Indian Journal of Engineering & Materials Sciences Vol. 6; April 1999, pp. 53-58

Influence of intake port design on diesel engine air motion characteristics

M Raghu & Prarnod S Mehta •

Internal Combustion Engines Laboratory, Department of Mechanical Engineering,

Indian Institute of Technology, Chennai 600 036, India

Received 26 May /998; accepted / January /999

The intake system has a major influence on the performance of diesel engines. In this work, the performance of intake system, for both directed and helical port, is carried out based on a mean flow consideration. A quasi-steady emptying- . filling method has been used in the present analysis to estimate the instantaneous mass flow rate for evaluation of intake swirl. The model validation for both directed and helical ports is carried out using available experimental data in literature. A detailed parametric study has been presented to highlight the role of such an analysis in intake port design and evaluation for required air motion characteristics.

In a four stroke direct injection (01) diesel engine, the mixing of fuel and air takes place inside the combus­tion chamber on injection of fuel near the end of compression stroke. Consequently, the time available for proper mixing of fuel and air and their burning is quite small. For accelerated mixing, either a finer at­omization of injected fuel or increase in air motion or both these features are utilized in modem engine de­signs. The generation of air motion is dependent on design of two engine sub systems, viz., intake port and the combustion chamber. The flow entering the cylinder is significantly affected by their geometries. An optimal selection of these components provides good fuel economy and low engine exhaust emis­sions I. Intake port being an important component of engine intake system, its influence on overall per­formance of an engine is of interest to designers.

The generation of swirl during intake is a complex three dimensional and unsteady process. Detailed in­cylinder flow structure in terms of velocity measure­ments have been made using hot wire anemome~·3 or laser doppler anemometr/. Of these two tech­niques, the LOA technique is currently in wide use due to its non intrusive nature and better accuracy in mapping complex flow fields . However, several in­vestigations have reported intake port swirl meas­urements by using paddle wheelS or impulse swirl meterS-8 in a steady flow test rig. These measurements provide only the mean swirl characteristics in terms of average angular speed of air rotation about an axis

·For correspondence

or the swirl ratio. For predicting intake generated swirl, a single re­

sultant direction for the flow entering the cylinder' past the valve9

-12 is generally implied. Mehta and

Chaturvedi 12 summarized the various basis of model­ling of intake port swirl ' and validated a scheme for port swirl calculation based on mean flow considera­tions. They corroborated their predictions with meas­ured swirl levels obtained from LOA experiments·.

In the present work, an attempt is made to further refine the procedure of air motion analysis of intake port: Also the effects of various intake port geometry related parameters for directed and helical ports are studied.

Formulation The sizing of an intake port depends upon the

maximum mass flow rate required at rated condition. The estimation of the gas exchange flow rate is a pre­requisite to an air motion calculation. Instantaneous mass flow rate thus obtained from gas exchange pro­cess enables calculation of the three orthogonal ve­locity components at the valve exit during intake. A quasi-steady emptying - filling method has been used in the present analysis. The intake generated swirl is computed by summing the angular momentum accu­mulated during the intake process. This momentum is equated to the product of the moment of inertia of the cylinder contents and the angular velocity of the swirl assuming a solid body rotation of the fluid mass. The measure of intake swirl is referred in terms of the swirl at intake valve closure.

54 INDIAN 1. ENG. MATER. SCI., APRIL 1999

The conservation equations of mass and energy along with the equation of state form the basis for estimation of trapped mass and the temperature and pressure of the cylinder contents.

Considering the intake flow to be compressible in nature, the mass flow rate is written as:

m=Cd Am p, l R~, (;, l (n ~ J[l-(;') ":' J1" ... (1)

with the choked flow condition as II

P ( 2 ) 11 - ) -< - -Po - n + 1

... (2)

where Po . To are the intake manifold pressure and temperature respectively, and Am is the minimum valve flow area. The variation of instantaneous valve flow area (Am) is estimated based on a procedure sug­gested in reference (13).

The estimated instantaneous mass flow rate is an input to the energy equation, that is, the First law of thermodynamics, and the state of equation to yield the c.hanges in cylinder temperature cn and pressure (P) respectively as:

d T = _1 [nCT;ll liJill _ T", rilex

) - T(ri~1l - lnex )-(n-I)E dV] dt nl R dt c

... (3) dp RdT pdv

... (4) dt Vdt vdt

where me is the instantaneous cumulative cylinder mass. Suffixes ill and ex represent intake and exhaust conditions respectively.

