Effects of Gasoline Viscosity and Injection Pressure

The Ninth International Conference on Modeling and Diagnostics for Advanced Engine Systems (COMODIA 2017), July 25-28, 2017, Okayama, Japan

C309

Effects of Gasoline Viscosity and Injection Pressure on the Performance and Emissions of a Multi-Cylinder Partially Premixed Combustion Engine *Bin Mao, Qiping Wang, Jialin Liu, Haifeng Liu, Zunqing Zheng, Mingfa Yao State Key Laboratory of Engines, Tianjin University No.92 Weijin Road, Nankai District, Tianjin 300072, China Key Words: Partially Premixed Combustion, Viscosity index improvement, Injection pressure, Soot emissions, Load limit

ABSTRACT Gasoline partially premixed combustion (PPC) is a promising combustion concept with high indicated thermodynamic efficiency (ITE) and low emission level. This study investigated the effect of gasoline viscosity index improvement on the multi-cylinder engine performance and the dependence of the PPC combustion on injection pressure over a wide range of engine speeds and loads. Results show that a small amount of viscosity improver can improve the viscosity of gasoline effectively without influencing the octane numbers. This is beneficial for elevating the mechanical efficiency of the fuel pump and improving the brake thermal efficiency (BTE) of the engine. For the part and medium load conditions, the requirement of common-rail pressure in PPC mode with engine-out NOx and soot emissions below Euro 6 levels is significantly lower than the calibration of conventional diesel combustion (CDC) mode with tailpipe Euro 6 emissions. For the high-load conditions, the low-speed operations are prone to achieve high premixing. While the highspeed operations are mainly mixing-controlled with similar NOx-soot trade-off relationship compared to CDC mode under moderate injection pressures. The fuel injection pressure becomes the most important driving factor for air-fuel mixing in PPC mode. The local equivalence ratio dominated by the injection pressure is more important than the global equivalence ratio for soot reductions. As a result, gasoline PPC load extension is very sensitive to rail pressure, and its injection pressure requirement is much higher than that of the Euro 6 CDC mode.

INTRODUCTION In the past decade, main progress of homogeneous charge compression ignition (HCCI) has been made when electronic control technology was becoming mature and fuel property was deliberately considered and designed. For example, gasoline is directly injected into a diesel engine [1] and some novel combustion concepts have been proposed based on this, such as low temperature gasoline combustion (LTGC) [2], gasoline direct injection compression ignition (GDCI) [3], and gasoline partially premixed combustion (gasoline PPC) [4]. These typical gasoline compression ignition (GCI) concepts have some differences from the perspective of in-cylinder fuel stratification, which leads to their difference in the requirement of common-rail pressures. LTGC uses port fuel injection to create a nearly homogeneous charge and uses direct injections to create sequential auto-ignition events. LTGC has the lowest injection pressure requirement. However, it tends to display similar challenges as HCCI because they are both kinetically controlled. GDCI increases the stratification level by directly injecting all the fuels during the compression stroke via multi-injection strategy. The

moderate stratified charge is to simultaneously reduce NOx and soot emissions while maintaining some level of control over the combustion process. For LTGC and GDCI, the mixing is dominated by ignition delay rather than injection event. Hence, there is a great scope for reducing the cost of the injection system compared to modern diesel engines by reducing the injection pressure [5]. However, their compression ratios are commonly lower than the conventional level in order to control the combustion noise and to achieve premixing charge. And their moderate combustion efficiency is another challenge to thermal efficiency improvement. Gasoline PPC is an important gasoline HCCI variant with full load range capability of clean and efficient combustion [6]. Gasoline PPC utilizes the highest level of fuel stratification and typically features direct injection relatively close to the top dead center (TDC). Hence an ideal combustion phasing can be guaranteed and a high compression ratio can be applied compared to LTGC. The injection pressure requirement of gasoline PPC is significantly lower than that of the diesel PPC [7], but it is still the highest in all the GCI modes. It is generally considered that the ratio of premixed Copyright @ 2017 by the Japan Society of Mechanical Engineers


