symmetrical valvetimings on the intake and exhaust. With these short dura-tion valve timings, a late intake valve closing will increasethe available in-cylinder pressure when the engine is tur-bocharged. The combustion timing is even more impor-tant due to the direct influence it has on boost pressureand therefore the intake temperature. A later combus-tion timing will increase the available exhaust enthalpyfor the exhaust turbine leading to higher boost pressure.This means that the intake temperature will also increase.There is then the freedom to adjust the intake charge tem-perature by the cooling circuit to obtain best overall perfor-mance depending on load and engine speed.
The maximum load in HCCI is often limited by the fastcombustion rate where the combustion induced noise is amajor factor. By increasing the air dilution in the cylinderwith turbocharging the audible noise is suppressed andtherefore the operating range can be extended. The pres-sure rise rate can be measured in bar/CAD or MPa/ms[19] and is used to impose a operating limit. But neitherof these catch the real combustion noise when the engineis supercharged. The Ringing Intensity (RI)– MW/m2 [20],on the other hand compares the pressure rise rate to peakcylinder pressure and engine speed and is therefore usedin these tests.
The engine tests in this study are performed at 2250 rpmand 400 kPa IMEPnet. This load and speed point waschosen due to the wide span of boost levels that can bereached here with alteration of the intake temperature. Italso has different operating limitations due to combustionstability, high pressure rise rate and turbocharger perfor-mance. Even if it is possible to operate this turbochargedHCCI engine at a higher load at this engine speed, it isunsuitable for a wide CA50 sweep due to the high pres-sure rise rate at early combustion timings. In the first setof results the CA50 position and intake temperature areswept. Then there is an external EGR and intake temper-ature sweep at locked CA50 timing, and finally, there isa CA50 sweep with and without EGR at the same intaketemperature. During these tests the intake valve timing isconstrained to latest possible opening position to increaseavailable in-cylinder pressure.
CA50 AND INTAKE TEMPERATURE SWEEP AT 2250RPM AND 400 KPA IMEPNET From this test there aresome additional graphs that can be found in the AP-PENDIX. In Figure 5 the intake manifold pressure (Pin) is
shown at the possible operating range in this test point.The temperature can be altered from 110
◦ C down to45
◦ C with the throttle controlled intercooler routing. Thetiming for CA50 can not be set to more than ∼ 7
o aTDCf
for stable engine operating at this test point. The impor-tance for robust CA50 control is evident when operatingthis turbocharged HCCI engine since it is dependent onthe boost pressure to decrease RI to manage high load.The resulting PCP reflects the wide intake pressure vari-ations and span from 6.6 to 8.2 MPa, see Figure 6.
CA50 [aTDCf]
Tin
take
[°
C]
Pin
[kPa] vs Tin
and CA50
1 2 3 4 5 6 7 840
50
60
70
80
90
100
110
150
155
160
165
170
175
180
185
190
Figure 5: Intake pressure, absolute
CA50 [aTDCf]
Tin
take
[°
C]
PCP [MPa] vs Tin
and CA50
1 2 3 4 5 6 7 840
50
60
70
80
90
100
110
6.8
7
7.2
7.4
7.6
7.8
8
Figure 6: Peak cylinder pressure
The influence of the intake temperature is strong onachievable boost level. Since less burned gas fraction (xb)is needed, see Figure 7, at higher intake temperatures fora given combustion phasing, the specific heat capacity inthe charge is reduced and thus the mass flow rate is in-creased. This increased mass flow to the turbochargerincrease the boost pressure.
The corresponding charge temperature at intake valveclosing (TIV C ) is plotted in Figure 8. To shift the CA50one degree, the TIV C changes approximately 8o, exceptat low temperatures where the CA50 timing is very sensi-tive and changes rapidly to the TIV C . The correspondingEVC position that is used to set the combustion phasingis seen in Figure 9.
