TECHNICAL PAPER

Influence of intake pipe length and diameter on the performanceof a spark ignition engine

Rodrigo Caetano Costa • Sergio de Morais Hanriot •

Jose Ricardo Sodre

Received: 5 July 2010 / Accepted: 29 October 2010 / Published online: 30 August 2013

� The Brazilian Society of Mechanical Sciences and Engineering 2013

Abstract This work shows the influence of intake pipe

length and diameter on the performance of a spark ignition

engine. A production four-stroke, four-cylinder, eight-

valve, 1.0-l engine was tested in a bench dynamometer,

fuelled by a blend of 78 % gasoline and 22 % ethanol.

Experiments were carried out in the engine speed range

from 1,500 to 6,500 rev/min. Three intake pipe lengths—

0.3, 0.6 and 0.9 m—and three intake pipe diameters—

0.044, 0.053 and 0.067 m—were investigated. The effects

of intake pipe geometry on intake air mass flow rate and

volumetric efficiency and the impacts on engine perfor-

mance parameters—torque, power, thermal efficiency and

specific fuel consumption—were evaluated. The results

revealed that, for low engine speeds, the intake pipe with

longer length and smaller diameter produced the best per-

formance. On the other hand, the intake pipe with shorter

length and larger diameter delivered the best engine per-

formance at high speeds.

Keywords Internal combustion engine � Intake

system � Performance � Volumetric efficiency � Fuel

consumption

List of symbols

D Diameter (m)

L Length (m)

SFC Specific fuel consumption (kg/kW h)

1 Introduction

The intake system influences air charge into the engine

cylinders, having a direct effect on volumetric efficiency

and output power. The primary function of the intake

system is to improve engine breathing capacity by keeping

pressure losses to a minimum. Intake systems are designed

to produce uniform air and fuel distribution between the

engine cylinders. Thus, optimization of the intake system is

essential to obtain maximum engine performance. The

volumetric efficiency can also be increased though opti-

mization of the intake valve opening, which normally starts

before top dead center (BTDC).

The engine volumetric efficiency depends on the iner-

tial, transient flow fields that occur in the intake system,

especially to take advantage of ram and tuning effects. The

intake system geometry and dimension are configured to

attend the project goals, such as high torque or high power

with low specific fuel consumption. The kinetic energy of

the air mass flowing into the cylinder is affected by the

intake pipe length and cross-sectional area. A previous

study by [3] shows that, for low engine speeds, the conduit

with shorter length allows for higher intake air charge to

the engine. A similar effect is noticed when a conduit of

Technical Editor: Demetrio Neto.

R. C. Costa

Department of Mechanical Engineering, Federal Institute

of Minas Gerais, Bambui, MG 38900-000, Brazil

e-mail: rodrigocaetano@cefetbambui.edu.br

S. M. Hanriot � J. R. Sodre (&)

Department of Mechanical Engineering, Pontifical Catholic

University of Minas Gerais, Belo Horizonte,

MG 30525-901, Brazil

e-mail: ricardo@pucminas.br

S. M. Hanriot

e-mail: hanriot@pucminas.br

123

J Braz. Soc. Mech. Sci. Eng. (2014) 36:29–35

DOI 10.1007/s40430-013-0074-2

small cross-sectional area is adopted, producing an inertial

effect even stronger. Intake systems with intake pipe of

variable length are employed especially to improve specific

fuel consumption and drivability, torque at low speed, and

power at high speed. Pipe length is varied by steps

according to engine speed to increase volumetric

efficiency.

Proper engine design must consider combustion cham-

ber geometry, materials, valvetrain, components such as

water pump and oil pump, and exhaust and intake systems.

This work investigates the influence of intake pipe length

and diameter on the performance of a spark ignition

engine. Experiments were conducted in a production four-

stroke, four-cylinder, 1.0-l engine to analyze intake air

mass flow rate, volumetric efficiency, torque, power, spe-

cific fuel consumption (SFC) and thermal efficiency. The

engine was fueled by a blend of 78 % gasoline and 22 %

ethanol (E22). Three intake pipe lengths and three intake

pipe diameters were used in the investigation.

