RESULTS AND DISCUSSION - Shodhgangashodhganga.inflibnet.ac.in/bitstream/10603/37646/15/15_chapter 6.pdf · achieves maxima at 0.338 m3/s at 429 N/m2 suction pressure in case of 8

“Studies on Radial Tipped Centrifugal Fan” 232

CHAPTER – 6

RESULTS AND DISCUSSION

6.1 Introduction

Extensive experimental investigations are made to get optimum design

methodology for radial tipped centrifugal fan. The experiments are made in five

phases as explained in chapter 5. Performance of each phase is critically evaluated.

Salient features of results obtained during this course of work are discussed herein.

6.2 Phase-I: Experimental Optimization of Finite Number of Blades under Varying Speed Conditions

The stage 1 of phase 1 is designed to study the influence of suction pressure

on performance of centrifugal fan. The design point parameters were 1150 Pa static

stage pressure rise at 0.417 m3/s volume flow rate and speed of impeller 2800 rpm

[30]. The number of blades is varied in four steps of 8, 12, 16 and 24. The suction

pressure is varied with the help of orifice plates of different diameter. Suction

pressure variation is carried out by using six orifice plates of diameter 80, 90, 110,

120, 130,150 mm and full closing of suction duct to get no flow conditions. Figure

5.14 of Chapter 5 shows experimental setup developed for phase 1 experiments and

measurements.

Here, suction pressure variation lies in the range of 167 N/m2 to 1364 N/m2.

Observation Tables are given in Annexure C for this phase of experiments. Figure 6.1

to Figure 6.4 shows graphical presentation of the distinctive results obtained under

first stage of experiments for 8, 12, 16 24 numbers of blades, respectively.

Chapter – 6: Results and Discussion


(a) Suction Pressure Vs. Discharge (b) Static Stage Pressure Rise Vs.

Discharge

(c) Shaft Power Vs. Discharge (d) Efficiencies Vs. Discharge

Figure 6.1 Phase I, Stage 1 Performance Curves for 8 Number of Blades

0

200

400

600

800

1000

1200

1400

0.000 0.200 0.400

Suction Pressure, N/m

2

Discharge, m3/s

0

200

400

600

800

1000

1200

1400

0.000 0.200 0.400Stage Pressure Rise, N/m

2

Discharge, m3/s

0

100

200

300

400

500

600

700

0.000 0.200 0.400

Shaft P

ower, W

Discharge, m3/s

0

10

20

30

40

50

60

70

80

0.000 0.200 0.400

Efficiency,

%

Discharge, m3/s

Static

Stagnation




Discharge



0

200

400

600

800

1000

1200

1400

0.000 0.200 0.400


2

Discharge, m3/s

0

200

400

600

800

1000

1200

1400


2

Discharge, m3/s

0

100

200

300

400

500

600

0.000 0.200 0.400

Shaft P

ower, W

Discharge, m3/s

0102030405060708090

0.000 0.100 0.200 0.300 0.400

Efficiency, %

Discharge, m3/s

Static

Stagnation




Discharge



0

200

400

600

800

1000

1200

1400

1600

0.000 0.200 0.400


2

Discharge, m3/s

0

200

400

600

800

1000

1200

1400


2

Discharge, m3/s

0

100

200

300

400

500

600

700

0.000 0.200 0.400

Shaft P

ower, W

Discharge, m3/s

0102030405060708090

100

0.000 0.100 0.200 0.300 0.400

Efficiency, %

Discharge, m3/s

Static

Stagnation




Discharge



Figure 6.1 (a), 6.2 (a), 6.3 (a) and 6.4 (a) presents the influence of suction

pressure on discharge for 8, 12, 16 and 24 number of blades, respectively. It is

observed that with increasing suction pressure, the discharge gradually increases and

achieves maxima at 0.338 m3/s at 429 N/m2 suction pressure in case of 8 numbers of

blades. Similarly it attain maxima at 0.338 m3/s discharge at 421 N/m2 suction

pressure for 12 numbers of blades, 0.348 m3/s discharge at 389 N/m2 suction pressure

for 16 numbers of blades and 0.345 m3/s discharge at 428 N/m2 suction pressure for

24 numbers of blades, respectively. The nature of the curve is parabolic in shape.

When suction pressure reduces, the system resistance reduces and hence fan is

0

200

400

600

800

1000

1200

1400

1600

0.000 0.200 0.400


2

Discharge, m3/s

0

200

400

600

800

1000

1200

1400

0.000 0.200 0.400

Stage Pressure Rise, N/m

2

Discharge, m3/s

0

100

200

300

400

500

600

0.000 0.200 0.400

Shaft P

ower, W

Discharge, m3/s

0102030405060708090

100

0.000 0.100 0.200 0.300 0.400

Efficiency, %

Discharge, m3/s

Static



capable to handle higher discharges. When suction resistance increases, the losses

increases and the discharge begin to drop with further increase in suction pressure.

The nature of the graph follows the well established performance characteristics of

centrifugal fan [9, 26 and 28].

Figure 6.1 (b), 6.2 (b), 6.3 (b) and 6.4 (b) represents the stage static pressure

rise as a function of discharge for 8, 12, 16 and 24 number of blades, respectively. It

seems to be stabilized at 1289 N/m2, in the discharge range of 0.179 to 0.255 m3/s for

8 numbers of blades, 1280 N/m2 at 0.142 m3/s discharge for 12 numbers of blades,

1313 N/m2 at 0.145 m3/s discharge for 16 numbers of blades and 1274 N/m2 at 0.144

m3/s discharge for 24 numbers of blades, respectively. Further increase in discharge

leads to decrease in static fan pressure rise.

Figure 6.1 (c) shows the performance of the fan in terms of shaft power

consumed with respect to change in discharge for 8 numbers of blades. Initially flat

curve is observed for shaft power up to discharge level of 0.255 m3/s. There after

shaft power increases proportionate to increase in discharge. While Figure 6.2 (c) for

12 numbers of blades shows that initially there is no discharge but 429 watts shaft

power is consumed due to system resistance. Thereafter shaft power reduces even

though with rise in discharge. This happens due to dynamic action of rotor. Impeller

disc friction losses and frictional resistance of shaft bearing may collectively be

responsible for such behaviour. There after shaft power rises with respect to increase

in discharge. In case of Figure 6.3 (c) for 16 numbers of blades, it is observed that

initially shaft power is nearly constant up to discharge level of 0.145 m3/s. Initial drop

in shaft power is not observed for 16 numbers of blades as seen in the case of 8 and

12 number of blades. While Figure 6.4 (c) for 24 numbers of blades shows that

initially flat curve is observed for shaft power up to discharge level of 0.144 m3/s.

Here also initial drop in shaft power is not observed for 24 numbers similar to 16

numbers of blades. This behaviour attributes that as number of blades increases,

impeller becomes bulky and higher torque is required to get momentum. Hence

reduction of disc friction loss will not contribute much to get shaft power reduction,

as seen in earlier cases.

Figure 6.1 (d), 6.2 (d), 6.3 (d) and 6.4 (d) depicts the variation of static and

stagnation stage efficiencies with respect to discharge for 8, 12, 16 and 24 numbers of

blades, respectively.



For 8 blades, the static and stagnation stage efficiencies evidently increases

quite sharply with increase in discharge up to 0.255 m3/s, and there after the static and

stagnation stage efficiency improves gradually to attain maxima of 55% and 76% at

0.255 m3/s and 0.315 m3/s discharge, respectively.

While for 12 numbers of blades, the static and stagnation stage efficiencies

increases with increase in discharge up to 0.180 m3/s, and there after the stagnation

stage efficiency improves gradually up to 80% at 0.289 m3/s discharge. While static

efficiency attains maxima of 56% at 0.180 m3/s discharge.

In case of 16 numbers of blades, static and stagnation stage efficiency

improves gradually to attain maxima of 61% and 82% at 0.257 m3/s and 0.306 m3/s

discharge, respectively.

