Calculation Analysis of Energy Saving Tools for Fan and Pump Applications
194
LAPPEENRANNAN TEKNILLINEN YLIOPISTO Faculty of Technology Energy Technology MASTER’S THESIS CALCULATION ANALYSIS OF ENERGY SAVING TOOLS FOR FAN AND PUMP APPLICATIONS Examiners: Professori Esa Marttila TkL Simo Hammo Supervisor: M.Sc. Jukka Tolvanen Lappeenranta x.x.2008 Teemu Taskinen Huhtiniemenkatu 12 as 6 53810 Lappeenranta Gsm. +358 40 568 1696
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Transcript of Calculation Analysis of Energy Saving Tools for Fan and Pump Applications
LAPPEENRANNAN TEKNILLINEN YLIOPISTO Faculty of Technology Energy Technology MASTER’S THESIS
CALCULATION ANALYSIS OF ENERGY SAVING TOOLS FOR FAN AND PUMP APPLICATIONS
Examiners: Professori Esa Marttila TkL Simo Hammo Supervisor: M.Sc. Jukka Tolvanen
Lappeenranta x.x.2008 Teemu Taskinen Huhtiniemenkatu 12 as 6 53810 Lappeenranta Gsm. +358 40 568 1696
ABSTRACT Lappeenranta University of Technology Faculty of Technology Energy technology Teemu Taskinen Calculation analysis of energy saving tools for fan and pump applications Master’s thesis 2008 189 pages, 92 figures, 31 tables and 7 appendices Examiners: Professor Esa Marttila Lic.Sc. (Tech.) Simo Hammo Keywords: FanSave, PumpSave, fan, pump, frequency converter This thesis analyses the calculation of FanSave and PumpSave energy saving tools calculation. With these programs energy consumption of variable speed drive control for fans and pumps can be compared to other control methods. With FanSave centrifugal and axial fans can be examined and PumpSave deals with centrifugal pumps. By means of these programs also suitable frequency converter can be chosen from the ABB collection. Programs need as initial values information about the appliances like amount of flow and efficiencies. Operation time is important factor when calculating the energy consumption and information about it are the length and profile. Basic theory related to fans and pumps is introduced without more precise instructions for dimensioning. FanSave and PumpSave contain various methods for flow control. These control methods are introduced in the thesis based on their operational principles and suitability. Also squirrel cage motor and frequency converter are introduced because of their close involvement to fans and pumps. Second part of the thesis contains comparison between results of FanSave’s and PumpSave’s calculation and performance curve based calculation. Also laboratory tests were made with centrifugal and axial fan and also with centrifugal pump. With the results from this thesis the calculation of these programs can be adjusted to be more accurate and also some new features can be added.
TIIVISTELMÄ Lappeenrannan teknillinen yliopisto Teknillinen tiedekunta Energiatekniikka Teemu Taskinen Pumppu- ja puhallinlaitteistojen energiansäästötyökalujen laskennan analysointi Diplomityö 2008 189 sivua, 92 kuvaa, 31 taulukkoa ja 7 liitettä Tarkastajat: Professori Esa Marttila TkL Simo Hammo Hakusanat: FanSave, PumpSave, puhallin, pumppu, taajuusmuuttaja Keywords: FanSave, PumpSave, fan, pump, frequency converter Diplomityön tarkoituksen on tarkastella energiansäästötyökalujen, FanSave ja PumpSave, laskentaa. Ohjelmilla voidaan tarkastella puhaltimien ja pumppujen taajuusmuuttajakäytön energian kulutusta ja vertailla sitä muihin säätötapoihin. FanSave –ohjelmalla voidaan käsitellä keskipakois- ja aksiaalipuhaltimia ja PumpSave –ohjelmalla keskipakoispumppuja. Ohjelmien avulla voidaan myös valita tarkasteltavalle puhaltimelle ja pumpulle sopiva taajuusmuuttaja ABB:n mallistosta. Ohjelmiin täytyy syöttää alkuarvoina tietoja laitteistosta, kuten virtauksen määrä ja hyötysuhteita. Laitteiston käyttöaika on tärkeä tekijä energian kulutusta laskettaessa ja sille voidaan määrittää kokonaiskesto ja profiili. Diplomityössä esitellään puhaltimien ja pumppujen toimintaan liittyvä perusteoria ilman tarkempia selvityksiä mitoitusperiaatteista. FanSave ja PumpSave –ohjelmat sisältävät monia eri säätötapoja virtaukselle. Nämä säätötavat esitellään työssä niiden toimintaperiaatteiden ja soveltuvuuden osalta. Myös puhaltimien ja pumppujen käyttöön olennaisesti liittyvät oikosulkumoottori ja taajuusmuuttaja esitellään työssä. Työssä vertaillaan energiansäästöohjelmien laskentaa puhaltimien ja pumppujen ominaiskäyrästöjen perusteella laskettuihin tuloksiin. Vertailupohjaa ohjelmien laskentaan saatiin myös suorittamalla laboratoriotestejä keskipakois- ja aksiaalipuhaltimilla sekä keskipakoispumpulla. Työssä saatujen tuloksien perusteella ohjelmien laskentaa voidaan säätää tarkemmaksi sekä lisätä siihen ominaisuuksia.
TABLE OF CONTENTS
1. INTRODUCTION ..............................................................................................................5 1.1 Background & objectives..............................................................................................5
1.2 Execution of the work...................................................................................................5
2. FANS ..................................................................................................................................6 2.1 Structure and operation of a centrifugal fan .................................................................6
2.1.1 Forward curved blades...........................................................................................7
2.1.2 Backward curved blades ........................................................................................8
2.1.3 Radial blades..........................................................................................................8
2.2 Structure and operation of an axial fan .........................................................................9
2.2.2 Axial fan with fixed blades..................................................................................10
2.2.3 Adjustable axial fans............................................................................................10
2.3 Fan theory ...................................................................................................................13
2.3.1 Specific fan power ...............................................................................................18
2.3.2 SFS ISO 5167 standard........................................................................................18
2.3.1 Characteristic curves for a fan .............................................................................21
2.4 Fan flow control methods ...........................................................................................21
2.4.1 No flow control....................................................................................................22
2.4.2 Rotational speed control ......................................................................................23
2.4.3 Outlet damper ......................................................................................................24
2.4.4 Slip coupling........................................................................................................25
2.4.5 Voltage control ....................................................................................................25
2.4.6 Two speed motor .................................................................................................26
2.4.7 Cycling on/off ......................................................................................................26
2.4.8 Inlet box damper ..................................................................................................26
2.4.9 Inlet vanes for centrifugal fans ............................................................................27
2.4.10 Pitch angle adjustment in axial fans ..................................................................28
3. CENTRIFUGAL PUMPS.................................................................................................28 3.1 Structure and operation of centrifugal pump ..............................................................29
3.2 Pump hydraulics .........................................................................................................30
3.2.1 Pump power demand and efficiency....................................................................30
3.2.2 System head and net positive suction head..........................................................31
3.2.3 Performance curves and affinity laws..................................................................34
3.3. Flow control methods in pumping .............................................................................36
3.3.1 Variable speed control .........................................................................................37
3.3.2 Throttling control.................................................................................................38
3.3.3 On/off –control ....................................................................................................39
3.3.4 Other control methods .........................................................................................39
4 SQUIRREL GAGE MOTOR ............................................................................................40 4.1 Structure of squirrel gage motor .................................................................................41
4.2 Motor losses................................................................................................................41
5. FREQUENCY CONVERTER .........................................................................................42 5.1 Drive losses.................................................................................................................44
6. FANSAVE AND PUMPSAVE ENERGYSAVING TOOLS..........................................44 6.1 FanSave.......................................................................................................................45
6.1.1 Initial values.........................................................................................................45
6.1.2 Calculations .........................................................................................................46
6.2 PumpSave ...................................................................................................................50
6.2.1 Initial values.........................................................................................................50
6.2.2 Calculations .........................................................................................................51
7. COMPARISON BETWEEN FANSAVE’S CALCULATION AND PERFORMANCE CURVE BASED CALCULATION .....................................................................................55
7.1 Centrifugal fans...........................................................................................................56
7.1.1 Radial blades........................................................................................................57
7.1.2 Fans with forward curved blades. ........................................................................62
7.1.3 Fans with backward curved blades ......................................................................67
7.2 Axial fans ....................................................................................................................76
8. LABORATORY TEST RUNS WITH CENTRIFUGAL FAN........................................89 8.1 Laboratory equipment.................................................................................................89
8.2 Measurements and comparison of results to FanSave’s calculation...........................91
8.2.1 Outlet damper ......................................................................................................91
8.2.2 Variable speed control .........................................................................................92
8.2.3 Two speed motor .................................................................................................94
8.2.4 Inlet box damper ..................................................................................................94
8.3 Energy consumption examination in operation time ..................................................96
9 LABORATORY TEST RUNS WITH AN AXIAL FAN .................................................96 9.1 Laboratory equipment.................................................................................................97
9.2 Variable speed drive ...................................................................................................97
9.3 Outlet damper .............................................................................................................98
9.4 Inlet damper ................................................................................................................99
10 PERFORMANCE CURVE BASED CALCULATION COMPARISON TO PUMPSAVE’S CALCULATION ......................................................................................101
10.1 VSD ........................................................................................................................101
10.2 Throttling ................................................................................................................115
11 LABORATORY TEST RUNS WITH A PUMP...........................................................129 11.1 VSD ........................................................................................................................130
11.2 Throttling ................................................................................................................133
12 SUMMARY...................................................................................................................137 12.1 Observation of the profitability of investment........................................................138
APPENDICES
3
SYMBOLS AND ABBREVIATIONS
Greek letters
β Diameter ratio
ε Expansibility factor
λ Friction-coefficient
ρ Density [kg/m³]
ζ Loss-coefficient
η Efficiency
κ Isentropic exponent
μ Dynaaminen viskositeetti [Ns/m]
υ Kinetic viscosity
ω angular speed [rad/s]
Roman letters
A Area [m²]
C Discharge coefficient
c Heat capacity [J/kgK]
D Pipe diameter [m], [mm]
E Energy [kWh]
d Orifice plate diameter [m], [mm]
g Gravitational factor [m/s²]
h Enthalpy [J/kg]
l
L quotient of a distance
n Rotational speed [rpm]
NPSH Net Positive suction head [m]
P Power [W], [kW]
p Pressure [Pa]
q Flow rate [m³/s], [kg/s]
R Universal constant for gas [J/kgK]
r Radius [m], [mm]
4
T Temperature [K]
u Tangential speed [m/s]
v Flow velocity [m/s]
Subscript
0 initial
1 inlet
2 outlet
e electricity
i flow amount
m mass
r friction
s isentropic
v volume
vsd variable speed drive
Acronyms
SFP Specific fan power
VSD Variable speed drive
5
1. INTRODUCTION
Industry uses lots of different types of pumps and fans in needs of different processes.
These devices run long times at fast speeds. All this needs lots of energy and use of pumps
and fans cover a significant part of industrys electricity bill. Therefore it is essential that
the use of pumps and fans is done as efficiently as possible.
Flowing in pumps and fans can be controlled in many different solutions but variable speed
drive has begun to become more and more common. This is caused by advancements in
technology and cost reduction in investment. With variable speed drive pumps and fans
can be used precisely matching the needs of the process. Therefore the whole process can
be controlled more accurately. Also the wear and noise on machinery is reduced. By using
variable speed drive energy can be saved and therefore also the emissions are reduced. In
favour of variable speed drive are also constantly increasing price of electricity and
tightening emission limits.
1.1 Background & objectives
The purpose of this thesis is to analyse the calculations of two Excel based energy saving
tools called FanSave and PumpSave by ABB Oy. With these tools the different flow
control methods for pumps and fans can be compared. These programs calculate the costs
of different flow control methods compared to variable speed drive.
For fans and pumps basic theory is introduced but any specific dimensioning of devices is
not processed because either FanSave or PumpSave do not include or demand so specific
information. Also parallel connection or series connection of fans and pumps is not dealt
with in this thesis because FanSave and PumpSave examine only one application at a time.
1.2 Execution of the work
Thesis starts with introduction to structure and operation of fans and pumps. Also the basic
theory of pumping and blowing is introduced. With FanSave axial and centrifugal fans can
6
be examined so these are the fan types studied in this thesis. PumpSave deals only with
centrifugal pumps so other pump types were left aside.
Flow during pumping and blowing can be adjusted in many different ways and the most
common ones of these are examined and compared to FanSave’s and PumpSave’s
calculation in different ways. All the other flow control methods and their operation
principles are also introduced.
Next part of the thesis consits of introduction of energy saving tools FanSave and
PumpSave. Calculations of both of the programs are presented in equations. Results from
the calculations of these programs are compared to calculations based on different types of
fans and pumps performance curves and to tests performed in laboratory. Laboratory tests
were performed in laboratories of Lappeenranta university of technology. All the flow
control methods included in the programs could not be tested due to limitations of
laboratory equipment and rareness of control methods. Results are presented in tables and
as curves for clarification. Thesis ends with conclusions and proposals for improvements
and corrections for the energy saving tools.
2. FANS
Fan is the most important part of a ventilation system. It is used to transfer air through
ducts and spaces. Air movement is caused by a rotating impeller which is installed with
blades. Fans are chosen by the purpose of use. Most important factors affecting the choice
are the volume flow and pressure difference generated by the fan. For the most powerful
fans also the changes in density and temperature of the fluid should be noticed when
estimating the power demand of the fan. (Seppänen, 121)
2.1 Structure and operation of a centrifugal fan
Flow enters the centrifugal fan in the direction of the axel and leaves the fan at
perpendicular to the axel. Three most common types of blades are forward curved blades,
backward curved blades and radial blades. (Seppänen, 121)
7
2.1.1 Forward curved blades
Characteristic for an impeller of a fan which has forward curved blades is that the impeller
is small-sized and has low rotational speed and noise level for certain pressure and volume
flow. Change in characteristic curve of a fan causes only a small change in pressure
difference between inlet and outlet of a fan. Solid matter particles can easily stuck to a
blade which is curved forward and that is why these types of blades are not used for
blowing combustion gases. (Puhaltimet teollisuudessa ja LVI-laitoksissa, II 1, Fan Types)
Pressure and power demand curve of a centrifugal fan with forward curved blades are
illustrated in figure 1. At low flow rates the operation of fan is unstable this is shown in
pressure curve which is descending at this operation area. Using fan in this operation area
is not suitable but at the operation range where pressure curve is ascending. Instability of
the system is shown also in the section angle of the system curve and the pressure curve
which is small. If high pressure generation is needed without changing the radial velocity
forward curved blades are a good option. Fan with forward curved blades can reach the
efficiency of about 60 %. (VLV-kone luentomoniste 1, s 67 ; Systemair)
Figure 1. Pressure curve for a centrifugal fan with forward curved blades. (VLV-kone luentomoniste 1, sivu
68)
8
2.1.2 Backward curved blades
Impeller which has backward curved blades it is characteristic to have high efficiency and
higher noise level than an impeller with forward curved blades. When fan’s characteristic
curve changes the volume flow doesn’t really change. Impeller with backward curved
blades is suitable for cases where low running costs are demanded. This type of impeller is
also sensitive for gathering particles from gas. For energy efficiency the backward curved
blades are better choice than forward curved blades or radial blades. (Puhaltimet
teollisuudessa ja LVI-laitoksissa, II 1, Fan Types)
Figure 2 illustrates pressure and power demand curves for a fan with backward curved
blades. This type of fan is able to function at wide operation range. Fan with backward
curved blades can reach the efficiency of 80 %. Power demand is at its highest at the same
area the efficiency is at its highest. (VLV-kone luentomoniste 1, 67 ; Systemair)
Figure 2. Pressure curve for a centrifugal fan with backward curved blades. (VLV-kone luentomoniste 1, s
68)
2.1.3 Radial blades
9
Fan with radial blades is the best choice for blowing gas which has lots of solid particles
because radial blades are better to resist particles sticking on them than curved blades.
Radial blades are straight pointing directly to the ring of the impeller. Fan equipped with
radial blades can gain maximum efficiency of 55 %. (Fan Types, Systemair)
2.2 Structure and operation of an axial fan
Axial fan is suitable for many applications because it is easy to install and it is small-sized
and cheap. Characteristic for axial fan is big volume flow with moderate pressure
development. Two phased axial fan has become more general due to its better pressure
development abilities. Two phased axial fan consists of two series-connected fans which
rotate to opposite directions. This solution generates almost three times as much pressure
as a single axial fan. When volume flow of air is big axial fan is a suitable choice for
almost any general ventilation applications. Pressure generation in fan can be squared in
function of impeller diameter. (Puhaltimet teollisuudessa ja LVI-laitoksissa, I 2-3, 10)
Erään aksiaalipuhaltimen, jonka potkurin siipikulma on suuri, painekäyrä on kuvassa 3.
Kuvassa nähdään sakkauskohta, jossa tehontarve käyrä saavuttaa huippukohdan.
Puhaltimen käynti sakkauskohdan vasemmalla puolella on epävakaata. Puhallinta tulisi
käyttää alueella, jossa painekäyrä on nouseva. Kuvassa 3 näkyvä tehontarvekäyrä on
riippuvainen impellerin siipiprofiilista eikä siksi päde kaikille aksiaalipuhaltimille.
Joillakin puhaltimilla tehontarve voi olla pienellä virtauksella lähes kaksinkertainen
sakkauspisteen tehontarpeeseen verrattuna. Tällöin on vaarana käyttömoottorin
ylikuormittuminen. Aksiaalipuhaltimen hyötysuhde on normaalisti maksimissaan n. 75 %,
mutta johtosiivillä se voidaan nostaa n. 85 %:iin. (VLV-kone luentomoniste 1, s 67)
10
Figure 3. Pressure curve for an axial fan. (VLV-kone luentomoniste 1, s 68)
2.2.2 Axial fan with fixed blades
Axial fan with fixed blades is suitable for duct installation. Lack of flow adjusting abilities
limits the fan’s suitability to applications where power consumption and noise level are not
important selection criteria. (Puhaltimet teollisuudessa ja LVI-laitoksissa, I 1)
2.2.3 Adjustable axial fans
There are types of axial fans which blade angle can be adjusted during operation or during
downtime. With downtime adjustment the adjustment needs the fan to be stopped but when
the fan operates at its optimum operation area its power consumption is small. It is suitable
for any industrial ventilation applications where axial fan’s pressure development abilities
are sufficient
Some axial fans have the possibility to adjust blade angles during operation. With this type
of adjustment the pressure and volume flow values can be rapidly changed to match
processes needs. Adjustment can be done with power or compressed air. This type of fan is
suitable for applications where maximum volume flow or maximum pressure is needed
only momentarily. (Puhaltimet teollisuudessa ja LVI-laitoksissa, I 1-2)
11
Centrifugal and axial fan have their own characteristic operational areas where they are
more suitable than other types of fans. Figure 4 illustrates the operational ranges of the
most common fan types. Uppermost operation area belongs to centrifugal fan which is
suitable for smaller volume flows but then again for higher volume flows and higher
pressure differences. Middle operation area belongs to two phased axial fan which can
produce average pressure difference in high flow rates. Lowest operation area belongs to
one phase axial fan. In practice these operational ranges can not be divided as clearly as in
figure 4 but they overlap each other more or less. Also in some situations some other fan is
more suitable to be used in some others nominal operation range.
Figure 4. Characteristic operation areas for different types of fans. (VLV-kone luentomoniste 1, s 68)
Typical efficiency curves for different types of fans are illustrated in figure 5. Total
efficiency of the fan means the efficiency calculated by the theoretic power demand which
is calculated by using the total pressure generated by the fan compared to the power
provided to the impeller. Efficiencies calculated for the curves in figure 5 are average
values from different types of fans so they may seem a little low. Upmost dash line
represents the efficiency of a centrifugal fan with backward curved blades and axial fan.
Lower dash line illustrates the efficiency of a centrifugal fan with forward curved blades.
Curve does not reach the highest volume flows because this type of fan is not usually
12
manufactured for such operation areas. Solid lines in figure 10 illustrate total efficiency of
blowing. This means how many percentages of electric power from the network can be
transformed into motion and pressure of fluid. Efficiency of the motor is presented as
second to upmost line. (VLV-kone luentomoniste 1, 68-69)
Figrue 5. Efficiencies of different types of fans. (VLV-kone luentomoniste 1, s 69)
Centrifugal fan’s impellers can be divided into three groups according to curvature of
blades. Blades can be bent backward or forward or they can be radial. Due to curvature the
wheels are described as F-wheel, B-wheel and R-wheel. Figure 6 illustrates these different
solutions for fan impellers. First impeller from the left is impeller with backward curved
wheels where angle is below 90°. One in the middle is impeller with radial blades.
Blade angle equals 90°. Last one illustrates forward curved blades and the blade angle is
over 90°. The best efficiency can be gained with backward curved blades as radial blades
give the most power. Forward curved blades are the best choice if maximum pressure
generation is necessary. (Puhaltimet teollisuudessa ja LVI-laitoksissa, II 1, VLV-kone
luentomoniste 1, s 57)
2β
13
Figure 6. Different shapes of fan impeller blades and velocity triangles. (VLV-kone s 57)
2.3 Fan theory
Many factors affects on power demand of a fan. According to equation 1 main factors are
volume flow and pressure difference between inlet and outlet of the fan. During blowing
density and temperature of gas changes and these factors must be noticed too. Following
part of this thesis deals with theory involved with fans.
Isentropic power demand for a fan can be calculated as equation 1 shows. In this case the
density of the gas is constant.
vs ΔpqP = (1)
where
isentropic power demand [W] sP
pressure rise in fan [Pa] pΔ
Isentropic power demand can be expressed as shown on equation 2 when density change in
gas has been noticed.
ρ
qΔpP ms = (2)
where
mass flow [kg/s] mq
14
Average value of density can be calculated with equation
2ρρ
ρ 10 += (3)
where
0ρ density of the gas before the fan [kg/m³]
1ρ density of the gas after the fan [kg/m³]
Change in enthalpy can be calculated as equation 4 or equation 5 shows.
s
phηρΔ
=Δ (4)
where
sη isentropic efficiency
Common value for isentropic efficiency for fans vary from 0,6 to 0,85.
s
shh
ηΔ
=Δ (5)
where
shΔ isentropic change in enthalpy [kJ/kg]
Common theory of fluid machinery is also valid for fans. According to this isentropic
change in enthalpy can also be presented with equation
( )1u2uhs ucucηh12
−=Δ (6)
where
hη hydraulic efficiency
15
tangential component for absolute speed at inlet edge [m/s] 1uc
tangential component for absolute speed at discharge edge [m/s] 2uc
tangential speed at inlet edge [m/s] 1u
tangential speed at discharge edge [m/s] 2u
Components and are 0 when the air sucked into the fan is for example outside air
and air is not blown into the inlet of the fan. Component depends on the shape of the
impeller blades, the blade angle and the tangential speed as shown in figure 9.
Hydraulic efficiency contains leakage from discharge outlet to suction inlet, secondary
twirls and wall friction.
1uc 1u
2uc
2β 2u
Angular speed for fan can be calculated as equation 7 shows when rotation speed of the fan
is known.
n (7) 2πω =
where
angular speed [rad/s] ω
n rotation speed of the fan [rpm]
Tangential speed at discharge edge of the impeller can be calculated as equation 8 shows
when radius of the impeller is known.
(8) 22 ωru =
where
radius of the impeller [m] 2r
Gas warms up when it flows through a fan because losses at the flow through the impeller
are transformed into heat. How much gas warms up can be calculated by marking the loss
power and the heat power equal.
