Experimental investigation on the influences of exhaust gas recirculation coupling with intake tumble on gasoline engine economy and emission performance

Energy Conversion and Management 127 (2016) 424–436

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Energy Conversion and Management journal homepage: www.elsevier.com/locate/enconman

Experimental investigation on the influences of exhaust gas recirculation coupling with intake tumble on gasoline engine economy and emission performance Jianqin Fu a,⇑, Guohui Zhu a, Feng Zhou a, Jingping Liu b, Yan Xia b, Shuqian Wang a a b

State Key Laboratory of Advanced Design and Manufacturing for Vehicle Body, Hunan University, Changsha 410082, China Research Center for Advanced Powertrain Technology, Hunan University, Changsha 410082, China

a r t i c l e

i n f o

Article history: Received 28 May 2016 Received in revised form 7 September 2016 Accepted 8 September 2016

Keywords: Gasoline engine Exhaust gas recirculation Intake tumble Combustion Thermal efficiency

a b s t r a c t To improve the economy and emission performance of gasoline engine under part load, the approach of exhaust gas recirculation coupling with intake tumble was investigated by bench testing. Based on a naturally aspirated gasoline engine, the sweeping test of exhaust gas recirculation rate was conducted in two intake modes (with/without intake tumble), and the parameters related to engine heat-work conversion process and emission performance were measured. Through comparing and analyzing the measured data, the effects of exhaust gas recirculation coupling with intake tumble on gasoline engine economy and emission performance were revealed. The results show that pumping loss decreases gradually while in-cylinder residual gas fraction increases linearly with the exhaust gas recirculation rate increasing; the high-pressure cycle efficiency ascends with exhaust gas recirculation rate increasing due to the decrease of heat transfer loss and exhaust gas energy loss. Thus, the improvement of indicated thermal efficiency is the superposition of double benefits of low-pressure cycle and high-pressure cycle. At 1600 r/min and 2.94 bar, the indicated thermal efficiency can be increased by 4.29%. With the increase of exhaust gas recirculation rate, nitrogen oxide emissions almost fall linearly, but hydrocarbon and carbonic oxide emissions have no obvious change in the effective range of exhaust gas recirculation rate. The biggest advantage of intake tumble is that it can extend the effective range of exhaust gas recirculation rate. As a result, the potential of energy conservation and emission reduction of exhaust gas recirculation is largely improved. Ó 2016 Published by Elsevier Ltd.

1. Introduction Due to the multiple advantages, such as high specific power, low noise and vibration, gasoline engine has been widely used as the power source of automobile. In China, about 80% of automobiles take the gasoline engine as power. Owing to the increasing energy crisis and environment pollution, energy conservation and emission reduction have become an important development strategy in the world [1]. Therefore, to promote the fuel economy and emission performance of gasoline engine is not only the requirement of energy conservation and emission reduction, but also the demand for the survival of traditional gasoline engine [2]. Only by continuously improving the combustion and emission performance of gasoline engine, can the traditional gasoline engine continue to hold the leading position in the future transportation [3].

⇑ Corresponding author. E-mail address: [email protected] (J. Fu). http://dx.doi.org/10.1016/j.enconman.2016.09.033 0196-8904/Ó 2016 Published by Elsevier Ltd.

To achieve these goals, a lot of experts and scholars have made unremitting efforts in the last few decades [4]. Generally, the effective thermal efficiency of current gasoline engine is relatively low especially under part load. From the analysis of gasoline engine thermal balance [5], one can see that the thermal efficiency is very low in low-load operating conditions. The main reasons come from two aspects. On one hand, gasoline engine has to suffer a higher throttling loss under part load, which brings a negative impact on the engine thermal efficiency since part of effective work is consumed to overcome the throttling loss [6]. On the other hand, the combustion duration of mixture gas becomes longer under low load [7], and it leads to the bad heatwork conversion efficiency of gasoline engine. Furthermore, for the automotive gasoline engine, in most situations it operates under part load rather than wide open throttle (WOT) conditions [8]. Therefore, to improve the thermal efficiency of gasoline engine under part load operating conditions is of great practical significance for automobile energy saving.

J. Fu et al. / Energy Conversion and Management 127 (2016) 424–436

425

Nomenclature

ak

_ m V_

P T Rg AK

qs

D cS P1 P2

j

TR cT n DMFL Q_ cP

r

N REGR m Vh

g

Hu

flow coefficient [–] mass flow rate [kg/s] volume flow rate [m3/s] pressure [bar] gas temperature [K] gas constant [J/(kg
K)] piston area [m2] gas density [kg/m3] bore diameter [m] gas flow velocity [m/s] upstream pressure [bar] downstream pressure [bar] isentropic exponent [–] tumble ratio [–] linear rotational speed [m/s] paddle wheel speed [m/s] mean paddle wheel diameter [m] energy flow [kJ] constant-pressure specific heat [kJ/(kg
K)] standard deviation [–] cycle number [–] EGR rate [%] mass [kg] displacement [L] efficiency [%] low heating value [kJ/kg]

eff exh int fue coo out in act the cyl std

effective exhaust intake fuel coolant outlet inlet actual theoretical cylinder standard state

Abbreviation EGR exhaust gas recirculation NA naturally aspirated WOT wide open throttle RGF residual gas fraction PMEP pumping mean effective pressure NMEP net mean effective pressure IMEP indicated mean effective pressure TDC top dead center MFB mass fraction burned COV coefficient of variation SI spark ignition NOx nitrogen oxide HC hydrocarbon CO carbonic oxide