2 Literature review

Engine performance depends on the air quantity admitted

per cycle, which defines the fuel amount to be burned. The

volumetric efficiency is determined by the air mass amount

admitted by the engine, which varies with atmospheric

condition, engine speed, displaced volume, intake and

exhaust pipe geometry, and flow resistance. Intake flow

resistance depends on intake pipe geometry and dimen-

sions, and increases with flow velocity. The air flow speed

is influenced by the intake pipe cross-sectional area and the

free flow area across the intake valves; thus, the intake

conduit is dimensioned for maximum air flow [5].

In order to optimize the air charge into the cylinder, the

intake valve is designed to open just before the piston

reaches the top dead center during the exhaust stroke. If the

intake valve opens too early and the intake pressure is

lower than the cylinder pressure, part of the burned gas can

be drawn through the intake system instead of the exhaust

system. A retard to close the intake valve after the bottom

dead center (ABDC) is applied to use the intake flow

inertia to increase the air charge and, thus, the volumetric

efficiency, known as RAM effect [2]. The RAM effect

depends on the intake pipe length and diameter. For longer

intake pipe lengths the inertial effects are stronger, as the

mass amount inside the pipe is larger, and the flow pressure

drop is higher. The inertial effects can be used to improve

the volumetric efficiency at low engine speeds [6].

Heisler [6] showed results from a 3.5-l, six-cylinder

spark ignition engine to demonstrate the influence of intake

pipe dimension on volumetric efficiency. Experiments with

pipes of variable length and diameter were described. For a

given tract length, the smaller the tract diameter the higher

will be the intake velocity and the lower will be the engine

speed at which the RAM charge pressure peaks. It was also

shown that, for a given tract diameter, a long passage

produces a much higher peak volumetric efficiency at a

relatively lower engine speed, whereas a short passage

produces a much lower peak volumetric efficiency at a

higher engine speed.

Benajes et al. [1] analyzed the intake system considering

it composed by two sub-systems. The first sub-system

consisted of pistons and valves which, due to periodical

movements act like excitation sources. The second sub-

system consisted of the intake pipe, which reacts to the

excitation source according to its geometry. This interac-

tion influences the transient flow condition at the valve port

and, consequently, affects the entire intake process.

Nowakowski and Sobieszczanski [8] developed a

numerical model to study the influence of the intake pipe

length on engine operating parameters. Experiments were

carried out in a four-stroke, two-cylinder engine tested in a

bench dynamometer to validate the model. An electronic

multi-point fuel injection system was adapted to the engine

to optimize fuel/air mixture equivalence ratio and ignition

timing. Intake pipes of 28 mm diameter and 220, 340, 480,

570, and 800 mm length were tested in the range from

1,500 to 5,500 rev/min, in 250 rev/min steps. The results

showed that the maximum torque was obtained for the pipe

of 800 mm length, at the speed of 2,750 rev/min. Never-

theless, for this pipe length and speeds above 3,750 rev/

min, torque values were lower than those obtained with

shorter pipes. The shorter the pipe length, the lower the

maximum torque and the higher the speed at which it

occurs.

An optimum design of a synthesis exhaust manifold has

been proposed and modeled by Galindo et al. [4], with the

aim to improve transient engine performance. A dual

walled air gap exhaust manifold and a pulse exhaust

manifold have been used in the analysis. The dual walled

exhaust manifold was used to analyze the thermal inertia

effects. The pulse exhaust manifold generates a directional

flow effect, increasing flow kinetic energy and avoiding

interference between the exhaust pressure pulses from

consecutive firing order cylinders. The use of the low

thermal inertia dual walled intake manifold positively

influenced the intake conditions, increasing boost pressure,

intake air mass flow rate, and brake torque during transient

engine operation, in comparison with the original manifold.

The use of the pulse exhaust manifold improved engine

volumetric efficiency under dynamic conditions, compared

to the original configuration.

Shannak et al. [12] investigated the influence of intake

pipe diameter on engine exhaust emissions. Tests were

conducted in a four-stroke, four-cylinder, gasoline-fuelled

30 J Braz. Soc. Mech. Sci. Eng. (2014) 36:29–35

123

engine with intake pipe diameters of 20, 25, 30, 35, 40 and

63 mm in the engine speed range from 1,000 to 4,000 rev/

min. The results showed that increasing pipe diameter

decreases unburned hydrocarbons (HC) and carbon mon-

oxide (CO) emissions. The decreasing rate of HC and CO

emissions were reduced with increasing pipe diameters.