16
( )vp0m
msvp0m
m
s
qc)1(P
ttqc1P
ρηη
ρηη
−=Δ→Δ=− (9)
where
mechanic efficiency of the fan mη
specific heat capacity for gas [kJ/kgK] pc
temperature rise [°C] tΔ
Because the power transferred to gas is product of pressure rise and volume flow can
equation 9 be written as equation
mp0c
ptηρ
Δ=Δ (10)
Temperature after the fan can also be calculated as equation 11 shows if the enthalpy
change in the fan can be calculated.
p01 c
hTT Δ+= (11)
where
gas temperature before the fan [K] 0T
gas temperature after the fan [K] 1T
Density of the gas can be resolved as equation 12 shows when temperature and pressure
are known.
RTp
=ρ (12)
where
17
R universal constant for gas [kJ/kgK]
Actual power transferred to gas can be calculated with equation
mpms
m qTchqηρ
ΔpqP Δ=Δ== (13)
where
P power transferred to gas [W]
Fan’s shaft power demand can be calculated with equation
maks η
PP = (14)
where
shaft power demand [W] aksP
Mechanic efficiency or the fan contains losses in bearings and transmission.
Demand for electric power can be resolved from equation
em
e ηηPP = (15)
where
electric power demand [W] eP
efficiency of the motor eη
Total efficiency of the fan can be calculated with equation
(16) ems ηηηη =
18
where
total efficiency η
2.3.1 Specific fan power
Comparison of fans is made easier with SFP number which means specific fan power. It
defines how much power is needed to move a cubic metre of air. In other words it
determines the efficiency of air ventilation system regarding to the needed electric power.
For a single fan SFP can be expressed with equation
v
total
qP
SFP = (17)
where
SFP specific fan power [kW/(m³/s)]
electric power input to fan application [kW] totalP
Term contains the power demand of frequency converter or some other adjusting
device as well as the power demand of the driving motor. So it takes into account all the
losses in application. For ventilation systems with multiple fans the SFP number is
calculated otherwise. (SFP-opas)
totalP
2.3.2 SFS ISO 5167 standard
Standard SFS ISO 5167 was applied for laboratory test results from centrifugal and axial
fan test runs. Standard was used to calculate volume flow when observing fan’s power
demand in different flow control methods. ISO 5167 standard contains instructions to solve
different quantities with iteration.
SFS ISO 5167 standard can be used for calculating different factors of flow. Mass flow
measured with an orifice plate the mass flow can be calculated with equation
19
ρπεβ
p2d41
Cq 2
4m Δ−
= (18)
Where
β diameter ratio
pressure difference over the orifice plate pΔ
ε expansibility factor
C discharge coefficient
d diameter of the hole in the orifice plate
Volume flow can be calculated by using density of the fluid as equation 19 shows.
mv qq ρ= (19)
Discharge coefficient C is calculated with equation
( ) ( )
( ) ( ) 3,11,1'2
'24
4
L7L103,0
D
65,3
7,0
D
682
M8,0M031,01
A11,01
*e123,0e080,0043,0Re10A0063,00188,0
Re10000521,0216,00261,05961,0C
11
ββ
β
β
βββ
−−−
−
−++⎟⎟⎠
⎞⎜⎜⎝
⎛++
⎟⎟⎠
⎞⎜⎜⎝
⎛+−+=
−− (20)
Where
1L quotient of the distance of the upstream tapping from the upstream face of
the plate and the pipe diameter
8,0
Re19000
⎟⎟⎠
⎞⎜⎜⎝
⎛=
D
A β (21)
Term in equation 20 is calculated with equation '2M
20
β−=
12 '
2'2
LM (22)
Where '2L quotient of the distance of the downstream tapping from the downstream
face of the plate and the pipe diameter
The orifice plate used has corner tappings so equation 23 is valid
0'21 == LL (23)
Expansibility factor ε is calculated with equation
( )⎥⎥⎥
⎦
⎤
⎢⎢⎢
⎣
⎡
⎟⎟⎠
⎞⎜⎜⎝
⎛−++−=
κββε
1
1
284 193,0256,0351,01pp (24)
When
75,01
2 ≥pp
Where
κ isentropic exponent
pressure before the orifice plate 1p
pressure after the orifice plate 2p
In the iteration the unchangeable factors of equation 18 are marked as . 1A
41
12
11D
p2dA
βμ
ρε
−
Δ= (25)
Changing factor in the iteration is marked . 1X
21
1D1 CAReX == (26)
For discharge coefficient C at the first round of iteration is recommended to use value
which can be found from table in appendix C. Now Reynolds’s number can be calculated
with equation 27. With factor the first value to volume flow can be solved with
equation
∞C
1X
1m DX4
q μπ= (27)
Second round in iteration is started by calculating value for the discharge coefficient C
with the Reynolds’s number. Iteration continues by calculating the value for factor .
Iteration rounds are repeated until the value for volume flow has converged enough.
1X
2.3.1 Characteristic curves for a fan
Fan’s characteristic curve is presented as pressure difference in the function of volume
flow. Characteristic curve can also contain curves or areas for fan’s efficiency, power
demand, rotation speed and noise level. With the characteristic curve operational values of
a fan can be solved at different operation points and in different control methods.
Throttling of the flow with dampers and rotational speed control are the most commonly
used control methods for volume flow.
When flow is throttled changes in fan’s operation can be solved by moving through fan’s
curve to a new operation point. With rotation speed control the moving is performed
through the system curve. Characteristic curves for different types of fans can be found
from appendixes.
2.4 Fan flow control methods
Adjusting the gas flow in fans to match the needs of process saves energy. Throttling was
most commonly used control method when energy was cheap but nowadays technology
22
has had to come up with new solutions to control air flow. Figure 7 illustrates
dependencies of power demand to volume flow in different flow control methods. Fans
compared are operating near their best efficiency with axel power ranging from 20 to 30
kW. At the total pressure 800 Pa this means that volume flow ranges from 20 to 30 m³/s.
Resistance curve of the duct is assumed to be a second-order parable. In figure 18 curve 1
illustrates a lossless power transferred to gas in rotational speed control, curve 2 illustrates
a centrifugal fan with rotational speed control, curve 3 illustrates an axial fan with pitch
angle adjustment, curve 4 illustrates a centrifugal fan with inlet vanes, curve 5 illustrates a
centrifugal fan with throttling control and curve 6 illustrates an axial fan with throttling
control. (Seppänen, 146-147)
Figure 7. Dependency of power demand to volume flow. (Seppänen, 147)
2.4.1 No flow control
This flow control method is used only with small applications and it can be used with
many fan types like propeller fans, axial fans and centrifugal fans. The impeller of the fan
is installed on the axel of the motor. Therefore there is no transmission loss. The motor is
exposed to the gas going through the fan so it must endure all the temperatures of the gas.
23
Rotational speed of the fan can not be changed in this flow control method so it is very
important to choose the right type of fan for the application. (Seppänen, 142)
2.4.2 Rotational speed control
Applications with small power consumption speed control are carried out with the use of
thyristors. Larger applications use frequency converters. Thyristor based speed control is
cheap to implement but the receivable savings are not significant since motor power is
small. Thyristor based speed control is feasible for applications such as heat extraction
where gas flow and pressure generation through fan are small. Other control methods for
rotational speed are squirrel gage motor, which has several windings for different
rotational speeds, using a v-belt drive and changing belt pulleys, using a transmission and
using a hydraulic clutch. (Puhaltimet teollisuudessa ja LVI-laitoksissa, I 7, Seppänen, 146)
Use of frequency control can produce significant saving in energy because the volume
flow of the gas can be exactly matched with process’ needs. Other benefits of using
variable speed drive are that noise level drops when revs drop and that fan efficiency stays
high at the whole control range. Frequency converters have been used for bigger
applications for some time now but are becoming more and more general also in smaller
applications. Figure 8 illustrates how fan’s operational values change when rotation speed
changes. (Puhaltimet teollisuudessa ja LVI-laitoksissa, I 7-8 II 6, Seppänen, 146)
24
Figure 8. Effects on characteristic curves of a fan when rotation speed changes. (Seppänen, 146)
2.4.3 Outlet damper
Outlet damper means flow control method where adjustable dampers are installed after the
fan to control the volume flow. This method is simple and cheap to implement but it
increases flow resistance in the duct and can cause noise level increase. Outlet damper as a
flow control method can be observed as a throttling control. (Puhaltimet teollisuudessa ja
LVI-laitoksissa, II 5)
Figure 9 illustrates the effect outlet damper flow control method has on the operation point
of the fan. At operating point A the dampers are fully open and at point B dampers start to
close. Closing the dampers produces always a new system curve when the operating point
moves to the intersection point of the pressure curve and system curve. The pressure curve
of the fan does not change. (Inlet and outlet dampers for centrifugal fans)
25
Figure 9. Fan operating points with outlet damper flow control method. (Inlet and outlet dampers for
centrifugal fans)
2.4.4 Slip coupling
Slip coupling as a flow control method means any type of clutch which can be used to
adjust torque from motor to fan. Slip coupling is usual only in big applications, power
ranging over 150 kW. There are many solutions available for slip coupling control. A
suitable clutch or coupling is selected according to the needed torque and rotational speed.
Clutches that doesn’t have any contact surface use eddy currents or magnetic fields for
transmission. Between clutches discs there can also be liquid to transmit power. When this
kind of hydraulic clutch operates the clutch discs are pressed together by the use of spring
tension causing pressure rise in liquid. The bigger the pressure is in the liquid the more
power is transmitted through the clutch. Using slip coupling is a lossy way to control
volume flow because the power generated by the motor isn’t transferred to the fan in its
totality. (Torque Limiters and Slip Clutches Information on GlobalSpec, Seppänen, 143)
2.4.5 Voltage control
26
Fan’s rotational speed can be changed by changing the voltage input to motor. This is
usually implemented by using a triac which is a semiconductor. Triac can conduct current
to both directions. It consists of two parallel connected thyristors which have shared grid
control. Controlling voltage with a triac is not fast because of its structure. (Karjalainen, 2)
2.4.6 Two speed motor
Two speed motor has two individual windings and it can be used in two different rotational
speeds. Two speed motor as a flow control method is best suitable when the application the
motor powers has two demand levels for power. It is cheaper to purchase and install than a
variable speed drive.
Two speed motors are available in variable torque, constant torque and constant
horsepower versions depending on the application. These motors are about 60 % more
expensive than single-speed motors. Big two speed motors are not very efficient at low
loads. (Electric Ideas Clearinghouse, 1-3)
2.4.7 Cycling on/off
When flow control method is cycling (on/off) the fan and motor are running at the speed
corresponding to the frequency of the input voltage or they are not operating. This flow
control method is suitable for example in removal of gas. For example when sensor detects
an overrun in limiting value fan starts and removes the gas from the space. It shuts down
when the air in the space is gas free again. Continuous starting and stopping of the fan
wears down the machinery.
2.4.8 Inlet box damper
Flow control method inlet box damper uses same kind of dampers as the control method
called outlet damper but inlet dampers are installed before the fan. These dampers causes
whirl in gas flow and due to this fan can not generate maximal pressure and volume flow.
Fan’s power demand is reduced also due to inlet dampers and the result is similar to
27
reducing rotational speed in fan which has no flow control. (Inlet and outlet dampers for
centrifugal fans)
Inlet dampers’ effect on fan’s operating point is illustrated in figure 10. When dampers are
closed the volume flow to the fan is decreased and so is the power demand. In operation
point C in figure X the dampers are fully open and when operation point moves to points D
and E as the dampers are closed new pressure and power demand curves are created.
System curve remains the same. (Inlet and outlet dampers for centrifugal fans)
Figure 10. Centrifugal fan’s operating points with inlet damper. (Inlet and outlet dampers for centrifugal
fans)
2.4.9 Inlet vanes for centrifugal fans
Inlet vanes are blades installed before the impeller of fan to direct the flow entering the
fan. Vanes can be adjustable and when flow is decreased also fan’s power demand
decreases. Inlet vanes are cheap and easy to install and almost maintenance free flow
control method. This flow control method is used in applications which include relatively
big pressures and flow velocities. Inlet vanes generate a swirl which is parallel to the
direction impeller rotates. Disadvantages for inlet vanes are weakening efficiency in wide
28
control range. Also the fan’s noise level doesn’t decrease because the fan rotates at the
same speed regardless of the adjustment. (Puhaltimet teollisuudessa ja LVI-laitoksissa, II
5-6 ; Inlet and outlet dampers for centrifugal fans)
2.4.10 Pitch angle adjustment in axial fans
Pitch angle adjustment for axial fan changes the outlet flow from the fan. Angles of the
fan’s blades can be changed during operation so adjusting the flow doesn’t need the fan to
be stopped. If the angle is too big the fan or duct can be damaged. If the angle is too small
the fan can not transfer enough power to gas. Characteristic curves in appendixes F1, F2,
F3 and F4 illustrates the effect pitch angle adjustment has on to the fan’s characteristic
curve, efficiency and power demand.(Inlet and outlet dampers for centrifugal fan ; inlet
vane control & dampers of any type ; Puhaltimet teollisuudessa ja LVI-laitoksissa, I 9)
3. CENTRIFUGAL PUMPS
Pumps used for transferring liquids are categorized into dynamic pumps, displacement
pumps and other pumps. Dynamic pumps use different types of impellers, which are
rotating wheels with blades, to generate kinetic energy for liquid. Displacement pumps use
for example a piston to move liquid from pump case to discharge pipe. Other types of
pumps are for example electromagnetic pumps and buoyancy pumps. In this thesis only
centrifugal pumps are studied since they are the most common type of pumps in industry
and the only type the PumpSave –program can calculate energy savings to.
The main characters of a pump can be discovered from the pump’s characteristic curves.
These curves are generated in a particular rotational speed and diameter of the impeller.
Curves state out pump’s head and efficiency in volume flow’s function. Pump’s head
means the total head which includes the heads demanded by suction inlet and discharge of
the pump. Characteristic curves for different types and sizes of centrifugal pumps can be
found at appendixes G1 to G17.
29
The density of the liquid does not affect on the head but it affects on the increase of the
pressure in the pump as equation 28 shows. (VLV-koneet luentomoniste 1, 81)
ghρp =Δ (28)
When designing a pumping system one main goal is to get the pumps working in their best
efficiency as much of the running time as possible. Figure 11 shows the typical operating
areas for different types of pumps.
Figure 11. Typical areas of operation for different pump types. (Luukkanen 2001, 45)
3.1 Structure and operation of centrifugal pump
Centrifugal pumps are the most commonly used pump type in industry. About 80 % of
process industry pumping needs is covered by centrifugal pumps. Main parts of a
centrifugal pump are impeller, shaft and casing. Impeller has backward curved blades.
Pump generates pressure to the liquid with the help of centrifugal force. Liquid discharges
30
out of the pumps case to a discharge pipe. One benefit of a centrifugal pump is its wide
operational range. Volume flow can vary from 1 cubic meter per hour to over 10000 cubic
meters per hour. Head can vary from less than a meter to about 3000 meters and the
temperature of the liquid can be up to 400 °C. Large centrifugal pumps can have efficiency
as high as 80 %. Figure 12 shows the main components of a centrifugal pump. (Luukkanen
2001, 44)
Figure 12. Structure of a centrifugal pump. (Viholainen 2007, 11)
3.2 Pump hydraulics
Next chapters introduce the basic theory and equations of pumping. Main purpose of
pumping is to transfer liquid from one place to another. Processes which need pumping
this often means that the liquid is needed to be pumped upwards through a pipeline. Pumps
function is to generate the needed pressure so this is possible.
3.2.1 Pump power demand and efficiency
Power input to liquid by pump is presented as equation 29 shows. Affecting factors to
power are volume flow and density of the liquid as well as the total system head and
gravitational constant.
31
HgqP v ⋅⋅⋅= ρ (29)
where
P power input to liquid [W]
volume flow [m³/s] vq
g gravitational constant [m/s²]
H total system head [m]
Pump receives power usually from a squirrel cage motor via shaft and transfers it forward
to liquid. Yet pumping is a lossy process so pumps demand for power can be expressed
with equation
pump
vpump
HgqP
ηρ ⋅⋅⋅
= (30)
where
pump’s demand for power [W] pumpP
efficiency of the pump pumpη
(Wirzenius 1978, 47-50)
3.2.2 System head and net positive suction head
While pumping liquid receives addition to total elevation head which is called the total
system head and marked as H. Total system head can be divided into static head and
dynamic head. Static head is the invariable part of the total head and it consist of geodetic
head and the head which is result of pressure difference between suction inlet and
discharge side of the pump. Geodetic head means the difference between liquid levels
between suction side and discharge side. Static head can be presented with equation.
ρg
ppHH 12
geost−
+= (31)
where
32
geodetic head [m] geoH
static head [m] stH
pressure at discharge outlet [Pa] 2p
pressure at suction inlet [Pa] 1p
Dynamic head is the part of the total head which is consists of pressure losses in the piping
and the volume flow of the liquid. Dynamic head can be presented with equation
∑+−= r
21
22
dyn Hg2
vvH (32)
where
dynamic head [m] dynH
flow velocity at suction inlet [m/s] 1v
flow velocity at discharge outlet [m/s] 2v
friction head [m] ∑ rH
Friction head means the part of total head which is due to friction. It can be presented with
equation
(33) 2vr kqH =
where
constant due to liquids characteristics k
Friction head can also be presented as equation 7 shows. In equation 34 parts of piping
which causes losses are taken into account. These parts are for example valves, curves and
connections. Also the roughness of the pipes coating is taken into account.
g2
vdlH
2
r ⎟⎠⎞
⎜⎝⎛ += ∑ζλ (34)
33
where
λ friction constant due to roughness of pipings coating
ζ friction constant due to valves, curves and connections in the piping
l total length of piping [m]
d diameter of piping [m]
Total head can be presented as equation 35 shows after all factors are taken into account.
∑+−+
−+=+= r
21
2212
geodynst Hg2vv
ρgpp
HHHH (35)
Net positive suction head or NPSH informs how much pressure must exist at pump’s
suction inlet in addition to liquids steam pressure so that the pump wouldn’t cavitate.
NPSH can be expressed with equation
geo,11,r
21vb1 HHg2
vg
pppNPSH −−+
⋅−+
=ρ
(36)
where
bp barometric pressure [Pa]
vp steam pressure in liquid [Pa]
1,rH head due to losses at suction inlent [m]
geo,1H static suction head [m]
geod,1H is negative in equation 36 if the pump is placed below the tank water level and has
positive suction. Figure 13 clarifies NPSH on suction lift. Cavitation means that pump’s
performance collapses producing lower head and flow than expected when steam bubbles
form in the liquid. Bubbles collapse instantly and cause shock waves to the impeller and
damage it. Strong corrosion will occur if cavitation continues for a longer period. Strongly
cavitating pump will fill up of steam and isn’t able to produce pressure. Cavitation can be
noticed from a rattling sound and vibration. Using several impellers behind each other or
34
connecting pumps in series reduces the possibility of cavitation essentially. (Volk 2005, 89
; VLV-koneet luentomoniste 1, 102-104)
Figure 13. NPSH. (Viholainen, 19)
3.2.3 Performance curves and affinity laws
Characteristic curves for pumps are usually presented in -coordinates where is
drawn with -curve also , and NPSH –curves. With these curves pump’s
operation can be observed when different variables change. Pump’s characteristic curves
are created for specific rotational speed and diameter of impeller.
H,q v
H,q v P,q v η,q v
H,q v -curve illustrates the relationship between volume flow and head, -curve
illustrates the relationship between power demand and volume flow,
P,q v
η,q v -curve
illustrates the relationship between efficiency and volume flow and NPSH –curve
illustrates the change in demand of minimum suction head when volume flow changes.
When pump’s rotational speed changes moves the original operation point of the pump to a
new characteristic curve provided by the new rotational speed. This happens through an
affinity parable which goes via origin. Figure 14 illustrates pump’s characteristic curves at
35
rotational speeds 0,8n, n and 1,2n. Pump’s other features are unchanged in other words
impellers diameter, blade angles and piping stays unchanged. (VLV luentomoniste 1, 96)
Figure 14. Effects of rotational speed change on to the operating point of the pump. (VLV-koneet
luentomoniste 1, 96)
Affinity parable can be presented with equation
2
v0
v0 q
qHH ⎟⎟
⎠
⎞⎜⎜⎝
⎛= , (37)
where
original head [m] 0H
original volume flow [m³/s] 0vq
Efficiency of the pump doesn’t change when moving via affinity parable. Next affinity
laws have been able to be assigned for pumps. (VLV luentomoniste 1, 96 - 97)
36
The relationship between volume flow and rotational speed is:
0v0
v
nn
= (38)
where
n new rotational speed [rpm]
original rotational speed [rpm] 0n
The relationship between head and rotational speed is:
2
00 nn
HH
⎟⎟⎠
⎞⎜⎜⎝
⎛= (39)
The relationship between power and rotational speed is:
3
00 nn
PP
⎟⎟⎠
⎞⎜⎜⎝
⎛= (40)
3.3. Flow control methods in pumping
In process the pumping demand changes in accordance with how the plant is operated so
pumping must adjust to the needs of the whole process. When selecting flow control
method for pump, three things must be paid attention to. First of all what kind of affect
does the control have on the volume flow on wanted pressure level. Secondly attention
must be paid on the affect on the efficiency of pumping and last is the change in power
demand. When adjusting flow in pumping it is important to be able to operate the pumps at
their best efficiency. Most commonly used flow control methods are throttling control,
changing rotational speed of the pump (variable speed control), cycling and organizing a
current from discharge to suction inlet. (Karttunen 2004, 252)
37
3.3.1 Variable speed control
It is possible to adjust the rotational speed of a pump while the motor rotates at a constant
speed. These methods are for example a hydraulic clutch, hydraulic clutch and moment
converter, inductive clutch or a mechanic variator which however is uncommon solution.
Some electric motors are possible to use at different speeds but variable speed control has
begun to establish more and more position at pump control in many processes. (Karttunen
2004, 252, VLV-luentomoniste 1, 99, Sarkomaa 1997, 6)
Hydraulic clutch can transfer motors power to pump at rotational speeds from 0 % to 80 %
of maximum speed. At speeds 80 % to 100 % of maximum the clutch is locked so that the
motor is connected via planetary transmission directly to the pump. Inductive clutch is an
electrical equivalent to hydraulic clutch. Other half of the inductive clutch contains a short-
circuit where currents induce if it moves regarding to magnetic field generated by a coil. In
order to generate a magnetic field direct current must be used. Inductive clutches
transmission ability depends on slipping and the magnitude of the current. Induced currents
and the power they generate are dependent on the slipping of the clutch. (Sarkomaa 1997,
7)
Using frequency converter in pumping isn’t a new thing but it has become more and more
competitive choice for flow control in pumping because of advancement in technology and
reduced investment costs. Changing the rotational speed of the motor the head or the
volume flow generated by the pump can be changed. Frequency converter controls the
electric motor, which shaft the pump’s impeller is connected to, and its rotation and at the
same time collects data from pumping. Figure 15 illustrates the main components of a
variable speed drive controlled pumping. (Volk 2005, 373)
38
Figure 15. Variable speed controlled pumping. (Viholainen, 31)
Pump’s operation point moves when the rotational speed of the pump changes. Operation
point is located at the intersection of pump and system curves. Figure 14 shows how
volume flow and head changes when rotational speed is changed.