Subscripts air air HP high pressure

Since exhaust gas recirculation (EGR) can effectively reduce the pumping losses of gasoline engine under part load, it became another energy saving technology for gasoline engine and attracted much attention in recent years [9]. As is well known, EGR is a common technique for IC engine combustion control and emission reduction [10]. A large number of studies demonstrate that EGR can effectively reduce the combustion temperature, thus it can restrain the generation of nitrogen oxide (NOX) [11]. For this reason, EGR plays an increasingly important role in engine combustion and emission control technology. Ivanicˇ et al. [12] investigated the effects of hydrogen enhancement on efficiency and NOx emissions of lean and EGR-diluted mixtures in a spark ignition (SI) engine, and concluded that the effect of EGR at equal dilution on NOx is substantially greater than the effect of excess air. Grandin and Angström [13] stated that both lean homogeneous operation and cooled EGR are possible replacements of fuel enrichment, and the cooled EGR allows the use of a three-way catalytic converter at all operating conditions, resulting in substantially lower tailpipe emissions of carbonic oxide (CO) and hydrocarbon (HC) compared to rich operation. Cairns and Blaxill [14] investigated the effects of combined internal and external exhaust gas recirculation on gasoline controlled auto-ignition. They found that introducing additional external exhaust gas can retard ignition, reduce the rate of heat release and limit the peak knocking pressure. Mardi et al. [15] conducted a numerical investigation on the influence of EGR in a supercharged SI engine fueled with gasoline and alternative fuels, and stated that 10% of EGR is the most desirable amount from the viewpoint of emissions and power. Papagiannakis [16] studied the effects of EGR and air inlet preheating on engine emission and performance, and demonstrated that combination of EGR and air inlet preheating could reduce BSFC

and CO without NO increase. Abd-Alla [17] made a review on the potential of EGR to reduce the exhaust emissions, particularly NOx emissions, and to delimit the application range of this technique. In recent years, EGR was commonly used also in gasoline engines together with other advanced techniques [18]. Tang et al. [19] studied the influences on combustion characteristics and performance of EGR vs. lean burn in a gasoline engine, and revealed the theoretic potential and practical limitations on pumping loss reduction via in-cylinder dilution in a SI engine. Zhang et al. [20] investigated the influence of EGR and oxygen-enriched air on diesel engine NO-Smoke emission and combustion characteristic, and achieved the optimal NO-Smoke emission at various conditions. Liu et al. [21] studied the effect of EGR coupling lean-burn gasoline engine on NOx purification of lean NOx trap based on experimental research and CHEMKIN software, and analyzed the effect of exhaust gases on NOx deterioration. Bozza et al. [22] studied the potentials of cooled EGR and water injection for knock resistance and fuel consumption improvements of gasoline engines. The presented results highlight that both the solutions involve significant BSFC improvements, especially in the operating conditions at medium engine speeds. Although lots of studies were carried out on EGR as well as EGR together with other technologies, the existing research mainly focused on the effects of EGR on engine combustion characteristics and emission performance. To the authors’ knowledge, the research on the combined effects of EGR coupling with intake tumble on gasoline engine working processes especially the heat balance and in-cylinder RGF is relatively less, and the experimental data about gasoline engine EGR related to tumble flow are still scarce. As a result, the relationships among EGR rate, intake tumble, gas exchange performance, in-cylinder RGF, energy

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distribution, etc, are still not clear in gasoline engines together with EGR and intake tumble. In addition, the optimum EGR rate for different engine operating conditions and different combustion conditions, e.g., EGR coupling with intake tumble, is also debatable [23]. Based on the above considerations, in this study the sweeping test of EGR rate was carried out on a gasoline engine in different intake modes (with/without intake tumble), and then a systematic study was made on the combined effects of EGR coupling with intake tumble on the working processes, combustion and emission performance of engine. In this study, the influences of EGR coupling with intake tumble on gasoline engine economy performance were investigated from gasoline engine working processes and incylinder heat balance, while the influences on emission performance were concerned on the volume concentration. In this way, the influence mechanisms of EGR coupling with intake tumble on gasoline engine working processes and the relationships among EGR rate, intake tumble, gas exchange performance, in-cylinder RGF, combustion and emission performance, in-cylinder energy distribution and indicated thermal efficiency were revealed. All these not only extended the study on gasoline engine EGR technology, but also provided the theory basis for gasoline engine energy saving and emission reduction.

For the purpose of improving the combustion performance and energy-saving potential of engine, a port blocker is installed in gasoline engine intake manifold near the intake port so as to enhance the tumble intensity of intake gas. To demonstrate the effects of port blocker on the flow performance of engine intake port, the intake flow measurements were carried out on the flow test bench in two kinds of port blocker state. Above all, the definitions and calculation formulas of flow coefficient and tumble ratio are introduced. The flow coefficient is defined as the ratio of the actual obtained mass flow rate of intake gas and the theoretical mass flow rate, and the mathematical expression is given as:

_ m

ð1Þ

_ act is the actual mass flow rate of where ak is the flow coefficient; m intake gas at standard conditions, and its calculation formula is given as follows:

_ act ¼ V_
m

Pstd Rg
T std

ð2Þ

where V_ is the volume flowrate measured by flow bench; Pstd and T std are the atmospheric pressure and temperature at standard conditions, respectively; Rg is the gas constant. The theoretical mass flow rate for a defined cross sectional area can be calculated via the following formulas:

_ the ¼ AK
qs
cS m AK ¼

p 4


D2cyl

ð3Þ ð4Þ

_ the is the theoretical mass flow rate; AK is the piston area; qs where m is the gas density under isentropic conditions; Dcyl is the cylinder bore diameter; cS is the gas flow velocity under isentropic conditions, the calculation formula of which is given as follows:

vffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi " #ffi u  j1 u 2j j P 2 cS ¼ t
R
T
1 P1 j  1 g std

qs ¼

ð5Þ

 j1 P1 P2
Rg
T P1

ð6Þ

where T is the measured gas temperature. The tumble ratio is defined as the ratio of the rotational air speed around the tumble axis to the axial speed of the air flow in cylinder, the mathematical expression of which is given as:

TR ¼

cT cA

ð7Þ

where T R is the tumble ratio; cT is the linear rotational air speed, which can be calculated via the following formula:

cT ¼ p
DMFL
n

ð8Þ

where DMFL is the mean paddle wheel diameter (as shown in Fig. 1 (b)); n is the paddle wheel speed. At the same time, the axial velocity of the air flow in cylinder can be calculated according to Formula (9):

cA ¼

2. Principles of EGR coupling with intake tumble

ak ¼ _ act mthe

where P 1 is the upstream pressure of valve; P2 is the downstream pressure of valve; j is the isentropic exponent. In addition, the gas density in Formula (3) can be calculated as follows:

V_ D2cyl


p4

ð9Þ

Fig. 1(a) is the schematic diagram of intake flow (including swirl and tumble) test setup, and Fig. 1(b) shows the paddle wheel geometry (tumble). All measurements were conducted at a constant pressure difference between tube and atmosphere. The gas pressure was measured and adjusted automatically by the flow bench controller, which also measured the air flow. Ambient pressure and ambient temperature were recorded by the pressure sensor and temperature sensor, respectively. Then the tumble ratios were obtained by evaluating the rotational speed of the paddle wheels (as shown in Fig. 1(a) and (b)). The measured results of flow coefficient and tumble ratio in two port blocker states are compared in Fig. 2(a) and (b), respectively. As illustrated in Fig. 2(a), for the port blocker closing, the flow coefficient always increases with valve lift increasing and the maximum value reaches up to 0.19. For the port blocker opening, the flow coefficient first increases and then remains unchanged with valve lift increasing, the maximum value of which is only 0.06. However, the situation is exactly opposite in the tumble ratio. As can be seen from Fig. 2(b), under the condition of port blocker closing, intake tumble ratio is very low or even can be ignored when the valve lift is lower than 6 mm. Nevertheless, when the port blocker is opening, it increases sharply and the maximum value reaches up to 5.9. Thus it can be seen, the port blocker has two kinds of effects on the engine performance. On one hand, the enhanced intake tumble is beneficial to improve the in-cylinder combustion process. On the other hand, the decreased flow coefficient means the increase of intake flow resistance, and thus the engine will suffer a higher pumping loss during the intake process. To better analyze the effects of EGR coupling with intake tumble on gasoline engine working processes, in this study the working cycle of engine is decomposed and then re-defined. For a four-stroke engine, the exhaust and intake strokes are defined as low-pressure cycle, while the compression and expansion strokes are defined as high-pressure cycle. In terms of low-pressure cycle, it not only influences the volumetric efficiency and in-cylinder RGF of engine, but also consumes part of piston work to overcome the pumping loss in the gas exchange process. Compared with lowpressure cycle, the high-pressure cycle has a more significant influence on engine economy and emission performance since it directly determines the combustion and heat-work conversion

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Swirl and Tumble Test Bench

Flow Bench

2 Tamb,P amb

4, 5

ΔP (Test Pressure)

nPa

1 9

14

5

3

10

7

8 6

9

Tumble

12

15

Swirl 13

11

10

(1) Cylinder Head; (2) Valve Lift Adjustment; (3) Valve; (4) Temperature Sensor; (5) Pressure Sensor; (6) Cylinder Liner; (7) Swirl Paddle Wheel; (8) Tumble Paddle Wheel; (9) Paddle Wheel Support; (10) Air Filter and Speed Sensor; (11) Air Flow Meter; (12) Compressor Unit; (13) Compensating Tank; (14) Control Unit and Display; (15) Shut Off Valve

(a) Schematic diagram of intake flow (swirl and tumble) test setup.

(b) Paddle wheel geometry (tumble) Fig. 1. Schematic diagram and structure principles of intake flow test setup.

process. When the EGR and intake tumble are introduced, both the low-pressure cycle and the high-pressure cycle are changed, thus the gas exchange process and heat-work conversion process of engine are impacted. When the engine exhaust gas is introduced to intake manifold, both the composition and proportion of engine intake gas are varied, and then the intake gas consists of fresh air and recirculated exhaust gas. Accordingly, the proportion (mass fraction) of recirculated exhaust gas in the total intake gas is defined as EGR rate,

REGR ¼

mEGR
100% mair þ mEGR

ð10Þ

where REGR is the EGR rate; mEGR is the mass of the introduced exhaust gas; mair is the mass of fresh air. As mentioned above, a completed working cycle (720 °CA) of four-stroke engine can be divided into a low-pressure cycle and a high-pressure cycle. The indicated mean effective pressure (IMEP) of high-pressure cycle can be expressed as:

R 180 IMEP ¼

pdV Vh

180

ð11Þ

where p is the in-cylinder instantaneous pressure; dV is the differential of cylinder working volume; Vh is the displacement of each cylinder.

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J. Fu et al. / Energy Conversion and Management 127 (2016) 424–436

8

Tumble ratio (-)

Flow coefficient (-)

0.20 0.15 Port blocker: Off

0.10

Port blocker: On

0.05 0.00

0

2

4

6

8

10

12

6 Port blocker: Off

4

Port blocker: On

2 0

14

0

2

Valve lift (mm)

4

6

8

10

12

14

Valve lift (mm)

(a) Intake port flow coefficient VS. valve lift

(b) Intake tumble ratio VS. valve lift

Fig. 2. Comparison of intake port flow characteristics and tumble ratio.

Accordingly, the pumping mean effective pressure (PMEP) of engine can be calculated as:

R 540 PMEP ¼

180

pdV

Vh

IMEP
V h mfue
Hu

ð13Þ

where gi HP is the high-pressure cycle efficiency of engine; mfue is the fuel-injection quantity of each cylinder; Hu is the low heating value of fuel. The indicated thermal efficiency of engine can be calculated as:

gi ¼

NMEP
V h mfue
Hu

ð14Þ

where gi is the indicated thermal efficiency of engine; NMEP is the net mean effective pressure of engine, the calculation formula of which is given as follows:

NMEP ¼ IMEP þ PMEP

ð15Þ

According to Formula (15), when the required engine NMEP is fixed, the lower PMEP corresponds to the lower IMEP. Therefore, to decrease the pumping loss can also reduce the required IMEP under the fixed operating conditions. Since the effects of EGR rate on the heat balance of engine were also concerned in this study, the calculation formulas of engine energy flows are given. The engine exhaust gas energy flow and its proportion in total fuel energy can be calculated as follows.

_ exh ðcP;exh T exh  cP;0 T 0 Þ Q_ exh ¼ m Q_

gexh ¼ _ exh
100% mfue Hu

ð16Þ ð17Þ

where Q_ exh is the engine exhaust gas energy flow; cP;exh and cP;0 are the constant-pressure specific heat of exhaust gas at the tested temperature and environment temperature, respectively; T exh and T 0 are the tested exhaust gas temperature and environment temperature, respectively; gexh is the percentage of engine exhaust gas _ exh is the mass flow rate of engine energy in total fuel energy; m exhaust gas, which can be calculated according to the mass flow rate of engine intake gas and fuel:

_ exh ¼ m _ int þ m _ fue m

ð19Þ

Q_

ð12Þ

The high-pressure cycle efficiency of engine can be calculated through the following formula:

gi HP ¼

_ coo
cP;coo ðT coo;out  T coo;in Þ Q_ coo ¼ m

ð18Þ

_ int is the mass flow rate of engine intake gas; m _ fue is the where m mass flow rate of engine fuel. The calculation formulas for engine coolant energy flow and its percentage in total fuel energy are given below.