4
CA50 [aTDCf]
Tin
take
[°
C]
xb vs T
in and CA50
1 2 3 4 5 6 7 840
50
60
70
80
90
100
110
0.4
0.42
0.44
0.46
0.48
0.5
0.52
0.54
Figure 7: Burned gas fraction, xb
CA50 [aTDCf]
Tin
take
[°
C]
TIVC
[°C] vs Tin
and CA50
1 2 3 4 5 6 7 840
50
60
70
80
90
100
110
210
220
230
240
250
Figure 8: Temperature at IVC
CA50 [aTDCf]
Tin
take
[°
C]
EVC [bTDCGE
] vs Tin
and CA50
1 2 3 4 5 6 7 840
50
60
70
80
90
100
110
52
54
56
58
60
62
64
66
68
Figure 9: EVC position
When the boost level is increased it often comes at acost, in the form of increased pumping losses that is rep-resented as pumping mean effective pressure (PMEP)in Figure 10. The achievable boost pressure hasto be weighted against the pumping losses from theturbocharger inefficiency and throttling losses over thevalves, to keep engine efficiency as high as possible.
By separating the pumping losses in two parts further in-formation is available: PMEP= Throttlingvalve + (Pex−Pin)
CA50 [aTDCf]
Tin
take
[°
C]
PMEP [kPa] vs Tin
and CA50
1 2 3 4 5 6 7 840
50
60
70
80
90
100
110
60
70
80
90
100
Figure 10: Pumping losses
Pex − Pin: The pressure loss over turbine/compressor isshown in Figure 11.Throttlingvalve: The pressure loss over inlet and exhaustvalves is shown in Figure 12.With a turbocharger one usually encounters higher backpressure than boost pressure. Since the turbocharger ison the small side here, it was not possible to have a laterCA50 timing than 3o (aTDCf ) the highest intake temper-ature. The reason to this is that the back pressure in-creased considerable which led to reduced possible loadrange. The high throttling losses originates from the shortduration valve timings. For example, at high HCCI load alate EVC is needed which lead to that the exhaust valveopening (EVO) can be close to BDC. This will increase thepumping work from the blow-down process.
CA50 [aTDCf]
Tin
take
[°
C]
Pex
− Pin
[kPa] vs Tin
and CA50
1 2 3 4 5 6 7 840
50
60
70
80
90
100
110
20
25
30
35
40
Figure 11: Pressure loss over turbine/ compressor
To decrease the engine noise the combustion timing canbe delayed but this can increases the probability for mis-fire. The coefficient of variance (CoV) of IMEPnet as seenin Figure 13 can be used as an indicator of combustionstability. Since the nature of HCCI combustion inhibitssome cycle to cycle variations of combustion timing, theCoV of IMEPnet need to be complemented by variationsin combustion phasing as well. In Figure 14 the standarddeviation (STD) of CA50 is plotted. Looking at the CoVgraph, a later CA50 timing is the major influence of low
5
CA50 [aTDCf]
Tin
take
[°
C]
Throttlingvalve
[kPa] vs Tin
and CA50
1 2 3 4 5 6 7 840
50
60
70
80
90
100
110
45
50
55
60
Figure 12: Valve throttling losses
combustion stability. But the STD of CA50 reveals thatthere is an issue when the intake temperature is low aswell. This is also noticeable when listening to the enginerunning. It has a more ruff engine sound at low intaketemperatures. At the lowest temperatures the burned gasfraction, as seen in Figure 7, is highest and the dilutionfrom boost pressure is lowest, making the influence fromsmall variations in TIV C more pronounced, see Figure 8.
CA50 [aTDCf]
Tin
take
[°
C]
CoV of IMEPnet
[%] vs Tin
and CA50
1 2 3 4 5 6 7 840
50
60
70
80
90
100
110
1.5
2
2.5
3
Figure 13: CoV of IMEPnet
The high pressure rise rate is the main limiting factor whenthe load is increased. In Figure 15 it can be seen that theCA50 position is the major control variable for the pres-sure rise rate. Recall that the intake pressure (Figure 5)increases with later CA50 position and high intake tem-perature. The pressure rise rate scale to the amount ofdilution but the effect of this is masked to some extent be-cause there is a need to burn more fuel at constant loadwhen the pumping losses increases.