Main benefits of variable speed control executed by a frequency converter are energy
saving, possibility to replace an existing flow control method with a frequency converter
without replacing the whole system, wide range of control without massive energy losses,
fine tuning possibility, lower loads on bearings and on pumps and the possibility to start
and stop the pump carefully. Variable speed control is the best choice for flow control in
cases where the flow of pumping has to be changed often. (Viholainen 2007, 33; Volk
2005, 374-375)
Other benefits of variable speed is that the pump’s efficiency doesn’t reduce during and
afnter adjustments especially if systems static head is small. With variable speed control
there is no need for a loss causing throttling valve. (Viholainen 2007, 32)
3.3.2 Throttling control
In throttling control flow is controlled by a throttling valve which is installed to pipeline
after the pump. It is not recommended to install throttling valve to a suction line because it
would cause possibility to cavitation. Throttling control is very common flow control
39
method in pumping because it is cheap and easy to implement. Using throttling control
though reduces efficiency in pumping. Pump produces bigger head when volume flow is
reduced. This growth in head is wasted energy. Throttling control is justified when the
pump is small and there is not much range for adjustments. It can also be used when the
pump operates with too little pressure compared to normal operating point. Figure 16
illustrates the components of throttling controlled pumping system. (Karttunen 2004, 253,
VLV-koneet luentomoniste 1, 99)
Figure 16. Throttling controlled pumping system. (Viholainen, 29)
3.3.3 On/off –control
Cycling, or on/off –control, means that the pump either runs or doesn’t run. This control
method can be used for example in situation where a tanks liquid level must be in between
certain limiting values. If the liquid level drops below a certain limiting value, the pump
starts and fills the tank to another limiting value. When it reaches the upper limiting value
pump stops. Starting and stopping the pump strains especially when the pump gets a full
load instantly. This could be resolved with variable speed control because the pump could
be started and stopped with slowly accelerating and slowing.
3.3.4 Other control methods
40
Liquids speeds tangential component can be changed by bending the blades of impeller.
This control method doesn’t cause much of losses but it also has very limited adjusting
possibilities. It also requires the pump to be stopped. It is a feasible control method for
axial pumps and in cases where flow resistance of piping is small.
Volume flow can be controlled also by blades which are installed just before the impeller.
This control method is complicated and expensive and it is feasible for only half axial or
axial pumps in cases where resistance from piping forms a significant part of the total
head. (VLV-luentomoniste 1, 100, Pumpputekniikka, I 18)
It is also possible to create a flow from pump’s discharge outlet to suction inlet but this
also isn’t a efficient way to adjust flow. This method is mainly used with propeller pumps.
In order to adjust a pump’s volume flow changes can be made to the impeller by changing
its diameter or by replacing the impeller to a whole new one which is geometrically
different than the original one. Of course it needs the pump to be stopped and it is feasible
only then when pump’s productive values are needed to be changed substantially and
permanently. (VLV-luentomoniste 1, 100, Pumpputekniikka, I 18)
4 SQUIRREL GAGE MOTOR
Popularity of the squirrel gage motor is based on its reasonable price, small maintenance
costs, working reliability and tolerance for tough environmental conditions. Squirrel gage
motors are manufactured for power ranging from 50 W to 10 MW. Rotational speed is
commonly 10, 12,5, 25 or 50 rounds per second. Rotational speed of a squirrel gage motor
depends on terminal pair number, resistance of the rotor, magnitude input voltage and
frequency and slipping of the load. Motor with two terminals running in 50 Hz electric
network runs at the speed of 3000 rpm, motor with four terminals runs at the speed of 1500
rpm and motor with six terminals runs at the speed of 1000 rpm. Previously when unusual
rotational speed has been required the choice has been a direct current motor but nowadays
frequency converter is preferred. Squirrel gage motor with power more than 11 kW has
efficiency from 90 % to 96 %. Smaller squirrel gage motors have efficiency below 90 %.
41
Squirrel gage motors are usually air cooled but big motors with power more than one
megawatt can also be water cooled. (Sarkomaa 1997, 8)
4.1 Structure of squirrel gage motor
Squirrel gage motor is simple and robust by its structure. The stator of a squirrel gage
motor has three-phase winding where alternate current is led. Current generates a rotating
magnetic field at the motor’s air gap and the magnetic field induces current and voltage to
the rotor which makes it rotate. (Montonen 1985, 16)
Squirrel gage motor is the most common type of electric motors in industrial use.
Nowadays lots of electric motors are being manufactured to be used with frequency
converter. Winding on the rotor of squirrel gage motor is done by pressure moulding or by
copper bars. It is called short circuit winding. /1/, /3/
4.2 Motor losses
Loss power of the motor consists of sine losses and harmonic losses. This is presented with
equation
harmlosslossloss PPP += sin (41)
where
lossP motor loss [kW]
sine losses [kW] lossPsin
harmonic losses [kW] harmlossP
Sine losses are formed from nominal loss of motor, loss caused by the rotational speed and
loss caused by the load.
42
lossnomnom
lossnomnom
lossnomloss PTTP
ffPP ⋅⎟⎟
⎠
⎞⎜⎜⎝
⎛⋅+⋅⎟⎟
⎠
⎞⎜⎜⎝
⎛⋅+⋅=
22
sin 65,015,02,0 (42)
where
motor nominal loss [kW] lossnomP
f output frequency of the drive [1/s]
nominal supply frequency [1/s] nomf
T actual torque of the motor [Nm]
nominal torque of the motor [Nm] nomT
Torque produced by the motor can be solved with equation
60
2 nTP ⋅⋅⋅= π (43)
Harmonic loss can be presented with equation 43. It is not dependent on the load but is
reversely dependent of the switching frequency of the frequency converter.
nomsw
harmloss Pf
P ⋅=9 (44)
where
switching frequency of the drive [1/s] swf
motor nominal power [kW] nomP
In calculation for the switching frequency value 3000 Hz can be used. (Ruuskanen, 51-52)
5. FREQUENCY CONVERTER
Frequency converter’s main function in industry is to adjust the rotation speed of AC-
motor. AC means alternate current. Adjustment is based in changing the voltage and
43
frequency of the current fed into the motor. With frequency converter the acceleration
speed and retardation speed of the motor can be changed to match the needs of the process.
Benefits of frequency converter drive use are decrement of strain in the electric network
which feeds the process and decrease in energy consumption in constant use. /1/, /2/
Frequency converters can be divided into direct frequency converters and frequency
converters with intermediate circuit. Direct frequency converters can be divided into
cyclone converters and matrix converters. Frequency converters with intermediate circuit
can be divided into commutated frequency converters, frequency converters with current
intermediate circuit and frequency converters with voltage intermediate circuit. Frequency
adjustment for squirrel gage motor is commonly handled by frequency converter with
intermediate circuit. Structure of this frequency converter is illustrated in figure 17.
(Hirvonen 2007, 1-2)
Figure 17. Block diagram of a frequency converter. (Montonen 1985, 20)
Rectifier changes the alternate current from the network into direct current. Rectifier can
be uncontrolled implemented with diode bridge or controlled thyristor rectifier. Inverter
which creates the alternating voltage of wanted frequency is placed next to the motor.
Function of the intermediate circuit is to filter the direct current provided by the rectifier
and to separate the rectifier and the inverter from each other and operates as a harmonic
filter. Other common components of frequency converters are network filter, cooling fan,
processor and some power electronic components. (Montonen 1985, 20-21)
Load commutated frequency converter contains two thyristor bridges which are linked
with intermediate circuit with chokers. Other thyristor bridge is connected to the network
and other to the electric motor. Bridge next to the network operates as a rectifier and the
bridge next to the motor operates as an inverter. In frequency converters with current
44
intermediate circuit the diode bridge produces the direct current to the intermediate circuit.
Voltage for the motor is controlled by modulating the pulse width because diode bridge is
uncontrollable. In frequency converter with intermediate circuit the bridge next to the
motor consists of commutation capacitors. They help to generate steadier torque than
produced by load commutated frequency converter. (Hirvonen 2007, 7-8)
5.1 Drive losses
Power loss of a frequency converter can be calculated with equation 45. Drive loss consists
of no-load basic loss, speed dependent no-load loss and load-dependent loss.
omdrivelossnnom
omdrivelossnnom
omdrivelossndriveloss PT
T55,0Pf
f1,0P35,0P ⋅⎟⎟⎠
⎞⎜⎜⎝
⎛⋅+⋅⎟⎟
⎠
⎞⎜⎜⎝
⎛⋅+⋅= (45)
where
drive loss [kW] drivelossP
drive nominal loss [kW] omdrivelossnP
(Ruuskanen, 53)
6. FANSAVE AND PUMPSAVE ENERGYSAVING TOOLS
With FanSave and PumpSave energy saving tools energy consumption when using
different flow control methods can be compared. Comparison is always performed between
variable speed drive and existing flow control method. Programs are based on Microsoft
Excel and contain also some Visual Basic code. Both programs are quite easy to use and
they have features that make the usage easier like fitting the view of the main window for
the screen and saving the performed calculation as a separate file. Programs also have
buttons for copying the suitable frequency converter model to clipboard and printing the
calculation in hand. Calculation can be performed in SI-units and in American units.
Programs also include a default model for the profile of the yearly operation time.
45
Calculation of energy costs is common feature in both programs and is performed the same
way. In both programs energy price, investment cost, interest rate and lifetime of the
application are input as initial values. As results programs provide consumed energy in
both flow control methods and possible energy savings as a subtraction. Programs are also
able to calculate yearly carbon dioxide emissions caused by the application. Production of
CO2 can be defined per kWh. Other economic results are payback period and present value
of the investment. Calculation of payback period is based on cumulative cash flow and
present value informs if the investment is profitable or unprofitable.
6.1 FanSave
FanSave compares energy consumption between variable speed drive controlled fan and
other flow control methods. In calculation program creates power demand curve in
function of volume flow which helps the comparison between different flow control
methods. FanSave includes axial and centrifugal fans. For centrifugal fan there is also three
different blade curvatures to choose from. These are forward curved blades, backward
curved blades and radial blades.
6.1.1 Initial values
Program needs as input values data from fan, transmission and motor. Also profile for
yearly operating time is needed to input for calculation of total energy consumption.
Existing flow control method and type of frequency converter must also be selected from
drop-down menus.
Needed fan data along with fan type and blade curvature are nominal volume flow,
generated pressure difference and efficiency. FanSave assumes that fluid is air with the
density of 1,2 kg/m³. It cannot be changed from the mainpage. If transmission is used its
efficiency can be input in transmission data. Choices for existing flow control method are
outlet damper, slip coupling, voltage control, two speed motor, cycling (on/off), inlet box
damper, inlet vanes for a centrifugal fan and pitch angle for axial fan.
46
Initial values for the driving motor are input voltage, power and efficiency. Voltage can be
chosen from 1-phase 115 or 230 V or 3-phase 230, 400, 460, 500 or 690 V. FanSave ads
10 % safety margin to the needed motor power which it calculates by using the fan data.
Next program needs information about the annual operation time. Total operation time and
percentual amount of time each flow rate is being used are needed for the calculation of
total energy consumption.
Last data about the appliances is what type of frequency converter is wanted to control the
fan and the motor. Drop-down menu gives five options for frequency converter type. They
are ABB component drive (ACS50), ABB component drive (ACS150), ABB general
machinery drive (ACS350), ABB standard drive (ACS550) and ABB industrial drive
(ACS800). Program presents suitable model for frequency converter when this selection is
made.
6.1.2 Calculations FanSave’s calculation is studied in this part of the thesis. All the equations and factors of
the program uses are introduced. FanSave assumes that the medium is air so the calculation
is observed based on that. Shaft power of the fan is calculated with equation 46. It contains
term which is calculated with equation 47. pK
ontransmissifan
vpaks ηη1,2
10ΔpqρKP
⋅⋅
⋅⋅⋅⋅= (46)
atm
fanp p
ηp0035,01K
⋅Δ⋅−= (47)
In practice term gets values which are very close to 1 so it doesn’t have much
relevance in what comes to results. Term would come up with values that have some
effect to results if the fan’s produced pressure difference would be abnormally high.
Equation 46 also contains term for fluid density at the numerator and value 1,2 at the
denominator. When flowing fluid is air these components evert each other.
pK
pK
47
FanSave does not take into account any changes in fluids temperature during blowing.
Therefore also the change in density is not correctly catered for. Density cannot be entered
to the program as an initial value but it can be changed at the calculation sheet. Calculation
could be added with possibility to take into account the temperature and density changes in
fluid.
Energy consumed at certain flow is written:
100
ttPE totali
ii⋅
⋅= (48)
Energy consumed during the yearly operating period is a sum of energies consumed at
each flow rates.
(49) ∑=
=flow.nom
%20iitotal EE
Possible energy saving is calculated by subtracting the total energy consumed by variable
speed drive from energy consumed by existing flow control method.
(50) total,vsdtotal,säätökWh,säästö EEE −=
Savings in currency is solved by multiplication of energy price and saved energy.
(51)
sähkökWh,säästö,€säästö hEE ⋅=
Specific fan power is calculated by FanSave with equation 52. Constant factors before
variables are to correct the unit of the result. FanSave deals with one fan at a time so the
SFP should be calculated for only one fan.
motorontransmissifan 0,01η0,01ηη01,0
p001,0SFP⋅⋅
Δ⋅= (52)
48
Power demand for most flow control methods in FanSave is calculated by using equation
53. For variable speed drive the efficiency of the frequency converter needs to be noticed
also as equation 54 expresses.
motorii xP
ηηη1,21000ΔpqρK
ontransmissifan
vp
⋅⋅⋅
⋅⋅⋅⋅⋅= (53)
freq.conv.motorontransmissifan
vpivsd,i ηηηη1,2
100000ΔpqρKxP
⋅⋅⋅⋅
⋅⋅⋅⋅⋅= (54)
Power demand for each flow control method at certain flow is dependent on a correcting
factor which gets values depending on the flow rate. These factors are presented in
Table 1. In table 1 for outlet damper the F-wheel stands for impeller with forward curved
blades, B-wheel for backward curved blades and R-wheel for radial blades. Correcting
factors are not dependent on the inintial values input to the program and they have been
defined empirically according to FanSave manual.
ix
Table 1. Correcting factors for power demand.
Outlet damper
Centrifugal
VSD F-wheel B-wheel R-
wheel Axial
Slip
coupling
Voltage
control
On/off Inlet
box
damper
Inlet
vanes
Pitch
angle
Nom.
flow
1,000 1,000 1,000 1,000 1,000 1,000 1,000 1,0 1,00 1,000 1,000
90 % 0,740 0,840 0,960 0,960 1,140 0,733 0,846 0,9 0,84 0,763 0,710
80 % 0,515 0,715 0,910 0,910 1,170 0,626 0,703 0,8 0,75 0,636 0,508
70 % 0,370 0,615 0,855 0,855 1,125 0,505 0,563 0,7 0,67 0,559 0,356
60 % 0,240 0,525 0,790 0,790 0,958 0,386 0,440 0,6 0,62 0,483 0,246
50 % 0,155 0,450 0,723 0,723 1,000 0,297 0,331 0,5 0,58 0,436 0,169
40 % 0,095 0,395 0,655 0,655 1,125 0,220 0,223 0,4 0,54 0,394 0,127
30 % 0,055 0,380 0,570 0,570 1,150 0,163 0,156 0,3 0,50 0,360 0,085
20 % 0,035 0,380 0,490 0,490 1,154 0,146 0,094 0,2 0,47 0,339 0,059
49
Factors for outlet damper flow control method with centrifugal fan are exactly the same
whether the fan has backward curved blades or radial blades. Outlet damper is the only
control method which is affected by the shape of the blades. Power demand curves for
every fan case studied with FanSave are always the same shape because of these factors.
These curves are illustrated in figure 18. Inlet vanes can only be used with centrifugal fan
and pitch angle adjustment is a feature of axial fans.
0
0,2
0,4
0,6
0,8
1
1,2
1,4
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Cor
rect
ing
fact
or
Variable speed drive Outlet damper F-wheel Outlet damper B-wheelOutlet damper R-wheel Outlet damper axial Slip couplingVoltage control Cycling (on/off) Inlet box damperInlet vanes Pitch angle
Figure 18. Power demand curves.
Power demand curves created by FanSave for different flow control methods in figure 18
are quite well matching to corresponding curves in figure 7. So at least the shape of the
curves is correct.
FanSave has an option to study the power demand of the fan application if no flow control
method is used. In this situation the power demand is calculated with equation 55.
motorontransmissifan
vpnocontrol ηηη1,2
1000ΔpqρKP
⋅⋅⋅
⋅⋅⋅⋅= (55)
50
FanSave calculates power demand for two speed motor controlled fan by replacing the
correcting factor with 0,25 when flow rate is 60 % of the maximum or lower. At higher
flow rates the power demand is calculated according to equation 60 and the factor
equals 1. This is studied more closely on the laboratory test section of the thesis.
ix
ix
FanSave’s calculation for on/off flow control method assumes that flow rate 20 % of the
maximum requires 20 % power of the maximum power and 30 % flow rate requires 30 %
power of the maximum etc. FanSave’s calculation for on/off flow control method would be
more accurate if it would be similar to PumpSave’s calculation which is dependent on the
total quantity of pumped water. When controlling fan with on/off flow control method the
fan either operates at full speed or doesn’t operate at all. With the profile of the operating
period the total quantity of air blown could be solved. This would make possible to solve
the time the fan needs to blow the total quantity when operating at full speed. Now the duty
time could be solved with the nominal operating point of the fan and the total energy
consumption could also be solved.
6.2 PumpSave
With PumpSave pump applications energy consumption can be compared between variable
speed drive control and few other flow control methods. Comparison is presented also as
bar diagram. PumpSave is designed for only centrifugal fans.
6.2.1 Initial values
PumpSave needs data from the pump, piping, driving motor, possible throttling valve,
exisiting flow control method and the annual operating time. Needed information about the
piping are the density of the liquid and static head. Density is 1000 kg/m³ by default and
static head determines the system curve in further calculations.
Needed data of the pump are nominal volume flow, nominal head, maximum head and
efficiency. These values can be found from the performance curve of the pump. If the
piping has a throttling valve after the pump the addition to the total head can be input
51
where it says head over open throttling valve. It contains also a help box for more
information.
For the driving motor exactly the same information is needed as in FanSave. Also the
profile and the total annual operating time are input the same way as in FanSave. For
existing flow control method there is a drop-down box like in FanSave. For pump
application the choices are throttling control, cycling (on/off) and hydraulic control.
Frequency converter type can be selected as in FanSave and the options are the same.
6.2.2 Calculations
This part of the thesis introduces all the equations and factors PumpSave uses for its
calculations. PumpSave solves pump’s power demand with equation 56. The program adds
a 10 % safety margin for motor power.
ontransmissipump
nomvp ηη3600
gHqρP
⋅⋅⋅⋅
= (56)
Consumed energy for certain flow is calculated as equation 57 shows. Power required for
certain flow is multiplied by the time pump was used in that flow
100
ttPE totali
ii⋅
⋅= (57)
Yearly energy consumption is calculated by adding all the energies used at all flow rates.
This is presented in equation
(58) ∑=
=flow.nom
%20iitotal EE
Possible energy saving is calculated by subtracting the energy used by variable speed drive
from the energy used by an existing flow control method. This is presented in equation
52
(59) total,vsdtotal,säätökWh,säästö EEE −=
Saving in currency is solved by multiplying the amount of saved energy by the price of
energy as equation
(60)
sähkökWh,säästö,€säästö hEE ⋅=
Payback period of the investment is calculated by means of cumulative cash flow.
Investment cost is divided by the yearly saving as equation
,€säästö
maksusintakai EIt = (61)
Throttling control
Power demand for throttling control for each flow rate is calculated as presented in
equation 62. Change in head produced by pump is taken into account by the term and
the change in pump’s efficiency is taken into account by the term . Calculation for
power demand and these variables are presented in equations 62, 63 and 64.
piH
pik
motorpumpi,pi
i,pii,vi,thi ηηk3600
gHqρP
⋅⋅⋅
⋅⋅⋅= , (62)
where
( nommax
2
v
i,vmaxi,pi HH
HH −⋅⎟⎟⎠
⎞⎜⎜⎝
⎛−= ) and (63)
v
v
i,vi,v
i,pi qq
q44,14,2q
k⎟⎟⎠
⎞⎜⎜⎝
⎛ ⋅−⋅
= (64)
53
On/off flow control method
For on/off flow control method the most essential thing is the total amount of pumped
water. By equation 65 can be calculated the amounts of pumped water at each flow rates.
totali
i,vi t%100tqQ ⋅⋅= (65)
All the pumped amounts are next added together so the total amount which is pumped can
be solved. This can be done with equation
∑=
=flow.nom
%20iitotal,v Qq (66)
Next PumpSave calculates the amount of water that would be pumped if the pump would
operate at full speed through the yearly operating time. This is presented in equation
totalvnom,v tqq ⋅= (67)
Now can be calculated the percentage of the total operating time which pump has to
operate at full speed to achieve the required amount of pumped water. This duty time can
be solved with equation
nom,v
total,vduty q
qt = (68)
Next PumpSave calculates the time in hours which the pump has to operate at full speed.
This is done with equation
(69) totaldutyoff/on ttt ⋅=
Finally the total energy consumed is calculated as equation
54
motor
pumpoff/onoff/on η
PtE
⋅= (70)
ydraulic flow control
ch
d to pum
lculated by equation
3. When pump operates at full flow the term equals 0,98.
H
Power demand for ea flow rate for hydraulic flow control is calculated by using equation
71. It contains term i,siH which is the system head term and it is calculated by equation 72.
Because clutch is used the power transmitte p’s shaft changes during flow
adjustment. This is taken into account by term i, which is ca ontransmissiη
iontransmissi ,7 η
ion,transmissimotorpump
isi,iv,i,hydr η1ηη3600
gHρqP
⋅⋅⋅
⋅⋅= (71)
( )lverottlingvaoveropenthstaticalminno
2
v
i,vstatici,si HHH
HH −−⋅⎟⎟⎠
⎞⎜⎜⎝
⎛+= (72)
⎟⎟
⎠
⎞
⎜⎜
⎝
⎛
⎟⎟
⎠⎠nom
⎞
⎜⎜
⎝
⎛⎟⎟⎞
⎜⎜⎝
⎛⋅⋅+⋅= max
flow%90flow%20flow%90flow%20i,ontransmissi HH
75,01,0...8,09,0...2,098,0η
(73)
ariable speed drive
orre
correcting factor for frequency converter’s
fficiency and it uses values shown in table 2.
V
Power demand for variable speed drive pump application is calculated by equation 74 for
each flow rate. Term i,siH in equation is the same factor for system curve as in hydraulic
flow control method. Term i,mik is c cting factor for driving motor’s efficiency. It is
calculated with equation 75. Term i,vsdik is
e
55
i,vsdii,mivsdmotorpump
isi,iv,i,vsdi kkηηη3600
gHqρP
⋅⋅⋅
⋅⋅⋅= (74)
nom
staticH
H2,02,0
v
iv,i,mi q
qk
⋅−
⎟⎟⎠
⎞⎜⎜⎝
⎛= (75)
Table 2. Correcting factors for the efficiency of the frequency converter.
Flow Factor
Nom. flow 1,00
90 % 1,00
80 % 1,00
70 % 1,00
60 % 0,99
50 % 0,99
40 % 0,99
30 % 0,98
20 % 0,97
Energy consumption of variable speed drive adjusted pump for each flow rate is calculated
with equation 76. Total energy consumption is addition of energy consumption of each
flow rate. This can be calculated as equation 77.