gcoo ¼ _ coo
100% mfue Hu

ð20Þ

_ coo is the mass flow where Q_ coo is the engine coolant energy flow; m rate of engine coolant; cP;coo is the specific heat of engine coolant; T coo;out and T coo;in are the coolant temperature at the outlet and inlet of engine, respectively; gcoo is the percentage of engine coolant energy in total fuel energy. 3. EGR sweeping test for gasoline engine A four-cylinder, four-stroke, naturally aspirated (NA) gasoline engine was employed as the research object in this study. The main technical parameters of the tested gasoline engine are listed in Table 1. Based on AVL test bench, the sweeping test of EGR rate was conducted in two kinds of port blocker states (Off and On) under engine part load conditions. Or rather, the EGR rate was continuously increased until the combustion process turned unstable. The schematic diagram of experimental setup for EGR sweeping test is depicted in Fig. 3, which also shows the layout of various kinds of sensors. According to this schematic diagram, the methods and processes of the EGR sweeping test were introduced. As shown in Fig. 3, both the lambda sensor and the sampling probe of emission analyzer were installed at the exhaust pipe before the catalyst; the air flow meter for engine intake gas was installed at the intake pipe before the air cleaner, and the air flow meter for the recirculated exhaust gas was located in the EGR pipeline after the EGR cooler; the in-cylinder pressure sensor was installed at the engine cylinder head, the signal of which was firstly transformed into voltage signal and then transferred to the combustion analyzer. Moreover, various pressure and temperature sensors were installed at the test system (see Fig. 3). The specifications of main test instruments and equipments as well as the precision are displayed in Table 2. The test conditions are listed in Table 3.

Table 1 The main parameters of the tested gasoline engine. Items

Content

Engine type Cylinder bore (mm) Stroke (mm) Displacement (L) Compression ratio Ignition sequence Max power (kW)/speed (r/min) Max torque (N
m)/speed (r/min) Intake type Cooling type

Inline 4 cylinder, 4 stroke 80.5 78.5 1.6 10.2 1–3–4–2 90/6000 154/5200 Naturally aspirated Water cooling

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J. Fu et al. / Energy Conversion and Management 127 (2016) 424–436

Battery Combustion analyzer

PC

Charge amplifier

Crank angle calculation

Cylinder pressure sensor

ECU

Fuel injector Air mass flow meter

Coil

Catalyst

Throttle Air filter EGR cooler

EGR exhaust flow meter Coolant temperature control system

Speed sensor

Electric dynamometer

Engine

Oil temperature control system

Temperature sensor location

Pressure sensor location

Lambda sensor location

Emission Sampling location

Fig. 3. Schematic diagram of experimental setup for EGR sweeping test.

Table 2 The specifications of main test instruments and equipments. Equipment name

Type

Precision

Electric dynamometer

INDY S22-2/05251BV-1 PUMA OPEN1.4.1 7351 CST PUMA

Torque: ±0.5% F.S Speed: ±1 r/min ±0.5% F.S ±0.12% F.S Resolution: 15 Bit Sampling rate: 1–200 Hz ±2 °C

Dynamometer control system Fuel consumption meter Data acquisition system

Fuel temperature control system Coolant temperature control system Oil temperature control system Mass air flow meter k analysis meter Combustion analyzer

753C 553 CONSYSCOOL 553-200 554 CONSYSLUBE 554

±2 °C

TP16A.00 ETAS INDISET ADVANCED PLUS 641

±1% F.S ±0.01 /

±2 °C

Table 3 Description of the test conditions. Items

Content

Environment temperature Relative humidity Fuel type Fuel temperature Coolant outlet temperature Oil temperature Intake pressure before air cleaner

25 ± 2 °C 30 ± 5% 93 RON 25 ± 2 °C 88 ± 2 °C 90–100 °C 3 to 0 kPa (Relative pressure)

parameters were measured in two kinds of port blocker state, including engine speed, torque, the flowrate of intake gas and recirculated exhaust gas, intake and exhaust parameters (pressure and temperature), in-cylinder transient pressure, engine emission parameters, etc. In each case, the in-cylinder pressure was recorded in 200 consecutive engine cycles, while other parameters were tested by the mean value method in a fixed time of 30 s. Then, the measured signals were processed as follows. The tested incylinder pressure signals were processed by using AVL combustion analyzer. By this means, the combustion and heat release rates as well as the combustion characteristic parameters were obtained. The tested intake and exhaust gas signals were coupled to the one-dimensional gas dynamic equations of engine gas exchange process. In this way, the in-cylinder RGF could be calculated [24]. The exhaust gas sample was firstly collected by the sampling probe and then processed by the emission analyzer. The measured signals of lambda sensor were analyzed by the excess air coefficient analyzer. As a result, the excess air coefficient was obtained. Table 4 describes the boundary conditions of the sweeping test of EGR rate in two kinds of port blocker states. As it can be seen, two kinds of operating conditions under part load were chosen for this study: one is 2000 r/min and 2.6 bar NMEP, and the other is 1600 r/min and 2.94 bar NMEP. In all the tests, the excess air

Table 4 Description of tested operating conditions (sweeping test of EGR). Case number

Description of test cases

Case 1

2000 r/min; port NMEP = 2.6 bar 2000 r/min; port NMEP = 2.6 bar 1600 r/min; port NMEP = 2.94 bar 1600 r/min; port NMEP = 2.94 bar

Case 2

In the test process, the ignition advance angle was slightly adjusted under each operating condition so as to achieve the best engine performance. Various engine operating and performance

Case 3 Case 4

blocker: off; k = 1.0, blocker: on; k = 1.0, blocker: off; k = 1.0, blocker: on; k = 1.0,

J. Fu et al. / Energy Conversion and Management 127 (2016) 424–436

coefficient was fixed at 1.0, and the EGR rate was regulated by adjusting the opening of EGR valve.

4. Results and discussions As mentioned above, both the low-pressure cycle and highpressure cycle of engine are influenced by the EGR and intake tumble. For this reason, the effects of EGR coupling with intake tumble on engine performance were analyzed from the two subcycles. Based on the tested data of EGR sweeping test, the gas exchange process, in-cylinder combustion process, heat-work conversion performance and heat balance are discussed under two port blocker conditions. In this way, the potential of energy conservation and emission reduction of EGR coupling with intake tumble on gasoline engine can be revealed.