In Figure 16 the combustion duration for CA10 to CA90 isshown. The CA50 timing is the most important factor butthere is some influence from the increased dilution withhigher boost pressures as well. The residual gas (Figure7) level has no direct influence on the combustion duration
CA50 [aTDCf]
Tin
take
[°
C]
STD of CA50 vs Tin
and CA50
1 2 3 4 5 6 7 840
50
60
70
80
90
100
110
1
1.2
1.4
1.6
1.8
2
Figure 14: Standard deviation of CA50
CA50 [aTDCf]
Tin
take
[°
C]
RI [MW/m²] vs Tin
and CA50
1 2 3 4 5 6 7 840
50
60
70
80
90
100
110
2
3
4
5
6
7
8
9
Figure 15: Ringing Intensity
here.
CA50 [aTDCf]
Tin
take
[°
C]
CA10 − CA90 [CAD] vs Tin
and CA50
1 2 3 4 5 6 7 840
50
60
70
80
90
100
110
5
5.5
6
6.5
7
7.5
8
8.5
9
Figure 16: Combustion duration, CA10 to CA90
The NOx emissions can be an issue when the load is in-creased in HCCI mode but the NOx level at this enginespeed and load is very low as seen in Figure 17. Alater combustion timing decreases the combustion tem-perature with further reduction in the NOx level. The in-creased dilution from increased intake pressure is also a
6
reason for the low NOx levels. In Figure 18 the corre-sponding lambda is plotted. Normally when the NOx levelis increased above a set limit [16] the lambda level is usedto operate the engine stoichiometric on cylinder to cylin-der basis.
CA50 [aTDCf]
Tin
take
[°
C]
EI NOx [g/kgfuel
] vs Tin
and CA50
1 2 3 4 5 6 7 840
50
60
70
80
90
100
110
0.05
0.1
0.15
0.2
0.25
0.3
0.35
Figure 17: NOx emissions
CA50 [aTDCf]
Tin
take
[°
C]
Lambda vs Tin
and CA50
1 2 3 4 5 6 7 840
50
60
70
80
90
100
110
1.6
1.7
1.8
1.9
2
2.1
2.2
Figure 18: Lambda
The engine total efficiency can be divided into fractions[21] as seen in Figure 19. The combustion efficiency is de-fined as the relationship between heat released (QhrMEP)and normalized fuel energy/ cycle (FuelMEP). The ther-modynamic efficiency is then the relationship betweenIMEPgross and QhrMEP. Gas exchange efficiency is de-fined as the relationship between IMEPgross and IMEPnet.Indicated efficiency is the relationship between IMEPnet
and FuelMEP.
By using suitable injection timings the combustion effi-ciency in Figure 20 is high at this test point. The maindeviation is from the combustion timing. The thermal ef-ficiency in Figure 21 on the other hand show its highestefficiency at low intake temperatures where the mass flowis low. At this low intake temperature there is a need foradvanced exhaust valve timing. The heat losses duringthe recompression phase can be an issue but this is more
Figure 19: Engine efficency break-down
pronounced at low engine speeds. There is also an in-crease in thermal efficiency around 80
◦ C. The gas ex-change efficiency in Figure 22 scales to the intake pres-sure and originates from the turbocharger performance(Figure 11) and throttling losses (Figure 12) associatedwith these type of short duration valve lifts at increasedload. The indicated efficiency Figure 23 is then the sumof the combustion, thermal and gas exchange efficiency.It is apparent that the intake temperature has strong influ-ence on total efficiency where the operating temperaturehas to be balanced against combustion noise and stability.
CA50 [aTDCf]
Tin
take
[°
C]
ηcombustion
[%] vs Tin
and CA50
1 2 3 4 5 6 7 840
50
60
70
80
90
100
110
98.4
98.5
98.6
98.7
98.8
98.9
99
99.1
99.2
Figure 20: Combustion efficiency
The next test is on how external EGR influences engineoperation. The EGR fraction is measured as the CO2 ratiobetween the exhaust and the intake manifold.
EGR AND INTAKE TEMPERATURE SWEEP AT 2250RPM AND 400 KPA IMEPNET The combustion timingduring these tests has a CA50 position at 4o (aTDCf ) ex-cept at the highest temperatures where no more than 3o
(aTDCf ) could be used due to increasing back pressureas described earlier. From this test there are some addi-tional graphs that can be found in the APPENDIX.