heat release is the critical factor influencing the soot emissions. However, studies show that the reactivity of gasoline has little effect on the ignition delay during highload operation [8]. By an up-to-date Cummins XPI injection system with a maximum rail pressure capability of 240 MPa, the injection and combustion process of full load PPC operation is narrowly separated, but its full load combustion has a more diffusion like heat release pattern. While it has to be noted that the heat release rate (HRR) of diffusion combustion part is high enough to rival the premixed HRR peak of medium and high loads with fairly Gaussian shapes. Hence full load range soot-free combustion can be achieved. The superiority of gasoline molecular structure with shorter molecular length is also considered to be the other contributor [4]. Furthermore, the short combustion duration further improves thermaldynamic efficiency. The lubrication of fuel pump and injector components relies on diesel fuel itself. Unfortunately, the viscosity of gasoline fuels is quite low compared with diesel fuel, which means gasoline generates only very thin hydrodynamic and elastohydrodynamic films. It is generally known that the up-to-date production GDI fuel system is just capable of the continuous operating pressure of 40 MPa. However, researchers commonly conduct PPC experiments on diesel engine without any modification on the fuel system. Gasoline and naphtha fuels are just added with diesel fuel lubricity additive on the order of parts per million (ppm). Many researchers [9] argue that it’s difficult to maintain a high fuel pressure with gasoline fuel at high mass flow demand probably due to cavitation, vapor lock, and internal leakage. Lewander et al. [10] reported that the maximum rail pressure for multi-cylinder gasoline PPC operation reduced nearly 40% compared to its design performance. Furthermore, the mechanical efficiency of gasoline PPC operation is lower than that of CDC mode. Tuner et al. [11] estimated that an FMEP increase of 15% related to fuel pump yielded 0.4% reduction in brake efficiency of a gasoline PPC engine. Owing to this, the above problems greatly hinder the transfer work from PPC concept into a production viable engine. Although the XPI fuel system has not demonstrated any particular problems for single cylinder PPC operation, little literature is yet reported about its mechanical performance and reliability on a multicylinder engine. Therefore, this article aims to achieve quantitative result about the effects of gasoline fuel viscosity on the mechanical efficiency of a diesel engine with a common-rail system. It is necessary to explore effective methods to improve gasoline PPC performance. Furthermore, the sensitivity of gasoline PPC operations to the variations of engine load, speed, and fuel injection pressures is presented. The relationship between the emission control and high injection pressure requirement is studied.

EXPERIMENTAL SETUP AND METHODS Engine and instrumentation An experimental investigation was conducted on a

six-cylinder heavy-duty diesel engine. Some modifications have been made. The original waste-gated turbocharger has been replaced by a two-stage turbocharger with an inter-stage cooler. The compression ratio is decreased from 17.5 to 16.8. In order to achieve a high EGR rate and an optimum trade-off performance between the gross indicated thermal efficiency (ITEgross) and pumping losses, a dual loop EGR (DL-EGR) system is applied. In this test, the split strategy of DL-EGR and rack position of variable geometer turbocharger (VGT) follow the results and suggestions from a previous study [13]. The key specifications of the engine are listed in Table 1. The schematic of the experimental setup is given in Fig. 1. A stock common rail injection system designed for diesel engines is used and it is the third-generation fuel system from Bosch. The key specifications of the fuel system are listed in Table 2. To ensure the repeatability and comparability of the measurements, the cooling water temperature was automatically controlled by a temperature controller to 85°C. Table 1. Engine specifications. Engine Type

DI In-line 6, water cooled

Bore×Stroke

113 mm ×140 mm

Connecting Rod

209 mm

Displacement

8.42L

Compression Ratio

16.8

Swirl Ratio

1.25

Number of Valves

4

Max Torque/Speed

1280 N.m/(1200-1700 r/min)

Rated Power/Speed

243 kW/2200 r/min

Table 2. Fuel system specifications. Injection System

Common rail

Injection Nozzle

0.163 mm×8-148°

Injector Hydraulic Flow (cc per 30sec@10 MPa)

690

High Pressure Pump Style

Bosch CPN2-18

Maximum Injection Pressure

180 MPa

Lubrication of the Pump

Oil

Pre-Supply Pump

Gear Pump

The cylinder pressure was measured using a pressure sensor (Kistler 6125A) with a corresponding charge amplifier and data acquisition system. The pressure data were taken in every 0.5°CA and the cylinder pressure was the ensemble average of 100 consecutive engine cycles. Gaseous emissions were measured using a gas analyzer (HORIBA MEXA 7100DEGR). Smoke was measured using an AVL 415S filter paper smoke meter and averaged between 5 samples of a 2 L volumes. According to the specification of the soot meter, the specific dry soot can be calculated from the filter smoke number (FSN) [12]. More information on the setup and procedures can be found in the reference [13]. For this study, the ringing intensity (RI) developed by Copyright @ 2017 by the Japan Society of Mechanical Engineers


Eng [14] was used to quantify the knock level: RI =

1 2γ

∗

2 𝑑𝑃 (0.05∗ ) 𝑑𝑡 𝑚𝑎𝑥

P𝑚𝑎𝑥

∗ (𝛾𝑅T𝑚𝑎𝑥 )0.5

(1)

Equation 1 shows that the RI is proportional to the square of the absolute maximum pressure-rise rate (in kPa/ms). The use of a time-based pressure-rise rate for computing ringing/knock is justified by the fact that the acoustic timescales are independent of engine speed, and 5 MW/m2 was considered to be the limit used for this work.

control the wear rate of the moving parts such as the plunger in the high pressure pump, a fuel lubricity improver (AFTON HiTEC® 4140F), originally designed for low sulfur diesel fuels, was added at a concentration of 500 ppm by volume into the two tested gasoline fuels. Table 3. Specific parameters for the tested fuels. Fuel RON MON Sensitivity Temp. 10% Evaporated (°C) Temp. 50% Evaporated (°C) Temp. 90% Evaporated (°C) Oxygenates ASTM D5599 (%) Specific Gravity ASTM D4052 (–) Kinematic Viscosity(mm2/s) Higher Heating Value (MJ/kg) Lower Heating Value (MJ/kg)