Varying intake pipe diameter did not affect carbon dioxide

(CO2) emission and exhaust oxygen (O2) concentration.

The authors concluded that increased intake pipe diameter

can be a useful strategy for emissions control.

Sekavcnik et al. [11] performed a series of numerical

simulations for a single, straight circular pipe. The varied

parameters were pipe length, mass flow amplitude and

duration of a transient disturbance at the pipe inlet. The

impulse disturbance at the pipe inlet was predefined as a

time-dependant boundary condition, which was simulated

by a rapid change of mass flow rate. Pipe length was shown

to be the most influential parameter for dynamic response

frequency. A good agreement was found for the linear

relationship between pipe length and sound speed to ei-

genfrequency ratio.

A design procedure of a cheap intake system adapted for

a race car engine has been presented by Pogorevc and Kegl

[10]. Two different geometries of the intake manifold were

taken into consideration, with the support of numerical

simulation. The simulated results showed that intake

manifold geometry influence mass flow rate distribution

among the engine cylinders. Moreover, an average differ-

ence of 4.3 % in the total intake air mass flow rate was

observed between the intake manifolds. Model and

experiments showed a good agreement for intake air mass

flow rate and volumetric efficiency variation with engine

speed.

Olczyk [9] investigated the specific mass flow rate dis-

tribution in pipes supplied with a pulsating flow. Special

attention was paid to the dynamic phenomena related to the

resonance occurring in a pipe for characteristic frequencies

depending on the pipe length. The test rig was composed

by a turbocharger, an air heating system in the turbine inlet,

a pulse generator and a set of valves to control the mass

flow rate. Two intake pipe lengths were tested: 0.544 and

1.246 m. Both pipes had a diameter of 0.042 m. The results

showed intensification and retreat of a reverse flow in the

neighborhood of the resonance frequency, which was

determined to be equal to 140 Hz for the short pipe. After

changing the pipe length, the resonance frequency was

displaced. For the long pipe, the resonance frequency was

placed in the neighborhood of 60 Hz. For both pipes a

divergence of the analyzed mass flow rates of up to 40 %

for the resonance frequency and a rapid decrease of the

mass flow rate at the inlet section were observed.

An investigation on the strategy to obtain the same port

flow distribution from a manifold design has been

presented by Tong et al. [13]. A two-dimensional array of

ten parallel channels that interconnect a distribution man-

ifold with a collecting manifold was used in the analysis.

Eight proposed strategies were analyzed by numerical

simulation. The most promising strategy, which showed the

best flow distribution performance, was enlargement of the

cross-sectional area of the distribution manifold.

3 Experimental methodology

A production 1.0-l, eight-valve, four-cylinder spark igni-

tion engine was used in the tests, which specifications are

described in Table 1. Engine development electronic con-

trol unit and interface software were employed, allowing

for the optimization of fuel/air mixture equivalence ratio

and ignition timing. For any speed tested the ignition

timing (Figs. 1, 2) was optimized by the minimum advance

for best torque (MBT) of by the detonation limit, which-

ever occurred first. Detonation was identified through an

accelerometer installed in the engine block. The exhaust

oxygen (O2) concentration was measured through a linear

lambda sensor located before the catalytic converter,

allowing for determination of fuel/air mixture equivalence

ratio. Fuel/air mixture equivalence ratio was optimized by

injecting the minimum fuel amount to produce the best

torque at each speed condition, remaining at 1.11 ± 0.01.