Finally from continuity equation, a resultant flow velocity (uo ) at the valve exit is obtained as

111 ill U =--

o pAm ... (5)

where p is the instantaneous charge density in the cylinder.

The direction of this velocity of flow entering the cylinder is taken to be the valve seat angle (13) as shown in Fig. 1. Also, a schematic representation of velocity components of the mean velocity Uo at the valve exit is shown in the plan view of the figure. The various velocity components given in terms of seat angle (13) and helix angle (y) are:

Axial component of velocity u. = Uo sin fJ ... (6)

Radial component of velocity ur= Uo cos fJ cos r (7)

Tangential component of velocity

u, = Uo cos fJ sin r (8)

Assuming that the axis of the valve and the cylin­der are paraIIel to each other, the law of conservation of angular momentum is written as

Rate of change of Rate of change of moment of angular momentum = momentum of the charge with of the cylinder respect to the cylinder axis contents

~[JO)] = mU o cosl3 dt

x[ecosy sin~ - e siny cos~ + Rv siny]

.. . (9) where ~ is port orientation angle. The three terms on RHS of the Eq. (9) represents the radial, tangential and pre swirl components respectively.

The above analysis requires input concerning cam and intake port. These input include the type of port, air velocity and its distribution at the valve exit, entry angle of the intake air, and the port eccentricity and orientation.

Results and Discussion The results of this study are presented in terms of a

non- dimensional swirl ratio defined as the ratio of angular speed of the charge to the angular speed of the engine.

Fig. I-Schematic representation of an offset intake valve geometry

RAGHU & MEHTA: INFLUENCE OF INTAKE PORT DESIGN ON DIESEL ENGINES 55

Model validation The model calculatioris for both directed and heli­

cal port designs for three engine speeds, viz., 960, 1440 and 2400 rpm are validated using the experi­mental data of Monaghan and Petti fer . These data were obtained on a direct injection diesel engine (bore: 0.1206 m, stroke: 0.1397 m, rated speed: 2400 rpm) having compression ratio of 15.9. In absence of a knowledge of the cam profile, a simple sinusoidal shape with a dwell period of 30 crank angle degrees has been assumed throughout this study.

The values of the other parameters concerning the intake ports, viz. valve diameter, seat angle, port ec­centricity and orientation are taken as 0.045m, 45°, 0.027m and 38° respectively. These values are in their recommended ranges given by Heywoodl3 and AI­coumanis et al. 14. For the helical port, ·a helix angle of 20° is chosen. An initial comparison of the model predictions with the experimental results at the three. engine speeds is shown in Table 1.

Despite the fact that the numerical values in certain cases are coincident, there appears to be no consistent

Table I--Comparison of predicted and experimental! swirl ratios at intake valve closure

Port

Directed

Helical

Engine speed (RPM)

960 1440 2400

960 1440 2400

Swirl ratio at intake valve closure

Experimental Predicted

2.6 3.15 3.6 3.02 2.0 \.99 3.4 4.61 3.5 4.41 3.3 2.91

correlation between predicted and experimental trends with speeds. This is attributed to the fact that the several model input values had to be assumed. To clarify this point a study on m"del sensitivity to vari­ous input variables is done and found that the model predictions are sensitive to the port orientation angle and which would possibly alter with engine speed due to change in the resultant flow field vector. On this premise, the appropriate, port orientation angles are chosen to make comparison between model predic­tions and experimental results and the corresponding results are shown in Table 2. These results suggest a very good correlation between the predicted and ex­perimental values for both directed and helical port designs. This particular model validation exercise also prove the model capability for a port design.

Parametric study Directed port

The influence of various parameters, considered to assess the predictability of the model, are discussed below. While changing a parameter, the values of the other p~rameters are kept at their reference base value as shown in Table 3.