Similarly, for 24 numbers of blades, the static and stagnation stage efficiencies

increase with increase in discharge up to 0.144 m3/s. They attain maxima of 69% and

88% at 0.144 and 0.295 m3/s discharge, respectively.

The static and stagnation stage efficiency behaviour slightly differs due to

differences in velocity head at different centrifugal fan flow sections. These results

are also quite in tune with published performance curves of centrifugal fan [6, 13, 28]

It is worth noting that the optimum performance of the fan is achieved at a

static pressure rise of 1060 N/m2 at discharge of 0.301 m3/s with 88% stagnation

efficiency. This means that this performance is quite away from the design point

performance which was targeted to be 0.417 m3/s discharge at 1150 N/m2 static stage

pressure rise. This underlines the need for critically evaluating the design guidelines

for suction sides.

Table 6.1 summarizes the optimum performance parameters obtained during

phase 1 under constant speed of rotation as 2800 rpm.

Table 6.1 Optimum performance parameters as a Function of Number of Blades

No. of Blades

Pressure Developed

Pressure Developed ηStatic ηStagnation Max.

Discharge

Δp Static N/m2

Δ p Stagnation N/m2 Max. Max. m3/s

8 1290 1473 55.1 75.7 0.338 12 1280 1408 55.7 79.8 0.335 16 1313 1594 60.5 88.2 0.347 24 1274 1424 69.1 87.6 0.345



The optimum number of blades of a radial impeller can only be truly

ascertained by experiments [9, 14, 20]. Study of these tabulated results clearly

indicates that the best performance is achieved with 16 numbers of blades. This may

be attributed to the fact that with 16 numbers of blades, the flow is probably guided

without separation and at lower frictional losses with proper incidence at impeller

inlet. Separation losses remain minima due to optimum blade guidance and least

deviation in direction and magnitude of the flow velocity [20, 112].

Experimental results of stage 1 have also revealed that design point stage

pressure rise of the order of 1150 N/m2 is obtained by 110 mm diameter orifice plate

suction resistance. To ascertain the optimisation process, the fan performance have

been evaluated under off design conditions by keeping the orifice diameter constant at

110 mm and varying the speed of fan through variac, in second stage of experiments.

S. Sundaram [44] observed that the optimization of number of blades of centrifugal

fan impeller involves a maximization problem of multivariable function with fluid

dynamic constraints. Experimental data based on a simple variation in blade number

alone, keeping other parameters constant, will not yield optimum blade numbers for a

global maximum hydraulic efficiency. Hence, this stage of experiments were specially

planned to optimize finite number of blades under varying speed conditions with the

number of blades varying in 4 steps as 8, 12, 16 and 24.

Figure 6.5 to Figure 6.8 shows graphical presentation of the distinctive results

obtained under second stage of experiments for 8, 12, 16 24 numbers of blades,

respectively.



(a) Suction Pressure Vs. Speed (b) Static Stage Pressure Rise Vs. Speed

(c) Shaft Power Vs. Speed (d) Efficiencies Vs. Speed

(e) Discharge Vs. Speed


0

200

400

600

800

1000

1200

0 1000 2000 3000


2

Speed, rpm

0

200

400

600

800

1000

1200

0 1000 2000 3000


2

Speed, rpm

0

100

200

300

400

500

600

0 1000 2000 3000

Shaft P

ower, W

Speed, rpm

0

10

20

30

40

50

60

70

80

0 1000 2000 3000

Efficiency, %

Speed, rpm

static

Stagnation

0.000

0.050

0.100

0.150

0.200

0.250

0.300

0 1000 2000 3000

Discharge, m

3 /s

Speed, rpm







0

200

400

600

800

1000

1200

0 1000 2000 3000


2

Speed, rpm

0

200

400

600

800

1000

1200

0 1000 2000 3000


2

Speed, rpm

050

100150200250300350400450500

0 1000 2000 3000

Shaft P

ower, W

Speed, rpm

0

10

20

30

40

50

60

70

80

0 1000 2000 3000

Efficiency, %

Speed, rpm

static

stagnation

0.000

0.050

0.100

0.150

0.200

0.250

0.300

0 1000 2000 3000

Discharge, m

3 /s

Speed, rpm







0

200

400

600

800

1000

1200

0 1000 2000 3000


2

Speed, rpm

0

200

400

600

800

1000

1200

0 1000 2000 3000


2

Speed, rpm

0

100

200

300

400

500

600

0 1000 2000 3000

Shaft P

ower, W

Speed, rpm

0

10

20

30

40

50

60

70

80

0 1000 2000 3000

Efficiency, %

Speed, rpm

static

stagnation

0.000

0.050

0.100

0.150

0.200

0.250

0.300

0 1000 2000 3000

Discharge, m

3 /s

Speed, rpm







0

200

400

600

800

1000

1200

1400

0 1000 2000 3000


2

Speed, rpm

0

200

400

600

800

1000

1200

0 1000 2000 3000


2

Speed, rpm

050

100150200250300350400450500

0 1000 2000 3000

Shaft p

ower, W

Speed, rpm

0

10

20

30

40

50

60

70

80

0 1000 2000 3000

Efficiency, %

Speed, rpm

static

stagnation

0.000

0.050

0.100

0.150

0.200

0.250

0.300

0 1000 2000 3000

Discharge, m

3 /s

Speed, rpm



The study of these results clearly indicates an increase in all the performance

parameters with respect to speed which is basically attributed to better whirling effect

available with increase in speed [13, 26].

The optimum values of the performance parameters of Phase I stage 2

experiments under varying speed conditions for each case of number of blades are

summarized in Table 6.2.

Table 6.2 Optimum Values of Performance Parameters at Varying Speed

Conditions

Number of

Blades Speed

Pressure Developed

Pressure Developed ηStatic ηStagnation

Max. Discharge

rpm Δ p static

N/m2 Δ p stagnation

N/m2 Max. Max. m3/s

8 2800 1047 1420 55.1 74.8 0.255 12 2800 986 1367 53.8 74.6 0.258 16 2800 1079 1456 56.8 76.7 0.257 24 2800 1044 1435 60.7 83.4 0.261

It is observed that the best performance in terms of static and stagnation stage

pressure developed is achieved with 16 numbers of blades, though the better

efficiency is achieved for 24 numbers of blades.

The results of both the stages of the phase I experiments, thus confirms that

the best performance is achieved with 16 number of blades. However it is worth

mentioning that these values are obtained at restricted orifice size of 110 mm and

design point performance as static pressure rise of 1150 N/m2, and Discharge as 0.417

m3/s, is not achieved in any stage even with this optimum number of blades as 16.

Dimensionless coefficients are used for design, comparison, and critical

assessment of all geometrically similar fans. Pressure coefficient, volume coefficient,

power coefficient and efficiency are important dimensionless coefficients.

Figure 6.9 to 6.12 presents the non-dimensional performance characteristics of

fan for 8, 12, 16 and 24 numbers of blades at all variable speeds. These Figures

represent the variation of stage efficiency, head coefficient and power coefficient as a

function of discharge coefficient.