In Figure 24 the available boost pressure decreases with
7
CA50 [aTDCf]
Tin
take
[°
C]
ηthermal
[%] vs Tin
and CA50
1 2 3 4 5 6 7 840
50
60
70
80
90
100
110
39
40
41
42
43
44
Figure 21: Thermal efficiency
CA50 [aTDCf]
Tin
take
[C
]
ηGE
vs Tin
and CA50
1 2 3 4 5 6 7 840
50
60
70
80
90
100
110
78
79
80
81
82
83
84
85
86
Figure 22: Gas exchange efficiency
CA50 [aTDCf]
Tin
take
[°
C]
ηindicated
[%] vs Tin
and CA50
1 2 3 4 5 6 7 840
50
60
70
80
90
100
110
31
32
33
34
35
36
37
Figure 23: Indicated efficiency
increasing EGR levels. This is a function of the increasedburned gas fraction, see Figure 25, where the charge willhave an increased heat capacity. This leads to a de-creased mass flow rate for a given combustion timing asseen in Figure 26.
The pumping losses are divided in two parts as describedearlier. In Figure 27 the influence of increased EGR showsome variation in turbocharger performance at lower in-
EGR [%]
Tin
take
[°
C]
Pin
[kPa] vs Tin
and EGR
0 5 10 15 20 25 3040
50
60
70
80
90
100
110
120
130
150
160
170
180
190
200
Figure 24: EGR sweep, Intake pressure
EGR [%]
Tin
take
[°
C]
xb vs T
in and EGR
0 5 10 15 20 25 3040
50
60
70
80
90
100
110
120
130
0.42
0.44
0.46
0.48
0.5
0.52
0.54
0.56
0.58
0.6
Figure 25: EGR sweep, Burned gas fraction
EGR [%]
Tin
take
[°
C]
Qm
[kg/s] vs Tin
and EGR
0 5 10 15 20 25 3040
50
60
70
80
90
100
110
120
130
0.02
0.025
0.03
0.035
0.04
Figure 26: EGR sweep, Mass flow
take temperatures. The largest decrease in throttlinglosses occur at low intake temperatures and high EGRlevels, see Figure 28. Here the mass flow is reducedwhich suites current valve timings.
The combustion stability seems to have an optimumaround an intake temperature of 60
◦ C when looking atthe CoV of IMEPnet in Figure 29. On the other hand, ifwe look at the standard deviation of CA50 in Figure 30the picture broadens. Here an increase of EGR leads toless combustion fluctuations except at low intake temper-
8
EGR [%]
Tin
take
[°
C]
Pex
− Pin
[kPa] vs Tin
and EGR
0 5 10 15 20 25 3040
50
60
70
80
90
100
110
120
130
20
25
30
35
40
45
50
Figure 27: EGR sweep, Pressure losses over turbine–compressor
EGR [%]
Tin
take
[°
C]
Throttlingvalve
[kPa] vs Tin
and EGR
0 5 10 15 20 25 3040
50
60
70
80
90
100
110
120
130
40
45
50
55
60
65
70
Figure 28: EGR sweep, Throttling losses
atures.
EGR [%]
Tin
take
[°
C]
CoV of IMEPnet
[%] vs Tin
and EGR
0 5 10 15 20 25 3040
50
60
70
80
90
100
110
120
130
1.2
1.25
1.3
1.35
1.4
1.45
1.5
Figure 29: EGR sweep, CoV of IMEPnet
Since the combustion noise, quantified here with RI, isof major interest during turbocharged HCCI operation, itis seen that increased external EGR has different effectson RI depending on intake temperature, see Figure 31. Athigher intake temperatures the RI level is stable even if theintake pressure is decreasing at the same time, see Fig-
EGR [%]
Tin
take
[°
C]
STD of CA50 vs Tin
and EGR
0 5 10 15 20 25 3040
50
60
70
80
90
100
110
120
130
1
1.2
1.4
1.6
1.8
Figure 30: EGR sweep, Standard deviation of CA50
ure 24. At lower temperatures RI is increasing with higherEGR levels. The corresponding combustion duration inFigure 32 reflects this behavior. There is no clear connec-tion between the combustion duration and the burned gasfraction shown in Figure 25. In normal engine operatingan increase of EGR level will delay the combustion timingdue to increased in-cylinder heat capacity, but here thecombustion timing is kept constant. In Figure 16 it couldbe seen that the combustion timing is the main parameterfor the combustion duration.
