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Fans in rail vehicle cooling systems.
Energy savings as a result of total systems
optimisation.
Karl-Oskar Joas, Christoph Kirschinger, Heidenheim, Germany
1

2
Contents
1. Introduction 3
2. History and technical development 3
3. Operation of cooling system fans 6
3.1 Drive and control 7
3.1.1 Mechanical / hydrodynamic drive 7
3.1.2 Hydrostatic drive 8
3.1.3 Electrical drive 10
3.2 Design and rating of the radiator 10
3.3 Cooling air demand, flow resistances, and safety margins 12
3.4 Arrangement of the fan 13
3.5 Fan operating points 14
3.6 Noise emission and sound insulation 15
4. Modern cooling systems and their design 16
5. Economical considerations 17
6. Summary 18
References 18

Summary
Fans are a key component in rail vehicle
cooling system. This fact has been recognised
by Voith Turbo GmbH & Co. KG,
Heidenheim, who consequently developed
a special range of cooling fans.
All fans from this range have modular
design and can be adapted to any application.
They ensure that cooling systems
run with maximum efficiency, offering
the customer excellent benefits.
The concept of a cooling system is determined
by the synergies between the
fan, the heat exchangers, the prevailing
operating conditions and the
driveline designs of individual rail vehicle
types, as well as by the way in which
they are installed. Intelligent arrangements
of the components and optimum
aerodynamic designs result in high
efficiencies of the fans and fan drives.
Initial investment costs for aerodynamic
layout schemes and higher purchasing
costs for high-quality components pay
off within just a few years.
1. Introduction
The increasing use of diesel engines
by railway operators since the 50s created
a high demand for powerful rail vehicle
cooling systems and thus also for
high-quality fans. A wide range of fans
evolved and was systematically improved
and adapted to the requirements
and installation characteristics of
the components to be cooled.
Fans are driven by auxiliary energy and
are used for dissipating waste heat from
power-transmitting machines. The
smaller the required amount of auxiliary
energy for the same quantity of dissipated
heat, the more economical the
overall system.
For best results it is important to examine
the entire aerodynamic system with
all the parts exposed to the air flow and
not just the fan alone. Intelligent speed
control is a prerequisite of a favourable
energy balance, too. And finally, aspects
such as cost, environmental conditions,
noise emission, and reliability
must be taken into account.
From this point of view, the main task
for project engineers of cooler manufacturers
is to optimally match the various
components necessary for overall
performance of the aerodynamic part
of the cooling system to each other.
A high knowledge of the technical conditions
in the vehicle and several iteration
steps when planning the aerodynamic
system are required for this,
since it is obviously impossible to create
optimum operating conditions for all
the components at the same time due
to the prevailing boundary conditions
and the available space. A compromise
made for the design of one of the components
often has a positive influence
on overall efficiency.
3

2. History and technical development
The development of axial-flow fans by
Voith Turbo, Heidenheim, began in 1930,
i. e. at the same time as the first turbo
transmissions were developed, with
research work by Marcinowski [1 to 3].
As early as 1932, Voith supplied a large
impeller for an axial-flow fan in a wind
tunnel. In the following years, development
concentrated on different impeller
blades and vanes. The different designs
formed a modular system that permitted
fans to be adapted to different operating
conditions in a highly flexible
way.
Later, the development of axial-flow
fans was supplemented by that of radial-flow
and cross-flow fans leading to
wind turbines and a multi-stage axial
compressor for research purposes.
This opened up a wide range of possible
applications for our fans:
- stationary applications in paper
machines,
- large-scale building ventilation,
-fans for underground mining,
- tunnel ventilation using fixed or
variable-pitch blades,
-fans in power stations, especially for
flue gas desulphurization plants and
wind tunnels,
-fans in vehicles, especially as a part
of Voith cooling systems for rail
vehicles, and in snow guns.
4
The versatile and comprehensive
modular system developed by
Marcinowski for axial-flow fans can be
used for both directions of rotation in
the following configurations:
- impeller with cap suitable for both
flow directions,
- same impeller as above, yet with
guide vanes for certain blade
designs,
- impeller with guide vanes (mainly
on downstream side) and downstream
centre fairing for certain
blade designs,
- impeller with guide vanes (mainly
on downstream side) and external
diffuser, and
- special versions derived from the
above.
The types of blading and blade designs
for different pressure stages are divided
into basic groups with defined hub/tip
ratios ν (ν = hub dia. / impeller blade tip
dia., table 1):
Type
Hub/tip ratio
d
ν =
D
d
D
∆p t = max. total pressure
difference
in Pa
Distributor
without with
N 0.315 ≈ 1000 –
M 0.5 ≈ 1500 ≈ 2000
H 0.63 ≈ 2000 ≈ 2500
Table 1: Hub/tip ratio and total pressure difference for certain types in
the range of standardized hub/tip ratios:
0,28 / 0,315 / 0,355 / 0,4 / 0,45 / 0,5 / 0,56 / 0,63 / 0,71
• The size ranges of hub and tip
diameter are based on the R 20
series of preferred numbers as per
DIN 323.
• The number of blades and their
angle of attack can be varied as
required (even for cast impellers
thanks to the modular system). The
same applies to the guide vanes.
•For the sake of simplicity the
principle of stating specific power
via flow coefficient ϕ and pressure
coefficient ψ has been transformed
into a work sheet (fig. 1) with fulllogarithmic
coordinate system
showing a family of straight lines for
- outer impeller diameter and
- speed (together with the resulting
tip speed).
• The work sheet (fig. 1) shows the
relation between the specific energy
Y (flow resistance ∆ρ referred to
an air density of ρ = 1 kg/m3 ) and
•
the volume flow VL for one type of
impeller and one hub/tip ratio.
Characteristic ranges found during
tests are used for the various hub/
tip ratios. Size, design, and further
technical features of a fan with a