0# diesel

92# gasoline

-

93.4 82.8 10.6

92# gasoline with 2.2% addition (vol. /vol.) 92.2 81.3 10.9

216.2

57.4

65.4

276.0

90.7

96.7

318.3

160.4

166.7

-

1.7

1.16

0.824

0.742

0.754

4.51 @20°C

0.96 @-20°C

1.35 @-20°C

45.8

46.4

46.4

42.8

43.4

43.4

Table 4. Typical characteristics of HiTEC® 5751 viscosity improver.

Fig. 1. Schematic of the experimental setup.

Test fuels and experimental procedures Commercially available China V 0# diesel fuel and 92# gasoline fuel were provided by Sinopec Corporation and Shell Corporation, respectively. The specific parameters of tested fuels used in this work are listed in Table 3. The viscosity of diesel and gasoline shows such a great difference. Thus, the HiTEC® 5751 olefin copolymer (OCP) viscosity improver, provided by Afton Chemical Corporation, was selected as an additive in this test. Table 4 shows the typical characteristics of this viscosity improver. This kind of viscosity improver is generally recommended for use when formulating crankcase and industrial oils. Fig. 2 shows the tested viscosity index of gasoline fuels as a function of the dosage rate of viscosity improver. HiTEC® 5751 may not be the best choice for gasoline viscosity improvement to fulfill the demand of PPC combustion, but as an additive, it can effectively change the viscosity of gasoline by a very low dosage rate and meet the requirements of the experimental design. Fig. 2 also shows that the increase rate of viscosity index speeds up obviously for the dosage rates higher than 3%. Due to inexperience in gasoline viscosity index improvement, a moderate dosage rate of 2.2% was applied in this test to ensure the safety of the fuel system. Table 3 shows that the viscosity improver offers an effective improvement without major influence on other physicochemical parameters, such as the boiling range and the lower heating value (LHV). In order to

Appearance

Clear to slightly hazy greenish tan viscous liquid

Specific Gravity@ 15.6/15.6°C

0.858

Density, lbs/gal.

7.14

Flash Point, °C (PMCC)

135 (minimum)

Viscosity @ 40°C, mm2/s

15,000

Viscosity @ 100°C, mm2/s

1240

Diluted Viscosity @ 100°C

13.0

SSI

50

Fig. 2.

Tested viscosity index of gasoline fuels as a function of the dosage rate of viscosity improver.

Firstly, the tests were performed at the European Stationary Cycle (ESC) operating speed of 1660 r/min with a constant gasoline fuel delivery of 70 mg/cycle in each cylinder. The fuel mass of diesel delivery was adjusted to maintain the injected total energy to be the Copyright @ 2017 by the Japan Society of Mechanical Engineers


same as that of gasoline fuel. The BMEP of this load is 0.83 ± 0.02 MPa. The injection pressures were set to 100 MPa and 150 MPa. The combustion phasing in terms of CA50 (the crank angle at which 50% completion of heat release) was set to 10°CA ATDC. A double injection strategy without EGR was used to modulate similar heat release patterns of the fuels. The difference in indicated thermodynamic efficiency between fuels can be minimized. The effects of gasoline viscosity index and viscosity improver addition on the fuel economy were quantified in comparison with those of diesel fuel. Then the effects of viscosity improver addition on the combustion process and emissions combined with EGR variation were investigated under a single injection strategy under a rail pressure of 100 MPa. The CA50 timing was set to 12°CA ATDC. Based on these investigations, the effects of viscosity index improver addition on PPC engine performance under different injection pressures can be revealed. Generally, the difference of mechanical efficiency in a fuel system is difficult to determine accurately when running different fuels. One common approach is to drive or motor the engine with a dynamometer (i.e. operate the engine without firing it) and measure the power which has to be supplied by the dynamometer to overcome all the friction losses. However, the major inaccuracy with this method is that the friction power of the fuel pump may be lower in the motored test than when the engine is firing. Another approach is calculating the mechanical efficiency based on the calculated net ITE and measured BTE. However, only one cylinder pressure is measured on this multi-cylinder test bench. The calculated indicated power will be affected by the uniformity between the cylinders. Therefore, the method of cylinder deactivation was used in this test. In detail, the method is to deactivate each cylinder once at a time and record the engine power while maintaining the fuel mass per cycle. Thus, the indicated power of each cylinder is the difference between the normal power and the recording power. Secondly, the test engine was operated with a wide load variation at 1660 r/min and a wide speed variation at 1.43 MPa BMEP at a fixed rail pressure of 100 MPa. The EGR rate was swept at each operating condition. Test conditions and operating boundaries are presented in Table 5 and Table 6, respectively. The engine performance and emissions by running diesel and gasoline with viscosity improver addition are compared. The purpose of this part was to investigate how the advantage of gasoline PPC in performance and emissions over CDC mode changes under different operating conditions. Thirdly, the gasoline PPC load ranges at three ESC operating speeds with Euro 6 compliant engine-out NOx and soot emissions under the rail pressures of 100 MPa, 130 MPa, and 150 MPa were presented. The PPC rail pressure requirement was compared with that of the Euro 6 CDC mode to emphasis the dependence of gasoline PPC operation on the injection pressure.

