A0320107

Research Inventy: International Journal Of Engineering And Science
Vol.3, Issue 2 (May 2013), PP 01-07
Issn(e): 2278-4721, Issn(p):2319-6483, Www.Researchinventy.Com
1
Design Analysis and Testing of a Gear Pump
E.A.P. Egbe
(Mechanical Engineering Department,
Federal University of Technology Minna, Nigeria)
Abstract - Nigeria depends heavily on importation of goods and machines. A shift from this trend requires the
development of locally available technology. The design analysis of a gear pump that aimed at delivering
4.0913x10-4m3/s (24.55litres/min) of oil was carried out in this work. Available technology was utilized in the
design and fabrication of the external gear pump. The design considered relevant theories and principles which
affect the performance of a pump. The parts of the pump were produced locally from available materials. The
performance of the pump was characterized and the test results showed a volumetric efficiency of 81.47 per cent
at a maximum delivery of 20litres/minute. The discharge dropped with increase in pressure head at a rate of -
0.344Litres/m.
Keywords - External gear, Involutes, Hydraulic, Pressure, Discharge.
I. INTRODUCTION
The design aimed at providing the down-stream sector of the petroleum industry and the small scale
industry with an indigenous pump that could deliver about 4.09123x10-4m3/s (24.55litre/minute) of hydraulic
oil. Pump is a device that enables mechanical energy to be imparted to a fluid and manifests in pressure energy
increase. Pumps have wide applications in science and engineering including, public water supply, irrigation,
up-steam/down-stream petroleum sector, auto-mobile, haulage equipment and chemical dosage. Gear pump is
the main choice of fuel system designers due to long life, low maintenance cost and high performance [1], [2].
The conventional centrifugal pumps must be primed first under service condition except when the
suction has a positive head. Situations arise in practice in which the suction to a pump has negative head. The
gear pump is self priming and not constrained by the type of suction head.
There are situations in which a fixed quantity of fluid is required per unit time or per revolution of the
pump. In water treatment for example, the amount of chemical dosage is closely regulated and the ability of the
dosing pump to supply a specified amount of chemical per pump revolution is critical. Chemical plants equally
require pumps that can deliver a fixed volume of fluid per unit revolution. Gear pumps belong to the group of
positive displacement pumps which are characterized with fixed volume discharge per unit revolution of the
pump.
The compactness of the gear pump makes it one of the few engineering equipment which any
developing country with kin interest in technology transfer should start with. Effective technology transfer can
only be realized when optimum application of available technology is combined with focus on products with
maximum local content.
II. LITERATURE REVIEW
Gear pump has a simple mechanism consisting two meshing spur or helical gears – the driver and the
idler. There are two major classes of gear pumps: the external type and the internal type. The former uses two
external spur gears while the latter uses one external spur gear and one internal spur gear.
The separation of gears on the suction side creates a partial vacuum which causes liquid to flow in and
fill the suction side. The liquid is carried to the discharge side between the rotating gear teeth and the fixed
casing. The meshing of the gears generates an increase in pressure which forces the liquid out through the
discharge line. In principle, either of the ports can become the discharge depending on the direction of rotation.
Tight side and top clearances between the gears and housing prevent the fluid from leaking backwards. The
amount of fluid pumped in one revolution depends on the amount of fluid that can be trapped within the gear

2
teeth space. Thus the discharge per unit revolution is a function of the size of the gears and the number of teeth
[3].
External gear pumps with spur gears are classified as low capacity pumps. When capacity higher than
912 litres per minute is required helical and herringbone gears are employed [4]. This work is limited to the
former.
The parameters of cutting tools such as pressure angle, diametral pitch, tooth addendum and dedendum have
been standardized and these standards form the foundation for gear design [5], [6]. The involute profile remains
the basic geometrical form for modern gearing. The 20o
pressure angle standard involute profile has a diametral
pitch, circular pitch, addendum, dedendum, and minimum clearance given respectively by:
Pd = n/D 1
Pc =πD/n = π/Pd 2
a= 1/Pd = Pc/π 3
dedendum = d= 1.25/Pd 4
clearance = c = 0.25/Pd 5
where n = number of teeth, D = circular pitch diameter.
Mott [5] indicated that the recommended working depth is 2/Pd. The following deductions can be made from the
foregoing conditions:
. The outside diameter of gear = D + 2a = D+ 2/Pd = D(n+2)/n 6
. Whole depth = a+d = 2.25/Pd = 2.25D/n 7
These conditions which have been tested by the American Gear Manufacturers Association (AGMA) were used
without further prove.
The torque required to drive an ideal pump at constant speed is given by:
T = )(
2
21
PP
D p


