CONVERGE CFD of gasoline spray,

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2013-01-0250

Experimental and Numerical Study of Water Spray Injection at Engine-Relevant Conditions

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Copyright 2012 SAE International

ABSTRACT

Water spray characterization of a multi-hole injector under

pressures and temperatures representative of engine-relevant

conditions was investigated for naturally aspirated and boosted

engine conditions. Experiments were conducted in an optically

accessible pressure vessel using a high-speed schlieren

imaging to visualize the transient water spray. The

experimental conditions included a range of injection

pressures of 34, 68, and 102 bar and ambient temperatures of

30 200C, which includes flash-boiling and non flash-boiling conditions. Transient spray tip penetration and spray

angle were characterized via image processing of raw

schlieren images using Matlab code. The CONVERGE CFD

software was used to simulate the water spray obtained

experimentally in the vessel. CFD parameters were tuned and

validated against the experimental results of spray profile and

spray tip penetration measured in the combustion vessel (CV).

With the validated CFD model, water spray injection into an

engine in-cylinder configuration was simulated. The CV

experimental results showed that collapsing spray plumes

were observed for higher temperature of the charge, showing

reduced spray tip penetration. The engine CFD results showed

that water injection at 90 BTDC showed better vaporization and decreased the formation of liquid wall film on piston

surface, cylinder head, and cylinder wall compared with those

for 60 BTDC water injection.

INTRODUCTION

Recently the water injection scheme has been used in high

performance engines equipped with a supercharger or

turbocharger. These engines entrain additional air and fuel that

is injected into the cylinder to maintain the stoichiometric ratio

and thus more power can be produced. Turbocharger also

heats the inlet air while compressing it, leading to less dense

air and thus altering the stoichiometric ratio to the rich side. In

high-pressure turbocharged engine operation during the

compression stroke, the air-fuel mixture sometimes

prematurely ignites before the spark plug ignition because of

the high inlet air temperature. This premature ignition may

cause severe engine damage. In order to avoid damages by

combustion detonation, water is injected in the engine cylinder

to cool down the inlet charge. Water can cool the inlet charge

during the compression stroke through its high latent heat of

vaporization and has an additional effect of total thermal mass

increase and thus reduces the peak combustion temperature.

However, inappropriate water injection in terms of injection

timing with respect to the fuel injection timing will locally

quench the flame, resulting in significant cycle-to-cycle

variation and substantially increase hydrocarbon emissions.

Combustion knock is a SI engine phenomenon and creates

high temperature regions in the cylinder; these regions are

mainly responsible for high NOx emissions. Water injection

has been used to inhibit NOx formation by injecting water in

the inlet manifold. Lanzafame [1] carried out experiments to

understand the effect of water injection in a single cylinder

CFR engine. With the range of water injection over the fuel

ratio from 1 to 1.25, experimental data showed a reduction in

the nitric oxide level by 50%.

Iwashiro et al. [2] suggested that direct (in-cylinder) water injection not only reduces the local temperature that causes

knock in a HCCI engine but also increases its thermal

efficiency by 2%. Moreover, it also extended the operation

range of IMEP from 460 kPa to 700 kPa, while maintaining

lower NOx levels.

The effect of water injection can also be significant during the

combustion process, where the injection of water reduces the

combustion product temperature and therefore reducing

thermal NOx formation substantially. Nicholls [3] used water-

fuel emulsion for controlling NOx emissions. According to

results observed, 10% water in gasoline caused a reduction in

nitrogen oxides by 10-20%. Conversely, 20% increase in

hydrocarbon emissions was found for water injection/fuel ratio

of 1.

Shah et al. [4] compared the effect of water injection with

EGR in terms of emission formation. As is similar with the

use of EGR, water injection provided a significant reduction in

NOx emissions but with a smaller increase in particulate

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matter (PM). At low and medium load conditions, it was

found that water injection is not as effective as EGR. When

water injection is accompanied with EGR, NOx emissions

were reduced significantly and particulate matter emissions

were also controlled. In addition, water injection increases the

charge humidity which has profound effect on NOx emission

as observed by Taxon et al. [5].

The objective of the current work is to study the effect of

different water injection timings during the compression

stroke on the inlet charge cooling. The CFD tool is used to

understand the vaporization profile and liquid film formation

over the crank angle and its effect on charge cooling and

reduction in peak temperature, while no combustion is carried

out during the simulation.

EXPERIMENTAL SETUP AND TEST

CONDITIONS

Combustion Vessel

An experimental study has been carried out in an optically

accessible constant-volume combustion vessel (CV) at

Michigan Technological University [5, 6] as shown in Figure

1. The Michigan Tech CV (MTU-CV) is a 1.0 L constant-

volume vessel which enables combustion studies including

spray, ignition, and flame investigations at a maximum

pressure of 350 bar with full-field, multi-axis, optical access.

The chamber is cubical with an interior of 100 mm per side.