The influence of valve lift on the intake swirl ratio at different engine speeds is shown in Fig. 2. It can be seen that for a given engine speed, increase in maxi­mum valve lift first results in an increase in swirl ra­tio and then decreases with further increase in the maximum valve lift. This particular trend has been found with other engine speeds as well. However, the maximum valve lift for the peak swirl ratio is varying at different speeds. While at lower engine speed, i.e., 960 rpm, the swirl ratio peaks for low maXImum

Table 2--Comparison of predicted and experimentaJl swirl ratios at intake valve closure

Port Engine speed (RPM) Port orientation angle (~) Helix angle (y) Swirl ratio at intake valve closure Experimental Predicted

960 30 2.6 2.56 Directed 1440 38 3.6 3.02

2400 38 2.0 \.99 960 22 20 3.4 3.38

Helical 1440 25 20 3.5 3.47 2400 45 20 3.3 3.21

Table 3--Reference base values of various parameters

I. d. w e p ~ y

0.0131 m 0.045 m 0.002475 m 0.027 m 45° 38° 20°

56 INDIAN J. ENG. MATER. SCI., APRIL 1999

e,---------------------~ DlncW port --- NO RPM

---- 1440 RPM - - 2400 RPM

~ ~",,, .,. - - . , , , -'

O~TTTrrn~"~TTTrrrrn~ 0.008 0.011 0.014 0 .017 0.020

Maximum valve lift. m

Fig. 2-Variation of intake swirl ratio with maximum valve lift

e~--------------------_, DlrecW port --- IICJO RPM

---- 1«0 RPM - - 2400 RPM

---

O~_rrT~~rr~"_.rrrT~

0.040 0.045 0.050 0.055 0.060 Intake valve diameter. m

Fig. 3-Variation of intake swirl ratio with intake valve diameter

valve lift of 0.00965 m, the swirl ratio at higher en­gine speeds peaks at higher maximum valve lifts of 0.012 m for 1440 rpm and 0.0157 m for 2400 rpm. This observation is in conformity with the results re­ported by Assanis and Polishak '5 . A change in swirl ratio values with maximum valve lift or engine speed is dependent on the total momentum inflow into the cylinder. This momentum inflow is the product of mass flow rate and the flow velocity. For valve lifts other than the optimum, that is for maximum swirl, the total momentum inflow is lower because of the reduced mass flow rate and inflow velocity. With in­crease in engine speeds, the swirl ratio has been found to be lower at low maximum valve lifts due to expected reduction in the charging rate.

As the intake valve diameter increases the swirl ratio starts decreasing as shown in the Fig. 3. This is because for the same mass flow rate, an increase in valve diameter increases the minimum valve flow area hence reducing the inflow velocity and the total momentum entering into the cylinder.

e~--------------------_, Directed port -- 1100 RPM

---- IUO RPM - - 2400 RPM

1

o ~ " " I' i " I " " I " i 'I " iii" " I 'Tn 25 30 35 40 45 50 55 60 65

VlJlve seat angle. deg·

Fig. 4--Variation of intake swirl ratio with valve seat angle

e,---------------------~ Dtnote4 port -- IHIO RPM

---- 1«0 RPM - - IUOO RPM

- -' ------

o . 0.024 0.026 0.028 0.030 0.032 0.034

Port eccentricity. m Fig. 5--Variation of intake SWITI ratio with port eccentricity

The effect of valve seat angle on swirl ratio is shown in Fig. 4. It is assumed that the intake flow enters the cylinder with an inclination equal to the valve seat angle. Hence, with increase in valve seat angle, there is reduction in tl}e radial component of velocity at the valve periphery for a given mean flow velocity. This results in a reduction of swirl ratio with increase in valve seat angle as seen in the figure. This finding is in conformity with the steady flow test re­sults of Stone and Ladommatos8

•

The change in swirl ratio due to change in valve seat width is found to be marginal as the valve seat width produces a very insignificant change in the minimum flow area during the intake process.

With increase in port eccentricity, the swirl ratio increases as shown in Fig. 5. The angular momentum generated by a directed port is directly proportional to the port eccentricity for a given mass f10w rate and velocity. This increases the inflow charge momentum about the cylinder axi s. Hence, an increase in port eccentricity tends to increase the intake swirl signifi-

RAGHU & MEHTA: INFLUENCE OF INTAKE PORT DESIGN ON DIESEL ENGINES 57

cantly at all speeds. Though higher port eccentricity values are advantageous to produce higher swirl, there is a limitation imposed by the space constraint on the cylinder head. Port eccentricity should be less than the difference between the cybnder and valve radii.