RESULTS AND DISCUSSIONS Effects of viscosity improver addition on engine performance Engine performance without EGR under a double injection strategy

(a) For the rail pressure of 100 MPa

(b) For the rail pressure of 150 MPa Fig. 3. In-cylinder pressures and HRRs of the three fuels under two rail pressures without EGR. Pre-ignition HRR (inset) at selected crank. Fig. 3 shows the in-cylinder pressures and HRRs of three fuels. The EGR valves were closed to eliminate the suppressing effect of EGR on (low temperature reaction) LTR. The pilot injection mass of diesel was kept to 3 mg. While those of gasoline were increased to 4 mg and 6 mg for the two rail pressures, respectively. The pilot-main intervals of diesel and gasoline were set to 10°CA and 12°CA, respectively. The start of injection (SOI) of main injection changed slightly to keep a constant CA50 timing. The heat release phasing of gasoline pilot injection is later than that of diesel as shown in Fig. 3(a) due to the higher temperature requirement for gasoline auto-ignition. The HRR peak of gasoline pilot injection is higher than that of diesel since the pilot HRR of gasoline partly coincides with the main HRR. The pilot heat release elevates the compression temperature before TDC, which shortens the ignition delay of gasoline main injection. Furthermore, typical LTR traces of gasoline pilot injections are presented in the inset of Fig. 3(b) for a high injection pressure, which is generally considered to produce some reactive species to decrease the ignition Copyright @ 2017 by the Japan Society of Mechanical Engineers


delay, such as methyl peroxide (CH3OOH), ketohydroperoxides (KETs), and hydrogen peroxide (H2O2) [15]. Therefore, the main heat release parts of three fuels show little difference. The similar heat release patterns of the three fuels ensure similar indicated thermodynamic efficiency.

(a) Lambda and combustion efficiency

(b) Pumping loss and pre-turbine temperature

(c) Mechanical efficiency and BTE Fig. 4. Engine performance by fueling diesel and gasoline with and without viscosity improver. Fig. 4 shows the engine performance by fueling diesel and gasoline with and without additive. Fig. 4(a) and 4(b) illustrate that the lambda, combustion efficiency, pumping loss, and pre-turbine temperature of the three fuels show the small difference for each rail pressure. This represents all tested fuels have similar turbocharging performance and corresponding intake conditions. The measuring procedures of mechanical loss in this test engine have been described in the previous section of

experimental procedures. As discussed above, the similar heat release patterns of the three fuels ensure similar ITEgross and pumping losses. Therefore, the difference in mechanical efficiency and BTE should be mainly attributed to the performance of the fuel pump and the lubricity of the fuels. Fig. 4(c) shows that the mechanical efficiency of diesel fuel decreases by 1.5% for a higher rail pressure. While the increase of rail pressure has little influence on the BTE of diesel fuel since the shortened combustion duration improves the ITEgross. For gasoline without additive, the mechanical efficiencies with injection pressure of 100 MPa and 150 MPa are lower than those of diesel fuel by 2.3% and 4.6%, respectively. This is because the poor lubricity of gasoline leads to the higher leakage level and higher pump driving powers. Correspondingly, the increased pump driving powers lead to BTE penalties of 1.3% and 1.8%, respectively. For gasoline with viscosity improver addition, the mechanical efficiency values increase by 1.64% and 1.8%, and the BTEs improve by 0.75% and 0.85%, respectively. This could be closely related to the reduced internal leakage and better lubrication of the sliding contact in the fuel pump. Therefore, the viscosity improver addition is an effective method to reduce the pump driving powers and to increase the mechanical efficiency. The BTE value of gasoline with additive under 100 MPa rail pressure is very close to that of diesel fuel, and the minor difference between them may be related to the slightly lower combustion efficiency and mechanical efficiency. Engine performance with EGR under a single injection strategy Fig. 5 shows the effects of viscosity improver addition on the gasoline PPC combustion process. The combustion phasing in terms of CA50 was fixed to 12°CA ATDC by changing the SOI while varying the EGR rate. The ignition delay is defined as the interval between the SOI and CA10 (the crank angle at which 10% completion of heat release). Fig. 5(a) shows that the viscosity improver addition decreases the PPC ignition delay. Then the combustion duration increases correspondingly. Fig. 5(b) shows that the heat release peak and in-cylinder pressure decrease slightly. It is generally considered that the chemical ignition phenomena under PPC conditions can be best described by using RON, MON, and their sensitivity [16]. Auto-ignition sensitivity of a gasolinelike fuel is the difference between its RON and MON. However, the viscosity improver addition slightly decreases the RON and MON numbers but has little influence on the sensitivity values as shown in Table 3. Under this circumstance, the lower ON numbers may be partly responsible for the shorter ignition delay with viscosity improver addition, but it is also probably due to physical reasons. Unfortunately, neither optical diagnostic and/or simulations were performed. Thus, the authors have to back up their theories using the available literatures. According to the studies [17], the in-cylinder regions with higher local equivalence ratio ignite first for gasoline PPC operations. However, mixtures portions in this region are also the coldest. Furthermore, the viscosity improver addition influences the fuel injection process, Copyright @ 2017 by the Japan Society of Mechanical Engineers