Nm 8
where Dp is the displacement of the pump in cubic metre per revolution of the driving gear. In practice there are
torque losses due to viscous and dry friction. There are also fluid losses due to leakage and compressibility of
the fluid. The performance and life expectancy of a pump depend on the properties of the liquid being pumped.
III. DESIGN ANALYSIS
The geometry of the relevant components determines the flow rate of the pump. Thus the first stage in
this work involved a geometrical design to determine the dimensions of all components to satisfy the target
discharge, followed by stress analysis to determine the most appropriate available material.
3.1 Geometrical design
The level of noise in gears is a function of the gear geometry, the clearance and the precision. The
relationships expressed in (1) to (7) were adopted in designing the geometries of all the components of the
pump. The discharge specification was 4.0913x10-4
m3
/s or 24.55litres/min - and the available hydraulic test rig
has an inlet port of 32mm diameter. At the preliminary stage of gathering relevant information on available
technology for this work it was found out that the available 20o
involutes gear cutter was limited to a minim of
12 number gear teeth.
DISCHARGE
dr
ar
Inlet
Figure 1: Cross section of the gear pump.

3
The volume of fluid displaced per revolution (denoted by Dp) is equal to volume of fluid trapped within
the space of gear teeth and housing (Figure 1). The trapped volume is given by:
Dp =
2
)(
22
brr da

9
where ra and rd are the addendum and dedendum radii respectively and b is gear face width. The geometry of the
gears in Figure 1 indicates that the addendum radius is given
by:
ra= a
D

2
=
n
nD
2
)2( 
10
Similarly the dedendum radius is given by:
rd = d
D

2
=
n
nD
2
)5.2( 
11
Substituting expressions for ra and rd into Equation 9 resulted in:
Dp = 2
2
8
)35.29]([
n
nbD 
12
The constraint on available 20o
involute cutter suggests that n must be greater than or equal to 12 and
the minimum possible value n=12 was used. The speed of the motor was specified as 1400rpm and pump
discharge as 4.082x10-4
m3
/s. The face width – b - and pitch circle diameter – D - remains unknown parameters
in (12). However two constraints were known on the face width. The face width must be longer than the
diameter of discharge port. This implied that b > 32mm. Design considerations indicates that face width should
be greater than 8/Pd but less than 16/Pd [5]. Substituting from (1) yielded:
n
D
b
n
D 168
 13
The upper limit is a critical value that must be avoided in order to prevent failure resulting from
dynamic forces due to misalignment and bending of the gear under load. Moreover the noise level increases
with face width. The first constraint was therefore used for initial design and the second constraint was used for
cross checking. Let b=38mm for initial design. Equation 12 becomes,



1400
)10(082.460
4
2
2
8
)35.29]([
n
nbD 
= 2
2
128
)35.2129](038.0[

D
14
Solving (14) for D yielded, D = 0.039974m; say 40mm. Substituting this value into Equation 13 in order to
verify complaint to the second constraint yields,
8x40/12=26.666< 38 =b < 16x40/12=53.333 15
This implied that the two constraints were satisfied and the final geometry of the gears became, b= 38mm,
D=40mm, n=12, Pd= 7.62/in, ra = 23.333mm and rd = 15.8333mm.
Moreover the thickness of a tooth is half the size of circular pitch and substituting into (2) produced a value of
5.235mm.
3.2 Stress analysis and material selection.
The dimensions of the gears which were obtained on basis of expected discharge were not altered under
consideration of stress. Stress analysis made it possible to select suitable materials under the operating
conditions.

4
Figure 2: Forces acting on the gears and component of forces.
The forces acting on the two gears under load are shown in Figure 2. The useful transmitted load that is
involved in transmission of power is the tangential component of force exerted by gear 1 on gear 2 and given
by:
Wt = Ft
12 16
The standard Lewis stress equation was modified to account for dynamic factor, geometry factor, and stress
concentration factor [7], [5] and the resulting load became,
Wt =
fd
v
KP
bYk 
17
where kv = dynamic factor, b =face width, Y= geometry factor, σ = stress, Pd = diametral pitch and Kf =
concentration factor. It is indicated that Kf for 20o
involute gear [7] is expressed by,
Kf = 0.18+
45.015.0
)()(
l
t
r
t
f
18
where standard gear root fillet radius is 0.3/Pd = 1mm, l= working depth = 2/Pd= 6.666mm and t=tooth
thickness = Pc/2 = πD/2n = 5.236mm.
Similarly kv is expressed by,
Kv =
)3(
3
v
19
where v is the pitch line velocity in m/s.
The design was considered satisfactory when the load computed from (17) was equal or greater than the
dynamic load on the gear.
The pump was expected to withstand a maximum discharge pressure of 10.2 x105
Pa. Recalling (8), the torque
applied on gear through the shaft was found to be,
T = )(
2
21
PP
D p