specified working point (L in fig. 1)
can be defined with the aid of the
work sheet and further dimensionless
characteristic diagrams (not
shown here) for various fans differing
from each other in the abovementioned
features by superposing
the two graphs.
The modular system developed by
Marcinowski was thus supplemented
by a tool for selecting the best fan for
a given application.
The influence of the number of blades
and the impeller clearance gap on fan
characteristics and noise emission
[1, 2] were measured as well. The clearance
gap depends exclusively on the
accuracy of manufacture of impeller
and shroud.
For rail vehicle applications, characteristic
diagrams were prepared that take
into account the arrangement of the
fans and the flow behaviour under the
different geometric conditions that prevail
in such vehicles. These empirical
findings were taken as a basis for the
proprietary software programs used for
calculating flow loss in cooling systems.
In view of increasingly stringent demands
made on cooling systems and
new developments in the field of radiator
elements, the fan system was further
perfected with the aim of achieving
a highly uniform flow velocity in the
annular surface of the fan while increasing
efficiency for common working
points. The result were two new blade
designs, one for an impeller without
guide vanes and one for an impeller
with outlet vanes.
These new blade designs permit deviations
from the standard outer diameters,
allowing higher flexibility of adaptation
to different installation conditions.
Other benefits include a further
optimization of the fan design and adaptation
to the fan drive used.
Y in Nm/kg
Speed n in 1/min
D in mm
2000
1800
1600
1400
1200
B
D = 710
3500
D = 800
3000
2850
2400
1000
VL in m3 5 6 7
/s
8 9 10
L
D = 900
2600
2000
1800
D = 1000
2200
Fig. 1: Detail of work sheet for determining fan diameter, fan speed,
and fan type.
B = Example of a reference point for superposition with the
characteristic range of a fan. The non-dimensional
characteristic diagram of a fan (not shown) is applied at
the reference points. The aim is to select a fan type
whose performance curve contains a specific
performance point L.
L = Example of a performance point for which a fan characteristic
is sought.
Tests have been made at the Voith
Aerodynamics Test Centre to examine
the air flow in cooling systems of rail
vehicles in tunnels. The effects of different
vehicle front end designs on air
flow around the vehicles were also
tested at different speeds. The arrangement
of air inlets and outlets was
optimized in tests, and pressures were
measured so that they could be taken
into account for fan design. For special
applications these findings help deciding
on the power reserves to be provided
[3].
5