(a) Head Coefficient Vs. Discharge Coefficient

(b) Power Coefficient Vs. Discharge Coefficient

(c) Efficiency Vs. Discharge Coefficient

Figure 6.9 Phase I, Dimensionless Performance Curves for 8 Number of Blades

00.0010.0020.0030.0040.0050.0060.0070.0080.009

0.000 0.100 0.200 0.300

Head Co

efficient, gh/N2 D

2

Discharge Coefficient, Q/ND3

0

0.5

1

1.5

2

2.5

0.000 0.050 0.100 0.150 0.200 0.250 0.300

Power Coe

fficient, P/ρ

N3 D

5


0

10

20

30

40

50

60

70

80

0.000 0.050 0.100 0.150 0.200 0.250 0.300

Efficiency, %


static

stagnation






Figure 6.10 Phase I, Dimensionless Performance Curves for 12 Number of

Blades

0.0062

0.0064

0.0066

0.0068

0.007

0.0072

0.0074

0.0076

0.000 0.050 0.100 0.150 0.200 0.250 0.300

Head Co

efficient, gh/N2 D

2


0

0.5

1

1.5

2

2.5

0.000 0.100 0.200 0.300

Power Coe

fficient, P/ρ

N3 D

5


0102030405060708090

0.000 0.050 0.100 0.150 0.200 0.250 0.300

Efficiency, %


static

stagnation






Figure 6.11 Phase I, Dimensionless Performance Curves for 16 Number of

Blades

00.0010.0020.0030.0040.0050.0060.0070.0080.009

0.000 0.100 0.200 0.300 0.400

Head Co

efficient, gh/N2 D

2


0

0.5

1

1.5

2

2.5

0.000 0.100 0.200 0.300 0.400

Power Coe

fficient, P/ρ

N3 D

5


0

20

40

60

80

100

0.000 0.050 0.100 0.150 0.200 0.250 0.300 0.350

Efficiency, %


static

stagnation






Figure 6.12 Phase I, Dimensionless Performance Curves for 24 Number of Blades

0.0066

0.0068

0.007

0.0072

0.0074

0.0076

0.0078

0.000 0.100 0.200 0.300 0.400

Head Co

efficient, gh/N2 D

2


0

0.5

1

1.5

2

2.5

0.000 0.100 0.200 0.300 0.400

Power Coe

fficient, P/ρ

N3 D

5


0.010.020.030.040.050.060.070.080.090.0100.0

0.000 0.100 0.200 0.300 0.400

Efficiency, %


static

stagnation



The non-dimensional behaviour for each case of 8, 12, 16 and 24 numbers of

blades is quite similar to that as discussed in stage 1. However its importance is to

arrive at optimum values of non-dimensional parameters for developing the future

design methodology. These optimal parameters are presented in Table 6.3.

Table 6.3 Stage I: Optimum Values of Non-Dimensional Performance Parameters

Number of Blades

Discharge Coefficient

Head Coefficient

Power Coefficient ηStatic ηStagnation

8 0.2823 0.0078 2.2747 55.1 75.7 12 0.2794 0.0074 2.1534 55.7 79.8 16 0.2894 0.0084 2.3034 60.5 88.2 24 0.2879 0.0076 2.0399 69.1 87.6

The results presented in Table 6.3 certainly advocate the facts that 16 numbers

of blades offers the best performance. The discharge coefficient, head coefficient and

power coefficient obtained for this case are 0.2894, 0.0084 and 2.3034, respectively.

These values can effectively be used to design a realistic fan for optimum

performance. In other words this information may prove to be a useful design base for

radial tipped forward swept centrifugal fan.

Further, it is worth to observe that in 1942, Kearton [25] has felt the need of

optimization of finite number of blades as he observed certain breaks and inactive

flow regimes in trailing of blade. Eck Bruno, Pfleiaderer and Stepanoff [9, 14, 32]

attempted to present correlations for determining finite number of blades in

centrifugal fans. S. Sundaram [44] in later stage categorically stated that the

optimization of number of blades of centrifugal fan to be a multivariable problem.

The present work has treated the problem of optimization as a multivariable

problem and the experimental optimization have been carried out under variable

suction, speed and number of blades conditions during two stage in this phase of

experimental studies. This multi variable experimental optimization process clearly

establishes 16 number of blades as the optimum number of blades for centrifugal fan.

These findings are quite close to the results of Songling Wang [123] who

numerically obtained optimum number of blades as 14.



6.3 Phase-II: Experimental Investigations on Slip Factor at Varying Number of Blades and Speed Conditions

Slip loss is defined as the ratio of actual and ideal values of the whirl

components at exit of impeller. It has significant effect on fan performance. It is

essential to estimate the slip factor correctly for proper design of centrifugal fans.

Stodola [31] had developed first useful method for slip factor approximation.

He correlated slip factor and finite number of blades. Stodola claimed that average

direction of discharge varies from the blade angle β2 due to number of blades and

relative circulation in vane to vane plane. Several correlations thereafter have been

proposed to estimate slip factor [9, 14, 75, 76, 78, 79, 83, 87]. The objective of

present study is to make comparative assessment of theoretically correlated and

experimentally evaluated slip factor at varied number of blades.

Theoretical velocity diagram for radial tipped blade exit is prepared with the

help of theoretical values of tangential velocity component U2, radial velocity

component Vr2 and flow angle θ2. Actual velocity diagram is prepared with the help of

measured components of absolute velocity V2’ at measured flow angle θ2

’ at impeller

exit. Theoretical component of tangential velocity U2 is also used to construct actual

velocity diagram. Actual component of radial velocity Vr2’ and actual blade angle β2

’

are received from actual velocity diagram. Construction of actual velocity diagram at

all test locations A to H for 16 numbers of blades is shown in Figure 6.13.

Figure 6.13 Theoretical and Actual Velocity Triangles at all Test Locations



Here the experimental value of the slip factor for a radial tipped centrifugal

fan is determined at various selected test locations around the circumference of the

impeller covering the entire width of impeller at varying number of blades.

Observation Tables for phase II experiments, showing measurements of slip

factors at various volute locations for 8, 12. 16 and 24 numbers of blades are given in

Annexure C.

The sub test locations slip factor values are averaged to get respective test

location slip factor value. Test location plan is given in Table 5.2 of chapter 5. These

slip factors are calculated for 8, 12, 16 and 24 numbers of blades.

Table 6.4 shows a typical variation of slip factor across the blade width at sub

location D for 16 numbers of blades along the blade width. The angle for actual

absolute velocity [V2'] with respect to the tangential velocity [U2] is θ2'= 370.

Table 6.4 Probe Readings at Location D for 16 Numbers of Blades

Location Hstag Hstat. Hdynamic Pd V2' Slip factor

mm of water N/m2 m/sec μ D1 20.45 1.81 18.64 182.65 16.47 0.693 D2 23.29 1.81 21.48 210.47 17.68 0.744 D3 24.85 1.55 23.3 228.31 18.41 0.775 D4 25.62 1.55 24.07 235.85 18.71 0.788 D5 24.59 1.55 23.04 225.76 18.31 0.771 D6 22.26 1.55 20.71 202.93 17.36 0.731 D7 17.6 2.07 15.53 152.17 15.03 0.633

Average 0.734

Study of Table 6.4 reveals that slip factor is not constant across the blade

width. Its variations are parabolic in nature. The boundary layer effect near shroud is

observed to be responsible for such behavior [50, 66].

It is also seen from experimental results that slip factor profile over blade

width at impeller exit shows negative parabolic profile. Hence design of exit blade

section must be parabolic to improve blade tip slip factor.

The average Experimental slip factor is 0.734 at location D for entire blade

width. Similarly, slip factor is measured at all volute sub locations for 8, 12, 16 and 24

number of blades.

Figure 6.14 to 6.17 graphically represents comparative study of experimental

and empirical slip factor values for each case of number of blades. Figure 6.18 shows



comparative analysis of empirical and experimental slip factors with respect to

number of blades.

Figure 6.14 Empirical and Experimental Slip Factors for 8 Blades


0.500

0.550

0.600

0.650

0.700

0.750

0.800

0.850

0.900

0.950

0 2 4 6 8 10

Slip Factor

Test Locations

Stanitz Balje Stodola Avg. Experimental

0.550

0.600

0.650

0.700

0.750

0.800

0.850

0.900

0 2 4 6 8 10

Slip Factor

Test Locations






0.550

0.600

0.650

0.700

0.750

0.800

0.850

0.900

0 2 4 6 8 10

Slip Factor

Test Locations


0.550

0.600

0.650

0.700

0.750

0.800

0.850

0.900

0.950

0 2 4 6 8 10

Slip Factor

Test Locations




Figure 6.18 Empirical and Experimental Slip Factors Vs Number of Blades

Further, results obtained under phase II experiments as depicted in Figure 6.14

to 6.18 have revealed that as the number of blade increases, the slip factor increases.

