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Effect of EGR routing on efficiency and emissions of a PPC engine
Applied Thermal Engineering 152 (2019) 742–750
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Applied Thermal Engineering journal homepage: www.elsevier.com/locate/apthermeng
Research Paper
Effect of EGR routing on efficiency and emissions of a PPC engine a,⁎
b
a
T
a
Nikolaos Dimitrakopoulos , Giacomo Belgiorno , Martin Tunér , Per Tunestål , Gabriele Di Blasiob a b
Faculty of Engineering, Division of Combustion Engines, Lund University, 21100 Lund, Sweden Istituto Motori, Consiglio Nazionale Delle Ricerche, 80125 Naples, Italy
H I GH L IG H T S
load favors mixed route EGR, low load favors short route EGR. • High both on NOx and soot production when moving from SR EGR to LR EGR. • Effect • Speed/load based EGR route switching has positive effect on gas exchange efficiency.
A R T I C LE I N FO
A B S T R A C T
Keywords: Partially premixed combustion Long route EGR EGR routing Multi cylinder engine Gas exchange efficiency
In order to significantly improve engine efficiency and reduce exhaust emissions at the same time, new radical combustion concepts have emerged. Gasoline partially premixed combustion (PPC) is one of them, with early results showing high gross indicated efficiency. To achieve that, PPC relies on high EGR (exhaust gas recirculation) use, with numbers that can reach up to 50%. Such a high amount of EGR poses a great demand on the gas exchange system, especially if it is not optimized for these requirements. A recent advancement that can provide high EGR rates especially under PPC conditions is the use of low pressure EGR, where gases are removed after the turbine and mixed with the intake air before the compressor. Experiments with the use of PPC and two different EGR routes were performed on a light duty Euro 6 2 L diesel engine. EGR sweeps between 100% use of long route to 100% short route under different conditions were performed. Gross indicated mean effective pressure (IMEPg) was kept around 10 bar, while four different speeds were used, 1200, 1800, 2400 RPM, as well as a reoccurring New European Driving Cycle (NEDC) speed-load point at 1500 RPM. To keep the fuel effects on combustion at a minimum, PRF 75 (Primary Reference Fuel) was used throughout the experiments. Results show that by combining EGR from both routes, generally, an optimum gas exchange efficiency can be found by splitting the EGR through both routes. This can be attributed to higher turbocharger efficiency due to better flow over the compressor regardless of engine load and speed. Emission wise, NOx emissions get an increase as EGR is moved from long route to short route, while soot emissions see an opposite trend for the same conditions. Based on these first results, a mixed EGR, or a long route system can be more beneficial for PPC type of engine applications.
1. Introduction Conventional diesel combustion can provide high brake efficiency both for heavy duty (HD) and light duty (LD) engines [1]. Still, due to the diffusion type of combustion, high emissions in the form of nitric oxides (NOx) and soot appear [2]. The most common method to control NOx emissions is the use of exhaust gas recirculation (EGR). By routing a percentage of the exhaust gases back into the intake, a reduction in oxygen concentration is achieved. Combined with the increased heat
⁎
capacity of the air-EGR mixture, a reduction in the peak combustion temperature is achieved, thus reducing the NOx formation [3,4]. On the other hand, soot emissions are dealt with the optimization of the airfuel spray interaction and the piston bowl shape. Despite the measures to reduce these emissions, to meet the emission legislations, manufacturers need to add emission aftertreatment systems as well. These systems increase not only the complexity of the engine design but also the total engine cost, with estimations of about $1600 for a light duty diesel aftertreatment system that complies with the Euro 6 emission
Corresponding author. E-mail address: [email protected] (N. Dimitrakopoulos).
https://doi.org/10.1016/j.applthermaleng.2019.02.108 Received 16 November 2018; Received in revised form 8 February 2019; Accepted 24 February 2019 Available online 25 February 2019 1359-4311/ © 2019 Elsevier Ltd. All rights reserved.
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Nomenclature ATDC BMEP CA50 CAD CDC CI Cp EGR HCCI HD HP IMEPg
IMEPn LD LP LR NEDC NOx PPC PRF RON RPM SR TDC VGT WLTP
after top dead centre brake mean effective pressure position in CAD of 50% of fuel mass burned crank angle degree conventional diesel combustion compression ignition specific heat capacity for constant pressure exhaust gas recirculation homogenous charge compression ignition heavy duty high pressure gross indicated mean effective pressure
net indicated mean effective pressure light duty low pressure long route New European Driving Cycle nitrogen oxides partially premixed combustion Primary Reference Fuel research octane number revolutions per minute short route top dead centre variable geometry turbine Worldwide harmonized Light vehicles Test Procedure
EGR percentages thus increasing the EGR tolerance of the engine [18]. Drawbacks include compressor wheel fouling from soot emissions as well as limited transient response [19–21]. The conclusions are that combining EGR routes can generally boost efficiency and reduce emissions. This is because, when combining routes, the total pumping losses are reduced, as well as the turbocharger is operated at a more efficient level. But optimization is necessary for each load and RPM operating conditions, in terms of portion of LR EGR to the total amount of EGR to achieve the full benefits from the combination of both routes. Because the PPC concept relies on high amount of EGR and high air dilution, it would be beneficial to be able to use both EGR routes in a dual loop arrangement, based on the engine’s operating point requirements. Prior work with simulations in HD engine [22] shows that combining both routes gives improved overall efficiency due to the higher turbocharger efficiency under most operating conditions while using a matched turbocharging system. This is due to the fact that depending on the load and the engine speed, the mass flow through the turbine can be controlled better, moving the operating point towards the higher efficiency regions of the compressor map. In this work the effect of combining EGR routes on a light duty engine is evaluated. The engine operates under partially premixed combustion (PPC) conditions, which is known to have a demand of high EGR rates. The potential beneficial effect of mixed route EGR on regulated diesel emissions, NOx and soot, as well as improved brake efficiency through optimization of the different engine efficiencies is assessed. Combining both EGR routes benefits the gas exchange efficiency compared to a single route, while emissions show better values when only long route EGR is used, if no limits are imposed on the exhaust oxygen content.