On each of the six faces of the cube are ports. In three of these

ports windows were installed providing unobstructed

orthogonal optical access to the combustion chamber. Optical

windows provide access and opportunity for high-speed

imaging to study water spray development. The top face port

houses the spark plug assembly and two fans in order to create

turbulence inside the vessel as shown in Figure 1. Another

face port houses the water injector assembly. On the eight

vertices of the combustion chamber there are instrument and

actuator access ports. In four of these ports are an intake and

two exhaust ports and a dynamic pressure transducer. The

pressure transducer is a Kistler 6001 piezo-electric dynamic

pressure transducer that is coupled to a Kistler 5044a charge

amplifier to measure the CV pressure.

Injection System

A schematic of the high-pressure GDI fuel supply system is

shown in Figure 2. A high-pressure fuel system based on

gasoline engine requirements was built in-house and is

capable of pressurizing fuel to 200 bar. The fuel supply system

is a high-pressure bladder type accumulator, which on one

side is pressurized with high-pressure nitrogen and with fuel

on the other side. In order to control the fuel temperature, an

injector window with a cooling jacket was designed and built

in-house. The chiller (Iso-Temp 3016D) has a capability to

control the temperature within a range of -22C to 35C

beyond which it works as a heater. However, the current test

Figure 1: Michigan Tech combustion vessel facility.

does not use the temperature control so no cooling system is

activated for all tests. It is assumed that the fuel temperature is

the same as the injector tip temperature during the injection

event.

Figure 2: GDI fuel supply system layout.

Gasoline Direct Injector

The injector utilized in the study is a commercial Bosch 6-hole

injector. Figure 3 shows the top view of the injector, where

symmetry between the spray plumes can be observed and

provides an idea of the side view of the fuel injection when the

injector is mounted in the combustion vessel.

Figure 3: Schematic view of GDI water injection: top view

(left) and side view (right).

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Because of the symmetry among the holes, only three plumes

can be observed and visualized in imaging presented in this

study (see right in Figure 3), as the remaining three plumes are

overlapped. During the experiment, the injector is mounted

horizontally such that the camera captures the side view of the

spray shown in the right of Figure 3. The obtained images thus

show only three plumes injecting from the tip of the injector.

An injection pulse width of 3 ms was used for all test

conditions. Technical specifications of the multi-hole GDI

injector utilized in the study are listed in Table 1.

Relationship of Temperature and Pressure at

Engine Conditions

The pressure and temperature conditions corresponding to the

crank angle positions can be obtained by considering the

following assumptions:

Considering ideal gas assumption and no irreversibility

Considering isentropic process, relation obtained between

pressure, temperature and compression ratio are used to

calculate pressure and temperature values at different

degrees before-top-dead-center (BTDC).

Compression ratio at various degrees BTDC is calculated

by using maximum compression ratio concept, maximum

compression ratio considered for this experiment was

CRmax=19.

There are equations used for the calculation of T1 temperature

influenced by the remaining residual gas [8].

Q* is the enthalpy decrease during isothermal combustion per

unit mass during the combustion as defined in Eq. 1,

(

)

(

)

Eq. (1)

where mf, ma, m and QLHV are fuel flowrate, air flowrate, total

flowrate and lower heating value of the fuel, respectively.

Furthermore, specific enthalpy for a stoichiometric mixture of

isooctane and air is given below in Eq. 2, including the

residual mass fraction:

( )

Eq. (2)

The residual gas mass fraction, xr, can be determined in terms

of pressure ratio of intake and exhaust (blow-down),

compression ratio, and specific enthalpy from Eq. (2) as given

below.

( )

[ (

) ]

Eq. (3)

where CRmax is the compression ratio, pe and pi are the intake

and exhaust pressures, respectively. Note that T1 decreases

with an increase in the residual gas. The residual mass

fraction, xr, increases as pi deceases below pe, decreases as

CRmax increases, and decreases as specific enthalpy increases.

Relationship between T1 and Ti (intake temperature) is given

by Eq. (4)

( ) ( )

Eq. (4)

The residual gas temperature, Tr can be determined below

from Eq. 5

(

)( )

(

)

Eq. (5)

Upon attaining p1 and T1 from the above equations, we can

calculate the pressure, p, and temperature, T, can be calculated

at any given crank angle using isentropic relations in Eq. 6.

( )

Eq. (6)

( )

where the crank angle can be determined by Eq. 7.

( )( )

Eq. (7)

Compression ratio, CRvar, is a function of crank angle. Specific

heat equation ratio,, is assumed to be 1.33, and initial pressure pi and temperature Ti are given as 1 bar and 300 K,

respectively.

Test Conditions

In this study, four different BTDC conditions (60, 75, 90 and

180 degrees) were selected for naturally aspirated and three

boosted cases. At each crank angle, three fuel injection

Table 1: GDI injector configuration

Parameter Value

Designed bend angle 10 deg

Diameter of all orifices, D 205 m

Length of all orifices, L 300 m

L/D ratio 1.463

Maximum injection pressure 200 Bar

Spray plume angle at 20 MPa ~15 deg

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pressures (34, 68, and 102 bar) were selected along with 5

repeated tests for the naturally aspirated case (Pi = 1.01 bar)

and 4 repeated tests for the boosted case (Pi = 1.36, 1.71, and

2.02 bar). The water injection duration was set as 3 ms for all

tests. This paper will present the results for the fuel injection

pressure of 68 bar, with CV pressure and temperature

conditions determined to match crank angle positions.