At less number of blades, the turbulence within blade passage increases due to lack of

flow guidance. When number of blades increases, the deviation of the exit absolute

velocity vector is minimized and hence it can be efficiently sensed by the designed

probe.

It is observed that experimental value of slip factor is found to be 3 to 12%

less with respect to various empirical co-relations. This difference is maintained for

almost all set of number of blades. Exceptions are seen at very low number of blades

due to sudden increase of turbulence between two blades. This behavior leads to

conclude that impeller designed for a large specific speed has a large inlet to exit

radius ratio and the curvature of the shroud is large to turn the flow from axial to

radial. Therefore unless the impeller is carefully designed, the flow in meridional

plane separates from the shroud. This supports work of Yedidiah Sh. [75] of

presenting a new model of slip factor that resolves basic discrepancies observed

between old theories.

It is also observed that slip factor is varying with number of blades. Hence it

can be said that slip factor is not only dependent of flow. Similar is supported by

0.550

0.600

0.650

0.700

0.750

0.800

0.850

0.900

0.950

0 5 10 15 20 25 30

Slip Factor

No. of Blades

Stanitz Balje Stodola Experimental



research work of R. Ajithkumar [22] who has concluded that slip factor is a function

of number of vanes, diameter ratio, and outlet blade angle and flow conditions after

impeller.

Lal and Vasandani [78] has studied slip factor effect on designing of impeller.

They concluded that slip factor reduces due to non-uniform velocity distribution at

impeller exit. This is also confirmed from Table 6.5 and Figure 6.13.

Table 6.5 Comparison of Experimental and Empirical Value of Slip Factors

Number of Blades

Experimental Slip Factor

Empirical Slip Factor by

Balje

Empirical Slip Factor by Stodola

Empirical Slip Factor by

Stanitz

8 0.773 0.639 0.608 0.753 12 0.731 0.726 0.738 0.835 16 0.771 0.78 0.804 0.876 24 0.785 0.841 0.869 0.918

Thus the present work has clearly highlighted the variation of slip factor

across the blade width as well as along the flow path of volute and the lacuna of

existing slip factor correlation to correctly predict the slip factor.

6.4 Phase-III: Assessment of Explicit Design Methodologies

This phase of experimental work is planned for experimental assessment of

fan performance which are designed and fabricated as per explicit design

methodologies. Forward curved radial tipped impeller centrifugal fans are fabricated

as per design methodology traced out by using fundamental principles of fluid flow

and other design methodologies suggested by Church A. H. and Osborne W. C.

The observation Tables which include fan inlet and outlet pressure, volute

casing pressure distribution at 90˚, 120˚, 180˚, 240˚ and 300˚, input power and average

air velocity, at different speed and damping conditions are recorded and are available

in Annexure C. Actual stage pressure head developed across the fan, average air

discharge, shaft power, static airpower developed, static and stagnation efficiencies

are calculated for each set of observation. Design point is also highlighted in

observation Tables.

The speed of impeller rotation was varied and kept constant at 500, 1000,

1500, 2000, 2500 and 2800 rpm as maximum as well as design speed of rotation. Inlet

volume flow is varied by damping discharge duct in terms of reduction of volute



discharge cross-section area. This damping is designated as 0% (means full opening

of discharge duct), 25%, 50%, 75% and complete blockage, i.e. 100% damping.

Performance of each design methodology is critically evaluated and discussed

in subsequent sections.

6.4.1 Fundamental design

Figure 6.19 to 6.23 shows the fundamental design performance obtained in

terms of discharge v/s stage pressure rise, air power, shaft power and static efficiency

and volute casing pressure distribution.

Figure 6.19 Static Stage Pressure Rise Vs Discharge (Fundamental Design)

0

200

400

600

800

1000

1200

0.000 0.100 0.200 0.300 0.400 0.500 0.600

Stag

e Pr

essu

re R

ise,

Pa

Discharge, m3/s

500 rpm1000 rpm1500 rpm2000 rpm2500 rpm2800 rpmDesign PointLinear (Design Point)



Figure 6.20 Air Power Vs Discharge (Fundamental Design)

Figure 6.21 Shaft Power Vs Discharge (Fundamental Design)

0

100

200

300

400

500

600

0.000 0.100 0.200 0.300 0.400 0.500 0.600

Air

Pow

er, W

Discharge, m3/s

500 rpm1000 rpm1500 rpm2000 rpm2500 rpm2800 rpmDesign Point

0

100

200

300

400

500

600

0.000 0.100 0.200 0.300 0.400 0.500 0.600

Shaf

t Po

wer

, w

Discharge, m3/s

500 rpm1000 rpm1500 rpm2000 rpm2500 rpm2800 rpm



Figure 6.22 Static Efficiency Vs Discharge (Fundamental Design)

Figure 6.23 Volute Casing Pressure Distribution (Fundamental Design)

0

10

20

30

40

50

60

0.000 0.100 0.200 0.300 0.400 0.500 0.600

Eff

icie

ncy,

%

Discharge, m3/s

500 rpm

1000 rpm

1500 rpm

2000 rpm

2500 rpm

2800 rpm

0

100

200

300

400

500

6001

2

34

5

Volute Casing Static Pressure (Pa) Distribution at 2800 rpm and 0 % Damping Conditions

Axis

12345

Press.(Pa)307347491497504

Deg.

90120180240300



Figure 6.19 shows the variation of stage pressure as a function of discharge

under different speed conditions. It is observed that at 25% damping, there is major

gain in stage pressure rise with slight reduction in discharge. Thereafter at further

increase in damping, there is reduction in flow rate without much affecting stage

pressure head.

Figure 6.20 shows variation of air power in watts against discharge at different

speed. The maximum airpower achieved is 255 W at 0.502 m3/s discharges. This also

indicates towards non-achievement of design point performance.

Figure 6.21 is presentation for variation of shaft power in watts with respect to

discharge, having speed as parameter. This is a basic performance curve to find

optimum range of operation for any identical fan. Here maximum shaft power is 544

W at 0.502 m3/s discharge. Maximum achieved discharge is little higher than design

point discharge 0.5 m3/s.

Figure 6.22 is another basic performance curve and it is to be evaluated along

with Figure 6.19 and Figure 6.21. This Figure is representing efficiency (at static

pressure based) as a function of discharge under different speeds. It is interesting to

note that maximum efficiency of 51% at 75% damping condition is achieved at 2500

rpm instead of design point 2800 rpm. This may be due to rise in hydraulic losses at

higher speed. Further it is seen that the maximum efficiency point in almost all cases

is achieved at 75% damping under all speeds. All observations are lying below design

point stage pressure rise hence it is not possible to mark best efficiency operating

range for present fan.

Figure 6.23 is graphical presentation of volute pressure distribution at 2800

rpm for 0% damping. At 0% damping it can be seen that velocity head is being

converted in pressure head uniformly during volute passage. This shows efficient

diffusion process of volute casing. It can also be seen that between 120˚ and 180˚,

diffusion rate is higher compared to other places.

It is concluded from above Figures, that fan inlet pressure measured very near

to impeller inlet is – 488 Pa while fan outlet pressure is only 20.6 Pa. Thus, total static

pressure rise of 509 Pa is achieved at 2800 rpm and 0% damping. The discharge

obtained is 0.502 m3/s and the efficiency at this point is only 38%. It is interesting to

note that only discharge is achieved as per design point.