limits [5]. New combustion concepts evolved through the years to overcome the inherit problem of emissions from a diesel engine. Under the umbrella of the low temperature combustion (LTC), most of these concepts suggest that a more premixed charge, coupled with a higher than usual amount of EGR, with values reaching up to 50% [6,7], or very lean mixtures, can simultaneously give low NOx and soot emissions. Another common characteristic of these newer concepts is the use of gasoline, gasoline-like fuels, or high-octane fuels and as a result they can also be classified as gasoline compression ignition (GCI) [8–11]. Due to the kinetic control of the combustion, all these concepts offer high efficiency, with recorded values of up to 55% gross indicated efficiency [6,7] for HD engines and between 48% and 52% for LD engines [12–14]. Still, while these concepts give high indicated efficiency, when the break efficiency is examined results are slightly better that the conventional diesel or gasoline engines. Friction is a crucial factor, because high combustion pressure is a common characteristic of LTC due to fast combustion, leading to high pressure rise rates. Also, a generally lower achievable maximum load compared to diesel engines gives lower overall mechanical efficiency. But pumping losses have an important contribution as well, considering the considerable amounts of EGR necessary to provide a prolonged ignition delay and keep temperatures and NOx production low [13,15,16]. The most common design to add EGR into the intake air is the use of High Pressure (HP) or Short Route (SR) EGR. It is used both in LD and HD engines, due to efficient packaging and good transient response. With that design, an amount of the exhaust gases is removed before the turbine in the exhaust manifold and mixed with the intake air after the compressor outlet. Drawbacks of this design is that it needs a positive pressure difference between exhaust and intake, to move the exhaust gases through the system. The hot exhaust gases need to be cooled down to around 100 °C before mixing with the air, increasing the load of the cooling system. Finally, the removal of the gases before the turbine decreases the available mass flow and energy flow to the turbocharger. A more recent design for EGR routing, that overcomes some of the drawbacks of the HP route, is the Low Pressure (LP) or Long Route (LR) EGR. With this design exhaust gases are extracted after the turbine and fed before the compressor. The benefits of such system are the lower temperature of the exhaust gas, the higher amount of energy available to the turbine and the smaller pressure difference needed to move the gases [3]. Another study suggests that with LR EGR, the mixing of exhaust gases with intake air is better, leading to better EGR distribution to the cylinders, while less energy is lost due to EGR cooling [17]. Also, an added benefit of LR EGR, can be the much cleaner exhaust gases, since it is commonly routed after the diesel oxidation catalyst and diesel particulate filter. EGR with low amount of soot and unburned hydrocarbons can reduce the in-cylinder soot formation at high loads and
2. Experimental setup For this study, a diesel light duty multi-cylinder engine, Volvo VEDD4, was used. The engine complies with the Euro 6 emission regulations and has the following specifications (Table 1). One distinct feature of the engine is the gas exchange system, which Table 1 Engine properties. Displacement Number of cylinders Bore × Stroke Compression ratio # of valves per cylinder Diesel Injection System Injector Max peak pressure rise rate (for this study) COVIMEP limit (%)
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1969 cm3 4 82 mm × 93.2 mm 15.8 [–] 4 [–] Common Rail (Maximum injection pressure: 2500 bar) Solenoid, 8-hole, < 160° umbrella angle 10 bar/CAD Below 5%
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consists of two turbochargers in serial configuration. The high-pressure turbo has Variable Geometry Turbine (VGT) technology for boost control while the low-pressure turbo uses a wastegate. Modifications include the use of an air to water intercooler in place of the traditional air to air intercooler, which gives better control for the intake air temperature and the use of a 6 kW air heater after the intercooler and before the intake manifold. The engine is fitted with a high pressure EGR system but to evaluate the capabilities of the boosting system under PPC operation and gain flexibility into controlling the EGR rates, an adjustable low pressure EGR loop was added in the exhaust, downstream the aftertreatment system, which, in this study is not installed. The low pressure EGR loop is cooled by a plate heat exchanger, to be able to get EGR temperatures as low as ambient and then the LP EGR path merges with the intake pipes just before the inlet of the first compressor. To be able to move exhaust gases through the low pressure EGR route a backpressure valve is fitted after the EGR bypass (Fig. 1). The exhaust aftertreatment system is removed from the engine, to be able to measure the raw emission output from the engine. The engine is coupled to a 300 kW electric motor to be able to operate the engine without having combustion as well as provide the necessary load for the engine to function. The control system of the engine consists of a Labview based inhouse developed software, that operates on the host-target principle. A dedicated real time computer, NI PXI-e8135 is used and is fitted with two FPGA cards for the fast sensors data acquisition and real time calculations of in-cylinder pressure, fuel rail pressure, torque and Crank Angle Degree (CAD)/Top Dead Center (TDC) signals. For the CAD signal an external Leine-Linde optical encoder is fitted to the engine, with a resolution of 0.2 degrees. In the real-time computer, a DAQ card, PXI6221 is used for the slow sensor sampling as well as the temperature logging. Injector actuation is possible with the use of NI driven modules. For the in-cylinder pressure acquisition, four AVL GH14D piezoelectric sensors are used and are mounted into specially drilled positions below the glow plugs. On the target side, the computer that is used runs the graphical interface that enables the engine operation by sending the control commands to the real-time computer and provides a graphical view of the output values of all the different sensors as well as the in-cylinder pressure trace. The fuel that was used was a mixture of iso-octane and n-heptane,
Table 2 Fuel properties. RON MON AFRs H/C ratio
PRF75 (Table 2). Primary reference fuel was used to minimize the fuel properties variations of different batches used and because the focus of this study was mainly on the EGR and gas exchange system interaction and less on the combustion behavior. The fuel was treated with 400 ppm of lubricity additive, Infenium R655, to prolong the lifetime of the high-pressure fuel pump as well as the rest of the fueling system, which is not adapted for gasoline operation. Fuel flow is measured with a Sartorious CPA10001 scale and logged into the labview control system. The emissions are analyzed using a Horiba Mexa 7100 system, with two separate CO2 lines so that the total CO2 percentage in the intake can be measured and calculate the EGR ratio.
Total EGR % =
% CO2 intake % CO2 exhaust
(1)
To be able to measure the CO2 percentage for long route individually, a second sampling point is used after the long route connection pipe. A portable emission analyzer (Horiba MEXA-554JE) is used to sample and log the recorded amount. Therefore, the long route EGR percentage can be defined as
LR EGR % =
% CO2 LR % CO2 exhaust
(2)
Soot emissions are measured with an AVL micro-soot sensor unit. Output of the equipment is in mg soot/m3 exhaust flow. All of the emissions are measured engine out and are converted to brake specific emissions in g/kWh, from the exhaust flow and the engine brake power as
bsEmission =
ṁ emission Pbrake
(3)
For the combustion data analysis, 150 engine cycles are saved and are processed offline with the use of Matlab software. Data is saved HP EGR cooler
HP EGR
To emission system
75 75 15.1:1 2.258:1
Wastegate VNT
4 3
Exhaust out LP EGR cooler
2 1
Air in
HP turbo LP turbo
Intercooler Fig. 1. Experimental setup, engine layout. 744
Air Heater
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3.3. Combustion phasing sweep (CA50)
after the engine has reached steady state conditions. These conditions are considered reached, when the oil, coolant, intake and exhaust temperatures have stabilized. Apparent heat release analysis is performed from the in-cylinder pressure trace using the equation:
In this study, the effects of the different EGR routing options will be evaluated. To be able to do that, different engine conditions are needed, to get different amounts of flow through the turbocharger system. To achieve that, four engine speeds are selected, that cover the low to medium engine speed range and four different loads that cover the low to medium load.
To better evaluate the effects of EGR on the gas exchange system, a CA50 sweep is performed at the two extreme points, at 20% EGR through the short route and 20% EGR through the long route. By retarding the combustion, a higher amount of energy can be added in the exhaust flow which can be beneficial for the overall efficiency of turbocharger, leading to a higher gas exchange efficiency. At the same time, the retarded combustion will reduce the gross indicated efficiency. This investigation wants to examine if there is a beneficial trade of in terms of net indicated efficiency. If the gas exchange efficiency increases more than the reduction in gross indicated efficiency, an overall higher net indicated efficiency can be realized in the end. The RPM are kept at 1800, the load is at 21.5 bar FuelMEP and the intake temperature is kept at 80 ℃. The fuel injection pressure is constant at 1100 bar. Five different CA50 are evaluated. These are at 7, 9, 11, 13 and 15 degrees ATDC. A note regarding the EGR percentage that is used in this study. In the literature, the different partially premixed concepts utilize considerable amounts of EGR, 50% EGR is a commonly used number, to reduce the NOx emissions and prolong the ignition delay. For this study only 20% of EGR was used. While that amount will certainly increase the NOx emissions, it is necessary for the different EGR routes to be able to provide the required EGR amount. During early testing it was found that with the current setup, EGR numbers higher that 20% through the short route give unstable turbocharger operation leading to pressure fluctuations. Therefore 20% EGR was used to keep stable engine operation.