3. Operation of cooling system fans
Fan design depends on the following
factors:
- the type of vehicle and thus the type
and arrangement of the cooling
system in the vehicle,
- the nature of the components to be
cooled and the type of coolant,
- the type of drive and control of the
fan and cooler group,
- the operating mode and design of
heat exchangers, and
- the total flow loss of the system.
It has become apparent over time that
certain arrangements, cooling system
types, and fan drives should be preferred
depending on vehicle type.
6
Fig. 2: Axial flow fan
For locomotives, the cooler group can
usually be arranged next to the component
to be cooled. For railcars and
trainsets with underfloor powerpacks a
greater distance may be necessary or
the cooler group may have to be arranged
under the roof of the vehicle.
Sometimes, the axial-flow fans (fig. 2)
mostly used in cooling systems are replaced
by radial-flow fans today if the
installation conditions and ratings are
favourable and noise is to be minimized.
In the majority of cases, however, axialflow
fans are adequate for the total
pressure loss mostly found today.
Moreover, axial-flow fans can be easily
adapted to the required cooling air
flows. With radial-flow fans, in contrast,
the possible pressure head cannot be
utilized. Cooling air flow is limited due
to the limited intake velocity of air in the
∆ptot in Pa
600
500
400
300
200
100
Fan characteristic
η = 0,58
η = 0,75
intake funnel so that efficiency levels
remain unsatisfactory, as can be seen
from the location of the working point
on the performance curves (fig. 3). The
working points where efficiency is
higher are further to the left or right on
the performance curve of the fan and
cannot be reached in practice.
Cross-flow fans are rarely used in rail
vehicles. Their advantages regarding
the arrangement of radiator and fan are
offset by an extremely low pressure
head and low efficiency, among other
things.
η = 0,74
0
0 0,3 0,6 0,9 1,2 1,5 1,8
V L in m 3 /s
Fig 3: Performance curve of a radial-flow fan without spiral
case in an underfloor-mounted cooling unit.
B = Operating point
B
η = 0,53
System characteristic

3.1 Drive and control
Today, the drive and control of the fan
cannot be separated any longer from
coolant temperature control. Both are
performed using advanced electronic
systems. The requirement is to keep the
operating temperature of the components
to be cooled as constant as possible
and to maintain close temperature
limits for certain operating conditions,
irrespective of variables such as drive
speed, loading condition, or cooling air
temperature.
For all liquid-cooled components, temperature
control of the system is realized
via speed control of the fan and
carefully matched to the thermal behaviour
of the system, taking into account
the overall concept of the cooling circuit.
This presupposes the availability
of information and knowledge about the
different characteristic operating conditions
of the components to be cooled.
With modern diesel engines, the thermal
characteristics of the start-up
phase and the effects of prolonged
idling are of particular importance. Additionally,
the partially conflicting requirements
of cylinder cooling and
charge-air cooling in the case of single-circuit
systems and the interaction
of the circuits in the case of two-circuit
systems must be considered. For engines
with charge-air-air cooling, in
particular, care must be taken to prevent
critical operating conditions such
as excessive cooling or overheating.
Further development of diesel engines
to improve efficiency and more stringent
exhaust gas regulations in conjunction
with the use of exhaust gas
turbochargers have resulted in an increase
in operating temperatures. The
higher coolant temperatures resulting
from this require higher safety in respect
of coolant circulation. Moreover,
the reduced temperature margin up to
the failure limit caused by the higher
operating temperatures makes it necessary
to define the control temperature
range even more accurately than
it was the case with older types of
diesel engine.
With diesel-hydraulic vehicles, waste
heat from the hydrodynamic transmission
and from the hydrodynamic brake,
if any, must be dissipated in addition.
The situation with regard to transformer
cooling is less complex and easier to
understand. Temperatures in the cooling
circuit are usually lower than in the
case of engine cooling, and they
change less quickly due to the high
transformer mass. The control is
matched to the operating conditions of
the transformer, preventing both excessive
temperatures and excess cooling
when ambient temperature is low. Coolant
circulation problems can occur
when the system is started at a very
low temperature and coolant flow
through the radiator element is impaired
or even inhibited by a highly viscous
coolant. This situation cannot be handled
by the temperature control alone
but requires selection of a suitable coolant.
Converter cooling requires a comparatively
high cooling air flow rate due to
the small difference between operating/
coolant temperature and ambient temperature.
In countries with a hot climate
the coolant temperature is only marginally
higher than the ambient temperature.
The high efficiency of present day
converters is due to the use of temperature-sensitive
components with a low
mass. Adherence to close temperature
limits and avoidance of excess tem-
peratures are essential for a long service
life and high efficiency of these converter
elements. This presupposes reliable
temperature control and quick
response of the fan drive control.
When the converter and transformer
are both cooled by the same fan, unfavourable
or extreme conditions may
require concessions to be made for one
of the two cooling circuits. This also applies
when further electrical components
in the cooler group are to be
cooled directly by the cooling air flow.
There are three different approaches
towards meeting the above demands
regarding cooling system control and
fan speed control for the fan drive.
3.1.1 Mechanical /
hydrodynamic drive
The simplest way of driving a fan - with
the cooler group arranged directly next
to the power pack - would be a purely
mechanical, rigid drive without control
options. Fan speed would then be coupled
to the speed of the prime mover,
usually a diesel engine. This is an inadequate
solution because of the fluctuations
in operating temperature that
result in changes in the required cooling
capacity due to different load stages
and ambient temperatures. Even the
addition of system components like
coolant temperature controllers or mixers
would not make this type of drive
efficient.
The shortcomings of cooling systems
with rigidly coupled fan and prime
mover led to the development of mechanical/hydrodynamic
and hydrostatic
drive systems for diesel-powered vehicles
that offer different control options
and high flexibility with regard to fan
speed.
7