= )102.10(
1400
60
2
10082.4 5
4
x
x


= 2.84Nm
Efficiency of torque transmission from motor to pump was taken as 70%. Therefore the motor torque, Tm =
2.84/.7 = 4.057Nm. However the load on a tooth, W is a function of torque transmitted and it is given by:
Wr =
2
D
W = T 20
Applying the maximum possible torque of 4.057Nm and known pitch circle diameter yielded,
Wt=
04.0
2
057.4
2

D
Tx
= 205N
Fr
12
Ft
12
Ff2
F12
Fe1
F21
Ff2
(a) (b)

5
The dynamic load on the gear is given by,
Wd = ][
3
)3(
t
W
v
21
where v= pitch line velocity =
260
2 DN
=
2
04.0
60
14002
= 2.932m/s
Substituting into (21) yielded a dynamic load value of 405.3533N
The following available materials were considered for use:
1) 0.2% C hardened steel.
2) 0.4%C hot-rolled steel
3) AISI 1020 cold rolled steel
4) Aluminium wrought (2024-T4).
The load transmission capacity, Wt, for each material was computed from Equation 17 as shown for .2%C
hardened steel. The value of geometry factor for 20o
involute gear with 12 teeth is 0.264 [7]. All known values
were substituted into (17) to give,
Wt =
fd
v
KP
bYk 
=
33.103.0
10426264.038.0506.0
6
x
xxxx
= 54.32479MN
The safety margin was obtained by dividing the load bearing capacity of the material - 54.32479MN - by the
dynamic load and it was found to be 134018 for .2%C hardened steel. Similar calculations were carried out for
the other materials and Table 1 presents the values.
Table 1: Load bearing capacity of available materials
Material σt (N/m2
Wd Wt
Safety margin
0.2% C hardened steel. 427x106
405.3533 54324700
134918
0.4%C hot-rolled steel 365x106
405.3533 46435300
114557
AISI 1020 cold rolled steel 414x106
405.3533 52670300
129936
Aluminium wrought (2024-T4). 331x106
405.3533 42110770
103886
Though aluminium demonstrated the least safety margin, followed by 0.4%C hot-rolled steel, the results
generally indicate that all available materials could be used.
The 0.4%C hot-rolled steel was used on the basis of cost and the fact that it was available in annealed state. This
implied that the material was not subjected to surface treatment before machining.
3.2.2 Shaft design
The shaft must be capable of resisting shear forces due to applied torque and that due to bending load. The
bending force was considered acting through the centre of the shaft.
The bending load was earlier found to be 205N. The loading configuration shown in Figure 3 indicate that the
reactions RA and RB are given by,
RA =RB = Fb/2 = 205/2 =102.5N
Figure 3: Bending forces and the bending moment diagram.

6
The maximum bending moment as shown on the diagram is,
Mmax. = RA x 3.6 x 10-2
= 3.69Nm.
The bending stress: σx = 3
.max
32
d
M
22
While stress due to torsion: 3
16
d
T

  23
The maximum shear stress on the shaft is given by:
)(
16 22
3.max
TM
d


 24
Variation in pressure from suction to discharge suggested that shaft would be subjected to fatigue load
and a factor of safety of four was used to account for fatigue. By applying the maximum shear stress theory of
failure, the shaft diameter was calculated from (24) as:
d3
= )(
816 22
TM
x
ty


25
Substitution of yield stress values of available materials into (25) produced the diameter values shown in Table
2.
Table 2: Determination of shaft diameter
Since the shaft was to operate in oil medium, corrosion was not considered and thus mild steel with a
standard diameter of 12mm was used.
3.2.3 Design of pump housing
The gears and housing are the most complex components of the pump. On the contrary the shafts are
the simplest component and easy to produce. The foregoing led to a sacrificial design of the shaft with reference
to the housing. The maximum possible torque in the mild steel shaft was computed from (23). That is,
Tsmax =
16
3
 d
=
162
3
x
d ty

=
162
10217)012.0(
63
x
x
= 36.81Nm
The maximum pressure associated with the torque of 36.81 was calculated with Equation 8. Therefore the
maximum pressure in cylindrical housing was computed from (8) as,
Tsmax = 36.81= )(
2
21
PP
D p