100tt
PE totalivsd,ivsd,i
⋅⋅= (76)
(77) ∑=
=flow.nom
%20ivsd,itot,vsd EE
7. COMPARISON BETWEEN FANSAVE’S CALCULATION AND PERFORMANCE CURVE BASED CALCULATION
56
This part of the thesis compares the calculations of FanSave energy saving tool to
calculation based on different sizes and types of axial and centrifugal fans. Compared flow
control methods are outlet damper and variable speed drive. FanSave does not take into
account any changes in temperature or density of the fluid. Used performance curves for
fans are created at constant density at and therefore changes in temperature or density were
not taken into account. FanSave calculates the power demand as a total power demand for
the whole application including driving motor and possible frequency converter. Changes
of motor’s efficiency were taken into account by equations 41, 42, 43 and 44 and changes
of frequency converter by equations 45 in characteristic curve based calculation.
Operational data for each motor was found from website of motor manufacturer called
Tamel.
Outlet damper flow control method was observed with performance curves by choosing a
nominal operating point from fan’s curve. Flow was then throttled 10 % at a time and new
operating point was found via fan curve. Operational values of the nominal point were
entered in FanSave as an initial values.
Reading values from performance curves is not absolutely precise so the results may
contain small errors on one way or the other. Characteristic curve based calculation gave
good image about the accuracy of FanSave’s calculation and its applicability for different
types of fans with different flow control methods. All the used performance curves for fans
are found at appendixes E1 to E14 and F1 to F9.
7.1 Centrifugal fans
Centrifugal fans for comparison were chosen from manufacturers called Fläktwoods and
Hürner Funken which provided good performance curves for their fans. Some curves
include correcting factors for fan’s efficiency when rotational speed changes. These factors
were taken into account when the total power demand was calculated.
57
In calculation comparison the fans are divided based on the curvature of the blades. Four
models with radial blades, six models with backward curved blades and four models with
forward curved blades were chosen for the comparison.
7.1.1 Radial blades
Nominal point for each fan was chosen at operational area where the flow rate is as large
as possible and the fan operates at good efficiency. Operation range was also pursued to
keep as wide as possible.
HF R 560 – 13 D/R
First fan in this series is 560-13 D/R which is powered by 5,5 kW electric motor. It
operates at the efficiency of 85 % at the nominal operating point
Table 3. Calculations for HF R 560 – 13 D/R.
Outlet Damper flow control n=1060 rpm VSD flow control
Flow q_v
[m³/s]
delta
p
[Pa]
eta_fan P_total
[kW]
P_FanSave
[kW]
q_v
[m³/s]
delta
p
[Pa]
n
[rpm]
P_total
[kW]
P_FanSave
[kW]
Nom.
flow
4,00 690 73 % 4,30 4,,29 3,70 760 1060 4,19 4,31
90 % 3,60 765 74 % 4,02 4,12 3,33 640 970 3,23 3,19
80 % 3,20 830 75 % 3,95 3,90 2,96 490 860 2,27 2,22
70 % 2,80 870 76,5 % 3,71 3,67 2,59 375 750 1,60 1,58
60 % 2,40 920 78 % 3,42 3,39 2,22 325 690 1,25 1,04
50 % 2,00 960 73 % 3,04 3,10 1,85 200 550 0,75 0,67
40 % 1,48 120 410 0,48 0,41
58
0
1
2
3
4
5
40 % 50 % 60 % 70 % 80 % 90 % Nom. flowFlow
Pow
er d
eman
d [k
W]
0,70,710,720,730,740,750,760,770,780,79
Effic
ienc
y
Calculated power demand for outlet damper Power demand for outlet damper by FanSaveCalculated power demand for VSD Power demand for VSD by FanSaveFan efficiency in outlet damper
Figure 19. Power demand and efficiency curves for HF R 560 – 13 D/R centrifugal fan.
Results in table 3 and in figure 19 show that for this fan FanSave’s calculation matches the
characteristic curve based calculation extremely well for outlet damper flow control. Fan’s
efficiency increases first when flow is choked for about 5 % and then decreases slowly
towards the end of the operating range. Efficiency of the motor remains high during the
adjustment.
Results for power demand for fan operating under variable speed drive does not correlate
as well as for fan with outlet damper. Power demand based on fan’s performance curve is a
bit higher than power demand calculated by FanSave at all the operation points. Fan’s
efficiency remains constant during the speed adjustment and motor’s efficiency decreases
about 35 % during the adjustment. Drives efficiency decreases about 8 %. All in all
FanSave’s calculation is very accurate for this fan case.
HF R 200 – 48 R
Second fan in this series is 200-48 R and it is powered by 22 kW motor. Motor operates at
88 % efficiency at the nominal operating point.
Table 4. Calculations for HF R 200 -48 R.
Outlet Damper flow control n=1640 rpm VSD flow control
Flow q_v
[m³/s]
delta
p
[Pa]
eta_fan P_total
[kW]
P_FanSave
[kW]
q_v
[m³/s]
delta
p
[Pa]
n
[rpm]
P_total
[kW]
P_FanSave
[kW]
59
Nom.
flow
2,16 5400 70 % 18,45 18,69 1,72 5900 1640 15,58 15,30
90 % 1,94 5600 71 % 17,00 17,94 1,55 5000 1430 12,26 11,32
80 % 1,73 5850 74 % 15,44 17,01 1,38 3900 1230 8,87 7,88
70 % 1,51 6100 71 % 14,46 15,98 1,21 3000 1060 6,23 5,62
60 % 1,30 6400 66 % 14,01 14,76 1,03 2200 940 4,34 3,67
50 % 1,08 6450 62 % 12,59 13,51 0,86 1500 820 2,91 2,37
40 % 0,69 950 630 1,93 1,45
0,00
5,00
10,00
15,00
20,00
40 % 50 % 60 % 70 % 80 % 90 % Nom. flowFlow
Pow
er d
eman
d [k
W]
0,55
0,6
0,65
0,7
0,75
Effic
ienc
y
Calculated power demand for outlet damper Power demand for outlet damper by FanSaveCalculated power demand for VSD Power demand for VSD by FanSaveFan efficiency in outlet damper Fan efficiency in vsd
Figure 20. Power demand and efficiency curves.
For outlet damper the power demand curves are not as well matching in this case than in
previous. Power demand calculated by FanSave is higher than power demand calculated
with fan’s performance curve through the whole operating range. Fan’s efficiency
increases first when the flow is throttled and it may be one reason for the difference in
power demands. Motor’s efficiency decreases only about one percentage during the
throttling.
For variable speed drive controlled fan the results for power demand are perfectly
matching. Fan’s efficiency decreases first slowly and more rapidly when flow rate enters
below 70 %. Motor’s efficiency decreases about 39 % and drives about 8 %. For this fan
case FanSave’s calculation is accurate for variable speed drive controlling and a bit
inaccurate for outlet damper controlled flow.
HF R 710 – 13 D/R
60
Next fan under inspection is 710-13 D/R which is powered by 75 kW squirrel cage motor.
It operates at the efficiency of 94,7 % at the nominal operating point.
Table 5. Results HF R 710 – 13 D/R.
Outlet Damper flow control n=3000 rpm VSD flow control eta_fan=58 %
Flow q_v
[m³/s]
delta
p
[Pa]
eta_fan P_total
[kW]
P_FanSave
[kW]
q_v
[m³/s]
delta
p
[Pa]
n
[rpm]
P_total
[kW]
P_FanSave
[kW]
Nom.
flow
13,00 2780 70 % 54,17 54,15 10,40 3400 1700 47,59 47,67
90 % 11,70 3120 75 % 51,11 51,99 9,36 2800 1530 35,73 35,27
80 % 10,40 3450 80 % 47,16 49,28 8,32 2100 1340 24,40 24,55
70 % 9,10 3680 80 % 44,08 46,30 7,28 1600 1170 16,85 17,50
60 % 7,80 3875 80 % 39,89 42,78 6,24 1220 1010 11,73 11,44
50 % 6,50 4050 77 % 36,21 39,15 5,20 850 830 7,51 7,39
0,0010,0020,0030,0040,0050,0060,00
50 % 60 % 70 % 80 % 90 % Nom. flowFlow
Pow
er d
eman
d [k
W]
0,640,660,680,70,720,740,760,780,80,82
Effic
ienc
yCalculated power demand for outlet damper Power demand for outlet damper by FanSaveCalculated power demand for VSD Power demand for VSD by FanSaveFan efficiency in outlet damper Fan efficiency in vsd
Figure 21. Power demand curves for HF R 710 – 13 D/R centrifugal fan.
For this fan the results for power demand in outlet damper flow control are quite well
matching. Difference between power demands increases little bit when fan’s operation
enters smaller flows. Efficiency of the fan increases first and then starts to decrease. Fan’s
efficiency remains good through the whole operating range.
Results for power demand when fan is operating under variable speed drive correlate
extremely well. Efficiency of the fan decreases only few percentages during speed
adjustment. Motor’s efficiency decreases about 13 % and drives efficiency about 7 %.
61
HF R 1000 – 13 D/R
Last fan equipped with radial blades is 1000-13D/R. It is connected to a 160 kW motor
which operates at the efficiency of 95,9 % at the nominal point.
Table 6. Calculations for HF R 1000 – 13 D/R.
Outlet Damper flow control n=3000 rpm VSD flow control eta_fan=58 %
Flow q_v
[m³/s]
delta
p
[Pa]
eta_fan P_total
[kW]
P_FanSave
[kW]
q_v
[m³/s]
delta
p
[Pa]
n
[rpm]
P_total
[kW]
P_FanSave
[kW]
Nom.
flow
24,50 3640 75 % 124,76 122,82 22,00 4000 1300 116,80 115,48
90 % 22,05 4000 81 % 114,25 117,91 19,80 3100 1150 82,42 85,46
80 % 19,60 4220 80 % 108,50 111,77 17,60 2400 1015 57,72 59,47
70 % 17,15 4450 79 % 101,41 105,01 15,40 1900 900 41,45 42,40
60 % 14,70 4680 78 % 92,67 97,03 13,20 1350 730 26,52 27,72
50 % 12,25 4780 77 % 80,08 88,80 11,00 960 620 17,00 17,90
0,0020,0040,0060,0080,00
100,00120,00140,00
50 % 60 % 70 % 80 % 90 % Nom. flowFlow
Pow
er d
eman
d [k
W]
0,72
0,74
0,76
0,78
0,8
0,82
Effi
cien
cy
Calculated power demand for outlet damper Power demand for outlet damper by FanSaveCalculated power demand for VSD Power demand for VSD by FanSaveFan efficiency
Figure 22. Power demand curves for HF R 1000 – 13 D/R centrifugal fan.
Results for power demand when fan operates under outlet damper flow control are quite
well matching as figure 22 show. Power demand calculated by FanSave decreases slightly
slower than power demand calculated with the fan’s performance curve. Fan’s efficiency
behaves similarly as in previous cases. It increases first when the flow is throttled and then
starts to decrease. Changes in efficiency depend on the chosen nominal operating point.
62
Good matching of the power demand curves point that this approach and this choice for
nominal point match FanSave’s calculation.
Results for the variable speed drive controlled fan are similar to previous cases. Power
demand curves between calculations are almost identical. Fan’s efficiency remains high
during the adjustment only dropping about a percentage. Motor’s efficiency decreases
nearly 10 % and drives about 8 % during the speed adjustment.
7.1.2 Fans with forward curved blades.
Next series of fans under the lens are models with forward curved blades. Like in the
section of fans with radial blades also in this series the study starts with the least powerful
fan and advances to more powerful models.
GTLF-3-025
First fan with forward curves blades is GTLF-3-025 and its impellers diameter is 250 mm.
Fan is powered by 4,8 kW motor which operates at 80 % of efficiency at the nominal
operating point.
Table 7. Calculations for GTLF-3-025.
Outlet Damper flow control n=3000 rpm VSD flow control eta_fan=58 %
Flow q_v
[m³/s]
delta
p
[Pa]
eta_fan P_total
[kW]
P_FanSave
[kW]
q_v
[m³/s]
delta
p
[Pa]
n
[rpm]
P_total
[kW]
P_FanSave
[kW]
Nom.
flow
0,91 2250 58 % 4,17 4,39 0,88 2250 3000 4,10 4,31
90 % 0,82 2240 57 % 3,81 3,69 0,79 1800 2700 3,01 3,19
80 % 0,73 2230 55 % 3,50 3,14 0,70 1420 2400 2,18 2,22
70 % 0,64 2220 53,5 % 3,15 2,70 0,61 1080 2100 1,53 1,58
60 % 0,55 2210 52 % 2,79 2,31 0,53 800 1800 1,06 1,03
50 % 0,46 2200 51 % 2,39 1,98 0,44 560 1500 0,72 0,67
40 % 0,36 2200 49 % 2,04 1,74 0,35 370 1200 0,50 0,41
30 % 0,27 2200 48 % 1,63 1,67 0,26 200 900 0,35 0,24
63
012345
30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Pow
er d
eman
d [k
W]
00,20,40,60,8
Effic
ienc
y
Calculated power demand for outlet damper Power demand for outlet damper by FanSaveCalculated power demand for VSD Power demand for VSD by FanSaveFan efficiency in outlet damper
Figure 23. Power demand and efficiency curves.
For VSD flow control the results for power demand are correlating through the whole
operating range. A little difference is generated at the smallest and highest possible flows.
Efficiency of the fan remains constant during the speed adjustment. Efficiency of the motor
decreases about 55 % and the efficiency of the drive about 8 % during the adjustment.
For outlet damper flow control the results for power demand are not as well matching as in
vsd flow control. There is no big difference generated between the calculations but the
power demand curves are in a bit different shape. Fan’s efficiency decreases about 10 %
during the adjustment. Motor’s efficiency decreases nearly 8 %. However the calculation
of FanSave for outlet damper flow control is quite well matching for this type of fan.
GTLF-3-040
Second studied fan is GTLF-3-040. It is a bit more powerful than the previous model and
has impeller the size of 400 mm of diameter. Driving motor for the fan is 9 kW squirrel
cage motor which reaches the efficiency of 87 % at the nominal operating speed.
Table 8. Calculations for GTLF-3-040.
Outlet Damper flow control n=1700 rpm VSD flow control eta_fan=65 %
Flow q_v
[m³/s]
delta
p
[Pa]
eta_fan P_total
[kW]
P_FanSave
[kW]
q_v
[m³/s]
delta
p
[Pa]
n
[rpm]
P_total
[kW]
P_FanSave
[kW]
64
Nom.
flow
2,20 1800 64 % 6,89 7,08 1,95 1800 1700 6,17 6,31
90 % 1,98 1800 65 % 6,13 5,95 1,76 1500 1540 4,72 4,67
80 % 1,76 1800 64 % 5,56 5,07 1,56 1050 1310 3,07 3,25
70 % 1,54 1800 62,5 % 5,01 4,36 1,37 900 1200 2,39 2,32
60 % 1,32 1800 62 % 4,37 3,72 1,17 600 1010 1,52 1,51
50 % 1,10 1800 59,5 % 3,84 3,19 0,98 450 850 1,08 0,98
40 % 0,88 1800 59 % 3,17 2,80 0,78 270 660 0,69 0,60
30 % 0,66 1800 58 % 2,51 2,69
012345678
30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom. flowFlow
Pow
er d
eman
d [k
W]
0,54
0,56
0,58
0,6
0,62
0,64
0,66
Effic
ienc
y
Calculated power demand for outlet damper Power demand for outlet damper by FanSaveCalculated power demand for VSD Power demand for VSD by FanSaveFan efficiency in outlet damper
Figure 24. Power demand and efficiency curves.
When flow is controlled by an outlet damper the results for power demand are similar to
previous case. Power demand calculated by FanSave decreases first more rapidly than in
calculation based on the performance curve of the fan. Decreasing slows down when the
operating area enters smaller flow rates giving the performance curve parabolic shape.
Power demand curve based on the performance curve is more linear. Efficiency of the fan
decreases only about 6 % during the adjustment. Motor’s efficiency decreases about 8 %.
Results for vsd controlled fan are matching almost perfectly and there are no big
differences in power demands between calculations. Efficiency of the fan remains constant
during the adjustment and motor’s efficiency decreases about 37 % when the rotating
speed decreases. Drives efficiency decreases about 9 %.
GTLF-3-056
65
Next fan model is GTLF-3-056. Its impeller diameter is 560 mm and the it is powered by
22 kW motor. Motor’s efficiency at nominal operating point is 88 %.
Table 9. Calculations for GTLF-3-056.
Outlet Damper flow control n=1300 rpm VSD flow control eta_fan=65 %
Flow q_v
[m³/s]
delta
p
[Pa]
eta_fan P_total
[kW]
P_FanSave
[kW]
q_v
[m³/s]
delta
p
[Pa]
n
[rpm]
P_total
[kW]
P_FanSave
[kW]
Nom.
flow
4,50 2000 65 % 15,71 15,66 4,50 2000 1300 16,08 15,98
90 % 4,05 2000 63,5 % 14,49 13,16 4,05 1580 1140 11,65 11,83
80 % 3,60 2000 63 % 13,01 11,20 3,60 1230 1025 8,29 8,23
70 % 3,15 1985 61,5 % 11,63 9,63 3,15 990 900 6,07 5,87
60 % 2,70 1975 60,5 % 10,16 8,22 2,70 720 770 4,08 3,84
50 % 2,25 1975 59,5 % 8,71 7,05 2,25 500 640 2,70 2,48
40 % 1,80 1960 58 % 7,24 6,19 1,80 330 520 1,80 1,52
02468
1012141618
40 % 50 % 60 % 70 % 80 % 90 % Nom. flowFlow
Pow
er d
eman
d [k
W]
0,54
0,56
0,58
0,6
0,62
0,64
0,66
Effic
ienc
y
Calculated power demand for outlet damper Power demand for outlet damper by FanSaveCalculated power demand for VSD Power demand for VSD by FanSaveFan efficiency in outlet damper
Figure 25. Power demand and efficiency.
Power demand curves for outlet damper controlled flow have the same shape as in
previous cases. Performance curve based power demand is higher than power demand
calculated by FanSave at all the operation points. Efficiency of the fan decreases about 7 %
during the flow adjustment. Motor’s efficiency remains quite good also decreasing only
about 6 %.
66
For vsd controlled fan the results for power demand are once again almost identical
between calculations. Motor’s efficiency decreases about 31 % and drives about 9 %.
GTLF-3-071
Last fan with forward curved blades is GTLF-3-071 and it is equipped with impeller the
size of 710 mm of diameter. The fan is powered by 22 kW squirrel cage motor which
reaches the efficiency of 91,7 % at the nominal operating point.
Table 10. Calculations for GTLF-3-071.
Outlet Damper flow control n=850 rpm VSD flow control eta_fan=69 %
Flow q_v
[m³/s]
delta
p
[Pa]
eta_fan P_total
[kW]
P_FanSave
[kW]
q_v
[m³/s]
delta
p
[Pa]
n
[rpm]
P_total
[kW]
P_FanSave
[kW]
Nom.
flow
7,70 1600 68 % 19,88 19,68 7,25 1590 850 18,69 18,52
90 % 6,93 1550 68 % 17,31 16,53 6,53 1230 760 13,13 13,71
80 % 6,16 1500 67 % 15,11 14,07 5,80 1000 680 9,62 9,54
70 % 5,39 1450 66 % 13,00 12,11 5,08 770 595 6,66 6,80
60 % 4,62 1420 65 % 11,13 10,33 4,35 560 505 4,37 4,45
50 % 3,85 1400 62 % 9,64 8,86 3,63 400 425 2,83 2,87
40 % 3,08 1400 58,5 % 8,25 7,77 2,90 255 330 1,73 1,76
30 % 2,31 1400 55,7 6,48 7,48
0
5
10
15
20
25
30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Pow
er d
eman
d [k
W]
0,5
0,55
0,6
0,65
0,7Ef
ficie
ncy
Calculated power demand for outlet damper Power demand for outlet damper by FanSaveCalculated power demand for VSD Power demand for VSD by FanSaveFan efficiency in outlet damper
Figure 26. Power demand and efficiency curves.
67
Results for outlet damper flow control are seemingly better matching for this fan than the
previous ones. There is very little difference between power demand curves at all points of
operating range. Efficiency of the fan decreases about 12 % during the flow adjustment.
Motor’s efficiency remains high during the adjustment decreasing only about 4 %.
When fan is operating under variable speed drive the results for power demand between
calculations are identical. Fan’s efficiency remains constant, motor’s efficiency decreases
about 23 % and the drives efficiency drops about 8 %.
7.1.3 Fans with backward curved blades
Last series of studied fans consists of fans with backward curved blades. First fan in the
series is the least powerful model and the last is the most powerful. Performance curves of
the fans are found at appendixes E9 to E14. Some curves have correcting factors for fans
efficiency during rotation speed adjustment. These factors were taken into account when
executing calculations.
GTLB-3-025
First fan of this series is GTLB-3-025. Diameter of the fan’s impeller is 250 mm and it is
connected to a 2 kW motor which operates at the efficiency of 71 % at the nominal
operating point.
Table 11. Calculations for GTLB-3-025.
Outlet Damper flow control n=4500 rpm VSD flow control
Flow q_v
[m³/s]
delta
p
[Pa]
eta_fan P_total
[kW]
P_FanSave
[kW]
q_v
[m³/s]
delta
p
[Pa]
n
[rpm]
P_total
[kW]
P_FanSave
[kW]
Nom.
flow
0,78 1320 70 % 2,07 2,07 0,59 1800 4500 1,99 2,00
90 % 0,70 1570 74 % 2,10 1,98 0,53 1470 4100 1,52 1,48
80 % 0,62 1780 75,7 % 2,07 1,88 0,47 1120 3600 1,09 1,03
70 % 0,55 1840 75,5 % 1,88 1,77 0,41 850 3130 0,80 0,73
68
60 % 0,47 2000 74 % 1,79 1,63 0,35 630 2670 0,58 0,48
50 % 0,39 2070 73 % 1,58 1,49 0,30 450 2300 0,44 0,31
40 % 1,35 0,24 295 1850 0,33 0,19
0
0,5
1
1,5
2
2,5
40 % 50 % 60 % 70 % 80 % 90 % Nom. flowFlow
Pow
er d
eman
d [k
W]
0,6
0,65
0,7
0,75
0,8
Effic
ienc
y
Calculated power demand for outlet damper Power demand for outlet damper by FanSaveCalculated power demand for VSD Power demand for VSD by FanSaveFan efficiency in outlet damper Fan efficiency in vsd
Figure 27. Power demand and efficiency.
Power demand calculated by characteristic curve of the fan is a bit higher than power
demand calculated by FanSave. Difference is not big and the curves are the same shape so
FanSave’s calculation matches the curve based power demand quite well. Fan’s efficiency
increases first when the flow is throttled and starts to decrease when flow rate reaches 70
% of the maximum. Motor’s efficiency decreases only one percentage during the
adjustment.
Results for fan operating under variable speed drive are matching better though little
difference is formed when operating point enters lower flow rates. This could be caused by
changes in fan’s efficiency which decreases about 10 % during the adjustment. FanSave
may not take into account any changes in fan’s efficiency when it is controlled by a
frequency converter. Motor’s efficiency drops about 38 % and drives about 3 %.
GTHB-3-040
Next model with backward curved blades is GTHB-3-040. Diameter of the impeller is 400
mm and it is powered by 8 kW motor. Efficiency of the motor is 81 % at the nominal
operating point.
69
Table 12. Calculations for GTHB-3-040.