4.1. Effects of EGR rate on gasoline engine combustion process As a significant parameter for combustion performance, the incylinder RGF is concerned firstly. As mentioned previously, in this study the in-cylinder RGF was obtained by the hybrid method of intake and exhaust gas signals measurement coupling with gas dynamics simulation, the detailed processes and principles of which can be seen in Ref. [24]. For the engine with EGR, the in-cylinder residual gas consists of two parts: one is the residual combustion gas of the last working cycle, and the other is the introduced exhaust gas from the EGR intake system. Fig. 4 illustrates the change rule of in-cylinder RGF under various operating conditions. As shown in this figure, there is almost a linear relationship between the in-cylinder RGF and EGR rate. In terms of in-cylinder RGF, the main influence factor is the EGR rate, while the effects of intake tumble can be ignored. Therefore, the in-cylinder RGF can be approximately controlled by regulating the EGR rate. After the introduction of EGR, the in-cylinder combustion and heat release processes are influenced. Fig. 5 shows the influences of EGR rate on the 10–90% combustion duration. As one can see, the 10–90% combustion duration ascends monotonically with the increase of EGR rate, and the increasing trend of 10–90% combustion duration turns more obvious under the higher EGR rate. This phenomenon can be explained as follows. To begin with, the introduced exhaust gas not only dilutes the concentration of mixture gas, but also reduces the in-cylinder combustion temperature, both of which can slow down the combustion velocity of mixture gas [25]. Furthermore, since the introduced exhaust gas is the combustion products of in-cylinder mixture gas, it has a strong negative effect on the chemical reaction kinetics of combustion process [26]. Moreover, Ref. [27] pointed out that the influence of

10-90% comb. duration (oCA)

430

40 30 20 2000 2000 1600 1600

10 0

0

5

10

15

r/min, r/min, r/min, r/min,

20

2.6 bar, Of f 2.6 bar, On 2.94 bar, Of f 2.94 bar, On

25

Fig. 5. Influences of EGR rate on 10–90% combustion duration.

in-cylinder residual gas on laminar flame speed is much larger than that of excess air dilution. Under the same engine operating conditions (speed and NMEP), the 10–90% combustion duration falls sharply when the port blocker is opening. Compared with intake tumble, the effects of engine operating conditions on the 10–90% combustion duration are much smaller. When the port blocker is turned off, the 10– 90% combustion duration in Case 3 (1600 r/min, 2.94 bar, off) is longer than that in Case 1 (2000 r/min, 2.6 bar, off). According to the previous study [27], it can be known that 10–90% combustion duration increases monotonously with engine speed, but decreases with engine load. It seems that this conclusion is not consistent with the above results. This is attributed to the differences of incylinder residual gas in the two cases. As can be seen in Fig. 4, the in-cylinder RGF in Case 3 is larger than that in Case 1, and it causes the 10–90% combustion duration in Case 3 is longer than that in Case 1. However, for the port blocker opening, the influence of RGF on the combustion velocity is weakened due to the enhanced intake tumble. Under the circumstances, the significance of operating condition becomes greater than the in-cylinder RGF. That is why the 10–90% combustion duration in Case 2 (2000 r/min, 2.6 bar, on) is longer than that in Case 4 (1600 r/min, 2.94 bar, on) when the port blocker is opening. To further analyze the influences of EGR rate on in-cylinder combustion process, the heat release rate and mass fraction burned (MFB) are concerned [28]. The instantaneous heat release rate and the MFB without intake tumble are displayed in Figs. 6 and 7, respectively. As shown in Fig. 6, the maximum heat release rate decreases obviously with the increase of EGR rate. When the EGR rate increases from 0% to 15.2%, the maximum heat release rate is decreased by about 1/3. Fig. 7 clearly shows the MFB of mixture

35

RGF (%)

30 25 20

2000 r/min, 2.6 bar, Off 2000 r/min, 2.6 bar, On 1600 r/min, 2.94 bar, Off 1600 r/min, 2.94 bar, On

15 10

0

5

10

15

20

25

30

EGR rate (%) Fig. 4. Influences of EGR rate on engine in-cylinder RGF.

30

EGR rate (%)

Fig. 6. Influences of EGR rate on heat release rate.

J. Fu et al. / Energy Conversion and Management 127 (2016) 424–436

Fig. 7. Influences of EGR rate on mass fraction burned.

431

Fig. 9. Influences of EGR rate on Max. pressure rise rate.

gas under four kinds of EGR rate. As illustrated, the 0–10% MFB changes little with the EGR rate. However, in terms of 10–90% MFB and 90–100% MFB, both of them differ from one another under different EGR rates. Or rather, the differences of MFB become obvious after the top dead center (TDC) of engine. Therefore, the EGR rate has a significant influence on the middle and late stage of in-cylinder combustion processes. Fig. 8 displays the relationship between the engine maximum cylinder pressure and EGR rate under various operating conditions. As one can see, EGR rate has little effect on the engine maximum cylinder pressure. The reasons are given as follows. From the viewpoints of in-cylinder combustion process, the EGR rate has two kinds of effect on the maximum cylinder pressure. On one hand, EGR can slow down the flame propagation speed and increase the combustion duration, which is a negative impact on the maximum cylinder pressure. On the other hand, as the EGR is introduced, the in-cylinder working medium increases, which causes the rise of compression pressure and initial combustion pressure, and this is a positive effect on the maximum cylinder pressure. In a certain range of EGR rate (the range depends on the engine operating conditions), the two kinds of influences are offset largely, thus the maximum cylinder pressure is not sensitive to the EGR rate. Compared with EGR rate, the influence of intake tumble on the maximum cylinder pressure is more obvious. Under the same operating conditions, the maximum cylinder pressure of engine can be promoted by intake tumble, since intake tumble can increase the heat release rate and then reduce the 10–90% combustion duration of fuel-air mixture (the descent rate is 10 °CA or more). Fig. 9 shows the maximum pressure rise rate of engine under different operating conditions. Unlike the maximum cylinder pressure, EGR rate has a great influence on the engine maximum pressure rise rate. As it can be seen, the maximum

where COV IMEP is the coefficient of variation of IMEP for 200 consecutive cycles; rIMEP is the standard deviation of IMEP; IMEP is the mean value of IMEP for 200 consecutive cycles; N is the tested cycle number, which is set to 200 in this study. Fig. 10 illustrates the influences of EGR rate and intake tumble on gasoline engine COVIMEP. As one can see, the COVIMEP almost remains unchanged in a certain range of EGR rate. However, when

Fig. 8. Influences of EGR rate on Max. cylinder pressure.