It can be said for fundamental design that, it is better for suction pressure

development but very much lacking in generating outlet pressure. This design can



achieve only design flow rate at maximum speed and 0% damping. Best operating

range is not possible to trace. Present efficient operating range for this fan is at 75%

damping position for all speed conditions. At off-design operating points, discharge

gets very much reduced. It is suggested that proper loss estimation and redesigning for

impeller outlet diameter can help to achieve design point performance. Positive

outcome of this design is that suction pressure at fan inlet is generated more than

twice to design point requirement.

6.4.2 Church design

Figure 6.24 to 6.28 shows the Church design performance obtained in terms of

discharge v/s stage pressure rise, air power, shaft power, efficiency on static pressure

and volute casing pressure distribution.

Figure 6.24 Static Stage Pressure Rise Vs Discharge (Church Design)

0

200

400

600

800

1000

1200

1400

1600

0.000 0.100 0.200 0.300 0.400 0.500 0.600

Stag

e Pr

essu

re R

ise,

Pa

Discharge, m3/s

500 rpm1000 rpm1500 rpm2000 rpm2500 rpm2800 rpmDesign PointLinear (Design Point)



Figure 6.25 Air Power on Static Pressure Vs Discharge (Church Design)

Figure 6.26 Shaft Power Vs Discharge (Church Design)

0

100

200

300

400

500

600

0.000 0.100 0.200 0.300 0.400 0.500 0.600

Air

Pow

er, W

Discharge, m3/s


0

100

200

300

400

500

600

700

800

0.000 0.100 0.200 0.300 0.400

Shaf

t Po

wer

, w

Discharge, m3/s




Figure 6.27 Static Efficiency Vs Discharge (Church Design)

Figure 6.28 Volute Casing Pressure Distribution (Church Design)

0

10

20

30

40

50

60

0.000 0.050 0.100 0.150 0.200 0.250 0.300 0.350 0.400

Eff

icie

ncy,

%

Discharge, m3/s

500 rpm

1000 rpm

1500 rpm

2000 rpm

2500 rpm

2800 rpm

0200400600800

10001200

1

2

34

5

Volute Pressure Distribution ( Pa) at 2800 rpm and 0 % Damping Conditions

Axis

12345

Press.(Pa)752765903883

1014

Deg.

90120180240300



Figure 6.24 presents features of stage pressure rise with discharge on abscissa,

at different speed conditions. Here design point stage pressure rise is achieved but

actual flow rate remains far away from designed. It is surprising to observe a big jump

in stage pressure rise between, 25 to 50% damping conditions.

Figure 6.25 is representing airpower generated against discharge of flow

received. This shows similar picture of non-achievement of design point performance

at any damping position. The maximum airpower achieved is 221 W at 25% damping

at 2800 rpm and 0.236 m3/s discharge.

Figure 6.26 presents expenditure of shaft power for amount of discharge

achieved at various speeds and damping conditions. Here 706 W shaft power is

consumed but it is contributing only 0.370 m3/s flow rate. Once again it is not

possible to establish efficient operating range for this fan.

Figure 6.27 is showing efficiency based on static pressure rise achieved during

this course of work against discharge obtained. Efficiency is increasing between, 50%

to 75% damping conditions for all impeller speeds. This design attains maximum 56%

efficiency at 75% damping and 2800 rpm by offering only 0.135 m3/s discharge. At

this point of operation, stage pressure rise is seen 1404 Pa. This is beyond design

point requirement.

Figure 6.28 shows volute casing pressure distribution at various angular

positions having 0% damping at 2800 rpm. It is observed that static pressure of

flowing fluid increases when it is marching towards volute exit. Between zero to 300

degree volute positions, static pressure rises from 752 Pa to 1014 Pa. This shows

efficient diffusion of fluid velocity and hence static pressure gets increased.

As per this design methodology, the experimental results are quite passive at

0% damping and design speed. The performance of this centrifugal fan at its best

efficiency point offers 1404 Pa static stage pressure rise, development of 189 W

airpower at 0.135 m3/s air discharge and consuming 270 W shaft power at 56% static

efficiency and 57% stagnation efficiency.

Performance has shown a need for redesigning of suction dimensions.

Impeller inlet dimensions along with blade width should increase so that higher flow

rate can be achieved. Increase in nose radius will help to reduce re-circulation of flow

and that way reduction in leakage losses. Church has used pressure coefficient K’. Its

value lies empirically between 0.5-0.65. This is used to calculate impeller outlet

diameter. This empirical concept requires reconsideration. However if design point



performance is neglected, the best operating region for this design lies between 50 to

75% damping conditions.

6.4.3 Osborne design

Figure 6.29 to 6.33 shows the Osborne design performance obtained in terms

of discharge v/s stage pressure rise, air power, shaft power, efficiency on static

pressure and volute casing pressure distribution.

Figure 6.29 Static Stage Pressure Rise Vs Discharge (Osborne Design)

0

200

400

600

800

1000

1200

1400

1600

1800

2000

0.000 0.200 0.400 0.600 0.800 1.000

Stag

e Pr

essu

re R

ise,

Pa

Discharge, m3/s

500 rpm1000 rpm1500 rpm2000 rpm2500 rpmDesign Point2800 rpmLinear (Design Point)



Figure 6.30 Air Power on Static Pressure Vs Discharge (Osborne Design)

Figure 6.31 Shaft Power Vs Discharge (Osborne Design)

0

200

400

600

800

1000

1200

1400

0.000 0.200 0.400 0.600 0.800 1.000

Air

Pow

er, W

Discharge, m3/s

500 rpm1000 rpm1500 rpm2000 rpm2500 rpmDesign Point2800 rpm

0

500

1000

1500

2000

2500

3000

0.000 0.200 0.400 0.600 0.800 1.000

Shaf

t Po

wer

, w

Discharge, m3/s




Figure 6.32 Static Efficiency Vs Discharge (Osborne Design)

Figure 6.33 Volute Casing Pressure Distribution (Osborne Design)

0

10

20

30

40

50

60

0.000 0.200 0.400 0.600 0.800 1.000

Stat

ic E

ffic

ienc

y, %

Discharge, m3/s


0

500

1000

15001

2

34

5

Volute Casing Pressure (Pa) Distribution at 2800 rpm and 0 % Damping Conditions

Axis

12345

Press.(Pa)728

1036949

10921028

Deg.

90120180240300



Figure 6.29 presents stage pressure rise against discharge at different design

and off design speed conditions. At 2800 rpm and 0% and 25% damping conditions,

fan differential pressure head generated is 1353 and 1796 Pa at 0.879 m3/s and 0.517

m3/s discharge, respectively. These are well above the design point requisition. It can

also be seen that 0-25% damping conditions offers efficient operating range.

Figure 6.30 presents airpower developed against discharge achieved at

different speed and damping conditions. 1190 W maximum airpower is received at

2800 rpm and 0% damping at 39% efficiency.

Figure 6.31 shows distinctive features of shaft power against discharge

achieved. It reveals that shaft power consumption is very much higher. It consumes

2410 W at 0% damping and 2800 rpm. But at 25% damping, shaft power is reduced to

1506 W to generate 0.517 m3/s discharge at 49% efficiency.

Figure 6.32 represent static efficiency achieved during this course of work

against discharge. This basic performance curve highlights that at 0% and 25%

damping, the static efficiencies achieved are 39% and 49%, respectively. This result

also supports that 0-25% damping conditions gives better efficiency region.

Figure 6.33 shows identical volute pressure distribution at 0% damping and

2800 rpm conditions. This seems to be uniform along the volute passage and

presenting proper diffusion of velocity.

The results obtained under this design are very much encouraging. Major

performance parameters like inlet pressure, outlet pressure, volume flow rate, stage

pressure rise are achieved on higher side of design point. This design is more suitable

where higher suction head is required. Here it can be summarized that fan designed as

per Osborne design method is good enough to meet design point performance. With

this design, flow regulation is also possible which can provide flexibility to run fan at

off design conditions, too. This is useful and essential in establishing best operating

range. Major limitation of this fan is geometrically bigger size compared to

fundamental and Church design. It is also consuming very high shaft power compared

to others two. To overcome these limitations, conservative loss estimation should be

made while redesigning the fan. Impeller size can be reduced by applying fan laws for

geometrically similar fans for required design point performance. The best operating

range for this fan is at 25% damping and 2800 rpm.