Outlet Damper flow control n=3500 rpm VSD flow control
Flow q_v
[m³/s]
delta
p
[Pa]
eta_fan P_total
[kW]
P_FanSave
[kW]
q_v
[m³/s]
delta
p
[Pa]
n
[rpm]
P_total
[kW]
P_FanSave
[kW]
Nom.
flow
2,38 2100 75 % 7,69 8,18 1,85 2850 3500 7,64 8,14
90 % 2,14 2450 79 % 7,66 7,86 1,67 2375 3170 5,83 6,02
80 % 1,90 2750 80,5 % 7,51 7,45 1,48 1880 2830 4,23 4,19
70 % 1,67 3100 80 % 7,45 7,00 1,30 1450 2500 2,99 2,99
60 % 1,43 3250 78 % 6,88 6,46 1,11 1070 2170 2,05 1,95
50 % 1,19 3300 76 % 6,00 5,92 0,93 715 1750 1,34 1,26
40 % 0,74 330 1300 0,76 0,77
30 % 0,56 240 1000 0,60 0,45
0
2
4
6
8
10
30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom. flowFlow
Pow
er d
eman
d [k
W]
0,680,70,720,740,760,780,80,82
Effic
ienc
y
Calculated power demand for outlet damper Power demand for outlet damper by FanSaveCalculated power demand for VSD Power demand for VSD by FanSaveFan efficiency in outlet damper Fan efficiency in vsd
Figure 28. Power demand curves for GTHB-3-040.
Results for power demand when fan is controlled by outlet dampers are rather well
matching. At highest volume flows power demand calculated by FanSave is a bit higher
but at lower flow rates performance curve based power demand is higher. Nevertheless the
differences are very small so the FanSave’s calculation is pretty accurate for this type of
fan. Efficiency of the fan increases first quite rapidly and then starts to decrease when flow
rate moves lower than 80 % of the maximum. Fan can not operate at lower flow rates than
50 % of the maximum flow rate.
70
Results for power demand when flow is controlled by a variable speed drive are correlating
extremely well. No big differences in power demands between calculations are formed at
any point of operation. Fan’s efficiency decreases 8 %, motor’s efficiency 53 % and drives
efficiency about 9 % during the speed adjustment.
GTHB-3-056
Next examined fan is GTHB-3-056. Its impellers diameter is 560 mm and it is connected to
a 17 kW motor. Motor operates at 86 % of efficiency at the nominal operating point.
Table 13. Calculations for GTHB-3-056.
Outlet Damper flow control n=1900 rpm VSD flow control eta_fan=84 %
Flow q_v
[m³/s]
delta
p
[Pa]
eta_fan P_total
[kW]
P_FanSave
[kW]
q_v
[m³/s]
delta
p
[Pa]
n
[rpm]
P_total
[kW]
P_FanSave
[kW]
Nom.
flow
4,45 2140 75 % 14,76 14,68 3,45 3070 2500 15,34 15,19
90 % 4,01 2650 80,5 % 15,34 14,10 3,11 2400 2240 10,95 11,24
80 % 3,56 2980 82 % 15,04 13,36 2,76 1920 2000 7,97 7,82
70 % 3,12 3200 80,5 % 14,39 12,55 2,42 1420 1740 5,39 5,58
60 % 2,67 3350 79 % 13,14 11,6 2,07 1070 1500 3,73 3,65
50 % 1,73 770 1280 2,52 2,36
40 % 1,38 500 1050 1,64 1,44
0
5
10
15
20
40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Pow
er d
eman
d [k
W]
0,70,720,740,760,780,80,820,84
Effic
ienc
y
Calculated power demand for outlet damper Power demand for outlet damper by FanSaveCalculated power demand for VSD Power demand for VSD by FanSaveFan efficiency in outlet damper Fan efficiency in vsd
Figure 29. Power demand curves for GTHB-3-056.
71
Results for power demand when fan operates under outlet damper fan control differ
between calculations. Based on the performance curve of the fan power demand is higher
through the whole operating range. This is caused by high pressure difference generated by
fan when the flow is throttled. Though the efficiency increases at first the power demand
also increases. In FanSave’s calculation there is no possibility of power demand rising
when flow is decreasing. Motor’s efficiency remains constant during the adjustment.
For variable speed drive controlled fan the results for power demand are almost identical.
Fan’s efficiency decreases only few percentages, motor’s about 28 % and drives efficiency
about 8 %.
GTHB-3-071
Next model in this series is GTHB-3-071. Diameter of the impeller is 710 mm and it is
connected to a 30 kW motor which operates at the efficiency of 93,2 % at the nominal
operating point.
Table 14. Calculations for GTHB-3-071.
Outlet Damper flow control n=1900 rpm VSD flow control eta_fan=84 %
Flow q_v
[m³/s]
delta
p
[Pa]
eta_fan P_total
[kW]
P_FanSave
[kW]
q_v
[m³/s]
delta
p
[Pa]
n
[rpm]
P_total
[kW]
P_FanSave
[kW]
Nom.
flow
8,0 2000 75 % 22,80 22,77 5,9 2900 1900 22,19 22,11
90 % 7,2 2400 81 % 22,80 21,89 5,3 2350 1720 16,33 16,36
80 % 6,4 2700 82,5 % 22,40 20,72 4,7 1800 1540 11,32 11,39
70 % 5,6 2900 83 % 20,94 19,47 4,1 1400 1320 7,91 8,12
60 % 4,8 3100 80 % 19,91 17,99 3,5 1000 1130 5,10 5,31
50 % 4,0 3200 76 % 18,06 16,46 2,95 700 950 3,29 3,43
40 % 2,4 480 790 2,14 2,10
30 % 1,8 270 600 1,29 1,22
72
0
5
10
15
20
25
30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Pow
er d
eman
d [k
W]
0,70,720,740,760,780,80,820,840,86
Effic
ienc
y
Calculated power demand for outlet damper Power demand for outlet damper by FanSaveCalculated power demand for VSD Power demand for VSD by FanSaveFan efficiency in outlet damper Fan efficiency in vsd
Figure 30. Power demand and efficiency curves.
Like in previous case also in this case power demand for outlet damper controlled fan
based on the characteristic curve is higher than power demand calculated by FanSave
through the whole operating area. In this case also the power demand of the fan does not
start to decrease immediately after the flow is throttled because of the high pressure
difference generated by the fan. Efficiency reaches its highest point when flow rate is 70 %
and then starts to decrease rapidly. Fan’s operating range reaches 50 % flow of the
maximum flow.
Results for variable speed drive controlled fan are seemingly better than fan with outlet
dampers. Power demand curves between calculations match perfectly. Efficiency of the fan
decreases two percents. Motor’s efficiency decreases about 37 % and drives about 17 %.
GTHB-3-100
Second to last fan in this series is GTHB-3-100. Diameter of the fan’s impeller is 1000 mm
and it is powered by 55 kW motor which reaches 94,2 % efficiency at the nominal
operating point.
Table 15. Results for GTHB-3-100.
Outlet Damper flow control n=1360 rpm VSD flow control eta_fan=82,2 %
Flow q_v delta eta_fan P_total P_FanSave q_v delta n P_total P_FanSave
73
[m³/s] p
[Pa]
[kW] [kW] [m³/s] p
[Pa]
[rpm] [kW] [kW]
Nom.
flow
15 2100 75 % 44,70 44,34 11,9 2850 1360 44,90 44,33
90 % 13,5 2500 81 % 44,70 42,57 10,7 2350 1230 33,63 32,81
80 % 12 2800 82,5 % 44,17 40,35 9,5 1800 1075 23,27 22,83
70 % 10,5 3000 83 % 42,57 37,91 8,3 1400 950 16,24 16,28
60 % 80 % 7,1 1020 810 10,62 10,64
50 % 76 % 6,0 730 690 6,94 6,87
40 % 4,8 480 550 4,27 4,21
0
10
20
30
40
50
40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Pow
er d
eman
d [k
W]
0,70,720,740,760,780,80,820,84
Effic
ienc
y
Calculated power demand for outlet damper Power demand for outlet damper by FanSaveCalculated power demand for VSD Power demand for VSD by FanSaveFan efficiency in outlet damper Fan efficiency in vsd
Figure 31. Power demand curves for GTHB-3-100.
For more powerful fans the results for outlet damper seem to be quite similar and this case
is not an exception. Based on the performance curve the power demand is higher than
power demand calculated by FanSave. Cause is once again the rising pressure difference
when the flow is choked. Operating range of the fan is quite narrow reaching only flow
rate of 70 % of the maximum. Driving motor’s efficiency remains high through the
adjustment.
Power demands between calculations for variable speed drive controlled fan are well
correlating for this fan also. Fan’s efficiency remains high during the adjustment. Motor’s
efficiency drops about 18,5 % and drives about 10 %.
GTLB-3-140
74
Last centrifugal fan with backward curved blades is GTLB-3-140 and its impellers
diameter is 1400 mm. The fan is powered by 75 kW motor which operates at 94,3 %
efficiency at the nominal operating point.
Table 16. Results for GTLB-3-140.
Outlet Damper flow control n=900 rpm VSD flow control eta_fan=85 %
Flow q_v
[m³/s]
delta
p
[Pa]
eta_fan P_total
[kW]
P_FanSave
[kW]
q_v
[m³/s]
delta
p
[Pa]
n
[rpm]
P_total
[kW]
P_FanSave
[kW]
Nom.
flow
25,70 1320 70 % 51,57 51,23 17,70 2360 900 53,56 52,81
90 % 23,13 1760 79 % 54,82 49,18 15,93 1920 815 39,66 39,08
80 % 20,56 2120 82,5 % 56,21 46,62 14,16 1500 720 28,10 27,20
70 % 17,99 2370 85 % 53,37 43,80 12,39 1200 640 20,22 19,39
60 % 15,42 2500 84 % 48,86 40,47 10,62 830 450 12,71 12,67
50 % 12,85 2650 82,5 % 44,01 37,04 8,85 600 460 8,36 8,19
40 % 7,08 380 370 5,08 5,02
0
10
20
30
40
50
60
40 % 50 % 60 % 70 % 80 % 90 % Nom. flowFlow
Pow
er d
eman
d [k
W]
0
0,2
0,4
0,6
0,8
1
Effic
ienc
y
Calculated power demand for outlet damper Power demand for outlet damper by FanSaveCalculated power demand for VSD Power demand for VSD by FanSaveFan efficiency in outlet damper
Figure 32. Power demand and efficiency curves.
Power demand for outlet damper controlled fan based on the performance curve of the fan
is a lot higher than power demand calculated by FanSave. This result is valid at all other
flow rates than at the nominal flow. Efficiency of the fan increases when the flow is
75
throttled at first and starts to decrease when flow rate goes below 60 % of the maximum.
Motor’s efficiency remains at 94 % through the adjustment.
When fan is operating under a variable speed drive the power demands between
calculations are correlating. Efficiency of the fan remains constant during the speed
change. Motor’s efficiency decreases about 22 % and frequency converter’s efficiency
decreases about 11 %.
Conclusions
For variable speed drive controlled fans the results from this part of the thesis show that
FanSave’s calculation is accurate. There are no big differences in power demands
generated between calculations for any of the fans. Surprisingly though FanSave uses same
correction factors for power demand for fans with radial and backward curved blades the
calculation is still accurate. Driving motor’s and frequency converter’s efficiency changes
during speed change is taken into account accurately and there is no need to adjust
FanSave’s calculation for variable speed drive controlled fan.
FanSave’s calculation is a bit inaccurate for fans with radial blades when they are
controlled with outlet dampers. For most of the fans FanSave calculates too high power
demand through the operating range. Difference between FanSave’s calculation and
performance curve based calculations are not big. Correcting factors which FanSave uses
for calculating power demand could be adjusted to a little bit lower level. For more
accurate adjustments more examinations with different types and sizes of fans should be
executed.
Results for fans with forward curved blades show that FanSave’s calculation gives a bit too
low values for power demand at most parts of the operating range when fan is controlled
by outlet dampers. Curves created from values of power demands are in a bit different
shape. Curve created from FanSave’s calculation is more parabolic and curve created from
power demand values based on fans performance curve is more linear. Based on these
observations FanSave’s correcting factors for power demand at lowest and highest flow
76
rates are accurate and little adjustment should be made when flow rate is between 40 and
80 % of the maximum.
For outlet damper controlled fans with backward curved blades FanSave’s calculation
produces a bit too low values for power demand at the possible operation range for most
observed fan cases. Difference between calculations seems to grow larger as the fan sizes
get bigger. So maybe in this case should be different correcting factors be used for more
powerful fans than smaller fans. More accurate results would also be achieved by adjusting
the existing factors.
No changes are needed to make for FanSave’s calculation for fan controlled by a variable
speed drive. Calculation for power demand is perfectly balanced and accurate for each fan
case and this conclusion is valid for each studied fan regardless of blade curvature.
7.2 Axial fans
This part of the thesis consists of comparison of power demand for different sizes of axial
fans to FanSave’s calculation. Power demand for each fan is calculated by using the fans
performance curves. These can be found from appendixes F1 to F9.
First studied flow control method is pitch angle adjustment and the fans under the lens
come from manufacturer called Hürner Funken. Nominal operating point for each fan was
chosen from operating area of high volume flow and pitch angle. Flow was decreased 10 %
at a time by moving to a suitable pitch angle curve. Fans were also studied by choosing the
nominal point from the area of best efficiency but these studies did not match FanSave’s
calculations as well as above-mentioned method. Results are presented in tables and
illustrated as power demand curves in figures.
Other studied fans come from manufacturers called Applied Energy and Ebm-Papst.
Applied Energy XLCA1000 fan was chosen for inspection in order to get information from
operation of a highly powerful fan. Ebm-Papst fans are inspected under outlet damper and
variable speed flow control methods. First fan of the series is the least powerful and the
77
next fan in line is more powerful than previous one. Operational values for motors were
looked up from motor manufacturer Tamels website.
HF A 630
HF A 630 is the first studied fan and it is the smallest one of the Hürner Funken axial fans.
It is driven by a 1,1 kW motor at the efficiency of 73 %.
Table 17. Results for HF A 630 axial fan.
q_v
[m³/s]
deltap
[Pa]
eta_fan Pitch
angle
[°]
P_motorloss
[kW]
eta_m P_total
[kW]
P_FanSave
[kW]
Nom.
flow
3,35 152 65 % 50 0,21 79 % 0,99 1,00
90 % 3,02 140 70 % 45 0,17 78 % 0,77 0,74
80 % 2,68 125 70 % 40 0,14 77 % 0,62 0,53
70 % 2,35 115 70 % 35 0,13 75 % 0,52 0,37
60 % 2,01 105 70 % 30 0,12 70 % 0,40 0,26
50 % 1,68 88 75 % 25 0,11 65 % 0,32 0,18
40 % 1,34 75 60 % 20 0,11 60 % 0,28 0,13
30 % - - - - - - - 0,09
20 - - - - - - - 0,06
0,000,200,400,600,801,001,20
40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Pow
er d
eman
d [k
W]
0 %
20 %
40 %
60 %
80 %
100 %
Effic
ienc
y
Calculated power demand Power demand by FanSaveFan efficiency Motor efficiency
Figure 33. Power demand and efficiency curves.
78
Power demand based on pump’s performance curve is higher through almost whole
operating range. At high flow rates the power demand is well matching between
calculations. Fan’s efficiency remains quite high through the operating range but motor’s
efficiency decreases nearly 20 %. This contributes to the difference in power demands.
HF A 800
This fan is driven by a 4 kW electric motor which operates at the efficiency of 86,5 % at
the nominal operating point.
Table 18. Results for HF A 800 axial fan. q_v
[m³/s]
deltap
[Pa]
eta_fan Pitch
angle
[°]
P_motorloss
[kW]
eta_m P_total
[kW]
P_FanSave
[kW]
Nom.
flow
6,85 245 70 % 50 0,32 88 % 2,71 2,80
90 % 6,17 225 75 % 45 0,27 87 % 2,12 2,00
80 % 5,48 205 75 % 40 0,25 86 % 1,74 1,43
70 % 4,80 200 78 % 35 0,23 84 % 1,46 1,00
60 % 4,11 165 75 % 30 0,22 81 % 1,12 0,69
50 % 3,43 145 70 % 25 0,21 77 % 0,92 0,48
40 % 2,74 105 60 % 20 0,21 70 % 0,69 0,36
30 % - - - - - - - 0,24
20 - - - - - - - 0,17
0,000,501,001,502,002,503,00
40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlowPo
wer
dem
and
[kW
]
0 %
20 %
40 %
60 %
80 %
100 %
Effic
ienc
y
Calculated power demand Power demand by FanSaveFan efficiency Motor efficiency
Figure 34. Power demand and efficiency curves.
79
Results for this fan are quite similar compared to the previous one. Power demand based
on the characteristic curve is higher through most part of the operating range. It seems like
FanSave does not take into account any changes in motor’s efficiency, because if it
remains constant during the adjustment the power demand curves correlate seemingly
better.
HF A 1000
Last fan from Hürner Funken is more powerful than previous ones. It is powered by 11 kW
squirrel cage motor which operates at the efficiency of 91 % at the nominal operating
point.
Table 19. Results for HF A 1000 axial fan. q_v
[m³/s]
deltap
[Pa]
eta_fan Pitch
angle
[°]
P_motorloss
[kW]
eta_m P_total
[kW]
P_FanSave
[kW]
Nom.
flow
12,90 390 70 % 50 0,65 92 % 7,84 8,10
90 % 11,61 365 75 % 45 0,55 91 % 6,20 5,76
80 % 10,32 330 75 % 40 0,49 90 % 5,03 4,12
70 % 9,03 310 75 % 35 0,45 89 % 4,19 2,89
60 % 7,75 280 75 % 30 0,42 87 % 3,32 2,00
50 % 6,45 225 70 % 25 0,40 84 % 2,48 1,37
40 % 5,16 190 65 % 20 0,39 79 % 1,90 1,03
30 % - - - - - - - 0,69
20 - - - - - - - 0,48
80
0,002,004,006,008,00
10,00
40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Pow
er d
eman
d [k
W]
0,0 %20,0 %40,0 %60,0 %80,0 %100,0 %
Effic
ienc
y
Calculated power demand Power demand by FanSaveFan efficiency Motor efficiency
Figure 35. Power demand and efficiency curves.
For this fan also the results for power demand are similar as with previous fans.
Characteristic curve based power demand is higher than power demand calculated by
FanSave through the whole operation range. Fan’s efficiency remains quite good through
the adjustment dropping about 10 % when entering the lowest possible flow rates.
Efficiency of the motor decreases about 13 % during the pitch angle adjustment.
Applied Energy XLCA1000
The last fan which operation was studied under pitch angle adjustment is Applied Energy
XLCA1000 axial fan. Driving motor for this fan produces 30 kW and has the efficiency of
93,2 % at the nominal operating point.
Operation of this fan was observed with different approaches and compared to results from
FanSave’s calculation. First approach was to choose the nominal point at operating area
where fan operates at relatively high volume flow and efficiency. Flow was decreased 10
% at a time and operating point was moved to a suitable pitch angle curve. Fan’s efficiency
decreased from 73 % to 65 % during the adjustment. This approach did not match
FanSave’s calculation at all. Power demand was a lot higher based on the fan’s
performance curve than in results of FanSave through the whole operating range.
Next approach was to choose the nominal operating point at the operating area where
pump operates at the best possible efficiency. For this pump this means 60 % flow of the
maximum possible flow. Flow was decreased 10 % at a time just like before. Every
81
operating point was chosen so that the pressure generation would be as low as possible and
the efficiency as high as possible. Therefore the power demand would be as low as
possible. Still a large difference is formed between the power demands of the calculations.
Also the operating range of the fan was very narrow.
Last approach matched the FanSave’s calculation the best. Now the nominal operating
point of the fan was chosen at area of high volume flow and pitch angle. The efficiency of
the fan is lower at first but when the flow is adjusted smaller it gets better. Results for this
case are presented in table 20 and as power demand and efficiency curves in figure 36.
Fan’s performance curve can be found in appendix F4.
Table 20. Results for XLCA1000. q_v
[m³/s]
deltap [Pa] eta_fan Pitch angle
[°]
eta_m P_total
[kW]
P_FanSave
[kW]
Nom.
flow
19,43 800 67,5 % 32 93 % 24,65 25,60
90 % 17,49 785 70,5 % 28 93 % 20,86 18,18
80 % 15,54 707 73,3 % 24 93 % 16,14 13,01
70 % 13,60 580 73 % 20 92 % 11,79 9,12
60 % 11,66 488 74 % 18 90 % 8,58 6,30
50 % 9,72 560 73,5 % 14 89 % 8,29 4,33
40 % 7,77 544 72 % 10 87 % 6,73 3,25
30 % 5,83 582 68 % 8 86 % 5,83 2,18
0,00
5,00
10,0015,00
20,00
25,00
30,00
30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom. flowFlow
Pow
er d
eman
d [k
W]
0,0 %
20,0 %
40,0 %
60,0 %
80,0 %
100,0 %Ef
ficie
ncy
Calculated power demand for pitch angle adjustment Power demand for pitch angle adjustment by FanSaveFan efficiency Motor efficiency
Figure 36. Power demand and efficiency curves.
82
Ebm-Papst fans
Next flow control methods observed are outlet damper and rotational speed control
handled by a frequency converter. Fan models come from manufacturer called Ebm-Papst
and they are relatively small powered and their operating ranges cannot match the smallest
flow rates calculated by FanSave. Each fan’s performance curve is marked with operation
points which were used in observing the operation of the fan and in calculation of power
demand. Outlet damper flow control method was studied by moving along fan curve to a
new operating point. Rotational speed adjustment was studied by tracing the new operating
point via system curve. Curves for fans are found in appendixes F5 to F9.
3G400
First observed fan in this series is 3G400 which is driven by 0,5 kW motor. It operates at
efficiency of 72 % at its nominal operating point.
Table 21. Results for 3G400.
Outlet Damper flow control n=1660 rpm VSD flow control
Flow q_v
[m³/s]
delta
p
[Pa]
eta_fan P_total
[kW]
P_FanSave
[kW]
q_v
[m³/s]
delta
p
[Pa]
n
[rpm]
P_total
[kW]
P_FanSave
[kW]
Nom.
flow
1,27 88 64 % 0,23 0,24 1,27 88 1660 0,24 0,25
90 % 1,14 125 62 % 0,30 0,28 1,14 70 1450 0,18 0,18
80 % 1,02 160 59 % 0,35 0,28 1,02 55 1300 0,14 0,13
70 % 0,89 180 55 % 0,37 0,27 0,89 42 1150 0,11 0,09
60 % 0,76 190 53 % 0,34 0,23 0,76 30 1000 0,08 0,06
83
0,0000,0500,1000,1500,2000,2500,3000,3500,400
60 % 70 % 80 % 90 % Nom. flowFlow
Pow
er d
eman
d [k
W]
0 %
10 %20 %
30 %
40 %
50 %60 %
70 %
Effic
ienc
y
Calculated power demand for outlet damper Power demand for outlet damper by FanSaveCalculated power demand for VSD Power demand for VSD by FanSaveFan efficiency in outlet damper
Figure 37. Power demand and efficiency curves for 3G400.
For outlet damper flow control method there is clear difference in power demands as can
be seen from figure 37. Based on pump’s performance curve the power demand is higher
than power demand calculated by FanSave through the whole operating range. Figure 37
contains also the efficiency curve for the fan in outlet damper flow control. Efficiency
decreases about 11 % during the flow adjustment. Efficiency in speed control remains
constant and therefore it is not illustrated in figure 37.
Power demands in variable speed control are well matching for this pump case. A little
difference starts to build up when operation enters smaller flow rates. All in all it is clear
that for this case FanSave’s calculation is more correct in variable speed control than in
outlet damper control.
3G990
Next fan is 3G990. It is powered by 1,1 kW electric motor which operates at 78 % of
efficiency at the nominal operating point.
Table 22. Results for 3G990.