EGR [%]
Tin
take
[°
C]
RI [MW/m²] vs Tin
and EGR
0 5 10 15 20 25 3040
50
60
70
80
90
100
110
120
130
3.5
4
4.5
5
5.5
6
6.5
Figure 31: EGR sweep, Ringing Intensity
Finally the indicated efficiency in Figure 33 shows no realgain to operate with external EGR at this load and speedpoint (and these CA50 timings) except at the point withlow intake temperature and high EGR level. But recallthat the RI level and the cycle to cycle variations are highhere.
Since the combustion timing is fixed during these tests thequestion rises of what happens if the combustion timingis swept with and without external EGR and the intaketemperature is kept constant.
9
EGR [%]
Tin
take
[°
C]
CA10− CA90 [CAD] vs Tin
and EGR
0 5 10 15 20 25 3040
50
60
70
80
90
100
110
120
130
6
6.2
6.4
6.6
6.8
7
7.2
Figure 32: EGR sweep, Combustion duration, CA10 toCA90
EGR [%]
Tin
take
[°
C]
ηindicated
[%] vs Tin and EGR
0 5 10 15 20 25 3040
50
60
70
80
90
100
110
120
130
30
31
32
33
34
35
36
Figure 33: EGR sweep, Indicated efficiency
CA50 SWEEP WITH AND WITHOUT EGR AT 2250 RPMAND 400 KPA IMEPNET To further expand this loadand speed point, a CA50 sweep is done with and with-out EGR at an intake temperature of 60
◦ C. This intaketemperature was chosen since it has acceptable RI lev-els, low cycle to cycle variations, a broad combustion tim-ing range and rather high total efficiency. The EGR level isset around 19%, this range of positive displacement couldbe reached without any increase in backpressure. In Fig-ure 34 the boost pressure is decreasing with EGR andso is also the throttling losses. The pressure differentialbetween the intake and exhaust (Pex − Pin) is not reallyaffected by the increased EGR here.
In Figure 35 the CoV of IMEP is reduced with EGR atlate CA50 timings, this trend is even more pronouncedwhen looking at the STD of CA50. Notable is also thatthe operating range is wider with EGR. The combustionduration, CA10 to CA90, shows no real influence of EGR.
The RI is higher at early CA50 positions when using EGRwhich reflects the relative lower dilution level, see Fig-ure 36. The other pressure rise expressions, dP/CAD anddP/ms, also reflect this at a smaller scale.
140
160
180CA50 sweep, Intake temperature ~60°C
P in [k
Pa]
0% EGR19% EGR
0
20
40
P ex−
Pin
[kP
a]
1 2 3 4 5 6 7 8 940
50
60
CA50 [aTDCf]
Thr
ottli
ng [k
Pa]
Figure 34: CA50 sweep, with and without EGR
0
2
4CA50 sweep, Intake temperature ~60°C
CoV
IME
P net [%
]
0% EGR19% EGR
1
2
3
Std
of C
A50
1 2 3 4 5 6 7 8 90
10
20
CA50 [aTDCf]C
A10
− C
A90
[CA
D]
Figure 35: CA50 sweep, with and without EGR
0
5
10CA50 sweep, Intake temperature ~60°C
RI [
MW
/m²]
0% EGR19% EGR
0
5
10
dP m
ax [b
ar/C
AD
]
1 2 3 4 5 6 7 8 90
5
10
CA50 [aTDCf]
dP m
ax [M
Pa/
ms]
Figure 36: CA50 sweep, with and without EGR
In Figure 37 the combustion efficiency is higher with EGRbut this effect is counteracted at early CA50 timings wherethe thermal efficiency is lower. The rise in thermal ef-ficiencies at late CA50 timings reflects the more stablecombustion position with EGR which has less cyclic fluc-tuations than without EGR. The gas exchange efficiencyis improved slightly with EGR due to reduced throttlinglosses as shown in Figure 34. Finally the indicated effi-ciency shows that it can be beneficial to operate with EGRdue to better combustion stability at late combustion tim-ings.