6.4.4 Comparative performance assessment of explicit design methodologies

Table 6.6 shows comparative performance assessment of fans running at

design speed of 2800 rpm and 0% damping conditions. Design point requirements and

operating points obtained under fundamental, Church and Osborne design

methodologies are also tabulated.

Table 6.6 Comparative Performance Assessment of Design Methodologies

Design Speed rpm Damp.

Fan Delta P

Pa

Avg. Air Discharge

m3/s

Static Air

Power W

Input Power

W

Eff. (Static)

%

Input Design Parameters 2800 0% 981 0.500 491 - -

Fundamental 2800 0% 509 0.502 255 680 38 Church 2800 0% 244 0.370 90 883 10 Osborne 2800 0% 1353 0.879 1190 3012 40

Based on experimental results obtained and evaluation made, it is recognized

that there exists a wide performance difference amongst fans under study. All fans are

not performing as per mark.

Fundamental and Church designs are failing to generate required discharge,

stage pressure rise and hence generating poor airpower. Their operating efficiencies

are also very much poor. Osborne design is working better amongst three design

methodologies but it is bulky and consuming very high shaft power. Osborne design

requires redesigning for controlling stage pressure rise and reduction in input power.

These improvements in design will lead to increase of stagnation efficiency at design

point flow and pressure rise.

It is found for fundamental and Church design that the fluid deviates and

becomes more non-uniform at impeller exit and volute casing is failing to guide and

diffuse fluid efficiently. Same is confirmed earlier by Prasad, Ganeshan and Prithviraj

[74]. Volumetric flow of fundamental and Church design is very much less as a result

of recirculation [92]. It is worth noting here that the design dimensions as per

Fundamental, Church and Osborne designs have wide variations. Similarly their

performance is also having wide variations.

This performance assessment has indicated that there is a need to develop

unified design methodology for radial tipped centrifugal fan to get real design point

performance.



6.5 Phase-IV: Unified Design Methodology and Comparative Performance Evaluation of Forward and Backward Curved Radial Tipped Centrifugal Fan

Experimental evaluation of fans stated in phase III has revealed that there

exists a wide performance variation. Hence successful outcomes of fundamental,

Church and Osborne designs are incorporated and unified design methodology for

radial tipped centrifugal fan is developed. Unified design procedure is developed from

fundamental concepts and involving minimum assumptions. Forward and backward

curved radial tipped impeller fans are fabricated as per this unified design

methodology. Their performance is measured as per standard test procedure IS: 4894-

1987, Indian Standard Specification for Centrifugal Fans. Experimental test

observation Tables for forward curved and backward curved radial tipped fans

fabricated as per unified design methodology are given in Annexure C. Result

analysis of the performance of forward and backward curved centrifugal fans are

discussed below.

6.5.1 Forward curved radial tipped (FCRT) centrifugal fan

Figure 6.34 to 6.38 shows the unified design forward curved radial tipped

(FCRT) centrifugal fan performance obtained in terms of discharge v/s stage pressure

rise, air power, input power, efficiency on static pressure and volute casing pressure

distribution.

Figure 6.34 Static Stage Pressure Rise Vs Discharge (Unified Design, FCRT Fan)

0200400600800

10001200140016001800

0.000 0.200 0.400 0.600 0.800 1.000

Stag

e D

elta

P,

Pa

Discharge, m3/s




Figure 6.35 Air Power on Static Pressure Vs Discharge (Unified Design, FCRT Fan)

Figure 6.36 Input Power Vs Discharge (Unified Design, FCRT Fan)

0

200

400

600

800

1000

1200

1400

0.000 0.200 0.400 0.600 0.800 1.000

Air

Pow

er, W

Discharge, m3/s

500 rpm

1000 rpm

1500 rpm

2000 rpm

2500 rpm

2800 rpm

Design Point

0

200

400

600

800

1000

1200

1400

1600

1800

2000

0.000 0.200 0.400 0.600 0.800 1.000

Inpu

t Pow

er, W

Discharge, m3/s




Figure 6.37 Efficiency (Static Pressure Based) Vs Discharge (Unified Design,

FCRT Fan)

Figure 6.38 Volute Casing Pressure Distribution (Unified Design, FCRT Fan)

0

10

20

30

40

50

60

70

80

0.000 0.200 0.400 0.600 0.800 1.000

Tot

al E

ffic

ienc

y ( S

tatic

), %

Discharge, m3/s


0

50

100

150

200

250

3001

2

34

5

Axis

12345

Press.(Pa)

268.4241.6161.1201.3201.3

Deg.

90120180240300

Unified Design - Forward Curved Radial TippedVolute Pressure Distribution ( Pa) at 2800 rpm and 0 % damping



Figure 6.34 represents stage pressure rise against discharge at different speed

conditions. It is very important to observe that it reflects true radial tipped blade

characteristics. Stage pressure rise remains nearly constant at all values of discharge

at different speeds. At 2800 rpm, fan differential pressure head generated is 1522 Pa

and discharge 0.865 m3/s. These values are well above the design point requisition. It

can also be seen that 0-50% damping positions is best operating range.

Figures 6.35 and 6.36 shows the variation of air power and input power as a

function of discharge, respectively. The input and air power increases with discharge.

This is well established characteristic of centrifugal fan [26, 28]. Design point lies

centrally in these performance graphs to offer design and off design operating range.

At 2800 rpm, 1316 W airpower is generated at the cost of 1888 watts input power

offering 70% efficiency.

Figure 6.37 shows the variation of static pressure based efficiency with respect

to discharge. This is basic performance characteristic curve, which highlights that at

0% to 50% damping, the best efficiency obtained are 70% at 2800 rpm. This result

also signifies that 0 to 50% damping position gives best efficiency region. It is

interesting to note that the present design offers design point performance in the terms

of efficiency and discharge.

Figure 6.38 shows identical volute pressure distribution at 0% damping and

2800 rpm conditions. This seems to be uniform along the volute passage and

presenting proper diffusion of velocity except 180 degree volute casing position due

to possible sudden increase in local velocities.

Here it can be summarized that designed fan based on unified design is good

enough to achieve desired performance. With this design, flow regulation is also

possible which can provide flexibility to run fan at off design conditions, too. The best

operating range for this fan is at 0-50% damping at 2800 rpm as shown in Table 6.7.

Table 6.7 Optimum Operating Range of Forward Curved Radial Tipped

Centrifugal Fan

Speed rpm Damp.

Fan Stage Pressure

Rise Pa

Avg. Air Discharge

m3/s

Static Air Power

W

Input Power

W

Static Eff. %

2800 0% 1522 0.865 1316 1888 69.7 2800 50% 1544 0.788 1216 1746 69.6



The results obtained under unified design are very much encouraging. Major

performance parameters achieved are on higher side of design point. This is useful in

establishing best operating range, which is essential for any turbo machine, especially

fan.

6.5.2 Backward curved radial tipped (BCRT) centrifugal fan

Figure 6.39 to 6.43 shows the unified design backward curved radial tipped

(BCRT) centrifugal fan performance obtained in terms of discharge v/s stage pressure

rise, air power, input power, efficiency on static pressure and volute casing pressure

distribution.