Outlet Damper flow control n=650 rpm VSD flow control
Flow q_v
[m³/s]
delta
p
[Pa]
eta_fan P_total
[kW]
P_FanSave
[kW]
q_v
[m³/s]
delta
p
[Pa]
n
[rpm]
P_total
[kW]
P_FanSave
[kW]
Nom. 7,64 40 72 % 0,54 0,54 7,64 40 650 0,56 0,56
84
flow
90 % 6,88 70 71 % 0,84 0,62 6,88 35 580 0,45 0,51
80 % 6,11 93 67 % 1,05 0,64 6,11 26 565 0,32 0,29
70 % 5,35 116 60 % 1,29 0,61 5,35 19 450 0,23 0,20
60 % 4,58 135 54 % 1,44 0,52 4,58 15 400 0,18 0,13
0,0000,2000,4000,6000,8001,0001,2001,4001,600
60 % 70 % 80 % 90 % Nom. flowFlow
Pow
er d
eman
d [k
W]
0 %10 %20 %30 %40 %50 %60 %70 %80 %
Effic
ienc
y
Calculated power demand for outlet damper Power demand for outlet damper by FanSaveCalculated power demand for VSD Power demand for VSD by FanSaveFan efficiency in outlet damper
Figure 38. Power demand and efficiency curves for 3G990.
Fan curve for this case is very steep and it causes the difference in power demands when
fan operates under outlet damper flow control method. Because the fan curve is so steep
the pressure difference generated by the fan increases rapidly. Also the fan’s efficiency
decreases about 18 % which also contributes to the difference between power demands.
For rotational speed control the power demands are quite well matching. Power demand
based on the fan’s performance curve is slightly higher than the power demand based on
FanSave. Efficiency in speed control decreases about 2 %.
3G800
Next fan is 3G800 and it is second to last in this series. It is driven by 1,8 kW motor at the
efficiency of 77 %.
Table 23. Results for 3G800.
Outlet Damper flow control n=1030 rpm VSD flow control
Flow q_v delta eta_fan P_total P_FanSave q_v delta n P_total P_FanSave
85
[m³/s] p
[Pa]
[kW] [kW] [m³/s] p
[Pa]
[rpm] [kW] [kW]
Nom.
flow
5,56 110 63 % 1,19 1,26 5,56 110 1030 1,22 1,29
90 % 5,00 149 60 % 1,51 1,44 5,00 90 940 0,92 0,95
80 % 4,44 182 58 % 1,69 1,48 4,44 70 810 0,67 0,66
70 % 3,89 205 50 % 1,93 1,42 3,89 54 720 0,49 0,47
60 % 3,33 220 45 % 1,97 1,21 3,33 40 630 0,36 0,31
0,000
0,500
1,000
1,500
2,000
2,500
60 % 70 % 80 % 90 % Nom. flowFlow
Pow
er d
eman
d [k
W]
0 %10 %20 %30 %40 %50 %60 %70 %
Effic
ienc
y
Calculated power demand for outlet damper Power demand for outlet damper by FanSaveCalculated power demand for VSD Power demand for VSD by FanSaveFan efficiency
Figure 39. Power demand curves for 3G800.
For outlet damper in this case the results are similar to previous cases. Based on the fan’s
performance curve power demand is higher than based on FanSave’s calculation through
the whole operating range. Power demand rises through the adjustment according to the
characteristic curve but according to FanSave the power demand starts to drop when flow
rate is decreased below 80 %. Efficiency decreases about 18 % during the flow adjustment.
During speed adjustment the power demands are well matching between calculations.
Efficiency of the fan decreases only about 2 % when adjusting the flow.
3G650
Second to last fan of the series is 3G650. It is driven by a squirrel cage motor which
produces 3 kW at the efficiency of 78 %.
86
Table 24. Results for 3G650.
er flow control n=1350 rpm VSD flow control Outlet Damp
Flow q_v
]
]
]
a]
]
P_total P_FanSave
[m³/s
delta eta_fan P_total P_FanSave
p
[Pa
[kW] [kW]
q_v delta n
[m³/s p
[P
[rpm [kW] [kW]
Nom. 5,19 74 % 1,01 0,94 5,19 1350 1,04 0,96
flow
105 105
90 % 4,67 180 71 % 1,50 1,08 4,67 85 1200 0,81 0,71
80 % 4,15 235 67 % 1,81 1,10 4,15 65 1050 0,67 0,50
70 % 3,63 280 63 % 1,99 1,06
60 % 3,11 325 58 % 2,15 0,90
0,000
0,500
1,000
1,500
2,000
2,500
60 % 70 % 80 % 90 % Nom. flowFlow
Pow
er d
eman
d [k
W]
0 %10 %20 %30 %40 %50 %60 %70 %80 %
Effi
cien
cy
Calculated power demand for outlet damper Power demand for outlet damper by FanSaveCalculated power demand for VSD Power demand for VSD by FanSaveFan efficiency in outlet damper Fan efficiency in VSD
Figure 40. Power demand and efficiency curves for 3G650.
or this fan in outlet damper the power demands between calculations differs even more
variable speed control fan’s operating range is very narrow for this fan. Power demand
G710
F
than in previous fan case. Fan’s power demand increases at smaller flow rates due to
higher pressure development and decrease of efficiency. Efficiency drops about 16 %
during the flow adjustment.
In
is slightly higher when calculation is based on fan’s characteristic curve than when power
demand is calculated by FanSave. This result is valid through the whole operating range.
Efficiency decreases 10 % during speed adjustment.
3
87
Last fan of the series is 3G710 and it is powered by 2,2 kW motor. It has 77 % efficiency
able 25. Results for 3G710.
er flow control n=1230 rpm VSD flow control
at the nominal operating point.
T
Outlet Damp
Flow q_v
]
]
]
a]
]
P_total P_FanSave
[m³/s
delta eta_fan P_total P_FanSave
p
[Pa
[kW] [kW]
q_v delta n
[m³/s p
[P
[rpm [kW] [kW]
Nom. 5,47 71 % 1,48 1,52 5,47 1230 1,52 1,55
flow
152 152
90 % 4,93 204 67 % 1,90 1,73 4,93 123 1100 1,14 1,15
80 % 4,38 257 63 % 2,27 1,78 4,38 98 970 0,84 0,80
70 % 3,83 75 880 0,61 0,57
0,000
0,500
1,000
1,500
2,000
2,500
70 % 80 % 90 % Nom. flowFlow
Pow
er d
eman
d [k
W]
58 %
60 %
62 %
64 %
66 %
68 %
70 %
72 %
Effic
ienc
y
Calculated power demand for outlet damper Power demand for outlet damper by FanSaveCalculated power demand for VSD Power demand for VSD by FanSaveFan efficiency in outlet damper
Figure 41. Power demand and efficiency curves for 3G710.
his fan has very limited operating range. For outlet damper results for power demand are
hen fan operates under variable speed drive the results for power demand correlate
onclusions
T
similar to previous cases. Fan reaches only 80 % of the maximum flow as the lowest flow
rate. Efficiency decreases about 8 % during the adjustment
W
between performance curve based calculations and FanSave’s calculations. Efficiency
remains at 71 % through the adjustment.
C
88
Results for pitch angle adjustment are similar for each fan case. Power demand calculated
hen controlling flow with outlet dampers the pressure difference generated by fan
hen fans are controlled by a variable speed drive the results for power demand between
by FanSave is smaller than power demand calculated with each fans performance curve.
This result is valid for each fan when flow rate is below selected nominal flow. Behaviour
of fan efficiency is also quite similar in all cases. It increases first when flow is decreased
and then starts to decrease when the operating range gets near its end. Anyway the changes
are not drastic and don’t have much effect on the power demand. Efficiency of the driving
motor is calculated by using equations 40, 41, 42 and 43. It decreases from 7 to 19 %
depending on the fan case. Changes in motor’s efficiency affect on the total power
demand. It is hard to say what changes in fan applications operation FanSave’s constant
factors for power demand covers. However these results show that the factors should be
revised. A little higher values would produce more truthful results for pitch angle flow
control method.
W
increases rapidly when flow is choked. Based on the performance curves of inspected fans
the power demand increased through the whole operating range in all cases but one. On the
other hand the operating ranges for the fans were so narrow that the power demand didn’t
have time to decrease until the operating range reached its limits. There was much less
changes in power demands for fans according to FanSave’s calculations. Power demand in
FanSave’s calculation increased at first a little bit when the flow was choked and then
decreased when flow rate went below 80 % of the maximum. Efficiencies of the fans
decreased quite a lot at each fan case which also increases the power demand. Constant
factors FanSave uses should be corrected so that the increase of power demand is more
rapid at first when the flow is choked and the turning point where the power demand starts
to decrease should be moved to operating area of smaller flow rate.
W
calculations are well matching for each fan case. Efficiencies of fans decrease only few
percents during the adjustment and thereby don’t have much effect on power demands.
Driving motors efficiencies decrease a lot more but factors of FanSave seem to take this
into account because there is not much difference in power demands.
89
Based on these examinations FanSave’s calculation is very accurate for variable speed
drive controlled axial fans and not so accurate for axial fans with outlet dampers or pitch
angle adjustment. When the calculation of power demand is inaccurate also the results for
total energy consumption of the application is incorrect. How much it differs depends on
the length and profile of the operation time.
8. LABORATORY TEST RUNS WITH CENTRIFUGAL FAN
8.1 Laboratory equipment
To compare calculation results of FanSave –energy saving tool, laboratory test runs were
performed at Lappeenranta university of technology. Fan which was used was a Ventur
GMT-100T centrifugal fan equipped with straight radial blade impeller. The manufacturer
of the fan didn’t have any information about the efficiency of the fan and there were no
efficiency curves created either. Frequency converter used in the tests was a Mitsubishi
Freqrol Z200 and the electric motor used in the tests was a Tamel SKg 71-2B squirrel cage
motor. Manufacturer states that motor’s efficiency is 75 %. This value was used in
FanSave’s calculation for comparison. Fans impeller was installed directly to the axel of
the motor so there were no transmission losses generated. Input voltage to the motor was
400 V at 50 Hz and the motor and the fan rotated at 2790 rpm. Duct consisted suction pipe
and discharge pipe which had an orifice plate for pressure measurement and a throttling
valve for adjusting the flow. Diameter of the hole in the orifice plate was 59,998 mm and
the diameter of the piping was 108,14 mm.
90
Figure 42. Ventur GMT 100 T centrifugal fans characteristic curve and dimensions (Ventur).
Flow could be adjusted with the choke valve at the discharge pipe and by adjusting the
rotational speed of the fan with the frequency converter. Also the inlet box damper flow
control was simulated by choking the inlet of the fan. Two speed motor as a flow control
method was also simulated by operating the fan at two different rotation speeds. Frequency
converter was used to adjust the speed.
Pressure difference transmitters were used to measure the total pressure difference
generated by the fan and the pressure difference at the both sides of the orifice plate.
Pressure difference transmitters were connected to the orifice plate as corner tapping. The
discharge pipe contained a thermoelement so the rise in the temperature could be
measured. Fluke hydra data logger was used to measure the temperature and the pressure
differences. Input power to the electric motor was measured with Fluke power logger 1735
measuring device. Rotational speed was calculated with the frequency shown on the
display of the frequency converter.
91
Figure 43. Laboratory equipment in the university laboratory.
8.2 Measurements and comparison of results to FanSave’s calculation
Measurement report is shown in appendix H1 for full results of the tests. Volume flow in
all the test runs was calculated by using iteration according to SFS ISO 5167 standard.
Tables for discharge coefficient and expansibility factor can be found in the appendixes C
and D. In order to be able to compare the laboratory test results and the results received
from FanSave’s calculation the initial values for FanSave were the same values produced
by the fan used in laboratory tests. These test results are comparable only for same size
centrifugal fans. More tests should be made with various size of equipment to attain all-
inclusive information about FanSave’s calculation.
8.2.1 Outlet damper
Throttling valve was used for adjusting the flow in throttling control and simulating the
outlet damper flow control method. Due to the structure of the valve the flow could only be
adjusted to about 70 % of the maximum flow. Fan was operated at its full capacity at first
and the flow was reduced 10 % at a time. Results of the throttling valve test can be red
from table 26.
Table 26. Calculations for outlet damper controlled fan.
Flow q_v [m³/s] delta p eta_fan P_motorloss eta_m P_total P_FanSave
92
[Pa] [kW] [kW] [kW]
Nom.
flow
0,0941 1377 47 % 0,072 79 % 0,349 0,345
90 % 0,0846 1372 45 % 0,070 79 % 0,330 0,331
80 % 0,0758 1418 43 % 0,068 79 % 0,318 0,314
70 % 0,0678 1451 40 % 0,068 78 % 0,313 0,295
0,26
0,28
0,3
0,32
0,34
0,36
70 % 80 % 90 % Nom. flowFlow
Pow
er d
eman
d [k
W]
0,36
0,38
0,4
0,42
0,44
0,46
0,48
Effi
cien
cy
Calculated power demand Power demand by FanSave Fan efficiency
Figure 44. Power demand curves for outlet damper controlled fan.
Results in table 26 and in figure 44 show that calculation of FanSave is well correlating
with results from laboratory test when flow rate is 80 % of the maximum or higher.
Entering lower flow rates test applications power demand curve straightens and starts to
differ from curve created by FanSave’s calculation. Main cause for this is the decreasing of
fan efficiency which speeds up a bit. One cause may be found from the constant factors
FanSave uses for both backward curved and radial blades equipped fans.
8.2.2 Variable speed control
Rotational speed and volume flow through the fan was adjusted by frequency converter.
Flow was decreased 10 % at the time starting with nominal flow. Throttling valve at the
discharge pipe was open fully during this test. Results of the tests are presented in table 27.
Testissä käytetyn taajuusmuuttajan nimellisteho oli 2,2 kilowattia, joten se ei toiminut sille
ominaisella toiminta-alueella, mistä selittyy sen heikohko hyötysuhde. Tämä huomioitiin
93
myös FanSave –ohjelman laskennassa, jossa taajuusmuuttajan hyötysuhteelle syötettiin
puhaltimen nimellisvirtausta vastaava arvo 91 %.
Table 27. Calculations for variable speed drive controlled fan.
FanSave
q_v
[m³/s]
deltap
[Pa]
eta_fan n
[rpm]
P_motorloss
[kW]
eta_m P_driveloss
[kW]
eta_drive P_total
[kW]
P [kW]
Nom.
flow
0,0914 1377 47 % 2970 0,072 79 % 0,034 91 % 0,383 0,382
90 % 0,0847 1185 50 % 2673 0,061 77 % 0,033 89 % 0,295 0,283
80 % 0,0759 1061 55 % 2376 0,052 74 % 0,031 86 % 0,231 0,197
70 % 0,0659 919 62 % 2079 0,045 69 % 0,030 83 % 0,173 0,140
60 % 0,0565 788 67 % 1782 0,040 62 % 0,029 79 % 0,136 0,092
50 % 0,0471 656 66 % 1485 0,037 56 % 0,028 75 % 0,112 0,059
40 % 0,0380 525 46 % 1188 0,036 55 % 0,028 74 % 0,107 0,036
30 % 0,0287 398 25 % 891 0,038 55 % 0,028 75 % 0,112 0,021
20 % 0,0190 218 5 % 594 0,075 51 % 0,031 83 % 0,184 0,013
0
0,1
0,2
0,3
0,4
0,5
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Pow
er d
eman
d [k
W]
00,10,20,30,40,50,60,70,8
Effi
cien
cy
Calculated power demand Power demand by FanSave Fan efficiency
Figure 45. Power demand curves for variable speed drive controlled fan.
Results show that the generated factors for power demand for this fan follows the factors at
the FanSave calculations well as the flow decreases to 60 % of the nominal flow. Flows
smaller than 60 % of the nominal flow the power demand starts to differ compared to
FanSave calculations. Power demand for laboratory test application is greater at 20 % and
30 % of the nominal flow than at 40 % of the nominal flow. This is because the efficiency
of the fan collapses rapidly. Efficiency curve is also shown in figure 45. Power demand
94
curve was generated to clarify the results and there is a clear difference between the shapes
of the power demand curves.
When the power demand differs from laboratory test application the amount of used
energy differs also. This affects on the amount of used energy during the operation time of
the fan. It also has effect on the energy costs. How much the energy consumption and the
costs differ depends on the size of the application and the operation profile during the
operation time.
8.2.3 Two speed motor
Two speed motor as a flow control method was simulated by using the frequency converter
and operating the fan and the electric motor at two different rotation speeds. FanSave
assumes that when the flow rate is 60 % or less of the nominal flow the shaft power
demand is 25 % of the maximum power demand.
Power demand for two speed motor controlled fan can be examined by using table 27
which includes the results from variable speed drive controlled test. According to FanSave
60 % flow rate needs 25 % power of the maximum. So 60 % flow demands about 0,107
kW power according to table 27 when power demand increase caused by frequency
converter efficiency is ignored. For this fan case 25 % power of maximum power equals
0,08 kW of power and it is not enough for 60 % flow. So for this case FanSave’s
calculation is inaccurate.
8.2.4 Inlet box damper
Inlet box damper flow control method was simulated as choking the inlet of the fan.
Results can be seen in table 28 and figure 45. Also the power demand curves were created
based on laboratory tests and FanSave’s calculations.
95
As can be seen from table 28 and power demand curves in figure 46 there is considerable
difference between power demand for laboratory equipment and a fan application
calculated by FanSave.
Table 28. Calculations for inlet damper controlled fan.
Flow q_v
[m³/s]
delta_p
[Pa]
eta_fan P_motorloss
[kW]
eta_m P_total
[kW]
P_FanSave
[kW]
Nom.
flow
0,0942 1377 47 % 0,072 79 % 0,349 0,345
90 % 0,0845 1185 39 % 0,070 79 % 0,329 0,290
80 % 0,0759 1065 33 % 0,068 79 % 0,317 0,259
70 % 0,0657 926 26 % 0,066 78 % 0,298 0,231
60 % 0,0564 795 20 % 0,064 77 % 0,284 0,214
50 % 0,0468 649 15 % 0,062 77 % 0,269 0,200
40 % 0,0377 536 10 % 0,061 76 % 0,259 0,186
30 % 0,0284 398 6 % 0,060 76 % 0,249 0,172
20 % 0,0188 225 2 % 0,059 75 % 0,235 0,162
00,050,1
0,150,2
0,250,3
0,350,4
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Pow
er d
eman
d [k
W]
0
0,1
0,2
0,3
0,4
0,5Ef
ficie
ncy
Calculated power demand Power demand by FanSave Fan efficiency
Figure 46. Power demand curves for inlet damper controlled fan.
Power demand calculated by FanSave is smaller through the operation area. This being
also the total energy consumption in operating time is smaller in FanSave calculation
depending on the length of the operating time and on the size of the application. Efficiency
curve is also shown in figure 46 and it collapses rapidly when flow is decreased.
96
8.3 Energy consumption examination in operation time
Energy consumption between FanSave’s calculation and results from laboratory tests are
compared in this part of the thesis. Annual operation time is chosen to be 5000 hours and
profile for the operation time was generated with FanSave’s default button for all other
flow control methods except for outlet damper control. Outlet damper control was created
its own profile for operation time because due to limitations of equipment power demand
for smallest flow rates was not possible to test.
Results from energy consumption comparison is presented as bar diagrams in figure 47.
Results show that there are no big differences between fan tested in laboratory and similar
fan calculated with FanSave. Though over 100 kWh difference is generated between
calculations for outlet damper and inlet damper control methods. For outlet damper energy
consumption is higher based on fan tested in laboratory and for inlet damper control
FanSave calculates higher energy consumption.
Energy consumption
FanSave
FanSave
FanSave FanSave
Lab
Lab
Lab
Lab
0
500
1000
1500
2000
kWh
FanSave 1617 636 1157 1048
Lab 1483 592 1317 990
Outlet damper [kWh]
Variable speed control
Inlet box damper [kWh]
Two-speed motor
Figure 47. Energy consumption comparison between FanSave and laboratory equipment.
9 LABORATORY TEST RUNS WITH AN AXIAL FAN
97
9.1 Laboratory equipment
Axial fan used in laboratory tests was connected to the same piping as the centrifugal fan
with the exception that the suction pipe was not connected to the fan. Fan sucked air from
the surrounding space. Used fan is a compact appliance consisting of a fan impeller, an
electric motor and a frequency converter with a remote control panel. Specific information
about the machine was not available so data from similar appliances was applied in
calculations. Data for the driving motor was used from Tamel Sg71-2A squirrel cage motor
which produces 0,7 kW and operates at the efficiency of 70 % at the nominal operating
point. Devices used in measuring were the same used for centrifugal fan.
Examined flow control methods are variable speed control, outlet damper and inlet
damper. Though dampers are seldom used with axial fans they were simulated because of
the opportunity to do so.
9.2 Variable speed drive
Efficiency of the frequency converter was assumed to be 90 % at the nominal operating
point of the application. Efficiencies were calculated by using the power losses calculated
with equations 41, 42, 43, 44 and 45. Results are presented as power demand and
efficiency curves in figure 48.
98
Variable speed drive
00,050,1
0,150,2
0,250,3
0,350,4
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Pow
er d
eman
d [k
W]
0,0000,1000,2000,3000,4000,5000,6000,7000,8000,9001,000
Effic
ienc
y
Measured power demand Power demand by FanSaveFan efficiency Motor efficiencyDrive efficiency
Figure 48. Curves for VSD control.
Results for power demand in figure 47 show that measured power is little bit higher than
power demand calculated by FanSave when flow rate is below 90 % of the maximum.
However the difference is very small so FanSave’s calculation is quite accurate for this
type of fan operating under variable speed drive. Calculated efficiency of the fan behaves
quite oddly reaching its highest values when flow rate is 30 and 40 % of the maximum.
Assumption of motor’s efficiency may have been too high because of the shape of the
efficiency curve in figure 48. First it decreases rapidly when speed decelerated and later
the decreasing is not so fast anymore. Drives efficiency decreases quite rectilinearly
through the operating range.
9.3 Outlet damper
Outlet damper flow control was simulated in laboratory by throttling the flow 10 % at a
time at the outlet of the fan. Driving motor operated at full 3000 rpm speed trough the test
run. Power loss caused by the frequency converter was not taken into account. Results are
presented as power demand and efficiency curves in figure 49.
99
Outlet damper
0
0,1
0,2
0,3
0,4
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Pow
er
dem
and
[kW
]
00,020,040,060,080,1
Effic
ienc
y
Measured power demand Power demand by FanSave Appliance efficiency
Figure 49. Curves for outlet damper adjustment.
Power demand for the application was constant through the test. Shape of the power
demand curve created by FanSave’s calculation is in totally different shape. Power
demands are matching when flow rate is 50, 60 or 100 %. Otherwise FanSave calculates
higher power demand for this type of fan application. Efficiency illustrated in figure 49 is
for the whole appliance. It is the result of multiplication of motor’s assumed efficiency and
fan’s calculated efficiency. It is decreasing because the power demand of the appliance
remains constant and the flow rate decreases. Pressure difference of the fan does not
increase much during the throttling.
9.4 Inlet damper
Inlet box damper flow control was simulated by throttling the flow from the intake of the
fan 10 % at a time. Driving motor operated at full speed through the test and the power loss
caused by the frequency converter was not taken into account. Results are presented as
power demand and efficiency curves in figure 50.
100
Inlet damper
0,0000,0500,1000,1500,2000,2500,3000,350
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Pow
er
dem
and
[kW
]
0,000
0,020
0,040
0,060
0,080
0,100
Fan
effic
ienc
y
Measured power demand Power demand by FanSave Appliance efficiency
Figure 50. Power demand and efficiency curves.