such as the liquid injection, the breakup, and the evaporative cooling. As a result, the increased proportion of regions with the most ignitable local equivalence ratio and the weakened evaporative cooling effect of these regions probably promotes the ignition process.

increment of soot can be efficiently suppressed by a higher injection pressure. While this operates against the original purpose of elevating the mechanical efficiency of a PPC engine. As we know, adding oxygenated additive into pure diesel can efficiently restrict soot generation and reduce CO emission for CI engines. Thus, oxygenated viscosity improver may overcome the trade-off relationship between emissions and performance [18].

Comparison of CDC and PPC modes in high loads For the operating conditions from medium load to full load at 1660 r/min The engine load described above is around 0.83 MPa BMEP. The following test is to extend the PPC operation to the full load condition with a constant rail pressure and to make comparisons of the engine performance and emissions with those of CDC mode. Test conditions and operating boundaries are presented in Table 5. (a) Ignition delay and combustion duration Table 5. Tested conditions and operating boundaries. Parameters Speed [r/min] Load [MPa] Tests Fuels Injection Strategy CA50 [°CA ATDC] Fuel Pressure [MPa] EGR [%] Boost Pressure [MPa]

(b) HRRs and in-cylinder pressures Fig. 5. Effects of viscosity improver addition on the PPC combustion process.

Fig. 6.

Effects of viscosity improver addition on PPC emissions.

Fig. 6 shows the effects of viscosity improver addition on the NOx, soot, CO, and HC emissions of gasoline PPC operations. The NOx and HC emissions are little affected by the viscosity improver addition. While the soot and CO emissions of PPC operations with addition are higher. This is because the soot and CO emissions are closely related to the locally rich fuel parcels which are sensitive to the ignition delays and air/fuel mixing process. In fact, an

Values 1660 0.95 1.43 1.91 EGR Sweep Diesel Gasoline + Viscosity Improver (2.2% by volume) Single Injection 15 100 5…38

5…24

5…20

0.184…0.200

0.244…0.251

0.286…0.294

Fig. 7 shows the comparisons of HRRs and incylinder pressures of CDC and PPC modes. For the three engine loads, the heat release peaks of CDC mode keep at relatively stable values around 200 J/°CA, which is mainly limited by the constant rail pressure of 100 MPa. For PPC mode, the ignition delay is significantly longer than that of CDC mode at 0.95 MPa BMEP, so its heat release pattern possesses a high premixing ratio. Its premixing heat release peak value is 3 times as high as that of CDC mode. The MPRR and RI of PPC mode are significantly higher than those of CDC mode but are still within their safe limits. However, the pressures and temperatures encountered at high-load conditions are sufficiently high to suppress the difference in fuel reactivity between diesel and gasoline fuels. PPC is losing its advantage in ignition delay over CDC with increasing load. Accordingly, the premixed peak and premixed ratio decrease sharply as the load increases. The combustion process is mainly mixing-controlled and driven by the injection process. Therefore, the CD, MPRR, and RI show the little difference between the two modes. For 1.91 MPa BMEP, the full load condition at this engine speed, the peak firing pressure reaches the limit of this engine, which becomes a limiting factor for combustion process optimization.

Copyright @ 2017 by the Japan Society of Mechanical Engineers


(a) For the load of 0.95 MPa BMEP

mode by 1.1%, 0.8%, and 1.3% from the medium load to full load, respectively. As previously discussed at 0.83 MPa BMEP, the two combustion modes share similar heat release patterns under the double injection strategy. The BTE of PPC is only lower than that of CDC mode by 0.5% mainly related to the lower mechanical efficiency and combustion efficiency. Under this circumstance, it can be inferred that the widened BTE gap at 0.95 MPa BMEP should be also related to another factor, namely increased heat transfer, since the peak mean combustion temperature of PPC mode is higher than that of CDC mode by 110 K shown in Fig. 7(a). As the difference in HRR decreases with a higher load, the influence of heat transfer on the thermal efficiency is decreasing. In addition, the combustion efficiency of the two modes shows a little difference at the high and full loads. With this in mind, the increasing trend of the BTE difference between the two combustion modes for a higher load is supposed to be related to the decreased mechanical efficiency of PPC mode as the fuel delivery increases. But this hypothesis about the fuel system performance should be further validated and will be the aim of a future study.