3.1. Constant RPM – Load, EGR sweep
4. Results - discussion
γ ∂p ∂Q ∂V 1 = p + V ∂θ γ − 1 ∂θ γ − 1 ∂θ
(4)
The in-cylinder pressure trace is used also to calculate the indicated mean effective pressure (IMEP). Both the gross (IMEPg) and the net (IMEPn) IMEP are calculated in order to estimate the gas exchange efficiency. Gas exchange efficiency is defined as:
ηge =
IMEPn IMEPg
In order to keep all the numbers with the same units, the amount of consumed fuel is expressed in pressure units, similarly to the expression of BMEP and IMEP and is named FuelMEP. A more thorough definition can be found in Appendix A. 3. Methodology
For this part the four different engine speed - load points were evaluated. They are summarized in Table 3. The first three load points represent points of high gross indicated efficiency that were evaluated in previous work [13]. The fuel energy is adjusted at 23.5 bar FuelMEP at low speed and 21.5 bar FuelMEP at medium and high speed, which results in an IMEPg of around 10 bar. Total amount of EGR is 20%, while a sweep is performed from the long route to the short route, resulting in 0%, 50% and 100% of EGR through the short route. No limit is imposed on lambda values that are expected to vary due to the different exhaust mass flow through the turbine, which will result in different amount of boost pressure. As a result, the emissions will vary as well. Intake temperature is kept constant at 75–80 °C to promote faster air fuel mixing and help with the ignition of the PRF75. The fourth load point represents an actual high frequency point in the NEDC/Worldwide harmonized Light vehicles Test Procedure (WLTP) test cycle for passenger vehicles. In this case the BMEP will be kept constant at 5 bar BMEP while the FuelMEP will vary around an average value of 13.5 bar. The total EGR percentage will be kept constant at 20%. The rest of the parameters are kept same as the other three load points.
4.1. Operation at different RPM and loads 4.1.1. Gas exchange efficiency Routing the EGR from long to short route shows an effect on gas exchange efficiency under all operating conditions. For the high engine load cases, generally, the gas exchange efficiency is higher than 90%, but there is a peak when EGR is split 50–50% between long and short route. On the other hand, at the low load case, the most efficient operation is when using the short route option (Fig. 2). An explanation regarding that behavior can be given from the fact that to get the necessary EGR flow through each route the exhaust backpressure should be high enough to move the EGR. For the long route the backpressure is created by a valve positioned on the exhaust pipe. For the short route, backpressure is created by the adjustment of the VGT position. This necessary backpressure is minimized in these cases when the EGR is split between both routes. Still at the low load NEDC case, a different trend appears. The highest gas exchange efficiency appears when gases are routed through the short route. An explanation for this is that at this low load point the exhaust mass flow is not high enough and requires higher amount of throttling from the backpressure valve for the long route to provide the necessary EGR amount both for the long route and the split route case, where the exhaust flow is reduced compared to the single route case.
3.2. Constant lambda EGR sweep In the previous part the air to fuel ratio (AFR) would vary based on the intake conditions. Still, under certain engine operating conditions, can be necessary the AFR to be higher than a specific value, to avoid producing an excess of emissions, such as soot. For this part the effects of EGR routing are investigated while keeping the AFR constant at lambda value of 1.9. The rest of the control parameters remain like the previous part; EGR percentage is kept constant at 20%, FuelMEP is 20.5 bar, injection pressure is 1100 bar, RPM is at 1800, intake temperature 80 ℃. EGR is swept from 100% through the long route, to 50–50% and finally to 100% through the short route.
Table 3 Engine operating conditions.
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# [–]
Speed [RPM]
Intake temp [°C]
FuelMEP [bar]
Rail pressure [bar]
1 2 3 4
1200 1800 2400 1500
77 80 75 75
23.5 21.5 21.5 ∼13.5
1100 1100 1100 800
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Regarding emissions, a similar trend is observed with NOx emissions. As the EGR is moved towards the short route the NOx emissions increase. Although the oxygen content, through the lambda value, remains almost the same, there is a small variation between the cases, together with a small variation of the total EGR percentage, from 20% to 20.5%. Mixing of EGR through the short route can be less homogenous compared to the long route case and can affect the combustion between the different cylinders. All these effects as well as the slightly earlier combustion phasing in terms of CA50, from 12 °ATDC at the long route to 11 °ATDC at the short route can explain the increase of NOx emissions (Fig. 8). Soot emissions show a minimal increase in terms of g/kWh but this is attributed to the reduced brake efficiency when moving from long route to split route and finally short route, because the emissions in terms of mg/m3 remain fairly constant in absolute values. 4.3. Effects of combustion phasing Fig. 2. Gas exchange efficiency for different RPM – load.
As the combustion is phased to later CA50 values, there is no large effect on the gas exchange efficiency. At the load point of 21.5 bar FuelMEP and with these engine settings the efficiency is quite high, and it seems that there is little to gain in terms of gas exchange efficiency by retarding the combustion (Fig. 9). As expected the gross indicated efficiency is affected significantly as the combustion phasing is retarded (increase in CA50). The effect of gas exchange efficiency on the net indicated efficiency is minimal, so it is not beneficial to retard the combustion under these operating conditions (Fig. 10). Emission wise, due to the late combustion, in cylinder temperatures go down with late CA50, leading to lower NOx emissions both in ppm and in specific emissions (g/kWh) (Fig. 11). The effect on the decreased temperature is higher than the effect of the decreased efficiency on the specific emissions and as a result the trend is that also the specific emissions show a decrease with the later combustion. Regarding the soot emissions, a different trend appears. While the soot emissions are low and at similar levels in terms of mg/m3 as the brake efficiency decreases the specific emissions increase (Fig. 12). Still due to the higher premixed part of the combustion as well as the higher temperatures they are low in absolute numbers.
4.1.2. Emissions When EGR is routed from long route to short route the trend regarding the NOx emissions is towards an increase (Fig. 3), in terms of specific emissions, while the soot emissions tend to decrease (Fig. 4). This can be attributed to the higher lambda values when operated with short route EGR (Fig. 5). As mentioned before, to maintain the necessary pressure difference between exhaust and intake and move the EGR through the short route, the VGT position is adjusted accordingly, which leads, as a result, to increased intake pressure. Higher intake pressure leads to higher air flow in the engine, which can explain the trend of higher NOx and lower soot emissions from the short route. Except of the case of the 1200 RPM where both NOx and soot emissions increase. This is a specific case and has to do with the fact that the lambda value decreases as EGR is routed through the short route. VGT adjustment is not necessary in this case, therefore the lower exhaust flow through the turbine gives reduced intake air pressure, leading to lower amount of available oxygen. As the air to fuel ratio becomes less lean and reaches a low limit of 1.2 that leads to higher combustion temperatures, resulting in higher soot and NOx emissions. Similar results appear also for the 1500 RPM case. As the EGR is moved from long to short route, no additional throttling is needed. As a result, lambda values decrease slightly (Fig. 5), leading to higher combustion temperatures and increase in the overall NOx emissions (Fig. 3). However, overall lambda values are high enough to lead to increased soot oxidation because of the higher combustion temperatures (Fig. 4). The difference in NOx and soot emissions as well as in gas exchange efficiency, for the previous cases can be summarized in Table 4.
4.4. Future work While the results of splitting the EGR routes provide an insight on the question if it beneficial to have a split route system, in terms of efficiency and emissions, more investigations can be performed to
4.2. Constant lambda operation As the EGR is swept from long route to short, the lambda value is quite stable around the target value of 1.9 ± 1.6%. While trying to maintain the oxygen content constant, there is an effect on gas exchange efficiency as EGR routing is moved from long to short route (Fig. 6). In this case the lowest efficiency can be found in the split between the two routes, while the highest efficiency is when all the EGR is passed through the long route. Again, this can be attributed to the necessary adjustment in the intake pressure to keep the target lambda value. This adjustment is done through the VGT part of the turbine which leads to higher exhaust pressure and thus reducing the gas exchange efficiency. Different EGR routing options do not influence gross indicated efficiency, but affect obviously the net indicated efficiency and as a result brake efficiency as well (Fig. 7).