The CV wall temperature was varied by electrical heaters at

the respective crank angle position condition, and pressurized

by low pressure nitrogen. Tests at temperatures less than

150C were considered as low temperature tests while high temperature tests were considered for temperatures above

150C. It was observed that ambient conditions for crank

angle positions of 75, 90, and 180 BTDC fall in the low

temperature region, while 60 BTDC falls in the high temperature region, as seen in Tables 2 and 3, respectively.

Table 2: Low temperature test matrix

Table 3: High temperature test matrix

Imaging Diagnostics

To visualize the evolution of the water spray, a conventional

schlieren imaging setup was used as shown in Figure 4. Two

schlieren mirrors, a light source, reflector, knife edge, and a

high speed camera were used for the schlieren imaging setup.

The concave schlieren mirrors were 152 mm in diameter and

had a focal length of 750 mm. A 65-W halogen lamp with a

slit size of 2 mm 5 mm was used as the light source. The light source was placed at the focal point of the first schlieren

mirror, this generated a collimated beam emerging from the

first schlieren mirror, which was directed towards the optical

combustion chamber. The collimated beam after passing

through the chamber was deflected by a 90 reflector onto the second schlieren mirror. This collimated beam from the

second schlieren mirror converges at the focal point of the

second mirror. The high speed camera (Photron FASTCAM

SA1.1) was placed at some distance from this focal point such

that the focus plane (plane passing through the injector tip) is

well captured. A knife edge is placed in a vertical position at

the focal point to block the non-refracted rays from passing

through the camera lens. The Photron camera was setup with

exposure time up to 2 s, frame rate up to 0.2 ms, and

resolution of 640 624 pixels. The camera was also equipped with a Nikon lens of focal length 50 mm.

Figure 4: Optical Setup for Schlieren.

Image Post-Processing

Matlab program was used to perform image processing on raw

schlieren images captured by the high-speed camera. First the

images were subtracted from the background image (with no

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spray). The obtained images were then filtered with 2D

median filter; this eliminated the unwanted noise from the

spray images. Then on applying simple threshold the images

were converted to binary images. Bwlabel, a predefined Matlab function, was used to find the biggest cluster of 1s in the binary image. Using Sobel and edge operator, a well-defined boundary around the biggest cluster was defined; this

spray boundary was then overlapped with the original spray

image. The horizontal distance between the injector tip and the

furthermost co-ordinate with value 1 was measured as

penetration length as suggested in SAE J 2715 [9].

Figure 5: Flowchart of post-processing schlieren image.

Raw image (far left), subtracted spray image, binary image,

and extreme eight points (far right) on overlapped spray

image.

RESULTS AND DISCUSSION

Schlieren images of transient water spray injection have been

analyzed to characterize the spray formation process and to

calculate the time-dependent spray penetration. Control

parameters include a range of ambient pressures, temperatures,

and injection pressures, while an injection pulse width of 3 ms

was kept for all test conditions.

Results from the injection pressure of 68 bar are presented

here at the condition of the ambient temperature ranging from

30 to 200C. Low temperature tests were conducted at the

temperatures less than 150C while the high temperature tests

above 150C. The ambient charge gas is nitrogen and charge pressure varies from 1 to 7 bar. Data presented here are spray

images, calculated penetration length as a function of time,

and penetration length at a fixed time after-start-of-injection

(ASOI).

Low Temperature Test Results

As stated earlier, ambient temperatures less than 150C were considered as low temperature tests. Time-dependent schlieren

images for naturally aspirated (Pi = 1.01 bar) and boosted (Pi =

1.36 bar) conditions are shown in Figures 6 and 7,

respectively.

The schlieren images presented are rotated 90 with respect to

the original raw images for comparison. Furthermore, the raw

spray was inclined slightly upwards because of a 10 designed

bend angle of the injector nozzle as defined in Table 1. As

seen in the Figure 6 and 7, all images show three distinct

plumes injected into the chamber at 0.2 ms ASOI. For low

density cases ( = 1.1 and 1.5 kg/m3), the left plume reaches the bottom of the chamber after 2.0 ms ASOI while the plume

at = 3.6 kg/m3 does not reach the bottom of the combustion vessel before 2.8 ms ASOI.

No particular trends in terms of spray structure formation are

observed for all cases regardless of charge temperature and

Figure 6: Spray images for naturally aspirated (Pi = 1.0 bar)

condition.