Measured power demand is seemingly higher through the whole operating range than
power demand calculated by FanSave. Measured power demand does not decrease very
much during the adjustment when power demand calculated by FanSave splits in half
when fan is operating at the lowest flow rate. Efficiency illustrated in figure 50 is for the
whole appliance. It is the result of multiplication of motor’s assumed efficiency and fan’s
calculated efficiency.
Conclusions
For variable speed control the results for power demands are quite well correlating
between measuring and calculation of FanSave. Because FanSave uses constant factors to
calculate power demand at different flow rates it is impossible to say how the program
assumes the behaviour of efficiencies for different devices.
For outlet damper flow control method performed test run does not give very realistic
results. Measured power demand is constant through the test run and curve created from
FanSave’s calculation is wavy and achieves high points at two different flow rates.
Increase in pressure difference is very little. This may be caused by the leaks in the
pipeline.
101
Power demand results for inlet damper flow control are not matching at all between
measured values and FanSave’s values. Measured power demand remains quite high
during the adjustment and decreasing of power demand calculated by FanSave is parabolic.
Changes in FanSave’s calculation based on these test runs are not necessary. FanSave’s
calculation for variable speed drive is proven to be accurate and damper flow control with
axial fans is extremely rare. A lot more test runs with different types and sizes of axial fans
should be performed to achieve comparable results for FanSave’s calculations.
10 PERFORMANCE CURVE BASED CALCULATION COMPARISON TO PUMPSAVE’S CALCULATION
In this part of the thesis PumpSave’s calculation is compared to different types and sizes of
centrifugal pumps. Calculation comparison is based on each pump’s characteristic curve.
Flow controls observed are throttling control and rotational speed control driven by a
frequency converter. Hydraulic control was left aside because it is not as common flow
control method than throttling or rotational speed control. Reading pump’s characteristic
curve is not absolutely precise so the calculations done based on the characteristic curves
may contain little errors.
PumpSave calculates power demand for throttling control according to equation 62 where
changes in head are noticed by term and changes in pump’s efficiency are noticed by
the term . Term is affected by pump’s nominal flow, throttled flow and
maximum and nominal head as equation 63 shows. Equation 64 shows that pump’s
efficiency changes are dependent on pump’s nominal flow and throttled flow. Calculation
performed with pump’s characteristic curve also examines how these factors change during
throttling and changes are compared to PumpSave’s calculation.
i,piH
i,pik i,piH
10.1 VSD
PumpSave calculates power demand for variable speed drive controlled pump application
as equation 74 shows. It pays attention to changes efficiencies of frequency converter and
102
driving motor when flow rate changes. PumpSave uses constant factors when calculating
change in frequency converter’s efficiency. These factors are presented in table 2.
PumpSave uses equation 75 to calculate change in driving motor’s efficiency. It depends
on flow rate, static head and pump’s nominal head. Static head affects also on system
curve which PumpSave calculates by equation 72. Static head and system curve was
created for each pump case separately. Operation range for each pump was attempted to
keep as large as possible. Therefore also the comparison to PumpSave’s calculation could
be done in wide operation range. Change in pump efficiency was observed with affinity
parable.
Comparison between characteristic curve based calculation and PumpSave’s calculation
was executed by choosing wide range of different types of pumps. Pumps were chosen
from manufacturer called Goulds Pumps which offer wide range of pumps for different
purposes. Company’s website provides a useful tool for selecting pumps. Review of pumps
starts from least powerful model and next pump in line is always more powerful than
previous one. Operational values for driving motor’s were looked up from electric motor
manufacturer Tamels website.
For efficiency changes of drive motor equations 41, 42, 43 and 44 were used. Frequency
converter’s efficiency change was calculated with equation 45. Pumps characteristic curves
can be found at appendixes.
Goulds 3298
First studied pump is Goulds CDS 5309-1 model 3298 which operates at high rotational
speed and reaches quite high head. Static head was chosen to be 15 m for this pump and
system curve was generated according to it. Driving motor was selected to operate at 37
kW power and 93,3 % efficiency at the nominal operating point.
103
0
5
10
15
20
25
30
35
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Pow
er d
eman
d [k
W]
0
20
40
60
80
100
120
Hea
d [m
]
Calculated pow er demand Pow er demand by PumpSave
Head Head by PumpSave
Figure 51. Power demand and head curves.
From power demand curves at figure 51 can be seen that the power demand curves
between calculations correlate almost perfectly together. Head curves in figure are equal to
system curves of calculations. These curves match also extremely well. Pump does not
reach the smallest flow rate calculated by PumpSave. Pump’s efficiency changes for better
at first when flow is decreased and decreases when operating point enters smaller flow
rates. However the changes are so small that they don’t have effect on total power demand.
0
0,2
0,4
0,6
0,8
1
1,2
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Effic
ienc
y
Motor eff iciency Motor eff iciency by PumpSave
Drive eff iciency Drive eff iciency by PumpSave
Figure 52. Efficiency curves.
Figure 52 shows driving motor’s and frequency converter’s efficiencies as curves between
calculations. Based on characteristic curve based calculations motor’s efficiency is better
104
through most part of the operation range. Efficiency starts to decrease more rapidly when
flow rate reaches about 40 % of the maximum. Based on PumpSave frequency converter’s
efficiency remains better than when efficiency is calculated using equation 45 through the
whole operating range. Curve created for drives efficiency calculated by PumpSave is
always the same shape due to the constant factors PumpSave uses.
GOUDLS IC/ICB/ICP/ICV
Next compared pump produces relatively small volume flow but high head. Static head
was chosen to be 35 m. Motor driving the pump is 55 kW producing squirrel cage motor
which operates at 94 % efficiency at nominal operating point.
0,000
10,000
20,000
30,000
40,000
50,000
60,000
70,000
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Pow
er d
eman
d [k
W]
0
50
100
150
200
250
Hea
d [m
]
Calculated pow er demand Pow er demand by PumpSave
Head Head by PumpSave
Figure 53. Power demand and head curves.
Results in figure 53 show that PumpSave’s calculation for power demand is well matching
at most points of operating range. When entering the smallest possible flow rates power
demand according to characteristic curve is slightly bigger than PumpSave’s calculated
power demand. This is caused by changes in pump’s efficiency. It collapses when flow and
rotation speed decreases enough. Pump’s efficiency curve is illustrated in figure 53.
PumpSave’s calculation doesn’t take into account any changes in pump’s efficiency when
changing rotation speed. Figure 52 also shows that head curves between calculations
correlate almost perfectly.
105
0,000
0,200
0,400
0,600
0,800
1,000
1,200
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Effic
ienc
y
Motor eff iciency Motor eff iciency by PumpSave
Drive eff iciency Drive eff iciency by PumpSave
Pump eff iciency
Figure 54. Efficiency curves.
Figure 54 illustrates efficiency changes among devices. In this case pump’s efficiency
curve is also included because pump’s efficiency changes considerably during speed
adjustment. This also has an affect on total power demand of the application. Driving
motor’s efficiency is better based on characteristic curve based calculation than in
PumpSave’s calculation through whole operating area. Situation is the opposite for
frequency converter. These changes in efficiencies does not really have an effect on the
total power demand.
Goulds 3638
Next pump produces big volume flow but lower head than previous models. Pump operates
at high 85 % efficiency at its nominal operating point. Static head and starting point of
system curve was chosen to be 5 meters. Driving motor for pump operates at 90 kW power
and 95 % efficiency at its nominal operating point.
106
010
2030
4050
6070
8090
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Pow
er d
eman
d [k
W]
0
5
10
15
20
25
30
35
40
Hea
d [m
]
Calculated pow er demand Pow er demand by PumpSave
Head Head by PumpSave
Figure 55. Power demand and head curves.
For this pump the results for power demand and head are well matching through whole
operating range between calculations. Pump does not reach the smallest flow rate
calculated by PumpSave. Pump’s efficiency decreases about 1,5 % during the speed
adjustment so it does not really affect on the total power demand. Pump’s lowest possible
operating point sets to about 40 % of the maximum flow.
0
0,2
0,4
0,6
0,8
1
1,2
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Effic
ienc
y
Motor eff iciency Motor eff iciency by PumpSave
Drive eff iciency Drive eff iciency by PumpSave
Figure 56. Efficiency curves.
Motor’s efficiency for this case is better through whole operating range when it is
calculated with equations 41, 42, 43 and 44 than when it is calculated by PumpSave.
107
Frequency converter’s efficiency starts to decrease sooner and faster when it is calculated
with equation 45 than when PumpSave calculates it. Even so the differences are quite
small so they only have a small effect on the power demands.
Goulds 3628
Next inspected pump is Goulds 3628. Static head was chosen to be 10 m. Driving motor
produces 132 kW at the efficiency of 95,8 %.
0
20
40
60
80
100
120
140
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Pow
er d
eman
d [k
W]
0102030405060708090100
Hea
d [m
]Calculated pow er demand Pow er demand by PumpSave
Head Head by PumpSave
Figure 57. Power demand and head curves.
Figure 57 illustrates that for this pump also the power demand and head curves correlate
almost pefcetly. Pump’s operating range reaches about 40 % of the maximum flow.
108
0
0,2
0,4
0,6
0,8
1
1,2
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Pum
pSav
e
Motor eff iciency Motor eff iciency by PumpSave
Drive eff iciency Drive eff iciency by PumpSave
Figure 58. Efficiency curves.
Efficiency curves created for figure 58 are similar to previous cases. Motor’s efficiency is
better through the operating range when it is calculated using the values from pump’s
characteristic curve than when it is calculated by PumpSave. For frequency converter the
situation is opposite. Pump’s efficiency decreases only about a percentage during the speed
adjustment.
Goulds 5500
Next pump is again more powerful than previous one. This pump is Goulds 5500 and it is
driven by a squirrel cage motor which produces 220 kW at 96,5 % efficiency. Static head
for this case was chosen to be 17 meters.
109
0,000
50,000
100,000
150,000
200,000
250,000
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Pow
er d
eman
d [k
W]
0
20
40
60
80
100
120
Hea
d [m
]
Calculated pow er demand Pow er demand by PumpSave
Head Head by PumpSave
Figure 59. Power demand and head curves.
Power demand curves shown in figure 59 are well matching for most parts of the operating
range. A little difference develops at the small flow rates because of the decrease in
pump’s efficiency. PumpSave doesn’t calculate any changes for pump’s efficiency so the
power demand calculated by PumpSave is smaller than power demand based on pump’s
characteristic curve. Head curves are identical between calculations.
0,000
0,200
0,400
0,600
0,800
1,000
1,200
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Effic
ienc
y
Motor eff iciency Motor eff iciency by PumpSaveDrive eff iciency Drive eff iciency by PumpSavePump eff iciency
Figure 60. Efficiency curves.
Efficiency curves for pump, motor and frequency converter are illustrated in figure 60.
Motor’s efficiency based on the values from pump’s characteristic curve remains better
through the whole operating range than when it is calculated by PumpSave. For frequency
110
converter the situation is yet again the opposite. Pump’s efficiency starts to decrease
quickly when flow is decreased to about 40 % of the maximum flow.
Goulds 3635
Next studied pump is Goulds 3635. Driving motor for the pump produces 280 kW and has
the efficiency of 97 % at the nominal operating point. Static head and the starting point of
the system curve are 16 meters.
0,000
50,000
100,000
150,000
200,000
250,000
300,000
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Pow
er d
eman
d [k
W]
0
20
40
60
80
100
120
Hea
d [m
]Calculated pow er demand Pow er demand by PumpSave
Head Head by PumpSave
Figure 61. Power demand and head curves.
Power demand curves and head curves are yet again very well matching between
calculations as can be seen from figure 61. Pump does not reach the smallest possible flow
rate calculated by PumpSave.
111
0,000
0,200
0,400
0,600
0,800
1,000
1,200
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Effic
ienc
y
Motor eff iciency Motor eff iciency by PumpSaveDrive eff iciency Drive eff iciency by PumpSavePump eff iciency
Figure 62. Efficiency curves.
Efficiency curves for devices are illustrated in figure 62. No difference in results for
motor’s and frequency converter’s efficiencies compared to previous cases. In this case the
motor’s and frequency converter’s efficiencies based on the characteristic curve are very
close to each other through the whole operating range. Pump’s efficiency decreases about 6
percentages during the flow adjustment.
Goulds 3652
Next pump is Goulds 3652 which can produce relatively large volume flow. Driving motor
for this pump produces 320 kW and operates at 97 % efficiency at the nominal operating
point. Static head is chosen to be 20 meters.
112
0,000
50,000
100,000
150,000
200,000
250,000
300,000
350,000
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Pow
er d
eman
d [k
W]
0
10
20
30
40
50
60
Hea
d [m
]
Calculated pow er demand Pow er demand by PumpSave
Head Head by PumpSave
Figure 63. Power demand and efficiency curves.
Figure 63 illustrates the power demand and head curves between calculations. Again the
results between calculations are well matching. Pump’s operating range stops at about 50
% of the maximum flow.
0,000
0,200
0,400
0,600
0,800
1,000
1,200
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Effic
ienc
y
Motor eff iciency Motor eff iciency by PumpSave
Drive eff iciency Drive eff iciency by PumpSave
Figure 64. Efficiency curves.
Efficiency curves for driving motor and frequency converter are shown in figure 64. In this
case the results are similar but differences between calculations are not as big as in some
other previous cases. Pump’s efficiency decreases about 3 % during the flow adjustment.
Goulds CWX
113
Second to last pump under the lens is Goulds CWX. It is driven by a squirrel cage motor
which produces 400 kW at the efficiency of 97 % at the nominal operating point. Static
head is 30 meters.
0
50
100
150
200
250
300
350
400
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Pow
er d
eman
d [k
W]
0
20
40
60
80
100
120
Hea
d [m
]
Calculated pow er demand Pow er demand by PumpSave
Head Head by PumpSave
Figure 65. Power demand and head curves.
Results as power demand and head curves are shown in figure 65. Power demand
calculated by pump’s characteristic curve is slightly smaller than power demand calculated
by PumpSave at the middle part of the operating range. Reason for this can be found from
the changes in efficiencies because the head curves are almost identical. Efficiency curves
are illustrated in figure 66.
0
0,2
0,4
0,6
0,8
1
1,2
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Effic
ienc
y
Motor efficiency Motor efficiency by PumpSaveDrive efficiency Drive efficiency by PumpSavePump Efficiency
Figure 66. Efficiency curves.
114
Results for motor’s and frequency converter’s efficiencies are similar to previous cases.
Based on the characteristic curve motor’s efficiency remains better and frequency
converter’s efficiency remains worse through the whole operating range compared to
efficiencies calculated by PumpSave. Pump’s efficiency increases first when rotation speed
decreases but starts to decrease when flow rate reaches about 60 % of the maximum flow.
Goulds 5500
Last observed pump is the most powerful pump of the series. It can produce the volume
flow of over 2800 m³/h at the best efficiency point. Static head was chosen to be 7,62
meters and driving motor produces 900 kW and operates at 97 % efficiency at the nominal
operating point.
0100200300400500600700800900
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Pow
er d
eman
d [k
W]
0102030405060708090
Hea
d [m
]
Calculated pow er demand Pow er demand by PumpSave
Head Head by PumpSave
Figure 67. Power demand and head curves.
Like previous cases the power demand and head correlate almost perfectly between
calculations as can be seen from figure 67. Pump does not reach the smallest possible flow
calculated by PumpSave.
115
0
0,2
0,4
0,6
0,8
1
1,2
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Effic
ienc
y
Motor eff iciency Motor eff iciency by PumpSave
Drive eff iciency Drive eff iciency by PumpSave
Figure 68. Efficiency curves.
Results for efficiencies are also similar compared to previous cases. Motor’s efficiency is
better through the operating range when it is calculated with the values from pump’s
characteristic curve than when it is calculated by PumpSave. For frequency converter the
situation is the opposite. Pump’s efficiency remains good during the adjustment only
decreasing about a percentage.
10.2 Throttling
Throttling control is examined next as a flow control method for centrifugal pumps. Pumps
under examination are from manufacturer called Afton pumps Inc. First pump has the
lowest power demand and the last one is the most powerful. This study is based on the
characteristic curves of the pumps. All pumps operate at the rotational speed of 3000 rpm.
When throttling the flow the new operation point of pump is found along the pump’s
characteristic curve. Pump’s power demand can be solved when flow, head and efficiency
is known. These factors at each operation point are compared to PumpSave’s calculation.
Results are shown as power demand, head and efficiency curves.
Static head is not taken notice and no imaginary system curves were created in
characteristic curves. Noticing static head does not affect on the calculation of PumpSave
when examining throttling control. PumpSave does not calculate any changes in driving
116
motor’s efficiency during throttling control. Changes in motor efficiency are taken notice
of in calculations based on pump’s characteristic curve.
Afton ILVS 1.5x2-7A
First pump under inspection is Afton ILVS 1.5x2-7A. Diameter of the pump’s impeller is
184 mm. Driving motor produces 4 kW of power and its efficiency at nominal operating
point is 85 %.
00,5
1
1,52
2,53
3,54
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Pow
er d
eman
d [k
W]
38
40
42
44
46
48
50
52
Hea
d [m
]
Calculated power demand Power demand by PumpSaveHead Head by PumpSave
Figure 69. Power demand and head curves.
Power demand and head curves for this pump are shown in figure 69. As can be seen from
the power demand curves there is difference in the values between calculations. According
to PumpSave the power demand decreases more rapidly than in calculations based on the
characteristic curve of the pump when the flow is choked. When flow is 40 to 80 % of the
maximum flow calculated power demand based on the characteristic curve is larger than
the power demand calculated by PumpSave.
Head curves are also illustrated in figure 69 and there is quite similar difference as in
power demand. Head based on the pump’s characteristic curve is larger at the whole
operating range. This has an affect on the power demand.
117
0
0,1
0,2
0,3
0,4
0,5
0,6
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Effic
ienc
y
Pump eff iciency Pump eff iciency by PumpSave
Figure 70. Efficiency curves.
Efficiency curves based on PumpSave’s calculation and pump’s characteristic curve are
illustrated in figure 70. According to PumpSave the efficiency remains higher during the
whole throttling than pump’s characteristic curve indicates. PumpSave does not take into
account the change in motor’s efficiency in throttling control. When using equations 41,
42, 43 and 44 to resolve power loss and efficiency of the motor it drops about 2 % during
the throttling. Pump does not reach the smallest possible flow calculated by PumpSave.
Afton ILVS 1.5x2-9H
Next pump is a bit more powerful than previous one producing larger volume flow and
higher head. It’s impellers diameter is 229 mm and driving motor operates at the power of
7,5 kW and 87,5 % efficiency rate.
012345678
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Pow
er d
eman
d [k
W]
50
55
60
65
70
75
Hea
d [m
]
Calculated power demand Power demand by PumpSaveHead Head by PumpSave
118
Figure 71. Power demand and head curves.
For this case power demand curve based on PumpSave’s calculation is very linearly
descending. Power demand curve based on pump’s characteristic curve on the other hand
differs dramatically from PumpSave’s curve when flow rate is 40 to 60 % of maximum.
This is caused by rapid collapse of efficiency at that operation area. Efficiency curves by
PumpSave and pump’s characteristic curve are created in figure 72. Characteristic curve
based power demand is higher than power demand by PumpSave through almost whole
operating area. This is supported also by change in pump’s produced head during
throttling. Head based on pump’s characteristic curve is also higher than head calculated
by PumpSave at all points of operation range. However the head curves are more
congruent than power demand curves.
0
0,1
0,2
0,3
0,4
0,5
0,6
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Effic
ienc
y
Pump eff iciency Pump eff iciency by FanSave
Figure 72. Efficiency curves.
Efficiency for this pump is clearly worse than efficiency calculated by PumpSave through
whole operating range. When flow is throttled from 60 % to 50 % of maximum the
efficiency collapses rapidly and causes a rise in power demand curve seen before.
Efficiency of the motor drops about 1,5 % during the throttling.
Afton ILVS 1.5x3-11H
119
Next pump under inspection is Afton ILVS 1.5x3-11H and its impellers diameter is 279
mm. Applicable motor for the pump is a squirrel cage motor which produces 30 kW power
and has an efficiency of 92,9 % at the nominal operating point.
0
5
10
15
20
25
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Pow
er d
eman
d [k
W]
102104106108110112114116118
Hea
d [m
]
Calculated power demand Power demand by PumpSaveHead Head by PumpSave
Figure 73. Power demand and head curves.
For this pump as can be seen from figure 73 the power demand calculated by pump’s
characteristic curve is higher than PumpSave’s power demand during the whole operating
range. The main factor causing this difference is the head generated by the pump. Pump’s
characteristic curve is steeply descending when moving in area of larger flows. When the
flow is choked to about 70 % of the maximum flow the pump’s characteristic curve is
almost horizontal and the head does not increase much. PumpSave’s curve is more linear
and the biggest difference in head is formed when flow is 70 % of maximum. The head by
pump’s characteristic curve is about 2 meters higher than the head calculated by
PumpSave.
120
0
0,1
0,2
0,3
0,4
0,5
0,6
0,7
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flow
Flow
Effic
ienc
y
Pump eff iciency Pump eff iciency by PumpSave
Figure 74. Efficiency curves.
Efficiency curves between pump’s characteristic curve and PumpSave’s are illustrated in
figure 74. Again there is clear difference how the efficiency changes during the throttling
adjustment. Pump’s efficiency starts to decrease much quicker according to characteristic
curve than PumpSave’s calculation. This contributes also to the difference in power
demands. Drive motor’s efficiency decreases only about one percent during the throttling.
Smallest flow the pump reaches is about 40 % of the maximum flow.
Afton ILVS 2x3-13H
Next pump is Afton ILVS 2x3-13H and it produces quite high head and a little bit higher
volume flow than the previous pump. Impellers diameter is 343 mm and the driving motor
operates at 55 kW power and efficiency of 94 %.
121
0
10
20
30
40
50
60
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Pow
er d
eman
d [k
W]
135140
145150155
160165
170175
Hea
d [m
]
Calculated power demand Power demand by PumpSaveHead Head by PumpSave
Figure 75. Power demand and head curves.
Results in figure 75 shows that pump’s characteristic curve which is shown as head curve
in figure X is steeper than the head curve calculated by PumpSave. This causes the power
demand to be higher than the power demand calculated by PumpSave at the range the
pump is able to operate.
0
0,1
0,2
0,3
0,4
0,5
0,6
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Effic
ienc
y
Pump eff iciency Pump eff iciency by PumpSave
Figure 76. Efficiency curves.
For this pump case the efficiency curves are quite congruent at the range the pump
operates. Efficiency of the motor connected to the pump decreases about one percent
during the throttling.
122
Afton ILVS 3x4-10
Next pump is Afton ILVS 3x4-10 with the impeller diameter of 254 mm. Motor connected
to the pump is a squirrel cage motor and it produces power 55 kW at the efficiency of 94
%.
05
10152025303540
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Pow
er d
eman
d [k
W]
70
75
80
85
90
95
Hea
d [m
]
Calculated power demand Power demand by PumpSave
Head Head by PumpSave
Figure 77. Power demand and head curves.
Differences in power demand and in head between PumpSave’s calculation and
characteristic curve based calculation are similar to previous pump cases. At the most
points of operation the head and therefore the power demand are higher based on
characteristic curve than based on PumpSave’s calculation. Pump’s power demand
descends below PumpSave’s calculated value at the smallest possible flow. Reason for this
is shown in figure 78 where efficiency curves are illustrated.
123
0
0,1
0,2
0,3
0,4
0,5
0,6
0,7
0,8
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Effic
ienc
y
Pump eff iciency Pump eff iciency by FanSave
Figure 78. Efficiency curves.