10
98
99
100CA50 sweep, Intake temperature ~60°C
η com
bust
ion [%
]
0% EGR19% EGR
35
40
45
η ther
mal [%
]
80
85
90
η GE [%
]
1 2 3 4 5 6 7 8 930
35
40
CA50 [aTDCf]
η indi
cate
d [%]
Figure 37: CA50 sweep, with and without EGR
DISCUSSION
When operating a turbocharged HCCI engine with NVO,the extra freedom to adjust the intake temperature is at-tractive from a combustion perspective where one can op-erate the engine at the most beneficial temperature. Forexample if the pressure rise rate is high we want a highintake pressure. Choosing a higher intake temperatureleads to higher intake pressure and the combustion noisecan be reduced. However, this has to be weighted againstthe increased pumping losses with reduction in efficiency.Too high intake pressure can result in loss of turbochargerperformance and this was the case at the highest intaketemperatures where the increasing backpressure limitedthe available combustion timing range. This can be cir-cumvented by appropriate intake valve timing which isconstrained in these tests. A shrinking combustion tim-ing range puts even higher demands on the control sys-tem, especially during transients where a broad combus-tion timing range is wanted.
At this speed and load point, 2250 rpm and IMEPnet 400kPa, the best efficiency was reached with low intake tem-peratures but at the same time there is a loss of com-bustion stability. The loss of combustion stability is notreally obvious when only using the CoV of IMEPnet as ameasurement. The standard deviation of CA50 visualizesthe combustion stability more clearly. By using externalEGR the combustion range and stability can be increased.EGR show no big influence on combustion duration andefficiency. The usage of an external EGR systems addedweight, cost and control complexity, has to be considered.
The combustion timing is the major variable during thesetests since it directly control the dilution level and the in-take temperature level that can be reached. This givesthe combustion timing a huge impact on test results inHCCI mode. This further emphasizes the importance ofrobust combustion timing control in turbocharged HCCImode. The prediction of the in-cylinder charge tempera-ture and properties at IVC can be used by the control sys-tem. Combined with a DI fuel system there is a possibilityto influence the combustion timing at a late stage.
To find the optimum operating intake temperature and
combustion timing one has to balance factors like pres-sure rise rate, efficiency, combustion stability/ range andemissions level against each other. This study is only per-formed at one engine speed and load point so the ques-tion is how the results are at other speed/ load points, howchanges in turbocharger, valve timings etc. influences theresult. This need to be investigated in future work.
CONCLUSIONS
From this test point at 2250 rpm and 400 kPa IMEPnet
• Available intake pressure scales directly with intaketemperature and combustion timing, it can thereforebe modulated with charge cooling and valve timings.
• An increase of intake pressure decreases peak pres-sure rise rate but has to be weighed against increas-ing pumping losses.
• The highest indicated efficiency at this study wasfound with the lowest intake temperature, but in prac-tice, the high pressure rise rate and cyclic variationsmakes it an unsuitable working area.
• Standard deviation of the combustion phasing is agood measurement of combustion stability that cancomplement CoV of IMEPnet for the cycle to cyclebehavior.
• The usage of external EGR is shown to improve com-bustion stability and extend usable combustion timingrange.
• Turbocharger performance and appropriate valve tim-ings are vital for improving efficiency and operatingrange in a NVO HCCI engine.
• Test results are heavily influenced by the combustiontiming.
11
ACKNOWLEDGMENTS
This project is supported by GM Powertrain and theSwedish Energy Agency
REFERENCES
[1] Kalghatgi, G. T, Risberg, P., Angstrom, H-E., ”Ad-vantages of fuels with high resistance to auto-ignition, low temperature, compression ignitioncombustion,” SAE Paper 2006-01-3385.
[2] Marriot, C. D., Reitz, R. D., ”Experimental investiga-tion of direct injection- gasoline for premixed com-pression ignited combustion phasing control,” SAEPaper 2002-01-0418.
[3] Urushihara, T., Hiraya, K., Kakuhou, A., Itoh, T.,”Expansion of HCCI operating region by the combi-nation of direct fuel injection, negative valve overlapand internal fuel reformation,” SAE Paper 2003-01-0749.
[4] Dec, J. E., Hwang, W., Sjoberg, M., ”An investiga-tion of thermal stratification in HCCI engines usingchemiluminescence imaging,” SAE Paper 2006-01-1518.