Experiments were carried out with the engine mounted

in an eddy current dynamometer, of maximum torque

235 N m and maximum power 110 kW. The dynamometer

was equipped with an accelerator automatic actuator, fuel

Table 1 Baseline engine conditions

Parameter Specification

Number of cylinders 4, in line

Bore 9 stroke (mm) 70.0 9 64.9

Displaced volume (cm3) 999.057

Compression ratio 12.00:1

Intake valve diameter (mm) 31.0

Exhaust valve diameter (mm) 26.5

Valve lift (mm) 9.0

Intake valve opens (�BTDC)a 2

Intake valve closes (�ATDC)b 221

Exhaust valve opens (�BTDC) 222

Exhaust valve closes (�BTDC) 1

Idle speed (rev/min) 850

Fuel/air mixture equivalence ratio 1.11 ± 0.01

Fuel E22

Lubricant SAE 15W40

a Before top dead centerb After top dead center

J Braz. Soc. Mech. Sci. Eng. (2014) 36:29–35 31

123

consumption meter, temperature sensors, pressure sensors

and relative humidity sensor. The dynamometer control

system allowed for control of coolant, lubricant and intake

air temperatures and ambient humidity, according to test

specifications. The dynamometer load cell featured reading

range 0 to 981 N, resolution 0.98 N and maximum error

0.5 % of reading. The rotational speed of the shaft con-

necting the dynamometer to the engine was measured by an

inductive magnetic sensor.

Platinum resistance PT-100 thermometers of reading

range from -200 to ?800 �C, resolution 0.1 �C and

maximum error 0.35 % of reading were used to measure

the engine inlet and outlet coolant water, lubricating oil,

and intake air temperatures. K-type thermocouples with

reading range from 0 to 1260 �C, resolution 0.1 �C and

maximum error 0.75 % of reading were employed to

monitor the exhaust gas temperature. Pressure transducers

were located at several points along the intake system. Fuel

consumption was continuously determined each 10 s dur-

ing the tests through a dynamic measurement electronic

scale, of 150 kg/h capacity, resolution of 0.001 g, and

maximum error of 0.12 %.

The engine was mounted in the dynamometer in a

similar angular position as it is mounted in a car chassis,

and was adequately aligned and leveled to avoid excessive

vibration and measurement error. Nine different intake

pipes were used, which diameter and length are shown by

Table 2. Prior to the experiments, the engine was tuned

according to FIAT 7-A6000 standard.

The experiments were carried out following NBR ISO

1585 standard, at wide open throttle. Data were acquired at

least 1 min after the steady-state condition was reached,

from 6,500 to 1,500 rev/min, in steps of 250 rev/min.

During the tests the engine outlet cooling water was kept at

82 ± 2 �C, lubricant temperature was over 100 �C, intake

air relative humidity was between 48 and 52 %, intake air

temperature was 20 ± 2 �C, ambient pressure was

0.910 bar, and fuel pressure was 3.50 ± 0.02 bar. To dis-

play the output power at the reference atmospheric condi-

tion, the measured power was multiplied by a correction

factor of 1.12, following recommendation from NBR ISO

1585 standard. Table 3 shows the total uncertainties asso-

ciated to the performance parameters, calculated according

to the methodology by Kline and McClintock [7]. The

results presented in the following section are the average of

three tests performed at each condition.

1000 2000 3000 4000 5000 6000 7000

SPEED (rev/min)

16.0

18.0

20.0

22.0

24.0

26.0

28.0IG

NIT

ION

TIM

ING

(°B

TD

C)

L= 0.300 mL= 0.600 mL= 0.900 m

Fig. 1 Ignition timing variation with intake pipe length and engine

speed

1000 2000 3000 4000 5000 6000 7000

SPEED (rev/min)

16.0

18.0

20.0

22.0

24.0

26.0

28.0

IGN

ITIO

NT

IMIN

G(°

BT

DC

)

D= 44 mmD= 53 mmD= 67 mm

Fig. 2 Ignition timing variation with intake pipe diameter and engine

speed

Table 2 Total uncertainties associated to engine performance

parameters

Parameter Uncertainty

Intake air mass flow rate (kg/h) 1.3

Volumetric efficiency (%) 0.9

Torque (N m) ±0.95

Power (kW) ±0.72

Specific fuel consumption (kg/kW h) ±0.006

Thermal efficiency (%) 0.7

Table 3 Intake pipe dimensions

Pipe # Diameter (mm) Length (mm)