Figure 7: Spray images for boosted (Pi = 1.36 bar) condition.

charge density. All plumes show pointed structures at the

leading edges such that three distinct travelling plumes at the

leading tips exist. This clearly indicates that the leading edge

is in liquid-phase. In addition, for the cases with almost

similar density but different temperature, i.e., naturally

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aspirated case at = 2.7 kg/m3 seen in Figure 6 versus the

boosted case at = 2.8 kg/m3 seen in Figure 7, the penetration length is similar at a given time for both cases, showing an

insignificant effect of temperature on the penetration length.

For the cases with different density but same temperature, i.e.,

naturally aspirated case at = 2.0 kg/m3 versus the boosted

case at = 2.8 kg/m3, the penetration length is slightly shorter at a given time for the high density case. As the inlet pressure

is increased in boosted conditions up to 2.02 bar, all spray

structures are similar to the images in Figures 8 and 9 except

for a decreased penetration due to the increasing charge

density.



As observed in all Figures 6-9 corresponding to low

temperature case, the spray shows a similar unique structure as

it penetrates in charged ambient gas. Three plumes penetrating

in three distinct directions with no plume to plume interaction

is observed. It is also observed that the penetration length of

the spray decreases with an increase in ambient density, i.e.,

boosted conditions. For naturally aspirated and boosted (Pi =

1.36 bar) engine conditions, the high impingement of spray on

the walls is seen during the time of injection before 3.0 ms

ASOI for all the ambient conditions. However for boosted (Pi

= 1.71 bar) and (Pi = 2.02 bar) engine conditions, wall

impingement is seen for the first two less dense ambient cases

while no impingement on walls is observed for their densest

ambient case during injection, up to 3 ms ASOI.

Impingement on the walls leads to formation of liquid mass

depository on the walls as seen in Figure 10. A black patch

circled is the liquid wall film formed on the front glass

window of the combustion vessel due to impingement of one

of the 6 plumes. The darker the patch, larger the film thickness

is. The patch fades away with time as the film vaporizes, but

even at 11.2 ms ASOI the patch is still visible due to the high

heat of vaporization of water, which slows down its

vaporization process.

Figure 10: Spray impingement on CV walls.

Liquid Penetration

Figure 11: Spray penetration by varying ambient conditions

for the low temperature tests. The specific conditions are the

same in Figures 6-9.

The influence of ambient pressure (or density) on spray

penetration is shown in Figure 11. Similar spray penetration

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profiles are observed for all low temperature cases, with a

reduction in liquid penetration observed with an increase in

ambient pressure.

In order to make sure the spray formation was repeatable over

time; 5 runs per injection pressure for naturally aspirated case

and 4 runs per injection pressure for boosted conditions were

performed. Figure 11 shows the mean values of liquid

penetration length. Greater the slope of the curve, greater the

momentum and penetration of the liquid droplets are. Boost

conditions increase the ambient density but do not

significantly change the penetration particularly for the low

boosted conditions. Table 4 shows the penetration length

variations and corresponding percentage variations of boost

cases with respect to naturally aspirated (NA) case at 1.8 ms

ASOI. This table shows the effect of charge density on

penetration length at the similar temperature range. As

expected, the penetration length shortens with increasing

density or pressure.

Table 4: Spray penetration comparison at 1.8 ms ASOI

High Temperature Test Results

As stated earlier, the high temperature tests are those above

150C, using nitrogen charged gas. Only one ambient

condition corresponding to engine like condition at 60 BTDC was considered for each engine type (naturally

aspirated/boosted) as shown in Table 3.

Figure 12 shows a sequence of water spray images by varying

the charge density but keeping the charge temperature around

200C. A significant change in the spray pattern is observed for the high temperature tests compared to the low temperature

tests as seen in Figures 6-9.

As seen in the Figure 12, low ambient density, i.e., = 3.71 kg/m

3 illustrates a collapsed spray pattern due to the flash-

boiling effect. The effect of the flash-boiling is enhanced in

terms of the vaporization process by reducing the ambient

pressure and increasing the temperature [5]. The concept of

flash boiling is to utilize the two-phase flow of fuel injection

to enhance atomization and vaporization, reduce the fuel

penetration, and therefore readily control the air-fuel mixing

and the resulting combustion process. With an increase in the

ambient density the separation among the plumes begins

earlier in time, i.e., close to the injector tip, leading to the

formation of three distinct separate plumes as seen in Figure

12. As the water is injected and rapidly depressurized to a

pressure below its saturation vapor pressure, sudden boiling

occurs and the liquid transforms into vapor. Thus, the vapor

formed is drawn towards the central axis of the spray

overcoming the radial momentum, resulting in collapsing or

merging of spray plumes.

Figure 12: Spray developments in the naturally aspirated

and boosted conditions of Pi = 1.36, 1.71, and 2.02 bar (from

left to right colum).

In these high temperature tests only one of the cases i.e., spray

collapsed case, has ambient pressure below injected water

saturation vapor pressure (around 6 bar for these conditions),

and thus collapsed plumes are observed. However the rest of

the three cases have a similar three-plume pattern like those of

the low temperature case. As observed, the tip of the spray

plumes for all the cases appear blunt resulting from mixing of

the liquid-phase and vapor-phase, indicating some evidence of

vapor formation. Table 5 shows the time at which some signs

of vapor formation were observed in Figure 12, indicating

vaporization starts slightly earlier for the naturally aspirated

case compared to the boosted cases.