Figure 6.39 Static Stage Pressure Rise Vs Discharge (Unified Design, BCRT Fan)

0

200

400

600

800

1000

1200

1400

1600

0.000 0.200 0.400 0.600 0.800 1.000

Stag

e Pr

essu

re R

ise,

Pa

Discharge, m3/s




Figure 6.40 Air Power on Static Pressure Vs Discharge (Unified Design, BCRT

Fan)

Figure 6.41 Input Power Vs Discharge (Unified Design, BCRT Fan)

0

200

400

600

800

1000

1200

0.000 0.200 0.400 0.600 0.800 1.000

Air

Pow

er, W

Discharge, m3/s

500 rpm

1000 rpm

1500 rpm

2000 rpm

2500 rpm

2800 rpm

Design Point

0

200

400

600

800

1000

1200

1400

1600

0.000 0.200 0.400 0.600 0.800 1.000

Inpu

t Pow

er, W

Discharge, m3/s

500 rpm

1000 rpm

1500 rpm

2000 rpm

2500 rpm

2800 rpm

Design Point



Figure 6.42 Static Efficiency Vs Discharge (Unified Design, BCRT Fan)

Figure 6.43 Volute Casing Pressure Distribution (Unified Design, BCRT Fan)

0

10

20

30

40

50

60

70

80

90

0.000 0.200 0.400 0.600 0.800 1.000

Tot

al E

ffic

ienc

y ( S

tatic

), %

Discharge, m3/s


0

50

100

150

2001

2

34

5

Unified Design - Backward Curved Radial TippedVolute Pressure Distribution ( Pa) at 2800 rpm and 0 %

damping

Axis

12345

Press.(Pa)1141541618160

Deg.

90120180240300



The results obtained with unified design for backward curved radial tipped fan

are very much encouraging similar to forward curved radial tipped centrifugal fan.

Major performance parameters achieved are on higher side of design point. This is

useful in establishing best operating range of a centrifugal fan.

Figure 6.39 represents stage pressure rise against discharge at different speed

conditions. It is very important to observe that it reflects true radial tipped blade

characteristics. Stage pressure rise remains nearly constant at all values of discharge

at different speeds. At 2800 rpm, fan differential pressure head generated is 1314 Pa

and discharge 0.777 m3/s. These values are well above the design point requisition. A

wide range is available for flow regulations.

Figures 6.40 and 6.41, shows the variation of air power and input power as a

function of discharge. The input and air power increases with discharge. Design point

lies centrally in these performance graphs to offer design and off design operating

range. At 2800 rpm, 1021 W airpower is generated at the cost of 1392 W input power

offering 73% efficiency.

Figure 6.42 shows the variation of static efficiency with respect to discharge.

This is basic performance characteristic curve, which highlights that at 0% to 50%

damping, the static efficiencies obtained are 73% and 75% at 2800 rpm. This result

also signifies that 0 to 50% damping position gives best efficiency region. It is

interesting to note that the present design offers design point performance in terms of

efficiency and discharge.

Figure 6.43 shows volute pressure distribution at 0% damping and 2800 rpm

conditions. This seems to be uniform along the volute passage and presenting uniform

diffusion of velocity except 240 and 300 degree volute casing position due to higher

velocities in this region. This may be possible due to recirculation shield provided at

scroll casing exit. Recirculation shield is covering 35 percent of exit area of scroll

casing in radial direction. The best operating range for this fan is at 0-50% damping at

2800 rpm as shown in Table 6.8.

Table 6.8 Optimum Operating Range of Backward Curved Radial Tipped Centrifugal Fan

Speed rpm Damp.

Fan Stage Pressure Rise

Pa

Avg. Air Discharge

m3/s

Static Air Power

W

Input Power

W

Static Eff. %

2800 0% 1314 0.777 1021 1392 73 2800 50% 1379 0.718 990 1322 75



Here it can be summarized that designed fan based on unified design is good

enough to achieve desired performance. With this design, flow regulation is also

possible which can provide flexibility to run fan at off design conditions too.

It is worth to mention that in the case of forward or backward curved radial

tipped centrifugal fan, the unified design thus stands experimentally validated.

6.5.3 Comparative Assessment of FCRT and BCRT centrifugal fans

Comparative performance assessment of forward and backward curved radial

tipped centrifugal fans at 2800 rpm is presented in Table 6.9.

Table 6.9 Optimum Operating Range of FCRT and BCRT Centrifugal Fan

Fan Speed rpm Damp.

Fan Stage Pressure Rise

Pa

Avg. Air Discharge

m3/s

Static Air Power

W

Input Power

W

Static Eff. %

Design Pt. 2800 0% 981 0.500 490 830 74

FCRT 2800 0% 1522 0.865 1316 1888 69.7 FCRT 2800 50% 1544 0.788 1216 1746 69.6 BCRT 2800 0% 1314 0.777 1021 1392 73 BCRT 2800 50% 1379 0.718 990 1322 75

It is concluded from comparative assessment that at same speed of 2800 rpm,

pressure head developed by the forward curve radial tipped centrifugal fan is higher

than backward curved radial tipped centrifugal fan. Hence if there is requirement of

higher stage pressure rise, it is better to go for forward curved centrifugal fan.

Similarly backward curved radial tipped fan consumes less power as compared to

forward curved radial tipped centrifugal fan. The efficiency for the backward curved

fan is comparatively higher than the forward curved fan at similar conditions as

confirmed in literature [10].

These results clearly show that fan based on unified design is good enough to

achieve desired performance. Major performance parameters achieved are on higher

side of design point. Flow regulation is possible between the operating ranges of 0 to

50% damping at 2800 rpm. This can provide flexibility to run fan at off design

conditions, too. This fan is consuming less input power compared to Osborne design

and at the same time gives better efficiency.

Thus, it may be stated that unified design methodology outlined during the

course of present work may be accepted as the experimentally validated design for



radial tipped centrifugal fan, which can confidently offer design point performance.

Further, it is most important to note that the performance of the fan achieved as per

unified design surpasses the efficiency target set for centrifugal fans by CSWD [4].

6.6 Phase V: Assessment of Theoretical and Experimental Losses

Centrifugal fan has to be designed for maximum efficiency. To achieve this

objective, there is a need for proper loss estimation. On account of the relatively

longer flow passages and greater turning of flow, the centrifugal stages, suffer higher

losses as compared to axial type.

There are various types of losses occurring when the fluid passes from inlet

duct to outlet duct of a centrifugal fan. The major losses are classified into three

categories as hydraulic or pressure losses, mechanical or power losses and volume

flow or leakage losses. Hydraulic losses reduce the available pressure head developed

by the impeller thereby reducing the Hydraulic efficiency. Mechanical losses are

encountered mainly due to disc Friction and friction between rotating shaft and the

journal bearing. Leakage losses reduce the quantity of fluid delivered per unit time

and hence reduce the volumetric efficiency. The stagnation efficiency of any machine

depends on the various losses occurring in the machine at different stages.

During this course of work, Centrifugal fan with 203 mm outside diameter

backward curved radial tipped impeller and associated scroll casing are fabricated.

Impeller is fabricated from mild steel sheet material and scroll casing is made

transparent by using acrylic sheet material. Fan is designed as per as per unified

design methodology and applying fan affinity laws. Static and velocity pressure heads

at design and off design speed conditions are measured by specially developed and

calibrated five holes probe. Other parameters are recorded simultaneously. Entry and

exit impeller velocity triangles are prepared at different peripheral locations. Slip

factor is calculated at all volute locations for different speed conditions.

Comparative assessment of theoretical estimation of hydraulic, leakage and

power losses suggested by various researchers and experimentally measured losses

are given in Annexure C. Velocity measurements and slip factor calculations for

backward curved radial tipped impeller fan are also given in Annexure C.

The measured dimensions, velocity and pressure parameters at inlet and outlet

conditions are given in Annexure C.



As per literature review, estimated losses proposed by different researchers

differ widely. Hence experimental investigations are essential to ascertain real

performance of radial tipped centrifugal fan.

Overall theoretical analysis of losses indicates that hydraulic losses are the

major source of all kind of losses occurring in radial tipped centrifugal fan. Leakage

and mechanical losses are comparatively less than hydraulic losses in terms of

percentile.