(b) For the load of 1.43 MPa BMEP

Fig. 8.

(c) For the load of 1.91 MPa BMEP Fig. 7. Comparisons of HRRs and in-cylinder pressures of CDC and PPC modes at the speed of 1660 r/min. Fig. 8 shows the engine performance comparisons between the CDC and PPC modes. The lambda and the pressure difference between the intake and exhaust indicate that the satisfactory performance of this air system contributes to a good gas exchange efficiency at these operating conditions. The PPC combustion efficiency is as high as that of CDC mode. However, the BTEs of PPC mode are still lower than those of CDC

Engine performance comparisons between CDC and PPC modes at the EGR rate of 20%.

Fig. 9 illustrates the NOx emissions versus EGR rates and the NOx-soot trade-off performance. As expected, the NOx emissions of PPC at 0.95 MPa BMEP are much higher than that of CDC mode, since the high premixing heat release patterns lead to higher combustion temperatures. For the higher loads, the NOx emissions of the two modes are similar to each other. As shown in Fig. 9(b), the PPC soot emissions at 0.95 MPa BMEP are extremely low over the whole range of NOx emissions. While the CDC soot emissions start increasing quickly when the NOx is decreasing from 3.0 g/kW-h. Generally, gasoline fuel has many soot emission advantages over diesel fuel, such as fewer poly-aromatics and poly-cycloparaffins, shorter average carbon length, better volatility, less liquid spray length, less impingement of the spray on the piston wall, and so on. These physicochemical advantages are supposed to be favorable for gasoline to obtain lower soot emissions at high load Copyright @ 2017 by the Japan Society of Mechanical Engineers


than diesel fuels. However, the NOx-soot performance of PPC mode under 1.43 MPa BMEP is much closer to that of CDC mode since the combustion process is mainly mixing controlled. Unexpectedly, for the 1.91 MPa BMEP, the trade-off performance of PPC mode almost coincides with that of CDC mode. This means the physicochemical advantages of gasoline show little help in emission control under this heat release pattern with high diffusion combustion ratio. As a matter of fact the ignition delay of PPC mode is longer than CDC mode by 3°CA, and there still exists a small amount of premixed combustion in the PPC heat release pattern as shown in Fig. 7(c). According to the authors, the PPC soot performance at the full load conditions may be due to: • Global lambda, less oxygen in the mixture. • Mixing period, from End of Injection (EOI) to CA10. • Local lambda, namely air/fuel mixing that closely related to injection pressure.

(a) NOx emissions versus EGR

according to [19], the soot emissions of gasoline PPC high-load operation (around 1.45 MPa IMEP) with a low compression ratio cannot be reduced to the values lower than 1 FSN even with a positive mixing period. Therefore, the mixing period is not the most important factor. According to the study by Manente et al. [4], the soot emissions can be reduced to lower than 0.3 FSN with Euro 6 compliant engine-out NOx emission levels at the full load condition with 2.6 MPa IMEP under the injection pressure of 240 MPa using the XPI fuel system. Thus, the most critical factor of high-load gasoline PPC operation is neither the lambda nor the mixing period, but the local lambda determined by injection pressure. As a matter of fact, diffusion like combustion cannot be fully avoided at full load PPC operations even with high rail pressure capability according to the study [4]. Injection pressure has great influence on the HRR of the mixing-controlled part and guarantees favorable local lambda. In other words, the soot benefit of gasoline physicochemical advantages over diesel fuel is closely related to the injection pressure for a high-load PPC operation. However, the high rail pressure requirement presents a great challenge to the common-rail system originally designed for diesel fuels. And the low viscosity of gasoline may lead to higher friction losses and potential reliability threats to a fuel system. Therefore, the development of high-performance cost-effective viscosity improver is a matter of urgency. For the operating conditions from low speed to high speed at 1.43 MPa BMEP. The following test is to show the effects of engine speed variation on the PPC performance and emissions and to make comparisons of the two combustion modes. Test conditions and operating boundaries are presented in Table 6. Table 6. Tested conditions. Parameters Load [MPa] Speed [r/min] Tests Fuels

(b) NOx-soot trade-offs Fig. 9. Engine-out NOx emissions and NOx-soot trade-offs at three speeds. As shown in the red numbers in Fig. 9(b), the global lambda levels of PPC full load operation are considered to be sufficient for soot emission control. This implies that lambda is not the primary cause of the PPC soot performance. It is generally believed that decreasing the overlap between injection and combustion process is an effective way to reduce soot emissions. Nevertheless, there are many ways to reduce the overlap, such as increasing the injector flow rate and lengthen the ignition delay using a low compression ratio. To take one example,

Injection Strategy CA50 [°CA ATDC] Fuel Pressure [MPa] EGR [%] IMP [MPa]