The core component of the mechanical/hydrodynamic
drive system is the
Voith variable-speed fan, which is available
in various versions. Its distinctive
feature is that it forms an integral unit
with a filling-controlled Föttinger coupling.
A pneumatic system is used for
filling control - and thus for control of
the fan speed - that uses the coolant
temperature of the components to be
cooled as output variable [4, 5].
The advantages of the Voith variablespeed
fan can be summarized as follows:
- The system permits a constant
coolant and component temperature
to be maintained.
-Fan speed can be varied within
certain limits and is only dependent
on diesel engine speed to a certain
extent.
- The efficiency maximum of about
90 % for the entire driveline is
reached at maximum fan speed,
i. e. when the coupling is full.
-Power consumption is matched to
the operating conditions.
- The system is self-sufficient.
- The system is considered robust,
long-lived, low-maintenance, and
vibration-resistant.
The disadvantages are:
- Lack of flexibility with regard to arrangement
in the vehicle due to the
mechanical drive.
- High cost and weight.
The hydrodynamic fan drive was therefore
replaced by the hydrostatic one.
8
3.1.2 Hydrostatic drive
The hydrostatic drive system for fans
and other auxiliary machines is considered
state-of-the-art for today’s dieselpowered
vehicles with all its hydraulic
components including electronic control.
The advantages of this system are
a high freedom of choice regarding the
arrangement of the components in the
vehicle and the favourable price of the
entire cooler group.
In principle it is possible to design a
hydrostatic drive in such a way that it is
only a hydraulic shaft without any possibility
of control. The simplest type of
control would be a pure on-off control.
Both these designs are simple and trouble-free
in respect of their operation, but
they do not come up to today’s demands
any longer and have therefore
become insignificant.
There are highly developed components
and regulating valves available
Temperature
sensors
Standby
valve
Diesel
engine
Variabledisplacement
pump
Control electronics
Control valve
Constantdisplacement
motor
Oil filter
Oil tank
Fig. 4: Voith flow system driving a cooling fan
today for variable-speed hydrostatic
drives. Several patented solutions have
emerged over the last few years both
for overall function and for control.
The use of a variable-displacement
pump that only delivers the oil flow actually
required for cooling has resulted
in a higher overall efficiency. Moreover,
the Voith flow system with variable-displacement
pump includes a stand-by
valve that guarantees availability of the
working pressure required for operation
of the pump at all drive speeds also in
the case of zero delivery (figs. 4 and
5). When the operating pressure signalled
by the control system exceeds
the stand-by pressure, the stand-by
valve opens completely and the pump
is controlled by the working pressure
of the hydrostatic system [6, 7].
The solution described here permits
power savings of about 10% on an average
compared to by-pass control. For a
diesel-hydraulic locomotive of 1,500 kW
Oil cooler
Cooling
fan

Capacity of hydrostatic pump
1,0
0,9
0,8
0,7
0,6
0,5
0,4
0,3
0,2
0,1
0,0
0%
10%
20%
By-pass principle
Voith flow system
Saving
30%
40%
50%
60%
Fan speed
Fig. 5: Hydrostatic fan drive during traction, comparison of by-pass
principle and Voith flow system.
70%
80%
90%
average savings in annual cycle approx. 10%
engine output this may correspond to
7,200 l of fuel per year (table 2).
Some of the advantages of hydrodynamic
drives mentioned above also
apply to the hydrostatic drive:
- constant operating temperature of
cooling circuits,
- adapted power requirement, and
-a self-sufficient system.
In addition to this, the fan speed can
now be freely selected within a wide
range.
100%
Water
Electronic
control
Diesel
engine
Oil
Particularly advantageous, however, are
the high power density and the possibility
to expand the system to drive components
such as compressors or generators
at a constant speed. For the latter
this is done using purely hydraulic control
with an accuracy of ± 2.5 % (fig. 6).
Assumption:
Generator
Compressor
Cooler
group
Fig 6: Hydrostatic drive of several components by a multiple pump unit.
• max. pump drive power 100 kW
• mean power savings 10 kW
• 3,000 operating hours per year
Energy savings: 10 kW * 3 000 h/Jahr = 30 000 kWh/year
Fuel savings: 30 000 kWh * 200 g/kWh = 6 000 kg/year
Table 2: Energy- and fuel savings
6 000 kg / 0.83 kg/l ≈ 7 200 l/year
9