Fig. 3. NOx emissions for different RPM – load. 746
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Fig. 4. Soot emissions for different RPM – load.
Fig. 6. Gas exchange efficiency for the constant lambda cases.
Fig. 5. Lambda value for different RPM – load.
Fig. 7. Gross, net and brake efficiency for the constant lambda cases.
Table 4 Percentage difference for NOx, soot and gas exchange efficiency between the best and the worst conditions. (NOx and soot are percentage reduction; gas exchange efficiency is percentage improvement). RPM × FuelMEP
NOx
Soot
Gas exchange efficiency
1200 × 23.5 1500 × 13.5 1800 × 21.5 2400 × 21.5
16.7% 39% 17% 50.5%*
56% 57% 28% 22%
1% 2.3% 4.5% 2.1%
better evaluate the benefits. In this study, the OEM turbocharging system was used, which was optimized for diesel operation with moderate amounts of EGR and short route operation. This posed a limitation on the achievable EGR percentage because EGR through the short route could not go higher that 20% whereas through the long route, EGR up to 40% could be used. Since PPC requires higher amounts of dilution, a better optimized system for higher EGR rates and possibly for split route operation will be valuable. PPC operation with higher amounts of EGR would help reduce NOx emissions even further although it could pose limitations on the maximum achievable load and possibly reduce further the overall gas exchange efficiency. Furthermore, compressor maps of the turbo system will give a further assessment on the conditions on which the system operates and how close these are to the maximum turbocharger efficiency, highlighting the improvement that can be
Fig. 8. Emissions, NOx and soot, for the constant lambda cases.
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Fig. 12. Effect of combustion phasing on specific soot emissions.
Fig. 9. Effect of combustion phasing on gas exchange efficiency.
reached and the system performance can be evaluated on a larger load – speed area. 5. Conclusions For this study, a four-cylinder light duty diesel engine was operated under PPC conditions to evaluate the effects of different EGR routing options on gas exchange efficiency, overall efficiency and emissions. These findings can be used in additional studies as a possible starting point, when examining further the effects of EGR on low temperature combustion, especially at much higher EGR rates, where the penalty on the gas exchange efficiency is going to be higher. (1) There is an effect on gas exchange efficiency from EGR routing even with a non-optimized turbocharger. Combining both EGR routes provides better results, in most of the cases, but no definite conclusion can be drawn. High loads favor a split between the two routes, while low loads perform better with short route. (2) If only the EGR percentage is kept constant while moving from long to short route, there is going to be an effect on oxygen content. This is a result of the different backpressure settings that are necessary to reach the target EGR percentage. If there is no lambda value target, oxygen mass will differ, which as a result will affect NOx and soot emissions. NOx emissions will tend to increase, while soot emissions decrease, due to higher lambda values when operated with short route EGR. (3) Keeping constant both lambda value and EGR percentage, affects negatively the gas exchange efficiency of the split route, leading to the lowest value compared to only short, or only long route. Emissions have the same trend as before. (4) Retarding the combustion (late CA50) has minimal effect on gas exchange efficiency and as a result negligible effect on net indicated efficiency therefore negatively affects the brake efficiency. Due to that, it has a negative effect on brake specific soot emissions. Brake specific NOx emissions reduce due to lower combustion temperatures.
Fig. 10. Effect of combustion phasing on Gross and Net indicated efficiencies.
Acknowledgments Fig. 11. Effect of combustion phasing on specific NOx emissions.
The authors would like to acknowledge the Competence Centre for Combustion Processes, KCFP (Project number 22485-4) and the Swedish Energy Agency for providing financial support for this study. Also, the authors would like to thank Dr. Kenan Muric, for providing the Matlab code that could calculate the specific heat of various gases based on the JANAF tables.
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Appendix A. Definitions of mean effective pressures and engine efficiencies
A.1. Fuel Mean Effective Pressure, FuelMEP Fuel Mean Effective Pressure is defined as fuel energy per cycle divided with the engine displacement. Fuel energy is the fuel mass per cycle multiplied by the fuel lower heating value.
FuelMEP =
mf × QLHV (A.1)
Vd
A.2. Gross Indicated Mean Effective Pressure, IMEPg Gross Indicated Mean Effective Pressure, IMEPg, is defined as the work to the piston during the compression and the expansion stroke divided with the engine displacement.
IMEPg =
1 Vd
180
∫−180 pdV
(A.2)
A.3. Net Indicated Mean Effective Pressure, IMEPn Net Indicated Mean Effective Pressure, IMEPn, is defined as the work to the piston during the entire four stroke cycle divided with the engine displacement.
IMEPn =
1 Vd
360
∫−360 pdV
(A.3)
In both cases, zero degrees is defined at the Top dead Center during the combustion part. A.4. Brake Mean Effective Pressure, BMEP Brake Mean Effective Pressure is defined as the useful work per cycle divided by the displacement.
BMEP =
P × nR Vd × N
(A.4) 3
where P is brake power [W], Vd is engine displacement [m ], N is engine speed [1/s] and nR is the number of crank revolutions per power stroke. With these definitions the different efficiencies can be defined as well. A.5. Gross Indicated Efficiency
ηgross =
IMEPg (A.5)
FuelMEP
A.6. Net Indicated Efficiency
ηnet =
IMEPn FuelMEP
(A.6)
A.7. Gas Exchange Efficiency
ηge =
IMEPn IMEPg
(A.7)
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A.8. Brake Efficiency
ηb =
BMEP FuelMEP
(A.8)
Appendix B. Supplementary material Supplementary data to this article can be found online at https://doi.org/10.1016/j.applthermaleng.2019.02.108.