Values 1.43 1170 1330 1660 EGR Sweep Diesel Gasoline + Viscosity Improver (2.2% by volume) Single Injection 15 100 5…33 0.220…0.233

5…33 0.237…0.247

5…25 0.244…0.251

Fig. 10 shows the comparisons of HRRs and incylinder pressures of CDC and PPC modes at 1.43 MPa BMEP of three engine speeds. Fig. 10(a) shows that a highly premixed heat release pattern can be successfully reached at the low engine speed. The non-knocking RI limit is generally considered to be 5 MW/m2, while the PPC mode has reached 5.1 MW/m2 at the speed of 1170 r/min. Hence care should be taken to avoid the fuel consumption increase due to the high pressure oscillation created after combustion if the CA50 timing is to be Copyright @ 2017 by the Japan Society of Mechanical Engineers


advanced. In addition, the peak firing pressure of gasoline PPC is higher than that of CDC mode even with the same CA50 timings. As the engine speed increases, the PPC operations are mainly diffusion-controlled combustion with less premixing ratio and longer combustion duration. And the CA90 timing (the crank angle at which 90% completion of heat release) increases for a higher speed, which is adverse for soot oxidation process.

Fig. 11.

(a) For the speed of 1170 r/min

(b) For the speed of 1330 r/min

(c) For the speed of 1660 r/min Fig. 10. Comparisons of HRRs and in-cylinder pressures of CDC and PPC modes at 1.43 MPa BMEP.

Engine performance comparisons between CDC and PPC modes at the EGR rate of 20%.

Fig. 11 shows the engine performance comparisons of the two modes at the EGR rate of 20%. Fig. 11 shows the lambda and combustion efficiency of the two modes are satisfactory. It has to be noted that the BTE performance at 1170 r/min and 1330 r/min are better than 1660 r/min. The speeds of 1330 r/min and 1660 r/min are the ESC operating speed A and speed B. While the 1170 r/min represents the 35% of the World Harmonized Steady-state Cycle (WHSC) normalized test speed. The average speed of WHSC duty cycle is much lower than that of ESC. In fact, down-speeding is an effective means to achieve higher closed cycle efficiency due to shorter combustion duration, higher open cycle efficiency associated with lower air flow, and higher mechanical efficiency by lowering the piston speeds. As to PPC operation, downspeeding is beneficial to not only thermal efficiency but also achieving high premixing HRR with good emission characteristics. Generally, down-speeding is usually accompanied with a new turbocharger matching. But in this case, this turbocharging system is matched under ESC test conditions. In addition, there is a point as engine speed is reduced where the increased time for combustion enhances heat transfer to the engine surfaces. For instance, as the speed decreases, the combustion durations of CDC as shown in Fig. 10 amount to 3.35ms, 3.01ms, and 3.05ms, respectively. And so does the PPC mode. Therefore, the BTEs of 1170 r/min are slightly lower than those of 1330 r/min. If a constant power is to be maintained as the engine speed is reduced, the engine torque should be further increased. Hence the peak firing pressure will be increased due to the increased fuel delivery and more premixed heat release as shown in Fig. 10(a) compared to CDC mode. Hence the friction reduction benefit of down-speeding on a gasoline PPC engine may start to erode. Furthermore, turbocharging and EGR system matching is also critical for down-speeding optimization on a PPC engine. Fig. 12 shows the NOx-soot trade-off performance of the CDC and PPC modes at 1.43 MPa BMEP under the three engine speeds. Fig. 12 shows that the PPC operation at 1170 r/min can achieve the best NOx-soot trade-off performance and achieve a great soot reduction, especially at the low NOx regions. Based on the data shown in Fig. 12, Fig. 13(a) shows the NOx emissions at the soot Copyright @ 2017 by the Japan Society of Mechanical Engineers


criterion of 0.01 g/kW-h for the two modes. And Fig. 13(b) shows the PPC soot emission reduction ratios compared to the CDC mode at the specific PPC NOx levels of each speed shown in Fig. 13(a). The PPC emission reduction ratio is calculated based on the CDC emission values. Fig. 13(a) clearly shows that both the CDC and PPC modes can achieve lower NOx emissions at a lower engine speed at a specific soot criterion. When compared to the CDC mode, the NOx and soot reduction ratios of low-speed PPC operation are significantly higher than those of high-speed conditions. It can be clearly seen that the speed variation have a great influence on the PPC emissions. From another angle, down-speeding should be suitable for PPC operation to fully achieve its emission advantage, and up-loading may compensate the power losses.

CDC mode at the specific PPC NOx values of each speed shown in Fig. 13(a) Fig. 13. Emissions of the two modes and PPC emission reduction ratios compared to CDC mode at the specific NOx and/or soot emission criteria.

Effects of rail pressure on the gasoline PPC load extension The following test is to present the PPC load ranges fulfilling the Euro 6 NOx and soot emission criteria under three rail pressures and to compare the rail pressure requirements between the gasoline PPC mode and the Euro 6 CDC mode. Table 7 shows the boundary conditions and emission criteria. Table 7.