1 44 600

2 53 300

3 53 600

4 53 900

5 67 600

32 J Braz. Soc. Mech. Sci. Eng. (2014) 36:29–35

123

4 Results and discussion

Figures 3, 4, 5 show the results from the tests with varying

intake pipe length and fixed diameter of 0.053 m (pipes 2,

3 and 4 in Table 3). The volumetric efficiency is given by

the ratio between the actual intake air mass flow rate (both

shown by Fig. 3) and the air mass charge into the cylinder

at a reference condition (25 �C, 1.013 bar). Thus, the

volumetric efficiency is directly influenced by the intake air

mass flow rate, although the differences caused by intake

pipe length are more easily noticed for volumetric effi-

ciency. Predominantly, for high engine speeds the smaller

pipe, of 0.300 m, produces the highest volumetric effi-

ciency levels, while in the low engine speed region the

longer pipe, of 0.800 m, delivers higher volumetric effi-

ciency levels. Peak volumetric efficiency was higher for the

longer pipe, which is in agreement with the results pre-

sented by Heisler [6]. However, the peak volumetricefficiency for all pipes happened at the same engine speed,

around 4,500 rev/min, probably influenced by the ignition

timing map (Fig. 1).

The intake air mass flow rate and the volumetric effi-

ciency curves (Fig. 3) resemble the power and torque

curves (Fig. 4), respectively. For speeds over 3,250 rev/

min, shorter pipe length presented higher torque and

power (Fig. 4). Below that speed the trend was inverted,

with longer pipe length producing higher torque, while

output power remained practically the same. These results

agree with those presented by Heisler [6] and by Nowa-

kowski and Sobieszczanski [8]. The high flow contact

area of long pipe lengths together with high speeds gen-

erate high flow resistance, thus decreasing the mass

amount into the cylinder and, consequently, the torque

and power developed. On the other hand, for low speeds

the flow resistance is reduced and the higher inertial

effect in the longer pipe increases the air mass into the

cylinder.

The specific fuel consumption (SFC) is given by the fuel

mass flow rate per unit power produced, while the thermal

efficiency is given by the ratio between the output power

and the total injected fuel energy content. That is why SFC

and engine thermal efficiency about opposite trends for the

three pipe lengths tested, as shown by Fig. 5. The SFC and

thermal efficiency do not show a clear behavior under the

influence of pipe length throughout the whole engine speed

range studied. That is a consequence of the optimizations

done on the fuel/air equivalence ratio and ignition timing

(Fig. 1), as cited in the previous section. The minimum

SFC and the maximum engine thermal efficiency are

observed at around 3,000 rev/min for any intake pipe

length.

The influence of intake pipe diameter on engine per-

formance was verified using three pipes of different

diameters and same length, 0.600 m (pipes 1, 3 and 5 in

1000 2000 3000 4000 5000 6000 7000

SPEED (rev/min)

30

60

90

120

150

180

210

240

270

INT

AK

EA

IRM

AS

SF

LOW

RA

TE

(kg/

h)

40

45

50

55

60

65

70

75

80

VO

LUM

ET

RIC

EF

FIC

IEN

CY

( %)

L= 0.300 mL= 0.600 mL= 0.900 m

Fig. 3 Intake air mass flow rate and volumetric efficiency variation

with intake pipe length and engine speed

1000 2000 3000 4000 5000 6000 7000

SPEED (rev/min)

10

20

30

40

50

60

70

80

PO

WE

R(k

W)

30

40

50

60

70

80

90

100

TO

RQ

UE

(N.m

)

L= 0.300 mL= 0.600 mL= 0.900 m

Fig. 4 Power and torque variation with intake pipe length and engine

speed

1000 2000 3000 4000 5000 6000 7000

SPEED (rev/min)

0.230

0.240

0.250

0.260

0.270

0.280

0.290

0.300

0.310

SF

C(k

g/kW

h)

30.0

32.0

34.0

36.0

38.0

40.0

42.0

44.0

46.0T

HE

RM

AL

EF

FI C

IEN

CY

(%)