References [1] T. Kogo, Y. Hamamura, K. Nakatani, T. Toda, et al., High Efficiency Diesel Engine with Low Heat Loss Combustion Concept - Toyota’s Inline 4-Cylinder 2.8-Liter ESTEC 1GD-FTV Engine, SAE Technical Paper 2016-01-0658, 2016, https://doi. org/10.4271/2016-01-0658. [2] J. Dec, A Conceptual Model of DI Diesel Combustion Based on Laser-Sheet Imaging*, SAE Technical Paper 970873, 1997, https://doi.org/10.4271/970873. [3] Ming Zheng, Graham T. Reader, J. Gary Hawley, Diesel engine exhaust gas recirculation––a review on advanced and novel concepts, Energy Convers. Manage. 45 (6) (2004) 883–900 ISSN 0196-8904, 10.1016/S0196-8904(03)00194-8. [4] N. Ladommatos, S. Abdelhalim, H. Zhao, Z. Hu, The Dilution, Chemical, and Thermal Effects of Exhaust Gas Recirculation on Diesel Engine Emissions - Part 1: Effect of Reducing Inlet Charge Oxygen, SAE Technical Paper 961165, 1996, https://doi.org/10.4271/961165. [5] F. Posada, A. Bandivadekar, J. German, Estimated Cost of Emission Control Technologies for Light-Duty Vehicles Part 2 - Diesel, SAE Technical Paper 2013-010539, 2013, https://doi.org/10.4271/2013-01-0539. [6] V. Manente, C. Zander, B. Johansson, P. Tunestal, et al., An Advanced Internal Combustion Engine Concept for Low Emissions and High Efficiency from Idle to Max Load Using Gasoline Partially Premixed Combustion, SAE Technical Paper 2010-01-2198, 2010, https://doi.org/10.4271/2010-01-2198. [7] D. Splitter, M. Wissink, D. DelVescovo, R. Reitz, RCCI Engine Operation Towards 60% Thermal Efficiency, SAE Technical Paper 2013-01-0279, 2013, https://doi. org/10.4271/2013-01-0279. [8] C. Kolodziej, J. Kodavasal, S. Ciatti, S. Som, et al., Achieving Stable Engine Operation of Gasoline Compression Ignition Using 87 AKI Gasoline Down to Idle, SAE Technical Paper 2015-01-0832, 2015, https://doi.org/10.4271/2015-01-0832. [9] K. Cho, E. Latimer, M. Lorey, D. Cleary, et al., Gasoline fuels assessment for Delphi’s second generation gasoline direct-injection compression ignition (GDCI) multi-cylinder engine, SAE Int. J. Eng. 10 (4) (2017) 1430–1442, https://doi.org/10.4271/ 2017-01-0743. [10] V. Manente, B. Johansson, P. Tunestal, Partially premixed combustion at high load using gasoline and ethanol, a comparison with diesel, SAE Technical Paper 200901-0944, 2009, https://doi.org/10.4271/2009-01-0944. [11] V. Manente, P. Tunestal, B. Johansson, W. Cannella, Effects of ethanol and different type of gasoline fuels on partially premixed combustion from low to high load, SAE Technical Paper 2010-01-0871, 2010, https://doi.org/10.4271/2010-01-0871. [12] V. Manente, B. Johansson, P. Tunestal, M. Sonder, S. Serra, Gasoline partially
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Contents lists available at ScienceDirect
Applied Thermal Engineering journal homepage: www.elsevier.com/locate/apthermeng
Research Paper
Effect of EGR routing on efficiency and emissions of a PPC engine a,⁎
b
a
T
a
Nikolaos Dimitrakopoulos , Giacomo Belgiorno , Martin Tunér , Per Tunestål , Gabriele Di Blasiob a b
Faculty of Engineering, Division of Combustion Engines, Lund University, 21100 Lund, Sweden Istituto Motori, Consiglio Nazionale Delle Ricerche, 80125 Naples, Italy
H I GH L IG H T S
load favors mixed route EGR, low load favors short route EGR. • High both on NOx and soot production when moving from SR EGR to LR EGR. • Effect • Speed/load based EGR route switching has positive effect on gas exchange efficiency.
A R T I C LE I N FO
A B S T R A C T
Keywords: Partially premixed combustion Long route EGR EGR routing Multi cylinder engine Gas exchange efficiency
In order to significantly improve engine efficiency and reduce exhaust emissions at the same time, new radical combustion concepts have emerged. Gasoline partially premixed combustion (PPC) is one of them, with early results showing high gross indicated efficiency. To achieve that, PPC relies on high EGR (exhaust gas recirculation) use, with numbers that can reach up to 50%. Such a high amount of EGR poses a great demand on the gas exchange system, especially if it is not optimized for these requirements. A recent advancement that can provide high EGR rates especially under PPC conditions is the use of low pressure EGR, where gases are removed after the turbine and mixed with the intake air before the compressor. Experiments with the use of PPC and two different EGR routes were performed on a light duty Euro 6 2 L diesel engine. EGR sweeps between 100% use of long route to 100% short route under different conditions were performed. Gross indicated mean effective pressure (IMEPg) was kept around 10 bar, while four different speeds were used, 1200, 1800, 2400 RPM, as well as a reoccurring New European Driving Cycle (NEDC) speed-load point at 1500 RPM. To keep the fuel effects on combustion at a minimum, PRF 75 (Primary Reference Fuel) was used throughout the experiments. Results show that by combining EGR from both routes, generally, an optimum gas exchange efficiency can be found by splitting the EGR through both routes. This can be attributed to higher turbocharger efficiency due to better flow over the compressor regardless of engine load and speed. Emission wise, NOx emissions get an increase as EGR is moved from long route to short route, while soot emissions see an opposite trend for the same conditions. Based on these first results, a mixed EGR, or a long route system can be more beneficial for PPC type of engine applications.
1. Introduction Conventional diesel combustion can provide high brake efficiency both for heavy duty (HD) and light duty (LD) engines [1]. Still, due to the diffusion type of combustion, high emissions in the form of nitric oxides (NOx) and soot appear [2]. The most common method to control NOx emissions is the use of exhaust gas recirculation (EGR). By routing a percentage of the exhaust gases back into the intake, a reduction in oxygen concentration is achieved. Combined with the increased heat
⁎
capacity of the air-EGR mixture, a reduction in the peak combustion temperature is achieved, thus reducing the NOx formation [3,4]. On the other hand, soot emissions are dealt with the optimization of the airfuel spray interaction and the piston bowl shape. Despite the measures to reduce these emissions, to meet the emission legislations, manufacturers need to add emission aftertreatment systems as well. These systems increase not only the complexity of the engine design but also the total engine cost, with estimations of about $1600 for a light duty diesel aftertreatment system that complies with the Euro 6 emission
Corresponding author. E-mail address: [email protected] (N. Dimitrakopoulos).
https://doi.org/10.1016/j.applthermaleng.2019.02.108 Received 16 November 2018; Received in revised form 8 February 2019; Accepted 24 February 2019 Available online 25 February 2019 1359-4311/ © 2019 Elsevier Ltd. All rights reserved.
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Nomenclature ATDC BMEP CA50 CAD CDC CI Cp EGR HCCI HD HP IMEPg
IMEPn LD LP LR NEDC NOx PPC PRF RON RPM SR TDC VGT WLTP
after top dead centre brake mean effective pressure position in CAD of 50% of fuel mass burned crank angle degree conventional diesel combustion compression ignition specific heat capacity for constant pressure exhaust gas recirculation homogenous charge compression ignition heavy duty high pressure gross indicated mean effective pressure
net indicated mean effective pressure light duty low pressure long route New European Driving Cycle nitrogen oxides partially premixed combustion Primary Reference Fuel research octane number revolutions per minute short route top dead centre variable geometry turbine Worldwide harmonized Light vehicles Test Procedure
EGR percentages thus increasing the EGR tolerance of the engine [18]. Drawbacks include compressor wheel fouling from soot emissions as well as limited transient response [19–21]. The conclusions are that combining EGR routes can generally boost efficiency and reduce emissions. This is because, when combining routes, the total pumping losses are reduced, as well as the turbocharger is operated at a more efficient level. But optimization is necessary for each load and RPM operating conditions, in terms of portion of LR EGR to the total amount of EGR to achieve the full benefits from the combination of both routes. Because the PPC concept relies on high amount of EGR and high air dilution, it would be beneficial to be able to use both EGR routes in a dual loop arrangement, based on the engine’s operating point requirements. Prior work with simulations in HD engine [22] shows that combining both routes gives improved overall efficiency due to the higher turbocharger efficiency under most operating conditions while using a matched turbocharging system. This is due to the fact that depending on the load and the engine speed, the mass flow through the turbine can be controlled better, moving the operating point towards the higher efficiency regions of the compressor map. In this work the effect of combining EGR routes on a light duty engine is evaluated. The engine operates under partially premixed combustion (PPC) conditions, which is known to have a demand of high EGR rates. The potential beneficial effect of mixed route EGR on regulated diesel emissions, NOx and soot, as well as improved brake efficiency through optimization of the different engine efficiencies is assessed. Combining both EGR routes benefits the gas exchange efficiency compared to a single route, while emissions show better values when only long route EGR is used, if no limits are imposed on the exhaust oxygen content.