= )(
1400
60
2
10082.4
4
P
x



Solving for P yielded a pressure of 13.22055MPa. This pressure could just cause failure of the shaft
and only a marginal difference would ensure a sacrificial failure of the shaft before the housing. A margin of 1.1
was adopted in this design. This implies an internal pressure of 13.22055x1.1 = 14.54MPas and the two stress
components in thin pressure vessels are, circumferential and axial stresses. The circumferential stress is more
critical in cylindrical pressure vessels and it is given by,
t
y
Pr
 26
Material σt (MN/m2
d (mm)
Mild steel 217 10.12
Structural steel 248 9.7
0.2% C hardened steel. 427 8.1
0.4%C hot-rolled steel 365 8.5
AISI 1020 cold rolled steel 414 9.8
Wrought iron 207 10.3

7
where P= internal pressure = 14.54MPa, r = internal radius of cylinder = ra+0.5c = 47.49mm and t= wall
thickness (hatched in Figure 1). Using steel (UNS-G10180 –HR) with a yield strength of 220.63MPa and a
design factor of 1.5, (26) yielded the thickness as,
t=
y

Pr
= 6
6
1063.220
04749.01054.14
x
xx
= 4.69x10-3
m; say 5mm.
IV. PUMP TESTING
The components of the external gear pump were assembled and the pump was coupled to an electric motor
(0.64kw, 1400rpm). The suction port was connected to the oil tank of a laboratory hydraulic rig and the
discharge to the inlet of test rig. The oil discharge (in litres/minute) of the pump was measured at different
pressure heads.
4.1 Test results and discussion
Figure 4 shows the variation in discharge with pressure head in metres. The gradual rate of drop in
discharge with increase in head was -.344 per metre. The drop in discharge with increase in pressure head was
due to increase in losses and pumps generally have this characteristic. The designed flow rate was
24.55litres/minute. The maximum discharge at zero head was 20litres/minute (Figure 4). The volumetric
efficiency of a pump is the ratio of the actual flow rate to the theoretical (or designed) flow rate. The test result
indicated a maximum volumetric efficiency of 81.47 per cent. This value is very high for a prototype. The
theoretical discharge assumed perfect geometry, absence of slip and friction losses. However the gears, shafts,
housing, wear plate, journal bearings and cover plate of this pump were machined manually and perfect
geometry cannot be expected from such processes. Moreover flow restriction at the discharge port was not ruled
out since the discharge port was smaller than the face width of the gears. Pressure build up on the discharge
would normally increase slip losses and thus a drop in volumetric efficiency.
CONCLUSION
The design analysis, fabrication and testing of an external gear pump was successfully carried out in
this work. This work indicated a good prospect for the design and fabrication of small machines/equipment
which will serve as a spring board for technological transfer and development of our country. The components
of this gear pump were fabricated by machining. Further work is required to investigate other processing routes
for mass production of external gear pumps in Nigeria.
REFERENCES
[1] Manring, N.D. and Kasaragadda, S.B. The Theoretical Flow Ripple of an External Gear Pump, Journal of Dynamic Systems,
Measurement, and Control, Transactions of the ASME, vol. 125, 2003, 396-404.
[2] Ragunathan, C., Manoharan, C. Dynamic Analysis of Hydrodynamic Gear Pump Performance Using Design of Experiments and
Operational Parameters, IOSR Journal of Mechanical and Civil Engineering, 1, (6), 2012, 17-23.
[3] Majundar, S.R. Oil Hydraulic Systems, Principles and Maintenance (Tata McGraw-Hill Publishing Corp. Ltd, New Delhi, 2001).
[4] Holland, F.A., and Chapman F.S., Pumping of liquids (Reinhold Publishing Corp., New York. 1966).
[5] Mott, R.L. Machine Elements in Mechanical Design, (Macmillan Publishing Comp. New York, 1992) .
[6] Kapelevich, A. and McNamara, T. Direct Gear Design for Optimal Gear Performance, SME Gear Processing and Manufacturing
Clinic/AGMA’s EXPO’ 03, Columbus, OH, 2003.
[7] Shigley, J.E. Mechanical Engineering Design, 3rd
Edition, McGraw-Hill Kogakusha Ltd, Tokyo, 1977.

A0320107

More Related Content

What's hot (17)

Viewers also liked (19)

Similar to A0320107 (20)

More from inventy (20)

Recently uploaded (20)

A0320107