Further experiments are carried out to measure hydraulic, leakage, mechanical

losses. Experimental and analytical results are tabulated and given in Annexure C.

When the speed of rotor is reduced, hydraulic losses are significantly reduced. There

is also reduction in leakage and mechanical losses. This leads to the conclusions that

low velocity fluid causes less centrifugal fan losses.

According to A. H. Church [26], among the impeller losses, turbulence in the

impeller vane inlet is the major source of loss. Most of the volute losses are due to

friction and turbulence in the discharge passages. As per W. C. Osborne [28],

impeller entrance loss is the maximum among the all other impeller losses. His

method to estimate volute casing losses is very conservative. It offers very little volute

losses compared to actual one. Eck Bruno’s [14] methodology finds that major

components of impeller losses are impeller entrance and retardation losses. While

D. J. Myles [33] said that leakage losses are much higher when compared to

theoretical estimations. Volute loss is one of the major sources of hydraulic losses and

their contribution is equally important to impeller losses.

According to A. H. Church, W. C. Osborne, Eck Bruno and D. J. Myles,

theoretical leakage loss is 2.52%, 14.47%, 15.79% and 31.75%, respectively for the

fan under consideration. Experimental value of leakage loss is found 4% of design

flow rate.

Similarly A. H. Church, W. C. Osborne, Eck Bruno and D. J. Myles observe

theoretical hydraulic losses as 38.27%, 25.26%, 26.89% and 31.45%, respectively,

while experimentally it is in the vicinity of 35.59%.

The mechanical losses as per A. H. Church, W. C. Osborne, Eck Bruno and D.

J. Myles are 16.32%, 1.57%, 10.18% and 10%, respectively. Experimentally it is

found as 12% of total power consumption. Duct friction losses are negligible as being

less than 1%.



The comparison of experimental and theoretical evaluations reveals that there

is no uniformity between experimentally determined and theoretically derived leakage

losses. Experimental hydraulic and mechanical losses are showing uniformity with the

major theoretical values obtained.

Overall analysis on losses indicates that hydraulic losses are the major source

of losses in radial tipped centrifugal blower/fan. Leakage and mechanical losses are

comparatively less. According to all researchers impeller entrance losses are major

source of total impeller losses.

Among all hydraulic losses, impeller losses are experimentally found to be

68% while volute casing losses are observed 31% of total hydraulic losses. This

confirms the study of Andre Kovats [27] and R. J. Kind [60]. Volute losses are also

significant and are one of the major sources of hydraulic losses and their contribution

is equally important to impeller losses. This acknowledges the work of Y. Senoo and

H. Hayami [98], stating that 30% or more kinetic energy at diffuser exit remains

unconverted to pressure energy.

When the speed of rotor is reduced, hydraulic losses are significantly reduced.

There is also reduction in leakage and mechanical losses. This leads to the

conclusions that low velocity fluid causes less centrifugal fan losses.

These facts lead to the conclusion that impeller and volute losses are to be

minimized to improve hydraulic efficiency of the centrifugal fans. Vaned diffuser will

help to reduce eddies in volute casing and hence reducing volute losses.

Comparison of theoretical and experimental losses is presented in Figure 6.44.

If we consider all type of losses as 100%, than contribution of leakage, hydraulic and

mechanical losses are shown in Figure 6.45 by pie chart for experimental losses.

Figure

The

minimized

help to redu

Furt

centrifugal

radial tippe

1

1

2

2

3

3

4

4

% o

f tot

al lo

ss

6.44 Comp

Figu

ese facts lea

to improve

uce eddies i

ther experi

fan under s

ed centrifug

0.00

5.00

0.00

5.00

0.00

5.00

0.00

5.00

0.00

5.00

A.H.C

Le

Hy

Me

“Studies o

parative An

ure 6.45 Dis

ad to the c

hydraulic e

in volute ca

imental inv

study. Phase

al fan, fabri

Church WOs

Comp

Leakage lo

eakage loss

ydraulic loss

echanical loss

on Radial Tipp

nalysis of T

stribution o

onclusion t

efficiency o

sing and he

vestigations

e V experim

icated as pe

W.C. sborne

Ec

parative a

oss Hydr

23%

Expe(% Sh

s

ped Centrifuga

Theoretical

of Experim

that impelle

of the centri

ence reducin

are made

ments were

er unified de

ck Bruno D

analysis of

raulic loss

69%

erimental hare in Tota

Chapter – 6:

al Fan”

and Exper

mental Losse

er and volu

fugal fans. V

ng volute lo

to find ou

made on 8”

esign metho

D.J.Myles E

f losses

Mechanica

8%

Lossesal Losses)

Results and D

rimental Lo

es

ute losses a

Vaned diffu

sses.

ut slip fact

” impeller d

odology.

Experimenta

al loss

Discussion

281

osses

re to be

user will

tor in a

diameter

al



Velocity measurements and slip factor calculations for backward curved radial

tipped impeller fan are given in Table 6.10. It also shows comparison of experimental

and empirical values of slip factor.

Table 6.10 Phase V: Experimental Observation Table with Comparison of Experimental and Empirical Value of Slip Factors (Z=16, N=2800 rpm)

Location Peripheral

velocity (m/s)

Flow angle

(degree)

Measured

Absolutevelocity

(m/s)

Whirl Velocity

(m/s)

Exp. Slip

factor

Empirical Values by

Balje Stodola Stanitz

A 31.43 21 17.66 16.49 0.52

0.796 0.804 0.876

B 31.43 44 20.77 14.94 0.48 C 31.43 23 22.26 20.49 0.65 D 31.43 19 24.71 23.36 0.74 E 31.43 20 26.97 25.34 0.81 F 31.43 22 27.18 25.2 0.80

Average slip factor 0.67

The average value of slip factor is found 0.67. This is less than the empirical

value of 0.8 considered for design purpose. Experimental value of slip factor is found

16.7% less than empirical value suggested by Stodola [31], 16% less than suggested

by Balje and 24% less than suggested by Stanitz [9]. Figure 6.46 shows comparative

analysis of empirical and experimental slip factor for 16 numbers of blades at speed of

2800 rpm.


0.45

0.5

0.55

0.6

0.65

0.7

0.75

0.8

0.85

0.9

0 1 2 3 4 5 6 7

Experimetal

Balje

Stodola

Stanitz



Slip factor has closer agreement at design point conditions but deviates at off

design conditions. This indicates that there is more deviation in actual and the

theoretical flow direction at off design conditions. This also confirms study of

Mohamad Memardezfouli and Ahmad Nourbakhsh [87] of comparing experimental

slip factors with the calculated theoretical values. Similarly wide difference was

noticed by Frank Kenyery [83] who concluded that slip factor varies with the flow

rate and even in the case of the nominal flow rate; values for the slip factor produced

by correlations could have errors as large as 52%.

As per conclusions of phase II experiments having 12”impeller diameter

centrifugal fan, there was 3 to 12% experimental slip factor deviation compared to

empirical co-relations. In present case of experiments having 8” impeller diameter,

deviation of slip factor ranges from 16% to 24% compared to empirical values. It

confirms that shorter blade passage height will produce more slip. More

deviation of exit velocities is observed in small size impellers. This is due to absence

of proper flow guidance within blade passage. Blade height is an important parameter

and should be considered while estimating slip factor [22, 81].

The present studies on determination of experimental values of slip factor for

two different designs of impeller with varying speed, flow, number of blades and

suction resistance clearly highlights the lacuna of predictive capabilities of existing

slip factor correlations in the literature and focuses towards a need of more accurate

model for evaluation of slip factor.

Documents

RESULTS AND DISCUSSION - Shodhgangashodhganga.inflibnet.ac.in/bitstream/10603/37646/15/15_chapter 6.pdf · achieves maxima at 0.338 m3/s at 429 N/m2 suction pressure in case of 8