Table 5: Sign of vaporization observed

BTDC Stage Penetration (mm)Percentage

change from NA

Naturally Aspirated 78.0 0.0

Boosted, Pi = 1.36 bar 77.7 0.3

Boosted, Pi = 1.71 bar 72.4 7.2

Boosted, Pi = 2.02 bar 66.6 14.6


Boosted, Pi = 1.36 bar 77.8 3.2

Boosted, Pi = 1.71 bar 78.6 2.2

Boosted, Pi = 2.02 bar 75.6 5.9


Boosted, Pi = 1.36 bar 87.7 0.4

Boosted, Pi = 1.71 bar 87.2 0.9

Boosted, Pi = 2.02 bar 79.4 9.8

At 1.8 0 ms ASOI

75

90

180

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Liquid Penetration

Due to the sudden formation of vapor for the collapsed spray,

a relatively large spray-plume width occurs at = 3.71 kg/m3 in Figure 12. This causes a reduction in spray penetration due

to transformation of nozzle axial momentum (related to

penetration) to radial momentum (related to the plume-width),

and due to the increased effect of the vapor and air drag

forces. Thus, for the collapsed spray case (lowest ambient

density among these cases), lower penetration is observed and

compared with the denser case seen in Figure 13. However

this effect will be substantially reduced by further increasing

the ambient density as seen in the case of the density above

6.3 kg/m3, which has been reported elsewhere [10].

Figure 13: Spray penetration comparison for high

temperature tests.

NUMERICAL SIMULATION

Break-up Models used in CFD Simulation

The numerical simulation of a multi-hole injector was carried

out using CONVERGE CFD software [12]. Various spray

break-up models are available in CONVERGE CFD code such

as the Kelvin-Helmholtz (KH) and Rayleigh-Taylor (RT)

instability mechanisms, Linearized Instability Sheet

Atomization (LISA) break up model, and Taylor Analogy

Breakup (TAB) model. In addition, CONVERGE has the capability of using both KH and RT models together, called

KH-RT spray break-up model. The use of KH-RT model is a

common practice while simulating high-pressure solid cone

sprays [12].

Figure 14: Liquid core approximation [11].

According to the KH-RT break-up model, a liquid core exits

near the nozzle tip as seen in Figure 14. Liquid blobs with the

diameter of nozzle are injected from the nozzle tip and these

blobs are shed into child droplets by primary aerodynamic

instability simulated by KH alone. This aerodynamic spray

break-up by KH alone takes place up to break-up length L, after which both KH and RT instabilities are responsible for

the spray break-up. In other words, the KH break-up

mechanism acts on the spray droplets throughout their

lifetime, while RT break-up mechanism acts once the break-up

length is achieved. The break-up length [12] is defined below:

,

where B1 is KH break-up time constant, l and g are the density of fuel and ambient gas, and ro is the radius of the

nozzle. B1 can be tuned (increased or decreased) to result in an

accurate prediction of spray penetration by comparing

experimental results. While B1 value varies from 5 to 100 for

various injectors, the typical value is 7 [12].

In the liquid core region, child droplets shed from liquid blobs

are subjected to rapid acceleration and RT instability becomes

a more dominant effect. These accelerated child droplets are

subjected to a rapid deceleration due to drag force exerted by

the ambient gases; this causes unstable RT waves to build up

on the surface of droplets. For RT waves to form on droplets,

the scaled wavelength of these waves (CRTRT) should be smaller than droplet diameter. When these RT waves have

grown for sufficient time (C /RT), it causes break-up of liquid droplets according to the RT mechanism [13]. CRT, RT

size constant, can be tuned (increased or decreased) to change

the outcome of the predicted RT break-up radius. While, C,

RT time constant, can also be tuned (increased or decreased)

to delay or promote RT break-up. The break-up parameters

selected for both low and high temperature tests are shown in

Table 6.

Table 6: KH-RT break-up parameters used for the present

study

Parameter Low Temp Case High Temp Case

L 7.0 7.0

C 1.0 0.1

CRT 0.6 0.25

Turbulence model used for all the simulations was rapid

distortion Reynolds Average Navier-Stroke (RNG) k- model [14]. During the spray process, droplets may collide with

others to form more fine droplets or may coalesce to form a

single droplet. The outcomes of these droplet collisions were

simulated by No Time Counter (NTC) method [15]. Dynamic

drag model, which determines the droplet drag dynamically

and takes into account the variations in the drop shape, was

chosen [16]. CONVERGE uses the Frossling correlation to

calculate the time rate of change of droplet radius due to water

vaporization [17]. A particle-based wall film model which is

available in CONVERGE was used to model the liquid

droplets interaction with solid surfaces.

g

ol1

rBL

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CFD Simulation Geometries

Two geometries were considered in the numerical modeling,

one representing CV and the second representing engine

geometry.