limits [5]. New combustion concepts evolved through the years to overcome the inherit problem of emissions from a diesel engine. Under the umbrella of the low temperature combustion (LTC), most of these concepts suggest that a more premixed charge, coupled with a higher than usual amount of EGR, with values reaching up to 50% [6,7], or very lean mixtures, can simultaneously give low NOx and soot emissions. Another common characteristic of these newer concepts is the use of gasoline, gasoline-like fuels, or high-octane fuels and as a result they can also be classified as gasoline compression ignition (GCI) [8–11]. Due to the kinetic control of the combustion, all these concepts offer high efficiency, with recorded values of up to 55% gross indicated efficiency [6,7] for HD engines and between 48% and 52% for LD engines [12–14]. Still, while these concepts give high indicated efficiency, when the break efficiency is examined results are slightly better that the conventional diesel or gasoline engines. Friction is a crucial factor, because high combustion pressure is a common characteristic of LTC due to fast combustion, leading to high pressure rise rates. Also, a generally lower achievable maximum load compared to diesel engines gives lower overall mechanical efficiency. But pumping losses have an important contribution as well, considering the considerable amounts of EGR necessary to provide a prolonged ignition delay and keep temperatures and NOx production low [13,15,16]. The most common design to add EGR into the intake air is the use of High Pressure (HP) or Short Route (SR) EGR. It is used both in LD and HD engines, due to efficient packaging and good transient response. With that design, an amount of the exhaust gases is removed before the turbine in the exhaust manifold and mixed with the intake air after the compressor outlet. Drawbacks of this design is that it needs a positive pressure difference between exhaust and intake, to move the exhaust gases through the system. The hot exhaust gases need to be cooled down to around 100 °C before mixing with the air, increasing the load of the cooling system. Finally, the removal of the gases before the turbine decreases the available mass flow and energy flow to the turbocharger. A more recent design for EGR routing, that overcomes some of the drawbacks of the HP route, is the Low Pressure (LP) or Long Route (LR) EGR. With this design exhaust gases are extracted after the turbine and fed before the compressor. The benefits of such system are the lower temperature of the exhaust gas, the higher amount of energy available to the turbine and the smaller pressure difference needed to move the gases [3]. Another study suggests that with LR EGR, the mixing of exhaust gases with intake air is better, leading to better EGR distribution to the cylinders, while less energy is lost due to EGR cooling [17]. Also, an added benefit of LR EGR, can be the much cleaner exhaust gases, since it is commonly routed after the diesel oxidation catalyst and diesel particulate filter. EGR with low amount of soot and unburned hydrocarbons can reduce the in-cylinder soot formation at high loads and
2. Experimental setup For this study, a diesel light duty multi-cylinder engine, Volvo VEDD4, was used. The engine complies with the Euro 6 emission regulations and has the following specifications (Table 1). One distinct feature of the engine is the gas exchange system, which Table 1 Engine properties. Displacement Number of cylinders Bore × Stroke Compression ratio # of valves per cylinder Diesel Injection System Injector Max peak pressure rise rate (for this study) COVIMEP limit (%)
743
1969 cm3 4 82 mm × 93.2 mm 15.8 [–] 4 [–] Common Rail (Maximum injection pressure: 2500 bar) Solenoid, 8-hole, < 160° umbrella angle 10 bar/CAD Below 5%
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consists of two turbochargers in serial configuration. The high-pressure turbo has Variable Geometry Turbine (VGT) technology for boost control while the low-pressure turbo uses a wastegate. Modifications include the use of an air to water intercooler in place of the traditional air to air intercooler, which gives better control for the intake air temperature and the use of a 6 kW air heater after the intercooler and before the intake manifold. The engine is fitted with a high pressure EGR system but to evaluate the capabilities of the boosting system under PPC operation and gain flexibility into controlling the EGR rates, an adjustable low pressure EGR loop was added in the exhaust, downstream the aftertreatment system, which, in this study is not installed. The low pressure EGR loop is cooled by a plate heat exchanger, to be able to get EGR temperatures as low as ambient and then the LP EGR path merges with the intake pipes just before the inlet of the first compressor. To be able to move exhaust gases through the low pressure EGR route a backpressure valve is fitted after the EGR bypass (Fig. 1). The exhaust aftertreatment system is removed from the engine, to be able to measure the raw emission output from the engine. The engine is coupled to a 300 kW electric motor to be able to operate the engine without having combustion as well as provide the necessary load for the engine to function. The control system of the engine consists of a Labview based inhouse developed software, that operates on the host-target principle. A dedicated real time computer, NI PXI-e8135 is used and is fitted with two FPGA cards for the fast sensors data acquisition and real time calculations of in-cylinder pressure, fuel rail pressure, torque and Crank Angle Degree (CAD)/Top Dead Center (TDC) signals. For the CAD signal an external Leine-Linde optical encoder is fitted to the engine, with a resolution of 0.2 degrees. In the real-time computer, a DAQ card, PXI6221 is used for the slow sensor sampling as well as the temperature logging. Injector actuation is possible with the use of NI driven modules. For the in-cylinder pressure acquisition, four AVL GH14D piezoelectric sensors are used and are mounted into specially drilled positions below the glow plugs. On the target side, the computer that is used runs the graphical interface that enables the engine operation by sending the control commands to the real-time computer and provides a graphical view of the output values of all the different sensors as well as the in-cylinder pressure trace. The fuel that was used was a mixture of iso-octane and n-heptane,
Table 2 Fuel properties. RON MON AFRs H/C ratio
PRF75 (Table 2). Primary reference fuel was used to minimize the fuel properties variations of different batches used and because the focus of this study was mainly on the EGR and gas exchange system interaction and less on the combustion behavior. The fuel was treated with 400 ppm of lubricity additive, Infenium R655, to prolong the lifetime of the high-pressure fuel pump as well as the rest of the fueling system, which is not adapted for gasoline operation. Fuel flow is measured with a Sartorious CPA10001 scale and logged into the labview control system. The emissions are analyzed using a Horiba Mexa 7100 system, with two separate CO2 lines so that the total CO2 percentage in the intake can be measured and calculate the EGR ratio.
Total EGR % =
% CO2 intake % CO2 exhaust
(1)
To be able to measure the CO2 percentage for long route individually, a second sampling point is used after the long route connection pipe. A portable emission analyzer (Horiba MEXA-554JE) is used to sample and log the recorded amount. Therefore, the long route EGR percentage can be defined as
LR EGR % =
% CO2 LR % CO2 exhaust
(2)
Soot emissions are measured with an AVL micro-soot sensor unit. Output of the equipment is in mg soot/m3 exhaust flow. All of the emissions are measured engine out and are converted to brake specific emissions in g/kWh, from the exhaust flow and the engine brake power as
bsEmission =
ṁ emission Pbrake
(3)
For the combustion data analysis, 150 engine cycles are saved and are processed offline with the use of Matlab software. Data is saved HP EGR cooler
HP EGR
To emission system
75 75 15.1:1 2.258:1
Wastegate VNT
4 3
Exhaust out LP EGR cooler
2 1
Air in
HP turbo LP turbo
Intercooler Fig. 1. Experimental setup, engine layout. 744
Air Heater
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3.3. Combustion phasing sweep (CA50)
after the engine has reached steady state conditions. These conditions are considered reached, when the oil, coolant, intake and exhaust temperatures have stabilized. Apparent heat release analysis is performed from the in-cylinder pressure trace using the equation:
In this study, the effects of the different EGR routing options will be evaluated. To be able to do that, different engine conditions are needed, to get different amounts of flow through the turbocharger system. To achieve that, four engine speeds are selected, that cover the low to medium engine speed range and four different loads that cover the low to medium load.