Boundary conditions and emission criteria in PPC load extension.

Parameters Speed CA50 Injection Pressure (single injection) NOx Soot

Fig. 12.

Values and limits 1330 r/min, 1660 r/min, and 1990 r/min 10°CA 100 MPa, 130 MPa, and 150 MPa 0.4 ±0.02 g/kW-h 0.01 ±0.001 g/kW-h

Engine-out NOx-soot trade-offs.

Fig. 14.

(a) NOx emissions at the soot emission of 0.01 g/kW-h (based on the data in Fig. 12).

(b) PPC soot emission reductions compared to that of

Effects of rail pressure on the gasoline PPC load extension and comparisons of the rail pressure requirement on the operating range extension between the CDC and PPC modes.

As shown in Fig. 14, the dashed blue lines illustrate the rail pressure map of Euro 6 CDC mode whose tailpipe emissions comply the emission standard with aftertreatment systems (SCR+DPF). The green lines represent the PPC load limits with Euro 6 compliant engine-out emission levels under the rail pressures of 100 MPa, 130 MPa, and 150 MPa respectively. Unfortunately, rail pressure of 180 MPa cannot be reached by running gasoline for this fuel system with a viscosity improver dosage rate of 2.2%. In general, the rail pressure calibrations of the Euro 6 CDC and the gasoline PPC over the whole map show a great difference. It can be seen that, in a sharp contrast to that of the Euro 6 CDC mode, the gasoline PPC load limit is sensitive to rail pressures, especially for the high-speed conditions. The gasoline PPC load range that can be covered by rail pressure of 100 MPa is significantly higher than that of the Euro 6 CDC mode, which means the rail Copyright @ 2017 by the Japan Society of Mechanical Engineers


pressure requirement of the gasoline PPC mode at the part and the intermediate load is significantly lower than the Euro 6 CDC mode. The red triangles show the crosspoints of the Euro 6 CDC and the gasoline PPC load boundaries under the rail pressures of 130 MPa and 150 MPa. The region under the red line (the yellow area) means the rail pressure requirement of PPC mode is lower than that of CDC mode. While the region above the red line (the grey area) means the just opposite. As for the grey area, the higher rail pressure in the gasoline PPC mode may lead to higher peak firing pressure and higher combustion noise. The red dashed arrow means the estimated general direction of the probable cross-points for higher injection pressures. Although the rail pressures higher than 150 MPa are not applied to the gasoline PPC mode in this test, it can be also inferred that the PPC rail pressure demand is higher than that of the Euro 6 CDC mode for above the load of 1.3 MPa BMEP and will increase sharply with the increase of load. According to Ref. [4], the rail pressure of 240 MPa has to be applied to gasoline PPC low-speed high-load operations for the low emissions. However, as to the high-speed full-load PPC operations, the ultra-high rail pressure requirements have become a great challenge to the reliability and stability of gasoline fuel supply system. Therefore, future investigations will likely involve better solutions to effectively improve the gasoline lubricity and to solve the trade-offs between the engine emissions and performance.

CONCLUSIONS This study investigated the effect of gasoline viscosity improvement on the multi-cylinder PPC engine performance and the sensitivity of gasoline PPC operations to the variations of engine load, speed, and fuel injection pressures. The main results are summarized as follows: A small amount of viscosity index improver addition can improve the viscosity of commercially available gasoline effectively without obviously influencing the gasoline octane numbers. The mechanical efficiency and BTE of the engine, by running the gasoline fuel with viscosity index improver, are increased by 1.64% and 0.74% under a rail pressure of 100 MPa, respectively. And the improvement is even more apparent for a higher rail pressure. For the part and medium loads, the PPC rail pressure requirement is significantly lower than that of CDC with Euro 6 compliant tailpipe emissions due to the emission benefit of gasoline PPC. But high premixing may increase heat transfer and has slightly negative effect on BTE. For the high loads, it is relatively easy to achieve high premixing at the low-speed conditions compared to the high-speed conditions. The high-speed full-load gasoline PPC operations with low injection pressures are mainly mixing-controlled with similar NOx-soot trade-off relationship compared to CDC mode. It can be inferred that the local lambda dominated by injection pressure is more important than the global equivalence ratio (i.e. boost pressure). Gasoline PPC high-load extension is sensitive to rail pressure, and the injection pressure

requirement of gasoline PPC is much higher than that of the Euro 6 CDC mode. Contact Information Professor, Dr. Mingfa Yao State Key Laboratory of Engines Tianjin University Tianjin 300072, P.R. China [email protected] Acknowledgements The authors would like to acknowledge the financial supports provided by National Natural Science Found of China (NSFC) through its Projects of 51576138 and 91541111 and provided by the National Basic Research Program of China (973 Program) through its Projects of 2013CB228402. The assistance of Mohsin Raza of Tianjin University with the final editing of the paper is also gratefully acknowledged.

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Copyright @ 2017 by the Japan Society of Mechanical Engineers