CV Geometry

A CAD model, cubic in shape, is shown in Figure 15 for the

CV simulation. Maximum cell size used was 3 mm, with total

number of cells of 39,000 during the start of simulation.

Adaptive mesh refinement (AMR), tool for generating

automatic adaptive mesh refinement, was activated for spray

velocity, temperature, and spray species, i.e., water. The level

of AMR embedding for these three parameters was set to 3.

Thus, cell size of 1 mm was generated where AMR was

activated.

Figure 15: CAD model CV simulation geometry.

In-Cylinder Geometry

A CAD model of the engine as shown in Figure 16, was used

for the simulation.

Figure 16: CAD model engine simulation geometry.

Simulation crank angle timing was from -180 to 180. During

this simulation both the inlet and outlet valves were kept

closed. Maximum cell size was 3 mm, with total number of

cells of 86,700. Adaptive mesh refinement was also activated,

with a level of embedding of 3 for spray velocity, temperature

and spray species. The piston surface has a bowl shape which

is typically adopted in the GDI engine [18]. The bore diameter

and stroke are 88 mm and 98 mm, respectively, with a

compression ratio of 10.4.

CV Simulation Results

Before performing the engine simulation, the CFD spray

break-up parameters for the CV simulation were validated

against the experimental results CV tests. Two cases, one from

high temperature and other from low temperature were

selected and simulated. The spray break-up parameters were

tuned to the values listed in Table 6 for each case such that the

CFD results have a good agreement of spray shapes and liquid

penetration with the experiment. With these tuned values

determined, spray simulation was extended to the engine

simulation. The following sections will present the

comparison of spray characterizations between the CFD and

experiments including the spray shapes and penetration.

Figure 17: Spray formation comparison: = 2.0 kg/m3,

114C and droplet radius size unit as meter.

The spray profiles from the experiment and simulation in

terms of droplet size distribution are shown in Figures 17 (low

temperature) and 18 (high temperature) as a function of time,

i.e., ASOI. The diameter of the droplets exiting from the

nozzle tip (at the start of injection) is approximately the same

as the diameter of the nozzle i.e., 205 m. These droplets

break further into smaller droplets through the KH-RT break-

up mechanism with time. For the high temperature case seen

in Figure 18 six individual plumes collapse towards the

injector central axis, whereas for the low temperature case

seen in Figure 17 they appear to be penetrating away from the

injector central axis. The simulation captures the vortex

formation at the tip of each plume for the high temperature

case since these vortexes were formed by allowing the RT

mechanism to act fast enough on the droplets.

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Figure 18: Spray formation comparison: = 3.7 kg/m3,

199C, and droplet radius size unit as meter.

Since both the RT time and size constants for the high

temperature case were tuned to lower values than the low

temperature case as listed in Table 6, the simulation promotes

the RT break-up significantly in the high temperature case.

This acceleration of the RT break-up causes the RT waves to

grow more rapidly on the droplets leading to faster breakup of

water droplets. However for the low temperature case the RT

time and size constant values were kept higher, thus RT break-

up was slow compared to the high temperature case, resulting

in the formation of larger droplet size. This can be observed

from Figures 17 and 18 where for the high temperature case

the droplet radius size distribution is mostly between 1 m to

30 m, whereas for the low temperature case it varies from 30

m to 100 m. Breaking up of droplets to finer droplets

increases the drag force that acts on them, which tends to

change their momentum and thus leads to vortex formation as

observed later in Figure 18. In addition, impingement on the

walls leads to formation to finer droplet size as observed in

Figure 17 at the time of 2.8 ms ASOI. However, no

impingement on the wall was observed for the high

temperature case.

Liquid Penetration

Simulated liquid penetration is determined based on the

fraction of total liquid mass in the whole domain [12]. The

liquid penetration is defined as the distance from the nozzle

exit that encompasses a certain percentage of the total

available liquid mass injected from this nozzle. In the present

work, this percentage was set to 98%. Simulated liquid

penetration is compared with the repeated experimental runs

as shown in Figures 19 and 20.

Overall, the predicted liquid penetrations are in good

agreement with the experiments for earlier injection time but

later they tend to divert over time for both cases. Although the

CFD predicts slightly higher penetration at longer spray

duration, the model well captures spray shapes observed in the

experiments seen in Figures 17 and 18.

Figure 19: Liquid penetration comparison at low

temperature. The conditions are the same as Figure 17.

Figure 20: Liquid Penetration comparison at high

temperature. The conditions are the same as Figure 18.

Percentage Distribution

Figure 21 shows the percentage distribution of liquid and

vapor phases of water spray for both the low and high

temperature cases as a function of time. High vaporization is

observed for the high temperature case. This can be expected

as the spray droplets are smaller in size compared to the low

temperature case.

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Figure 21: Liquid and vapor percentage distribution.