[5] Thring, R. H., ”Homogeneous-charge compression-ignition (HCCI) engines,” SAE Paper 892068.
[6] Christensen, M., Hultqvist, A., Johansson, B.,,”Demonstrating the multi-fuel capability of a homo-geneous charge compression ignition engine withvariable compression ratio,” SAE Paper 1999-01-3679.
[7] Haraldsson, G., Tunestal, P. , Johansson, B.,Hyvonen, J., ”HCCI combustion phasing in a multi-cylinder engine using variable compression ratio,”SAE Paper 2002-01-2858.
[8] Lang, O., Salber, W., Hahn, J., Pischinger, S., Hort-mann, K., Bucker, C., ”Thermodynamical and me-chanical approach towards a variable valve trainfor the controlled auto ignition combustion Process,”SAE Paper 2005-01-0762.
[9] Martinez- Frias, J., Aceves, S. M., Flowers, D.,Smith, R., Dibble, R., ”HCCI engine control by ther-mal management,” SAE Paper 2000-01-2869.
[10] Haraldsson, G., Tunestal, P. , Johansson, B.,Hyvonen, J.,”HCCI closed- loop combustion controlusing fast thermal management,” SAE Paper 2004-01-0943.
[11] Persson, H., Agrell, M., Olsson, J-O. ,Johansson,B., Strom, H.,”The effect of intake temperature usingnegative valve overlap,” SAE Paper 2004-01-0944.
[12] Christensen, M., Johansson, B., Amneus, P.Mauss, F., ”Supercharged homogeneous chargecompression ignition,” SAE Paper 980787.
[13] Christensen, M., Johansson, B., ”Supercharged ho-mogeneous charge compression ignition (hcci) withexhaust gas recirculation and pilot fuel,” SAE Paper2000-01-1835.
[14] Hyvonen, J., Haraldsson, G., Johansson, B., ”Su-percharging HCCI to extend the operating rangein a multi-cylinder VCR-HCCI engine,” SAE Paper2003-01-3214.
[15] Olsson, J-O., Tunestal, P., Johansson, B., ”Boostingfor high load HCCI,” SAE Paper 2004-01-0940.
[16] Johansson, T., Johansson, B., Tunestal, P., Aulin,H., ”HCCI operating range in a turbo-charged multicylinder with VVT and spray- guided DI,” SAE Paper2009-01-0494.
[17] Sjoberg, M., Dec, J.E., Hwang, W., ”Thermody-namic and Chemical Effects of EGR and Its Con-stituents on HCCI Autoignition,” SAE Paper 2007-01-0207.
[18] Olsson, J-O., Tunestal, P., Ulfvik, J., Johansson, B.,”THe effect of cooled EGR on emissions and perfor-mance of a turbocharged HCCI engine,” SAE Paper2003-01-0743.
[19] Andreae, M.M, Cheng, W.K., Kenney, T., Yang,J., ”On HCCI Engine Knock,” SAE Paper 2007-01-1858.
[20] Eng, J.A., ”Characterization of pressure waves inHCCI combustion,” SAE Paper 2002-01-2859.
[21] Hyvonen, J., Wilhelmsson, K., Johansson, B., ”Theeffect of displacement on air-diluted, multi-cylinderHCCI engine performance,” SAE Paper 2006-01-0205.