Fig. 10. Influences of EGR rate on COV of IMEP.

pressure rise rate declines monotonously as the EGR rate increases. This is because the combustion duration is increased and heat release rate is slowed down (see Figs. 5 and 6) with EGR rate increasing. When the port blocker is opening, the maximum pressure rise rate increases sharply due to the enhanced intake tumble. Thus it can be seen, the maximum pressure rise rate is very sensitive to both the EGR rate and intake tumble. Moreover, the combustion stability is also concerned as it is very sensitive to the EGR rate and intake tumble [29,30]. Usually, the coefficient of cyclic variation (COV) was used to evaluate the combustion stability in cylinder, which includes IMEP and combustion period parameter, etc. [31]. As the most frequently used parameter in evaluating engine cyclic variations, the coefficient of variations of IMEP (COVIMEP) was discussed in this study, which is defined as follows.

COV IMEP ¼ rIMEP =IMEP

ð21Þ

rIMEP ¼

rffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi XN  2 IMEP ðiÞ  IMEP =N i

ð22Þ

IMEP ¼

N X IMEPðiÞ =N

ð23Þ

i

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J. Fu et al. / Energy Conversion and Management 127 (2016) 424–436

the EGR rate exceeds a threshold value, the COVIMEP ascends sharply. Thus, one can conclude that the EGR rate has little influence on the COVIMEP in a certain range, but when the EGR rate is higher than a threshold value, the in-cylinder combustion process will turn terrible and unstable [32]. In the meantime, the threshold value becomes the focus of attention. In this study, the maximum EGR rate at which the COVIMEP is lower than 10% is defined as the critical EGR rate. Accordingly, the EGR rate range between 0% and the critical EGR rate is defined as the effective range of EGR rate, since the COVIMEP is relatively low in this range of EGR rate but it increases sharply when the EGR rate is larger than the critical value. As can be seen from Fig. 10, when the port blocker is opening, the effective range of EGR rate turns larger. More specifically, the effective range of EGR rate can be extended to 20% or more by intake tumble. Consequently, the potential of energy conservation and emission reduction can be further improved as the effective range of EGR rate is extended.

Fig. 12. Influences of EGR rate on engine PMEP.

Fig. 11 displays the relationship between the required volumetric efficiency and EGR rate under various operating conditions. As illustrated, when the EGR rate is lower than 10%, the required volumetric efficiency under various operating conditions almost keeps unchanged. Under the circumstances, both the EGR rate and intake tumble have little effects on the required volumetric efficiency. However, when the EGR rate is higher than 10%, things are changed. Taking the operating condition of 1600 r/min and 2.94 bar (Case 3 and Case 4) as an example, the required volumetric efficiency still keeps unchanged as intake tumble is introduced, but it increases with EGR rate increasing when the port blocker is turned off. It means that the required fuel-air mixture can be reduced by intake tumble. As a critical parameter for evaluating the gas exchange performance of engine, the PMEP is discussed. Fig. 12 shows the influences of EGR rate and intake tumble on engine PMEP. As one can see, the PMEP descends monotonously with the increase of EGR rate under the four kinds of operating conditions. Taking Case 2 (2000 r/min, 2.6 bar, port blocker opening) as an example, when the EGR rate increases from 0% to 24%, the engine PMEP decreases from 0.69 bar to 0.61 bar, and the descent rate comes up to 11.6%. Under the same operating conditions, the engine PMEP can be reduced by closing the port blocker (as shown in Case 1 and Case 3), because the intake tumble results in a larger intake resistance. However, since intake tumble can extend the effective range of EGR rate, the maximum decrease rate of PMEP still appears in

the operating condition with port blocker opening (as shown in Case 2 and Case 4). Then, the influences of EGR coupling with intake tumble on the heat-work conversion efficiency of gasoline engine are discussed. As can be seen from Fig. 13, under the four kinds of operating conditions, the high-pressure cycle efficiency ascends with EGR rate increasing in a certain range. For the sake of analysis, Table 5 lists the key values of high-pressure cycle efficiency and indicated thermal efficiency. Taking Case 4 (1600 r/min, 2.94 bar, port blocker opening) as an example, the high-pressure cycle efficiency can be improved by a maximum of 2.43% (as shown in Table 5). However, when the EGR rate goes up to a certain value, the high-pressure cycle efficiency falls sharply. This is because the EGR rate is so high that the in-cylinder combustion stability is influenced [33,34], which can also be explained from Fig. 10. Therefore, there is no practical significance to choose a very high EGR rate to improve the engine thermal efficiency. In a certain range of EGR rate, intake tumble has little effect on the high-pressure cycle efficiency. However, this is not yet the case in the range of high EGR rate. The influence mechanisms of EGR rate on the high-pressure cycle efficiency will be further analyzed from the in-cylinder heat balance in the next section. Fig. 14 shows the influences of EGR rate and intake tumble on engine indicated thermal efficiency. By comparing Fig. 13 with Fig. 14, one can find the sequence of curves in the two figures is different from one another. According to Formulas (13)–(15), it can be seen that the difference between Figs. 13 and 14 gives the PMEP. Since the PMEP differs from one another obviously in various operating conditions (see Fig. 12), it causes the change in the sequence of curves in the two figures. Under the four kinds of operating conditions, the indicated thermal efficiency first increases and then decreases with the EGR rate increasing. Similar to the

Fig. 11. Influences of EGR rate on volumetric efficiency.

Fig. 13. Influences of EGR rate on high-pressure cycle efficiency.