Engine Simulation Results

In the engine simulation, two different cases of injection

timings were considered, i.e., 60 and 90 BTDC, in order to characterize the effect of injection timing on spray

vaporization. Since the engine with a compression ratio of

10.4, the maximum pressure and temperatures are 25.9 bar and

740 K, respectively, at the time of TDC. The engine

simulation runs from -180 to 180 with inlet and exhaust valves closed at the fixed engine speed of 2000 rpm. Figure 22

shows the water spray formation process for various crank

angles when the water injection begins at 90 BTDC. Since the injection duration is 3 ms which is the same as the CV

simulation, the injection is complete at 54 BTDC at the engine speed of 2000 rpm. A spray cloud is formed near the

piston head, while no stagnation of water is observed on the

piston surface as the piston approaches 180 ATDC. A spray impingement on the bowl of the piston head can be seen

around 70 BTDC while most of spray bounces back towards the cylinder head. Droplet sizes decrease substantially where

the temperature reaches maximum. However, there is a

discernible amount of liquid drops present when the piston

reaches at 180 ATDC. During the injection period from 90 to

36 BTDC, relatively large drops persist over the longer period since the cylinder temperature is relatively low.

A different spray formation is observed for the 60 BTDC injection as shown in Figure 23. The spray cloud is observed

near the piston region and stagnation of water inside the piston

cavity is also seen. The water droplets of 90 BTDC injection

penetrates at a higher velocity than 60 BTDC injection because of its injection into a less denser medium and also low

increment in cylinder pressure during its time of injection. In

fact, the 90 BTDC injection is observed to have high impingement on the piston surface, allowing most of the

droplets to rebound by breaking into smaller droplets by

impingement. These smaller droplets further impinge on the

piston head and lose most of its momentum by that time,

giving rise to spray cloud formation near cylinder head. In

contrast, for the 60 BTDC injection water droplets lose most of its momentum before its first impingement as it is injected

into a more dense medium and due to high increment in

cylinder pressure during its time of injection. The relative

injector-piston distance is also less compared to the 90 BTDC injection, thus most of the droplets on their impingement slide

over the piston surface leading to the formation of wall film.

Wall Wetting

In order to characterize water impingement impact on the

walls, calculation has been performed to visualize how much

water spray hangs within the cylinder and how much sticks on

the walls during the piston movement. Figures 24 and 25

visualize the portions of water moving midair in the cylinder

as colored with blue and sticking on the walls as colored with

red. The wall surface includes the piston surface, cylinder

head, and cylinder wall. Note that dense red color indicates

thicker wall film or large mass deposit on the walls.

A significant difference of wetting characteristics can be

observed. A larger portion of spray sticks on the cylinder head

and on the periphery around the piston bowl for 90 BTDC injection while a larger portion of wetting can be seen inside

the piston bowl for 60 BTDC injection. A large amount of

water sticks to the piston surface for 60 BTDC injection spray, whereas maximum amount of liquid is found stuck on

the cylinder head for 90 BTDC injection. This implies that wall wetting can be decreased by injecting earlier during the

compression stroke. Earlier injection takes advantages of more

available cylinder volume for spray atomization. For example,

injection at 90 BTDC decreases the maximum film mass by

52.8% compared to injection at 60 BTDC.

Water Spray Distribution

Figure 26 shows the vapor and liquid phase mass distribution

of water as a function of crank angle at the speed of 2000 rpm.

Red color represents the vapor phase while both green and

brown colors represent the liquid phase. In the liquid phase,

green represents liquid droplets non-adhesive to the walls

whereas brown refers to liquid droplets adhesive to the walls.

It shows that a large amount of liquid film mass is formed for

the 60 BTDC injection and most of this film mass (around 90%) resides on the piston surface. In contrast, a majority of

the film mass (around 60%) is on the cylinder head for the 90

BTDC injection. At the crank angle of 40 ATDC in the 60 BTDC injection there is a drastic fall in film mass, indicating

the film mass formed on the piston begins to vaporize rapidly.

There is gradual increase in vapor formation and gradual

decrease of total liquid phase for the 90 BTDC injection case,

but a late steep increase in vapor phase is observed for the 60 BTDC injection case. Table listed in Figure 26 shows the total

area distribution over the crank angle for both the cases.

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Figure 22: Water injection at 90 degree BTDC.

Figure 23: Water injection at 60 degree BTDC.

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Figure 24: Wall film formation at 90 degree BTDC injection. Blue color indicates droplets non-adhesive to walls while red color

indicates droplets adhesive to walls.

Figure 25: Wall film formation at 60 degree BTDC injection. Blue color indicates droplets non-adhesive to walls while red color

indicates droplets adhesive to walls.

Figure 26: Liquid and vapor phase distribution over crank angle.