CONTACT
Thomas Johansson, M.Sc.ME
E-mail: [email protected]
Lund University, Faculty of Engineering, Department ofEnergy Sciences, Division of Combustion Engines, P.O.Box 118, SE-221 00 Lund, Sweden
12
DEFINITIONS AND ABBREVIATIONS
aTDCf : After Top Dead Center, firingBDC: Bottom Dead CenterCA10: Crank angle 10% burnedCA50: Crank angle 50% burnedCA90: Crank angle 90% burnedCB: Cylinder BalancingCAD: Crank Angle DegreesCoV: Coefficient of VariationDI: Direct InjectiondP: Pressure derivateEI: Emission IndexEGR: Exhaust Gas RecirculationEVC: Exhaust Valve ClosingEVO: Exhaust Valve Openηcombustion: Efficiency, combustionηGE : Efficiency, gas exchangeηindicated: Efficiency, indicatedηthermal: Efficiency, thermalFuelMEP: Fuel Mean Effective PressureHCCI: Homogeneous Charge Compression IgnitionIMEPgross: Indicated Mean Effective Pressure, grossIMEPnet: Indicated Mean Effective Pressure, netIVC: Intake Valve ClosingIVO: Intake Valve OpeningNA: Naturally AspiratedNOx: Nitrogen OxideNVO: Negative Valve OverlapPex: Pressure ExhaustPin: Pressure IntakePCP: Peak Cylinder PressurePMEP: Pumping Mean Effevtive PressureQemisMEP: Emission Mean Effevtive PressureQhrMEP: Heat Release Mean Effevtive PressureQlossMEP: Heat losses Mean Effevtive PressureRI: Ringing IntensitySI: Spark IgnitionSTD: Standard deviationTDC: Top Dead CenterTDCf : Top Dead Center, firingTDCGE: Top Dead Center, gas exchangeTIV C ; Temperature, intake valve closingVVT: Variable Valve Timingxb: Burned gas fraction
13
APPENDIX
Complementary graphs from CA50 and intake tempera-ture sweep at 2250 rpm and 400 kPa IMEPnet
CA50 [aTDCf]
Tin
take
[°
C]
Qm
[kg/s] vs Tin
and CA50
1 2 3 4 5 6 7 840
50
60
70
80
90
100
110
0.025
0.03
0.035
0.04
Figure 38: Mass flow
CA50 [aTDCf]
Tin
take
[°
C]
dP max [bar/CAD] vs Tin
and CA50
1 2 3 4 5 6 7 840
50
60
70
80
90
100
110
3
3.5
4
4.5
5
5.5
Figure 39: Pressure rise, bar/ CAD
CA50 [aTDCf]
Tin
take
[°
C]
dP max [MPa/ms] vs Tin
and CA50
1 2 3 4 5 6 7 840
50
60
70
80
90
100
110
4
4.5
5
5.5
6
6.5
7
7.5
8
Figure 40: Pressure rise, MPa/ ms
Complementary graphs from EGR sweep at 2250 rpmand 400 kPa IMEPnet
EGR [%]
Tin
take
[°
C]
dP max [bar/CAD] vs Tin
and EGR
0 5 10 15 20 25 3040
50
60
70
80
90
100
110
120
130
4
4.2
4.4
4.6
4.8
5
Figure 41: EGR sweep, Pressure rise dP/ CAD
EGR [%]
Tin
take
[°
C]
dP max [MPa/ms] vs Tin
and EGR
0 5 10 15 20 25 3040
50
60
70
80
90
100
110
120
130
5.4
5.6
5.8
6
6.2
6.4
6.6
Figure 42: EGR sweep, Pressure rise dP/ ms
EGR [%]
Tin
take
[°
C]
PMEP [kPa] vs Tin
and EGR
0 5 10 15 20 25 3040
50
60
70
80
90
100
110
120
130
60
70
80
90
100
110
120
Figure 43: EGR sweep, Pumping losses
14
EGR [%]
Tin
take
[°
C]
EI NOx [g/kgfuel
] vs Tin
and EGR
0 5 10 15 20 25 3040
50
60
70
80
90
100
110
120
130
0.05
0.1
0.15
0.2
0.25
0.3
0.35
0.4
Figure 44: EGR sweep, NOx emissions
EGR [%]
Tin
take
[°
C]
ηcombustion
[%] vs Tin and EGR
0 5 10 15 20 25 3040
50
60
70
80
90
100
110
120
130
98.8
98.85
98.9
98.95
99
99.05
99.1
99.15
99.2
99.25
Figure 45: EGR sweep, Combustion efficiency
EGR [%]
Tin
take
[°
C]
ηthermal
[%] vs Tin and EGR
0 5 10 15 20 25 3040
50
60
70
80
90
100
110
120
130
37
38
39
40
41
Figure 46: EGR sweep, Thermal efficiency
EGR [%]
Tin
take
[°
C]
ηGE
vs Tin
and EGR
0 5 10 15 20 25 3040
50
60
70
80
90
100
110
120
130
76
78
80
82
84
86
Figure 47: EGR sweep, Gas exchange efficiency
15