4.2. Effects of EGR rate on gasoline engine thermal efficiency and heat balance

433

J. Fu et al. / Energy Conversion and Management 127 (2016) 424–436 Table 5 Influences of EGR coupling with intake tumble on engine heat-work conversion performance. Operating conditions

High-pressure cycle efficiency at 0% EGR rate (%)

Maximum highpressure cycle efficiency (%)

Maximum increment of high-pressure cycle efficiency (%)

Indicated thermal efficiency at 0% EGR rate (%)

Maximum indicated thermal efficiency (%)

Maximum increment of Indicated thermal efficiency (%)

Case 1: 2000 r/min, 2.6 bar, port blocker closing Case 2: 2000 r/min, 2.6 bar, port blocker opening Case 3: 1600 r/min, 2.94 bar, port blocker closing Case 4: 1600 r/min, 2.94 bar, port blocker opening

37.4

37.8

1.07

29.8

30.3

1.68

37.3

37.9

1.61

29.5

30.4

3.05

36.9

37.2

0.81

30.3

30.9

1.98

37.0

37.9

2.43

30.3

31.6

4.29

Fig. 14. Influences of EGR rate on indicated thermal efficiency.

high-pressure cycle efficiency, intake tumble has little effect on the indicated thermal efficiency in a certain range of EGR rate, but things are changed in the range of high EGR rate since intake tumble can extend the effective range of EGR rate. Compared with the high-pressure cycle efficiency, the effects of EGR rate on engine indicated thermal efficiency are more apparent. This is because the improvement of engine indicated thermal efficiency comes from two aspects: one is the descent of engine PMEP in lowpressure cycle and the other is the ascent of heat-work conversion efficiency in high-pressure cycle. In other words, the improvement of engine indicated thermal efficiency is the superposition of double benefits of low-pressure cycle and high-pressure cycle. Taking Case 4 (1600 r/min, 2.94 bar, port blocker opening) as an example, the indicated thermal efficiency can be increased by 4.29% through adjusting the EGR rate (as shown in Table 5). At the same time, due to the influence of pumping loss, engine has a higher indicated thermal efficiency at high load (Case 3 and Case 4) than that at low load (Case 1 and Case 2). Thus it can be seen, in terms of engine indicated thermal efficiency, the load is the key influence factor and has a higher effect than engine speed; but in terms of highpressure cycle efficiency, the speed is more important than the load. Moreover, since the indicated thermal efficiency reaches its maximum value at the critical EGR rate (see Fig. 14), the performance parameters at the critical EGR rate are focused in the following study. To further analyze the change rule of engine indicated thermal efficiency, the influences of EGR coupling with intake tumble on engine thermal balance are discussed [35]. According to the thermal balance theory of engine [36], the fuel energy released through combustion can be divided into several parts: effective work, exhaust gas energy, coolant energy and remainder heat loss. In this

study, only the exhaust gas energy and coolant energy are considered since the proportion of remainder heat loss is very low. The percentages of exhaust gas energy and coolant energy in total fuel energy are depicted in Figs. 15 and 16. From the two figures one can find that both the percentages of exhaust gas energy and coolant energy monotonously decrease with the increase of EGR rate. As can be seen from Table 6, the exhaust energy percentage can be decreased by 10.76% at most in the effective range of EGR rate. Although the mass flow rate of exhaust gas is increased, the decrease of exhaust gas temperature plays a more vital role and finally it results in the descent of exhaust gas energy percentage. Compared with exhaust energy percentage, the decrease rate of coolant energy percentage is more obvious, which reaches up to 15.48% in the effective range of EGR rate (see Table 6). From Fig. 5 one can see that the in-cylinder combustion duration turns longer with EGR rate increasing, and it seems to be good for promoting the in-cylinder heat transfer loss. However, the incylinder heat transfer loss is still decreased because the maximum combustion temperature turns lower [37]. Therefore, a conclusion can be arrived that the percentages of in-cylinder heat transfer loss and engine exhaust gas energy loss can be reduced by introducing EGR, and this conclusion can also be used to explain the relationship between engine indicated thermal efficiency and EGR rate (see Fig. 14). 4.3. Effects of EGR rate on gasoline engine emission characteristics Since the combustion process is influenced by the approach of EGR coupling with intake tumble, the engine emission characteristics are also changed. Fig. 17 shows the volume concentration of HC emission under various operating conditions. At the same time,

Fig. 15. The percentage of exhaust energy in total fuel energy.

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J. Fu et al. / Energy Conversion and Management 127 (2016) 424–436

Fig. 17. Influences of EGR rate on HC source emission.

Fig. 16. The percentage of coolant energy in total fuel energy.

the HC emissions at the critical EGR rate and at 0% EGR rate are listed in Table 7. As one can see, the volume concentration of HC emission is almost unchanged within the effective EGR range (the range depends on engine operating conditions). However, when the EGR rate is greater than the critical value, the HC source emission sharply increases. Under the four kinds of operating conditions, when the port blocker is opening, the critical value of EGR rate for HC emission turns larger. In other words, the effective range of EGR rate in which the HC emission almost keeps unchanged can be extended by intake tumble. This conclusion is very similar to the change rule of COVIMEP, as shown in Fig. 10. Fig. 18 shows the influences of EGR rate coupling with intake tumble on the volume concentration of CO emission. As can be seen from the figure, the volume concentration of CO emission fluctuates a little in the effective range of EGR rate. However, when the EGR rate is beyond the effective range, the CO emission ascends dramatically, which is similar to the variation trend of HC emission. This is owing to the combustion deterioration under the higher EGR rate, e.g., flameout, incomplete combustion [30,38]. After the EGR is introduced, the concentration of fuel-air mixture in cylinder is diluted, and the recirculated exhaust gas has a strongly inhibitory effect on the flame propagation velocity. All these factors result in the decrease of combustion temperature and combustion velocity (see Figs. 5 and 6) [39]. However, there is no apparent relationship between the CO emission and combustion velocity in the effective range of EGR rate. In other words, so long as there is no combustion deterioration due to the introduced EGR, the EGR rate has very little influence on the HC and CO emissions. For each operating condition (except Case 1), the critical EGR rate makes the CO emission reach the lowest value. Moreover, the

minimum value of CO emission can be decreased by intake tumble, as shown in Table 7. Fig. 19 displays the volume concentration of NOX emission under four kinds of operating conditions. As illustrated, the volume concentration of NOX emission falls rapidly with the increase of EGR rate. This is because the in-cylinder combustion temperature is decreased due to the introduction of EGR [40]. It seems that the operating condition has no relation with the NOX emission, especially at higher EGR rate [41]. However, the NOX emission is very sensitive to the intake tumble. When the port blocker is turned on, the NOX emission has an obvious rise. The reason is stated as follows. According to Ref. [42], the generation of NOX is mainly determined by the oxygen concentration and maximum combustion temperature. Because the intake tumble can improve the in-cylinder combustion velocity (see Fig. 5) and then enhance the maximum combustion temperature [43], it becomes a favorable factor for the generation of NOX. However, since intake tumble can extend the effective range of EGR rate, the minimum NOX emission still appears in the operating conditions with port blocker opening (as shown in Fig. 19). When the EGR rate reaches 10%, the NOX emission can be reduced by more than 50% under the four kinds of operating conditions. Taking Case 2 (2000 r/min, 2.6 bar, port blocker opening) as an example, if the EGR rate is increased to the critical value (19.6%), the NOX emission will be decreased from 3205.4 ppm to 396.7 ppm (see Table 7), which indicates a great potential for NOX emission reduction [44]. Thus, in terms of emission performance, the EGR is the most useful way to reduce NOX emission. Fig. 20 illustrates the influences of EGR rate coupling with intake tumble on the volume fraction of CO2 emission. It can be