The port orientation angle is the entry angle of the inflow with respect to the line joining the valve and cylinder axis and it does.not influence the mass flow rate. The influence of the port orientation angle on swirl ratio is shown in Fig. 6. With increase in port orientation angle, the swirl ratio increases with the maximum value occurring at port orientation angle of 90° as expected. Further increase of the port orienta­tion angle beyond 90° reduces the swirl ratio, with peak swirl ratio and symmetricity about <I> equals 90°. This is because the turning moment distance of the inflow about the cylinder axis (e sin <I> ) has the lowest

7,----------------------, DIncW port -- NO RPM ---- lUG RPM - - lUOO RPM

O~TTTr~~~TT~~~~,M

o 30 do gO 1 0 160 Port orientation angle. deg

Fig. 6--Variation of intake swirl ratio with port orientation angle

value of zero for the zero port orientation angle and has a maximum at the port orientation angle of 90°.

Helical ports Fig. 7 shows the effect of maximum valve lift on

swirl ratio from a helical port. The swirl ratio first

e.-----------,

, , ,

/ /

/

/ /

/

,­/

Belleal port

-- NO RPM ---- lUG RPM - - 2.00 RPM

O~Trnn~Trnn~Tn~Trrl 0.008 0.011 0.01" 0.017 0.020

Maximum valve lift. m Fig. 7-Variation of intake swirl ratio with maximum valve lift

8~----------,

-- IlISO RPM ---- lUG RPM - - UOO RPM

04TnT~~TnTnTn~nTnT~ o 30 'eo gO 120 150

Helix angle. deg

Fig. 8--Variation of intake swirl ratio with helix angle of helical port

Table 4--Summary of the influence of intake port and valve parameters on intake swirl at engine speed of 2400 RPM

Parameter Range Swirl ratio at IVC Directed port Helical port

Lift (m) 0.00965 - 0.0181 Increases (143.9%) then decreases Increases (102.4 %) then (9.86%) decreases (9 .28%)

Valve diameter (m) 0.043 - 9.055 Decreases (27.61 %) Decreases (17.65%)

Valve seat angle (deg) 30 - 60 Decreases (107.4%) Decreases (106 .1 %)

Valve seat width (m) 0.002 - 0.003 Increases (0.3%) Increases (0.7%)

Port eccentricity (m) 0.026 - 0.032 Increases (22.4%) Increases (15 .9%)

Port orientation angle (deg) 20 - 90 Increases (191 %) Increases (104.4%)

Helix angle (deg) (For helical port) 10 - 90 Increases (90.24%)

58 INDIAN J. ENG. MATER. SCI., APRIL 1999

increases up to an optimum depending on the engine speed and maximum valve lift conditions and then again starts decreasing. The magnitude of swirl in case of helical ports for any parameter is higher com­pared to the corresponding level in a directed port. Similar observation has been made by Pozniak and Rydzewski '6 in their experimental study.

Analysis of the influence of various other helical port parameters on swirl ratio showed that a trend in its variation is more or less similar to that found with a directed port except that the helical ports have higher magnitudes due to the additional angular mo­mentum generated about the port axis prior to entry into the cylinder.

Fig. 8 shows the effect of helix angle on swirl ratio for a helical port. The swirl ratio is increasing with the helix angle mainly because of an expected in­crease in preswirl component of the swirl.

Table 4 summarises the influence of various port and valve parameters on intake swirl ratio for di­rected and helical ports at an engine speed of 2400 rpm.

Conclusions The model for the prediction of intake swirl ratio

has been validated for both directed and helical port successfully. The model response for all the port and valve parameters have been carried out and has been found to be satisfactory. The main parameters whick significantly influence the intake swirl ratio are valve lift, valve diameter, seat angle, port eccentricity, port orientation angle, cam dwell and the helix angle in case of a helical port." The effect of valve seat width is insignificant.

AckilOwledgements One of the authors (PSM) wishes to acknowledge

the financial support received from the Department of Science and Technology, Government of India for this work.

Nomenclature Dv valve diameter, m

Dm mean seat diameter, m Dp port diameter, m D, valve stem diameter, m e port eccentricity, m Lv instantaneQus valve lift, m

m intake mass flow rate, kg/s n ratio of specific heats p cylinder pressure, Pa R specific gas constant, J/kg it( Rv intake valve radius, m T cylinder temperature, K uo resultant mean flow velocity, mls u. axial component of velocity, mls

u, radial component ofvelociIty, mls u, tangential component of velocity, mls

V cylinder volume, mJ

w valve seat width, m

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