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Percentage Distribution

Figure 27 presents the percentage distribution of liquid and

vapor phases of water spray as a function of crank angle. A

better vaporization profile is observed for earlier injection

such that 50% of vaporization for the 90 BTDC injection is

observed 100 degrees before the 60 BTDC injection achieves

its 50% vaporization. For the 90 BTDC injection,

impingement happens at 75 BTDC which approximately triples its vaporization rate. This increment in vaporization is

observed due to the break-up of water droplets into more finer

droplets (radius from 70 m 30 m to 30 m 1 m) as

seen in Figure 23. For the 60 BTDC injection, a drastic fall in

film mass is observed at 40 BTDC as it begins to vaporize rapidly, which can be seen by sudden increase in vapor

percentage for the 60 BTDC injection at that crank angle.

Although the 60 BTDC injection shows high vaporization

percentage at 180 BTDC, the 90 BTDC injection has a better vaporization profile for most of the crank angles as seen

in Figure 27.

Figure 27: Liquid and vapor percentage comparison for two

different injection timings as a function of crank angle.

Pressure and Temperature Traces

Figure 28 shows the effect of water injection on the cylinder

pressure. With water injection, a fall in peak pressure is

observed as water is converted to steam during the

compression stroke. This fall in peak pressure is due to the

loss of energy of intake charge, as this energy is being

absorbed to convert water to steam (heat of vaporization).

Figure 29 shows the effect of water injection on cylinder

temperature. The temperature of the charge falls down due to

heat transfer from the charge to water droplets. The fall in

peak temperature is signified for the 90 BTDC injection as it

has better vaporization compared to the 60 BTDC injection.

Table 7 shows the peak temperature for three different cases

considered.

Figure 28: Pressure traces for no water injection,

60 BTDC, and 90 BTDC.

Figure 29: Temperature trace curve for no water injection,


Table 7: Peak charge temperature for no water injection,

60 BTDC, and 90 BTDC

Case Temperature (K)

No water injection 740

90 BTDC injection 680

60 BTDC injection 690

Figure 30 shows the effect of water injection on cylinder

ambient gas density. The 90 BTDC injection has the highest peak density because of more water moving midair in the

ambient gas, while the 60 BTDC injection has less water moving midair in ambient gas and more on the wall film. As

there is a pressure loss with water injection, the compression

work required to compress the charge also decreases. The 90 BTDC injection which has better vaporization during the

compression stroke may be expected to have the lowest peak

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pressure. According to ideal gas equation, pressure is directly

proportional to density and temperature, but density and

temperature are inversely proportional to each other, thus,

pressure trend for different water injection timing cannot be

predicted but will always be expected to be below the pressure

curve for no water injection case.

Figure 30: Density trace curve for no water injection,


SUMMARY

Water spray experiments were carried out in a constant-

volume combustion vessel for various engine relevant

conditions varying from 180 to 60 BTDC simulating a naturally aspirated and three different boosted engine cases.

For the low charge temperature below 150C, all six water plumes from a multi-hole GDI injector penetrated in six

different directions away from the injector axis. However, for

the high charge temperature of 200C, all six plumes appear to

collapse towards the injector axis for one case, i.e. = 3.71 kg/m

3 due to the flash-boiling effect.

The CONVERGE CFD package was used to simulate and

validate the water spray with experiment results. The

validation was achieved by tuning KH-RT break-up

parameters to match the spray tip penetration and spray shapes

experimentally observed. With this tuned model, simulations

of water spray injection were carried out for two different

cases in the engine in-cylinder configuration; water spray

injection at 60 and 90 BTDC. The results are summarized below:

Earlier water injection at 90 BTDC during the compression stroke shows better vaporization profile

compared to at 60 BTDC. Water injection at 90 BTDC achieved 50% of water evaporation 100 degree crank

angle earlier than the injection at 60 BTDC when the engine speed is 2000 rpm.

Earlier water injection also decreases the tendency of formation of liquid wall film, which decreases the

vaporization. 28% increase in wall film mass was

predicted for injection at 60 BTDC compared to injection

at 90 BTDC.

Decrease in cylinder peak pressure and charge temperature was simulated while increase in charge

density was observed when the water is injected into the

cylinder. Approximately 8% and 6.7% reduction in peak

temperature observed for the 90 and 60 BTDC injection timings, respectively, compared to without water

injection.

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u_th_khrt_model.html

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CONTACT INFORMATION

Seong-Young Lee,

MEEM, Michigan Technological University

815 R.L. Smith Bldg.

1400 Townsend Drive,

Houghton, MI 49931, USA

Phone: 906-487-2559

Email: [email protected]

ACKNOWLEDGMENTS

Acknowledgements are given to Nostrum Energy for financial

support.

DEFINITIONS/ABBREVIATIONS

ASOI After Start Of Injection

ATDC After Top Dead Center

BTDC Before Top Dead Center

CFD Computational Fluid

Dynamics

CFR Cooperative Fuel Research

CV Combustion vessel

EGR Exhaust gas recirculation

GDI Gasoline Direction Injection

HCCI Homogeneous Charge

Compression Ignition

IMEP Indicate Mean Effective

Pressure

KH Kelvin-Helmholtz

PM Particular Matter

RT Rayleigh-Taylor

Documents

CONVERGE CFD of gasoline spray,