Table 6 Influences of EGR coupling with intake tumble on engine heat balance. Operating conditions

Exhaust energy Exhaust energy percentage at 0% EGR percentage at critical EGR rate (%) rate (%)

Descent rate of exhaust energy percentage (%)

Coolant energy Coolant energy percentage at 0% EGR percentage at critical EGR rate (%) rate (%)

Descent rate of coolant energy percentage (%)

Case 1: 2000 r/min, 2.6 bar, port blocker closing Case 2: 2000 r/min, 2.6 bar, port blocker opening Case 3: 1600 r/min, 2.94 bar, port blocker closing Case 4: 1600 r/min, 2.94 bar, port blocker opening

26.1

24.8

4.98

25.7

24.5

4.67

25.1

22.4

10.76

24.2

20.9

13.64

24.7

23.5

4.86

21.3

19.3

9.39

24.5

22.3

8.98

23.9

20.2

15.48

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J. Fu et al. / Energy Conversion and Management 127 (2016) 424–436 Table 7 Influences of EGR coupling with intake tumble on engine emission characteristics. Operating conditions

HC emission at 0% EGR rate (ppm)

HC emission at critical EGR rate (ppm)

CO emission at CO emission at NOX emission at NOX emission at critical EGR rate 0% EGR rate 0% EGR rate critical EGR rate (%) (%) (ppm) (ppm)

CO2 emission CO2 emission at at 0% EGR rate critical EGR rate (%) (%)

Case 1: 2000 r/min, 2.6 bar, port blocker closing Case 2: 2000 r/min, 2.6 bar, port blocker opening Case 3: 1600 r/min, 2.94 bar, port blocker closing Case 4: 1600 r/min, 2.94 bar, port blocker opening

646.4

732.2

0.51

0.51

2643.3

1109.2

14.53

14.59

659.7

822.0

0.39

0.38

3205.4

396.7

14.54

14.67

909.7

981.4

0.62

0.46

2365.9

550.7

14.46

14.54

976.6

1147.6

0.43

0.29

2970.1

318.0

14.59

14.67

Fig. 18. Influences of EGR rate on CO source emission.

Fig. 20. Influences of EGR rate on CO2 source emission.

5. Conclusions

Fig. 19. Influences of EGR rate on NOX source emission.

seen from this figure, the volume fraction of CO2 emission almost keeps unchanged with the EGR rate increasing in the effective range of EGR rate. In this range of EGR rate, intake tumble results in a little increase of CO2 emission at 1600 r/min and 2.96 bar NMEP, while the effect of intake tumble can be ignored at 2000 r/min and 2.6 bar NMEP. Nevertheless, when the EGR rate is larger than the critical value, CO2 volume fraction begins to decrease, which is due to the increase of CO concentration (see Fig. 18). Obviously, the critical value of EGR rate turns larger by intake tumble. It also can be seen from Table 7, the volume fraction of CO2 emission at the critical EGR rate is nearly the same as that at the initial state (EGR rate = 0%). Therefore, the effect of EGR rate on the CO2 emission can be ignored in the effective range of EGR rate, while the effective range of EGR rate is influenced by intake tumble.

In this study, the sweeping test of EGR rate was carried out on a gasoline engine in two intake modes (with/without intake tumble), and then the combined effects of EGR coupling with intake tumble on the working processes, combustion and emission performance of engine were discussed. Through the research of this paper, the following conclusions can be drawn. The in-cylinder RGF is mainly influenced by the EGR rate while the effects of intake tumble can be ignored. Thus, it can be approximately controlled by regulating the EGR rate. The maximum cylinder pressure is not sensitive to the EGR rate, but it increases obviously as the intake tumble is introduced. The maximum pressure rise rate is influenced by both the EGR rate and intake tumble, which deceases with the EGR rate increasing but increases with the introducing of intake tumble. According to the change rules of COVIMEP and the indicated thermal efficiency, the critical EGR rate and effective range of EGR rate were obtained. Without intake tumble, the critical EGR rate is 10% at 2000 r/min and 2.6 bar, but it can be extended to 20% or so by using intake tumble. Therefore, the biggest advantage of intake tumble is that it can extend the effective range of EGR rate. In this way, the potential of energy conservation and emission reduction of EGR is largely improved. Under the four kinds of operating conditions, the indicated thermal efficiency first increases and then decreases with the EGR rate increasing. Intake tumble has little effect on the indicated thermal efficiency in a certain range of EGR rate, but its effect turns more obvious in the range of high EGR rate. Finally, the maximum value of indicated thermal efficiency still appears in the operating condition with intake tumble. The improvement of engine indicated

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thermal efficiency by EGR comes from two aspects: one is the descent of engine PMEP and the other is the ascent of high-pressure cycle efficiency. The improvement of high-pressure cycle efficiency by EGR is due to the decrease of the percentages of in-cylinder heat transfer loss and engine exhaust gas energy loss. In the effective range of EGR rate, the EGR rate has little influence on the HC and CO emissions since the in-cylinder combustion is not deteriorated. But when the EGR rate is beyond the effective range, both the HC and CO emissions will increase sharply. The NOX emission almost falls linearly with the increase of EGR rate, but it increases obviously when intake tumble is introduced. However, the minimum NOX emission still appears in the operating conditions with intake tumble since the effective range of EGR rate is extended. When the EGR rate is increased to the critical value (19.6%), the NOX emission will be decreased from 3205.4 ppm to 396.7 ppm. The CO2 emission almost keeps unchanged in the effective range of EGR rate, but is decreases sharply when the EGR rate is beyond the effective range.

Acknowledgements This research work is jointly sponsored by the National Key Technology Support Program (No. 2014BAG09B01 and No. 2014BAG10B01), and the National Natural Science Foundation of China (No. 51506050 and No. 51376057). The authors appreciate the anonymous reviewers and the editor for their careful reading and many constructive comments and suggestions on improving the manuscript.

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