L= 0.300 mL= 0.600 mL= 0.900 m

Fig. 5 Specific fuel consumption and thermal efficiency variation

with intake pipe length and engine speed

J Braz. Soc. Mech. Sci. Eng. (2014) 36:29–35 33

123

Table 3), as shown by Figs. 6, 7, 8. The effects of intake

pipe diameter are higher on engine volumetric efficiency

than on the intake air mass flow rate (Fig. 6). For speeds

above 3,000 rev/min the intake pipe of larger diameter

generally produces the highest volumetric efficiencies,

while, for lower speeds, the pipe of larger diameter pro-

duces the lowest volumetric efficiencies. Small pipe

diameter generates faster air flow speed, thus improving the

inertial effects and the intake air charge. On the other hand,

at high engine speeds, the high pressure drop effect of

small pipe diameter overcomes the inertial effect. In this

case, the adoption of a large pipe diameter decreases

pressure drop, therefore increasing the volumetric effi-

ciency. The use of large pipe diameter is also expected to

reduce exhaust CO and HC emissions, as reported by

Shannak et al. [12]. Peak volumetric efficiency happened at

lower speeds with increasing pipe diameter, as observed by

Heisler [6].

Figure 7 shows that the effects of pipe diameter on

engine torque and output power seems to be lower than

those presented by pipe length variation (see Fig. 4). It can

hardly be noticed that the larger intake pipe diameter

produced slightly higher torque and power in the high

speed region, showing an inverted trend at low speed. The

torque and power curves produced by pipe diameter and

engine speed variation are also resembled by those of

volumetric efficiency and intake air mass flow rate,

respectively (Fig. 6).

Figure 8 shows that there is not a clear influence of

intake pipe diameter on SFC and thermal efficiency, as it

has previously been noticed for pipe length variation

(Fig. 5). This result is also attributed to optimization of the

fuel/air ratio and the ignition timing (Fig. 2), as explained

before. Minimum SFC and maximum thermal efficiency

were observed between 2,500 and 3,500 rev/min, depend-

ing on pipe diameter.

5 Conclusions

Intake pipe length and diameter can influence engine per-

formance. Long intake pipe produces higher volumetric

efficiency, torque, and power at low engine speeds, while,

for high engine speeds, short pipe length produces higher

engine performance. For the engine tested, the minimum

SFC and the maximum thermal efficiency were noticed

around 3,000 rev/min. Large pipe diameter produce higher

volumetric efficiency at high engine speeds and lower

volumetric efficiency at low speeds, but its effect on engine

torque and output power are hardly noticed. Intake pipe

length and diameter does not show a clear influence on

specific fuel consumption and on engine thermal efficiency.

Overall, intake pipes with long length and small diameter

1000 2000 3000 4000 5000 6000 7000

SPEED (rev/min)

30

60

90

120

150

180

210

240

270IN

TA

KE

AIR

MA

SS

FLO

WR

AT

E(k

g/h)

40

45

50

55

60

65

70

75

80

VO

LUM

ET

RIC

EF

FIC

IEN

CY

(%)

D= 0.044 mD= 0.053 mD= 0.067 m

Fig. 6 Intake air mass flow rate and volumetric efficiency variation

with intake pipe diameter and engine speed

1000 2000 3000 4000 5000 6000 7000

SPEED (rev/min)

10

20

30

40

50

60

70

80

POWER

(kW)

30

40

50

60

70

80

90

100

TORQUE(N.m)

D= 0.044 mD= 0.053 mD= 0.067 m

Fig. 7 Power and torque variation with intake pipe diameter and

engine speed

1000 2000 3000 4000 5000 6000 7000

SPEED (rev/min)

0.230

0.240

0.250

0.260

0.270

0.280

0.290

0.300

0.310

SF

C(k

g/kW

h)

30.0

32.0

34.0

36.0

38.0

40.0

42.0

44.0

46.0T

HE

RM

AL

EF

FIC

I EN

CY

(%)

D= 0.044 mD= 0.053 mD= 0.067 m

Fig. 8 Specific fuel consumption and thermal efficiency variation

with intake pipe diameter and engine speed

34 J Braz. Soc. Mech. Sci. Eng. (2014) 36:29–35

123

are recommended to improve engine performance at low

engine speeds, while, for high engine speed, it is suggested

the use of intake pipes with short length and large diameter.

Acknowledgments The authors thank FIAT Automobiles, FAP-

EMIG, CNPq and CAPES for the financial support to this project.

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