To better evaluate the effects of EGR on the gas exchange system, a CA50 sweep is performed at the two extreme points, at 20% EGR through the short route and 20% EGR through the long route. By retarding the combustion, a higher amount of energy can be added in the exhaust flow which can be beneficial for the overall efficiency of turbocharger, leading to a higher gas exchange efficiency. At the same time, the retarded combustion will reduce the gross indicated efficiency. This investigation wants to examine if there is a beneficial trade of in terms of net indicated efficiency. If the gas exchange efficiency increases more than the reduction in gross indicated efficiency, an overall higher net indicated efficiency can be realized in the end. The RPM are kept at 1800, the load is at 21.5 bar FuelMEP and the intake temperature is kept at 80 ℃. The fuel injection pressure is constant at 1100 bar. Five different CA50 are evaluated. These are at 7, 9, 11, 13 and 15 degrees ATDC. A note regarding the EGR percentage that is used in this study. In the literature, the different partially premixed concepts utilize considerable amounts of EGR, 50% EGR is a commonly used number, to reduce the NOx emissions and prolong the ignition delay. For this study only 20% of EGR was used. While that amount will certainly increase the NOx emissions, it is necessary for the different EGR routes to be able to provide the required EGR amount. During early testing it was found that with the current setup, EGR numbers higher that 20% through the short route give unstable turbocharger operation leading to pressure fluctuations. Therefore 20% EGR was used to keep stable engine operation.
3.1. Constant RPM – Load, EGR sweep
4. Results - discussion
γ ∂p ∂Q ∂V 1 = p + V ∂θ γ − 1 ∂θ γ − 1 ∂θ
(4)
The in-cylinder pressure trace is used also to calculate the indicated mean effective pressure (IMEP). Both the gross (IMEPg) and the net (IMEPn) IMEP are calculated in order to estimate the gas exchange efficiency. Gas exchange efficiency is defined as:
ηge =
IMEPn IMEPg
In order to keep all the numbers with the same units, the amount of consumed fuel is expressed in pressure units, similarly to the expression of BMEP and IMEP and is named FuelMEP. A more thorough definition can be found in Appendix A. 3. Methodology
For this part the four different engine speed - load points were evaluated. They are summarized in Table 3. The first three load points represent points of high gross indicated efficiency that were evaluated in previous work [13]. The fuel energy is adjusted at 23.5 bar FuelMEP at low speed and 21.5 bar FuelMEP at medium and high speed, which results in an IMEPg of around 10 bar. Total amount of EGR is 20%, while a sweep is performed from the long route to the short route, resulting in 0%, 50% and 100% of EGR through the short route. No limit is imposed on lambda values that are expected to vary due to the different exhaust mass flow through the turbine, which will result in different amount of boost pressure. As a result, the emissions will vary as well. Intake temperature is kept constant at 75–80 °C to promote faster air fuel mixing and help with the ignition of the PRF75. The fourth load point represents an actual high frequency point in the NEDC/Worldwide harmonized Light vehicles Test Procedure (WLTP) test cycle for passenger vehicles. In this case the BMEP will be kept constant at 5 bar BMEP while the FuelMEP will vary around an average value of 13.5 bar. The total EGR percentage will be kept constant at 20%. The rest of the parameters are kept same as the other three load points.
4.1. Operation at different RPM and loads 4.1.1. Gas exchange efficiency Routing the EGR from long to short route shows an effect on gas exchange efficiency under all operating conditions. For the high engine load cases, generally, the gas exchange efficiency is higher than 90%, but there is a peak when EGR is split 50–50% between long and short route. On the other hand, at the low load case, the most efficient operation is when using the short route option (Fig. 2). An explanation regarding that behavior can be given from the fact that to get the necessary EGR flow through each route the exhaust backpressure should be high enough to move the EGR. For the long route the backpressure is created by a valve positioned on the exhaust pipe. For the short route, backpressure is created by the adjustment of the VGT position. This necessary backpressure is minimized in these cases when the EGR is split between both routes. Still at the low load NEDC case, a different trend appears. The highest gas exchange efficiency appears when gases are routed through the short route. An explanation for this is that at this low load point the exhaust mass flow is not high enough and requires higher amount of throttling from the backpressure valve for the long route to provide the necessary EGR amount both for the long route and the split route case, where the exhaust flow is reduced compared to the single route case.
3.2. Constant lambda EGR sweep In the previous part the air to fuel ratio (AFR) would vary based on the intake conditions. Still, under certain engine operating conditions, can be necessary the AFR to be higher than a specific value, to avoid producing an excess of emissions, such as soot. For this part the effects of EGR routing are investigated while keeping the AFR constant at lambda value of 1.9. The rest of the control parameters remain like the previous part; EGR percentage is kept constant at 20%, FuelMEP is 20.5 bar, injection pressure is 1100 bar, RPM is at 1800, intake temperature 80 ℃. EGR is swept from 100% through the long route, to 50–50% and finally to 100% through the short route.
Table 3 Engine operating conditions.
745
# [–]
Speed [RPM]
Intake temp [°C]
FuelMEP [bar]
Rail pressure [bar]
1 2 3 4
1200 1800 2400 1500
77 80 75 75
23.5 21.5 21.5 ∼13.5
1100 1100 1100 800
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Regarding emissions, a similar trend is observed with NOx emissions. As the EGR is moved towards the short route the NOx emissions increase. Although the oxygen content, through the lambda value, remains almost the same, there is a small variation between the cases, together with a small variation of the total EGR percentage, from 20% to 20.5%. Mixing of EGR through the short route can be less homogenous compared to the long route case and can affect the combustion between the different cylinders. All these effects as well as the slightly earlier combustion phasing in terms of CA50, from 12 °ATDC at the long route to 11 °ATDC at the short route can explain the increase of NOx emissions (Fig. 8). Soot emissions show a minimal increase in terms of g/kWh but this is attributed to the reduced brake efficiency when moving from long route to split route and finally short route, because the emissions in terms of mg/m3 remain fairly constant in absolute values. 4.3. Effects of combustion phasing Fig. 2. Gas exchange efficiency for different RPM – load.
As the combustion is phased to later CA50 values, there is no large effect on the gas exchange efficiency. At the load point of 21.5 bar FuelMEP and with these engine settings the efficiency is quite high, and it seems that there is little to gain in terms of gas exchange efficiency by retarding the combustion (Fig. 9). As expected the gross indicated efficiency is affected significantly as the combustion phasing is retarded (increase in CA50). The effect of gas exchange efficiency on the net indicated efficiency is minimal, so it is not beneficial to retard the combustion under these operating conditions (Fig. 10). Emission wise, due to the late combustion, in cylinder temperatures go down with late CA50, leading to lower NOx emissions both in ppm and in specific emissions (g/kWh) (Fig. 11). The effect on the decreased temperature is higher than the effect of the decreased efficiency on the specific emissions and as a result the trend is that also the specific emissions show a decrease with the later combustion. Regarding the soot emissions, a different trend appears. While the soot emissions are low and at similar levels in terms of mg/m3 as the brake efficiency decreases the specific emissions increase (Fig. 12). Still due to the higher premixed part of the combustion as well as the higher temperatures they are low in absolute numbers.