Efficiency curves between PumpSave’s calculation and characteristic curve are quite
congruent for this pump case. When the flow is throttled to about 45 % of maximum the
efficiency curves cross and for smaller flows the efficiency based on the characteristic
curve is better than the efficiency calculated by PumpSave. This shows up also in power
demand curves at flow rate of 40 % in figure 77 where power demand based on
characteristic curve is lower than power demand calculated by PumpSave
Afton ILVS 4x6-11H
Next pump under inspection is Afton ILVS 4x6-11H. Diameter of the impeller is 279 mm.
Driving motor operates at 90 kW power and at the efficiency of 95 %.
124
0102030405060708090
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Pow
er d
eman
d [k
W]
80
85
90
95
100
105
110
Hea
d [m
]
Calculated pow er demand Pow er demand by PumpSaveHead Head by PumpSave
Figure 79. Power demand and head curves.
Power demand and head curves are illustrated in figure 79. Once again the power demand
and head curves based on the pump’s characteristic curve go through higher values than
curves based on PumpSave’s calculation. Efficiency curves are illustrated in figure 80.
Difference between head and efficiency curves are the reason the power demand based on
characteristic curve differs from the power demand calculated by PumpSave.
0
0,1
0,2
0,3
0,4
0,5
0,6
0,7
0,8
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Effic
ienc
y
Pump eff iciency Pump eff iciency by PumpSave
Figure 80. Efficiency curves.
Shapes of efficiency curves are similar to previous pump cases. Efficiency based on
characteristic curve starts to decrease sooner than efficiency based on PumpSave’s
calculation when flow is throttled. This contributes to the difference in power demands.
125
Driving motor’s efficiency stays high during throttling decreasing only few decimals of
percent.
Afton ILVS 3x4-13L
Next observed pump is Afton ILVS 3x4-13L with the impeller of 343 mm of diameter.
Pump’s driving motor produces 110 kW at the efficiency of 95,8 %.
0102030405060708090
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Pow
er d
eman
d [k
W]
020406080100120140160180
Hea
d [m
]
Calculated pow er demand Pow er demand by PumpSaveHead Head by PumpSave
Figure 81. Power demand and head curves.
Head curves for this pump are almost perfectly matching between pump’s characteristic
curve and calculation of PumpSave as can be seen from figure 81. However the power
demands between these two are not so well matching. Power demand based on calculations
done by characteristic curve is higher than power demand calculated by PumpSave at
almost whole operating area. Only at the highest and lowest possible flow the power
demands are matching. Difference is caused by difference in efficiencies between pump’s
characteristic curve and PumpSave’s calculations.
126
0
0,1
0,2
0,3
0,4
0,5
0,6
0,7
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Effic
ienc
y
Pump eff iciency Pump eff iciency by PumpSave
Figure 82. Efficiency curves.
Figure X illustrates the efficiency curves created by pump’s characteristic curve and
PumpSave’s calculations. Efficiency indicated by pump curve decreases more quickly than
efficiency calculated by PumpSave. Drop in efficiency is not so rapid when pump is
operating at small flow rates and characteristic curve based efficiency is better than
efficiency calculated by PumpSave at the smallest possible flow. Driving motor’s
efficiency decreases less than a percent during throttling.
Afton ILVS 4x6-13H
Last studied pump is Afton ILVS 4x6-13H and it is the most powerful pump of the series.
Pump’s impeller has the diameter of 343 mm and it is driven by a squirrel cage motor
which produces 132 kW of power at the efficiency of 95,8 %.
127
0
20
40
60
80
100
120
140
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Pow
er d
eman
d [k
W]
135
140
145
150
155
160
165
170
Hea
d [m
]
Calculated pow er demand Pow er demand by PumpSaveHead Head by PumpSave
Figure 83. Power demand and head curves.
Results between calculations are not surprising at this point. Values for power demand and
head based on pump’s characteristic curve are once again higher than PumpSave’s
calculated values. Difference in power demands is affected by difference in efficiencies
which is shown in figure 84.
0
0,1
0,2
0,3
0,4
0,5
0,6
0,7
0,8
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.flowFlow
Effic
ienc
y
Pump eff iciency Pump eff iciency by PumpSave
Figure 84. Efficiency curves.
Effciency curves based on pump’s characteristic curve and PumpSave’s calculations are
quite congruent as figure 84 shows. Still the efficiency calculated by PumpSave is better
than efficiency based on characteristic curve at the whole operating range of the pump.
Driving motor’s efficiency remains high through the throttling.
128
Conclusions
PumpSave’s calculations for throttling control are fairly competent based on these
examinations. In throttling control PumpSave starts the throttling at the best efficiency
point of operation and the efficiency does not get better at any point of throttling. So it is
important to check with the pump’s characteristic curve what the point of nominal flow for
the pump is where the throttling of the flow is planned to start.
PumpSave’s calculation presents a bit too low values for pump’s characteristic curve
through the whole operating area. Each pump’s characteristic curve is shown in figures as
a head curve. For one pump the curves between PumpSave’s calculation and the pump’s
characteristic curve are almost perfectly matching but for other cases the head values differ
more or less. This also contributes to the calculated power demand which was higher based
on calculations done by pump’s characteristic curve than calculated power demand by
PumpSave for each pump through nearly whole operation range.
Results for pump applications power demand presented in previous chapter is supported by
results acquired from efficiency examination. Comparison of efficiency for each pump
case showed that PumpSave’s calculation produces higher values than pump’s
characteristic curve indicates. Each pump’s efficiency starts to decrease much sooner
according to pump’s characteristic curve than in PumpSave’s calculation when throttling
the flow.
Based on acquired results PumpSave’s calculated values for efficiency and head should be
corrected a bit. Curves created by means of these values should be steeper. Pump’s
characteristic curve should rise more rapidly and efficiency curve should decrease more
rapidly when throttling the flow. For more exact results much more pump cases should be
studied so the calculation would be more universal for all possible pumps.
For one of the studied pumps its efficiency collapses so much at one stage of throttling that
its power demand at that point rose higher than at previous operating point where the flow
rate was larger. This is of course a special scenario and PumpSave can not predict it but
129
these kinds of situations can happen. All in all PumpSave gives fairly good prediction for
pump applications total power demand and energy consumption.
11 LABORATORY TEST RUNS WITH A PUMP
In this part of the thesis PumpSave’s calculation is compared to results gathered from
laboratory testing. Flow control methods studied were throttling control and rotational
speed control by frequency converter. Tests were done in Lappeenranta university of
technology where is a specific laboratory for testing pumps. Pump used in test was made
by G.A. Serlachius Oy model DC 80/260. It was powered by Strömberg HXUR 328A2 B3
electric motor which produces 15 kW. Speed of the motor was controlled with ACS 800
variable speed drive. A layout of the laboratory equipment is illustrated in figure 85.
Figure 85. Laboratory equipment. (Viholainen, 60)
A venturi tube was used to measure the speed and volume of the flow. Pressures at pump
inlet and outlet were measured to make possible to calculate the total head of the pump.
Measured data was lead to Fluke Hydra Data Logger. Also the torque provided by the
motor was measured with separate device. Rotational speed, input frequency and shaft
power could be red from the panel of the variable speed drive.
130
11.1 VSD
First flow control method under observation is variable speed drive. Pump was first driven
at full speed and adjustments were done by decreasing the rotational speed so that flow
produced by the pump was decreased 10 % at a time. Rotational speed was adjusted by
regulating the input frequency at the panel of the frequency converter. All the valves in the
pipeline were fully open during the speed adjustment. Results and the most important
variables are presented in table 29. Power demand and head between calculations are
illustrated as curves in figure 86. Pump’s efficiency curve is illustrated in figure 87.
Table 29. Results of test run for VSD flow control.
PumpSave
q_v
[m³/s]
p_imu
[Pa]
p_paine
[Pa]
deltap_vent.
[Pa]
Vääntö
[Nm]
n
[rpm]
f
[Hz]
H
[m]
eta_p
[%]
P_displ
[kW]
P_calc
[kW]
P
[kW]
H[m]
0,0354 82500 177000 50413 60,0 1500 50 9,63 35,5 9,60 9,43 9,41 9,63
0,0319 88125 171000 40700 47,5 1311 43,7 8,45 40,5 6,48 6,52 7,39 8,28
0,0283 92500 162000 32098 38,0 1158 38,6 7,08 42,7 4,55 4,61 5,71 7,06
0,0248 96250 156000 24975 30,6 1035 34,5 6,09 44,7 3,26 3,32 4,32 5,99
0,0212 99375 150000 18130 24,2 915 30,5 5,16 46,3 2,29 2,32 3,24 5,07
0,0177 102500 145500 12580 18,8 801 26,7 4,38 48,3 1,55 1,58 2,34 4,28
0,0142 105000 141000 8140 14,5 699 23,3 3,67 48,2 1,04 1,06 1,65 3,64
0,0106 106875 138000 4625 11,2 609 20,3 3,17 46,4 0,69 0,71 1,12 3,14
0,0071 108125 136500 2035 8,5 537 17,9 2,89 42,1 0,47 0,48 0,71 2,79
131
0
2
4
6
8
10
12
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom. FlowFlow
Pow
er d
eman
d [k
W]
0
2
4
6
8
10
12
Nos
toko
rkeu
s [m
]
Akseliteho tm:n näytöltä Laskettu akselitehoAkseliteho by PumpSave PumpSave akseliteho hyötysuhteen ka:n mukaanNostokorkeus Nostokorkeus by PumpSave
Figure 86. Shaft power curves and head curves.
From curves illustrating shaft power in figure X can be seen that the shaft power red from
the display of the VSD correlates with the calculated power demand. Calculated power
demand is calculated on the basis of the measured values. Apposed to that shaft power
calculated by PumpSave differs from measured and calculated shaft power through almost
whole operating range. This is caused by changes in pump’s efficiency which PumpSave
does not take into account. PumpSave is fed with the efficiency of the pump at the nominal
operating point and it remains constant at each point of operation. There is also a power
demand curve in figure X which illustrates power demand calculated by PumpSave and it
is calculated with an average efficiency of the operation range. This curve matches the
curves created with results from laboratory tests much better than curve created with
efficiency at the nominal operating point. With this curve the biggest difference to curves
from laboratory tests is formed when pump operates at maximum flow. This is because
pump’s efficiency is worse than average in actual situation. Figure X shows also the head
produced by the pump and head calculated by PumpSave. Values are well matching
through the operation range though the head values calculated by means of the pressure
difference of the pump are a bit higher at most part of the operation range. However the
difference is so small that it hardly affects the total power demand of the application.
132
0
0,1
0,2
0,3
0,4
0,5
0,6
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.Flow
Effic
ienc
y
Figure 87. Pump efficiency curve.
Figure 87 illustrates pump’s efficiency curve during the speed adjustment. Efficiency
reaches its maximum value when the flow rate is 50 % of the maximum. At the nominal
flow pump’s efficiency is at its lowest being only 35,5 %. This takes PumpSave’s
calculation on the wrong track. Efficiency is not constant when pump’s rotation speed is
changed and this causes difference on results for power demand..
Next is examined situation where point of nominal flow is also the best efficiency point.
Nominal point is chosen to be 50 % of the maximum flow based on the previous study.
Results of power demand and pump efficiency are illustrated in figure 88.
00,20,40,60,8
11,21,41,61,8
40 % 50 % 60 % 70 % 80 % 90 % Nom. flowFlow
Pow
er d
eman
d [k
W]
0,38
0,4
0,42
0,44
0,46
0,48
0,5
Effic
ienc
y
Akseliteho näytöltä Laskettu akseliteho Akseliteho by PumpSavePump efficiency Poly. (Pump efficiency)
133
Figure 88. Shaft power and efficiency curves.
Results show that all the power demands are now well matching. This way of inspecting
pump’s operation matches best PumpSave’s calculation. This also was proved in section of
the work where PumpSave’s calculation was compared to calculation based on pump’s
characteristic curve. Figure X also shows points and trend line of pump’s efficiency.
11.2 Throttling
Other flow control method which could be studied at the laboratory was throttling control.
Pump operated at full speed through the test and at the nominal point all the valves on the
pipeline were fully open. Flow was adjusted by throttling the last valve before the upper
tank so that flow was decreased 10 % at a time. Results of calculations are shown in table
30 and as shaft power and head curves in figure 89. Efficiency curves of the pump are
illustrated in figure 90.
Table 30. Results for throttling control.
PumpSave
q_v
[m³/s]
p_imu
[Pa]
p_paine
[Pa]
deltap_vent.
[Pa]
Vääntö
[Nm]
H
[m]
eta_p
[%]
P_displ
[kW]
P_calc
[kW]
P
[kW]
H[m] eta_p
[%]
0,0354 82500 177000 50413 60,0 9,63 35,5 9,51 9,43 9,41 9,63 35,5
0,0319 86875 238500 40700 58,0 15,46 53,1 9,12 9,11 10,89 12,38 35,5
0,0283 91875 270000 32098 53,9 18,16 59,5 8,46 8,47 11,62 14,84 35,5
0,0248 95625 288000 24975 50,5 19,61 60,1 7,95 7,93 11,94 17,01 34,4
0,0212 99375 304500 18130 46,2 20,91 59,9 7,28 7,26 12,02 18,89 32,7
0,0177 101875 319500 12580 42,8 22,18 57,3 6,72 6,72 11,91 20,48 29,8
0,0142 104375 330000 8140 31,9 23,00 52,2 6,15 6,14 11,67 21,78 25,9
0,0106 106250 336000 4625 35,4 23,42 43,8 5,57 5,56 11,32 22,80 20,9
0,0071 107500 340500 2035 31,6 23,75 33,3 4,95 4,96 10,88 23,52 14,9
134
0
2
4
6
8
10
12
14
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.FlowFlow
Pow
er d
eman
d
0
5
10
15
20
25
Hea
d [m
]
Akseliteho näytöltä Laskettu akselitehoAkseliteho by PumpSave NostokorkeusNostokorkeus by PumpSave
Figure 89. Shaft power and head curves.
Shaft power curves in figure 89 show that values calculated by PumpSave differ drastically
from measured and calculated power demand. This is caused by pump’s efficiency which
is rather low when operating at full speed. Efficiency starts to increase rapidly when flow
is throttled and reaches its maximum point at 70 % of the maximum flow. Efficiency input
as an initial value for PumpSave does not get any better at any point of throttling but starts
to decrease when flow is about 80 % of the maximum. Efficiency curves are illustrated in
figure 90 and are the main reason to the difference between shaft powers.
Figure 89 shows also that head curves differ from each other quite a bit. Difference
between curves is larger when pump is operating at larger volume flows and decreases
when operating point enters the area of smaller flow rates. In reality pump produces higher
head than PumpSave calculates at almost whole operating range. In this case it reduces the
difference of shaft powers.
135
0
0,1
0,2
0,3
0,4
0,5
0,6
0,7
20 % 30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.Flow
Flow
Effic
ienc
y
Pump efficiency Pump eff iciency by PumpSave
Figure 90. Efficiency curves.
Next is studied a situation which matches PumpSave’s calculation much better. Nominal
operating point of the pump is equal to best efficiency point. In performed throttling
control test best efficiency point was found at 70 % of the maximum flow. So this point is
chosen to be the nominal operating point and at this point flow is throttled 10 % at a time.
Results are shown in figures 91 and X where shaft power, head and efficiency curves are
illustrated.
0
1
2
3
4
5
6
7
8
9
30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom. FlowFlow
Shaf
t pow
er [k
W]
0
5
10
15
20
25
Hea
d [m
]
Calculated shaft pow er Shaft pow er by PumpSave Head Head by PumpSave
Figure 91. Shaft power and head curves [m]
136
In this case values for shaft power are quite well matching between calculations. Shaft
power red from the panel of the frequency converter matched the calculated power demand
perfectly so it was left a side from the figure 91. Head curves are also almost identical in
this case as can be seen from figure 91.
0
0,1
0,2
0,3
0,4
0,5
0,6
0,7
30 % 40 % 50 % 60 % 70 % 80 % 90 % Nom.FlowFlow
Effic
ienc
y
Pump eff iciency Pump eff iciency by PumpSave
Figure 92. Efficiency curves.
Efficiency curves for this case are illustrated in figure 92. Now the curves are considerably
closer to each other through the whole operating range than previously. Naturally this
contributes to the good results gained for shaft power.
Conclusions
Results of laboratory testing are quite similar to results gained from comparing
PumpSave’s calculation to characteristic curve based calculation. PumpSave’s calculation
works correctly when the pump’s nominal operation point equals the best efficiency point.
In variable speed control PumpSave does not take into account any efficiency changes
during rotation speed adjustment. Based on first introduced test run, pump’s efficiency
changes through the operation range quite a bit. When setting the nominal operating point
equal as the best efficiency point the results for efficiencies match PumpSave’s calculation
better. Program could be installed with some correcting factors for pump’s efficiency at
different flow rates but this would require more testing and studying of different types and
sizes of pumps in laboratory and with characteristic curves. These kind of correcting
137
factors PumpSave and FanSave uses are always some kind of a compromise among
different sizes of applications.
For throttling control the results from laboratory test runs are well matching when the
nominal operation point where the throttling starts equals the best efficiency point.
PumpSave’s calculation does not match the results at all if the throttling is begun at the
descending area of the efficiency curve.
When using PumpSave to observe the power demand and energy consumption of different
applications the user should be aware how the pump’s efficiency behaves when moving on
to some other operating point than the nominal point. It is also good to remember that the
characteristic curves for each pump are based on an ideal situation and in real life those
values are hard or impossible to reach.
12 SUMMARY
For powerful fans it would be good to include FanSave with possibility to calculate change
in density and in temperature of gas during blowing. Change in gas’s temperature affect
the power demand of the fan when change is big enough. Adding this feature into the
program would require tests with different sizes of fans. At different flow rates density
changes different amount so if the changes to the program would be made with constant
factors which FanSave uses a lot different types of fans should be tested to find the average
factors valid for various fans. One way could also be to add the option to input as initial
values the temperatures for incoming and outgoing flow. FanSave could then solve the
average density based on these values.
FanSave’s manual does not contain any information about the SFP. Also the calculation of
SFP in FanSave is incorrect. This could easily be corrected to match calculation presented
in equation 17. Also the manual of the program should be added information about the SFP
number.
138
Previous studies for different sizes of pumps controlled by a variable speed drive prove
that PumpSave’s calculations are almost perfectly accurate. Power demands and heads for
each cases are almost identical for each pump case. For some cases pump’s efficiency
changes several percentages during the flow adjustment and has affect on the power
demand. Especially when operating point enters the lowest flow rates PumpSave assumes
that in variable speed control pump’s efficiency remains as a constant.
Motor’s efficiency differs from PumpSave’s calculation when it is calculated with
equations presented in chapter X. through the operating range for each pump motor’s
efficiency calculated by PumpSave is worse than efficiency calculated using pump’s
characteristic curve. If PumpSave’s calculation would be changed to match calculation
presented in equations 41, 42, 43 and 44 more information about the driving motor’s
operation would be required. At least torque and rotational speed at each operation point
should be calculated.
12.1 Observation of the profitability of investment
FanSave and PumpSave examines the investment of the variable speed drive with present
value which has been subtracted the investment cost. If the present value is positive the
investment is profitable. Present value is calculated with equation
( ) ( )
Hi1
JAi1
SNA n
n
1tt
t −+
++
= ∑=
(78)
where
i interest rate
n utilization time [a]
t time
annual net savings or profit [€] tS
JA residue value [€]
H investment cost [€]
139
Payback period is calculated directly as the quotient of the investment cost and the annual
saving. This method is called interest free payback period. Particularly for the larger
investments it would be appropriate to calculate the payback period with interest. This
could be calculated by using equation 78 or 79. This feature could also be added to both of
the programs. All the factors are already included.
( )∑=
=−+
pn
1tt
t 0Hi1
S (79)
where
pn payback time with interest [a]
( )
( )i
iSH
i t
+
−⎟⎟⎠
⎞⎜⎜⎝
⎛−−
=1ln
ln1lnn (80)
140
REFERENCES
Electric Ideas Clearinghouse. North Seattle Community College Two-Speed
Motors Case Study. 1993. [Viitattu 20.11.2007] Saatavilla www-muodossa:
http://www.p2pays.org/ref/36/35282.pdf
Fan Types. Environmental Protection Agency. [Viitattu 9.10.2007] Saatavilla
www-muodossa: http://www.epa.gov/eogapti1/module5/fans/types/types.htm
Fläktwoods Oy. SFP-opas. [Viitattu 21.11.2007] Saatavilla www-muodossa:
http://www.flaktwoods.fi/476d6be3-be6e-42e9-bd82-6152ff71a7aa
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muodossa: http://www.gouldspumps.com/pss.html
Hirvonen Sampsa. Taajuusmuuttajan vaihtaminen pakkaajarobottiin. Insinöörityö.
Helsingin ammattikorkeakoulu Stadia. Sähkötekniikka. 2007.
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muodossa: http://www.damper-designs.co.uk/
Invertteri. Lappeenrannan teknillinen yliopisto. Sähkötekniikan osasto. [Viitattu
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29.2.2008] Saatavilla www-muodossa:
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141
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Luukkanen Petteri. Pumpunvalitsimet integroidussa toimintaympäristössä.
Diplomityö. LTY. Kemiantekniikan osasto. 2001.
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142
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Torque Limiters and Slip Clutches Information on GlobalSpec. GlobalSpec.
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s/Slip_Clutches?SrchItem=1&frmqry=slip%20coupling
143
144
Appendix A: Mainpage of FanSave
145
Appendix B: Mainpage of PumpSave
146
Appendix C: Discharge coefficient
147
Appendix D: Expansibility factor
148
Appendix E: Centrifugal fans
Appendix E1: HF R 560-13 D/R
149
Appendix E2: HF R 200-48 R
150
Appendix E3: HF R 710-13 D/R
151
Appendix E4: HF R 1000-13 D/R
152
Appendix E5: GTLF-3-025
153
Appendix E6: GTLF-3-040
154
Appendix E7: GTLF-3-056
155
Appendix E8: GTLF-3-071
156
Appendix E9: GTLB-3-025
157
Appendix E10: GTHB-3-040
158
Appendix E11: GTHB-3-056
159
Appendix E12: GTHB-3-071
160
Appendix E13: GTHB-3-100
161
Appendix E14: GTLB-3-140
162
Appendix F: Axial fans
Appendix F1: HF A 630
163
Appendix F2: HF A 800
164
Appendix F3: HF A 1000
165
Appendix F4: Applied Energy XLCA1000
166
Appendix F5: Ebm-Papst 3G400
167
Appendix F6: Ebm-Papst 3G990
168
Appendix F7: Ebm-Papst 3G800
169
Appendix F8: Ebm-Papst 3G650
170
Appendix F9: Ebm-Papst 3G710
171
Appendix G: Pumps
Appendix G1: Goulds CDS 5309-1 model 3298
172
Appendix G2: Goulds CDS IC307-1 model IC/ICB/ICP/ICV
173
Appendix G3: Goulds 3638
174
Appendix G4: Goulds 3628
175
Appendix G5: Goulds 5500
176
Appendix G6: Goulds 3635
177
Appendix G7: Goulds 3652
178
Appendix G8: Goulds CWX
179
Appendix G9: Goulds 5500
180
Appendix G10: Afton ILVS 1.5x2-7A
181
Appendix G11: Afton ILVS 1.5x2-9H
182
Appendix G12: Afton ILVS 1.5x3-11H
183
Appendix G13: Afton ILVS 2x3-13H
184
Appendix G14: Afton ILVS 3x4-10
185
Appendix G15: Afton ILVS 4x6-11H
186
Appendix G16: Afton ILVS 3x4-13L
187
Appendix G17: Afton ILVS 4x6-13H
188
Appendix G18: Serlachius
189
190