4.1.2. Emissions When EGR is routed from long route to short route the trend regarding the NOx emissions is towards an increase (Fig. 3), in terms of specific emissions, while the soot emissions tend to decrease (Fig. 4). This can be attributed to the higher lambda values when operated with short route EGR (Fig. 5). As mentioned before, to maintain the necessary pressure difference between exhaust and intake and move the EGR through the short route, the VGT position is adjusted accordingly, which leads, as a result, to increased intake pressure. Higher intake pressure leads to higher air flow in the engine, which can explain the trend of higher NOx and lower soot emissions from the short route. Except of the case of the 1200 RPM where both NOx and soot emissions increase. This is a specific case and has to do with the fact that the lambda value decreases as EGR is routed through the short route. VGT adjustment is not necessary in this case, therefore the lower exhaust flow through the turbine gives reduced intake air pressure, leading to lower amount of available oxygen. As the air to fuel ratio becomes less lean and reaches a low limit of 1.2 that leads to higher combustion temperatures, resulting in higher soot and NOx emissions. Similar results appear also for the 1500 RPM case. As the EGR is moved from long to short route, no additional throttling is needed. As a result, lambda values decrease slightly (Fig. 5), leading to higher combustion temperatures and increase in the overall NOx emissions (Fig. 3). However, overall lambda values are high enough to lead to increased soot oxidation because of the higher combustion temperatures (Fig. 4). The difference in NOx and soot emissions as well as in gas exchange efficiency, for the previous cases can be summarized in Table 4.
4.4. Future work While the results of splitting the EGR routes provide an insight on the question if it beneficial to have a split route system, in terms of efficiency and emissions, more investigations can be performed to
4.2. Constant lambda operation As the EGR is swept from long route to short, the lambda value is quite stable around the target value of 1.9 ± 1.6%. While trying to maintain the oxygen content constant, there is an effect on gas exchange efficiency as EGR routing is moved from long to short route (Fig. 6). In this case the lowest efficiency can be found in the split between the two routes, while the highest efficiency is when all the EGR is passed through the long route. Again, this can be attributed to the necessary adjustment in the intake pressure to keep the target lambda value. This adjustment is done through the VGT part of the turbine which leads to higher exhaust pressure and thus reducing the gas exchange efficiency. Different EGR routing options do not influence gross indicated efficiency, but affect obviously the net indicated efficiency and as a result brake efficiency as well (Fig. 7).
Fig. 3. NOx emissions for different RPM – load. 746
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Fig. 4. Soot emissions for different RPM – load.
Fig. 6. Gas exchange efficiency for the constant lambda cases.
Fig. 5. Lambda value for different RPM – load.
Fig. 7. Gross, net and brake efficiency for the constant lambda cases.
Table 4 Percentage difference for NOx, soot and gas exchange efficiency between the best and the worst conditions. (NOx and soot are percentage reduction; gas exchange efficiency is percentage improvement). RPM × FuelMEP
NOx
Soot
Gas exchange efficiency
1200 × 23.5 1500 × 13.5 1800 × 21.5 2400 × 21.5
16.7% 39% 17% 50.5%*
56% 57% 28% 22%
1% 2.3% 4.5% 2.1%
better evaluate the benefits. In this study, the OEM turbocharging system was used, which was optimized for diesel operation with moderate amounts of EGR and short route operation. This posed a limitation on the achievable EGR percentage because EGR through the short route could not go higher that 20% whereas through the long route, EGR up to 40% could be used. Since PPC requires higher amounts of dilution, a better optimized system for higher EGR rates and possibly for split route operation will be valuable. PPC operation with higher amounts of EGR would help reduce NOx emissions even further although it could pose limitations on the maximum achievable load and possibly reduce further the overall gas exchange efficiency. Furthermore, compressor maps of the turbo system will give a further assessment on the conditions on which the system operates and how close these are to the maximum turbocharger efficiency, highlighting the improvement that can be
Fig. 8. Emissions, NOx and soot, for the constant lambda cases.
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Fig. 12. Effect of combustion phasing on specific soot emissions.
Fig. 9. Effect of combustion phasing on gas exchange efficiency.
reached and the system performance can be evaluated on a larger load – speed area. 5. Conclusions For this study, a four-cylinder light duty diesel engine was operated under PPC conditions to evaluate the effects of different EGR routing options on gas exchange efficiency, overall efficiency and emissions. These findings can be used in additional studies as a possible starting point, when examining further the effects of EGR on low temperature combustion, especially at much higher EGR rates, where the penalty on the gas exchange efficiency is going to be higher. (1) There is an effect on gas exchange efficiency from EGR routing even with a non-optimized turbocharger. Combining both EGR routes provides better results, in most of the cases, but no definite conclusion can be drawn. High loads favor a split between the two routes, while low loads perform better with short route. (2) If only the EGR percentage is kept constant while moving from long to short route, there is going to be an effect on oxygen content. This is a result of the different backpressure settings that are necessary to reach the target EGR percentage. If there is no lambda value target, oxygen mass will differ, which as a result will affect NOx and soot emissions. NOx emissions will tend to increase, while soot emissions decrease, due to higher lambda values when operated with short route EGR. (3) Keeping constant both lambda value and EGR percentage, affects negatively the gas exchange efficiency of the split route, leading to the lowest value compared to only short, or only long route. Emissions have the same trend as before. (4) Retarding the combustion (late CA50) has minimal effect on gas exchange efficiency and as a result negligible effect on net indicated efficiency therefore negatively affects the brake efficiency. Due to that, it has a negative effect on brake specific soot emissions. Brake specific NOx emissions reduce due to lower combustion temperatures.
Fig. 10. Effect of combustion phasing on Gross and Net indicated efficiencies.
Acknowledgments Fig. 11. Effect of combustion phasing on specific NOx emissions.
The authors would like to acknowledge the Competence Centre for Combustion Processes, KCFP (Project number 22485-4) and the Swedish Energy Agency for providing financial support for this study. Also, the authors would like to thank Dr. Kenan Muric, for providing the Matlab code that could calculate the specific heat of various gases based on the JANAF tables.
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Appendix A. Definitions of mean effective pressures and engine efficiencies
A.1. Fuel Mean Effective Pressure, FuelMEP Fuel Mean Effective Pressure is defined as fuel energy per cycle divided with the engine displacement. Fuel energy is the fuel mass per cycle multiplied by the fuel lower heating value.
FuelMEP =
mf × QLHV (A.1)
Vd
A.2. Gross Indicated Mean Effective Pressure, IMEPg Gross Indicated Mean Effective Pressure, IMEPg, is defined as the work to the piston during the compression and the expansion stroke divided with the engine displacement.
IMEPg =
1 Vd
180
∫−180 pdV
(A.2)
A.3. Net Indicated Mean Effective Pressure, IMEPn Net Indicated Mean Effective Pressure, IMEPn, is defined as the work to the piston during the entire four stroke cycle divided with the engine displacement.
IMEPn =
1 Vd
360
∫−360 pdV
(A.3)
In both cases, zero degrees is defined at the Top dead Center during the combustion part. A.4. Brake Mean Effective Pressure, BMEP Brake Mean Effective Pressure is defined as the useful work per cycle divided by the displacement.
BMEP =
P × nR Vd × N
(A.4) 3
where P is brake power [W], Vd is engine displacement [m ], N is engine speed [1/s] and nR is the number of crank revolutions per power stroke. With these definitions the different efficiencies can be defined as well. A.5. Gross Indicated Efficiency
ηgross =
IMEPg (A.5)
FuelMEP
A.6. Net Indicated Efficiency
ηnet =
IMEPn FuelMEP
(A.6)
A.7. Gas Exchange Efficiency
ηge =
IMEPn IMEPg
(A.7)
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A.8. Brake Efficiency
ηb =
BMEP FuelMEP
(A.8)
Appendix B. Supplementary material Supplementary data to this article can be found online at https://doi.org/10.1016/j.applthermaleng.2019.02.108.
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