The hydrostatic drive differs from the
hydrodynamic one in the following
ways:
• limitation of operating conditions, e.
g. by the media used, their starting
conditions, or maximum and minimum
working temperatures,
• higher maintenance requirements,
• acceptable service life when certain
design limits are observed,
•low failure rate despite extremely high
demands made on control accuracy,
• requires modern control in order to be
able to also achieve high efficiency
values also under partial load.
3.1.3 Electrical drive
Electric fan drives are mainly found in
electric vehicles.
The fan is mounted on the shaft of the
electric motor and depends on the motor
speed. The hub diameter and
impeller design and its installation conditions
often depend on the dimensions
of the motor.
With the fan powers involved, threephase
motors are used exclusively. A
possible type of control is therefore on/
off control, perhaps superposed by pole
changing. On the other hand, more and
more vehicles have frequency-controlled
auxiliaries converters today. However,
these are usually controlled by the
demands of several consumers, which
leaves only a limited range of power
adjustment for the fans. On/off control
with pole changing, in contrast, can
depend exclusively on the component
to be cooled.
Also typical of three-phase motor drives
are the limited capacity of frequencycontrolled
auxiliaries converters and
10
the dependence of fan power consumption
on cooling air temperature and
density.
The fact that power consumption depends
on air density applies to all fans
and speeds. A high drive power is
needed in particular when cooling air
temperature at the fan is substantially
lower than the (high) temperatures, for
which the cooling system was rated.
From +40°C on 1,000 m above sea level
to -10°C and 0 m above sea level, for
instance, the air density increases from
1 kg/m 3 to 1.35 kg/m 3 , i. e. by 35%. Both
the electric motors and the supply systems
must therefore be adequately
rated and protected.
ϑ 10K
∆ 2…3K
∆ L30K
T
ϑ A
ϑ 4
ϑ 1b
1
s KM
ϑ 3
ϑ 1a
ϑ E = ϑ 1,2
1a
VL, tLE Fig. 7: Simplified schematic representation of temperature behaviour with a cross feed radiator
without temperature overlap.
1…4 = Reference points in cooling circuit
VKM =Volume of cooling medium flow
VL =Volume of cooling air flow
ϑn =Cooling medium temperatures
tLn =Cooling air temperatures
ϑ10K =average cool-down cooling medium 10 K
3.2 Design and rating of the
radiator
The purpose of a cooling system is to
dissipate waste heat. The heat is transported
from the component to be
cooled to the radiator by the coolant.
The radiator is required for heat exchange
and is designed for the required
heat flows in conjunction with the mass
flows of the coolant and the cooling air
as well as their temperatures. It therefore
influences fan design to a considerable
extent.
The maximum and minimum coolant
flow and its temperature depend on the
component to be cooled. The cooling
air flow should be as small as possible.
It should be selected in such a way that
t L2
t L3
t Lm
t L4
s L
2
3
4
V KM, ϑ E
ϑ A
1
1b
∆L30K = average warm-up cooling medium 30 K
sKM = Travelling distance of cooling medium in
cooling circuit
sL = Travelling distance of cooling air through
cooling circuit
T = Temperature

temperature overlaps between coolant
and air outlet temperature are prevented,
i. e. that the coolant temperature
at radiator outlet is about 2 to 3 K
higher than the mean temperature of
the hot cooling air (fig. 7). The purpose
of this is to achieve an economical radiator
volume. If, under extreme conditions,
temperature overlap cannot be
avoided (fig. 8), care must be taken to
ensure that suitable fundamentals and
procedures of radiator calculation are
applied and that the inner and outer
flows through the radiator can be specially
matched to this situation as is the
case in a counterflow crossfeed radiator
(fig. 9).
Despite the fact that cooling air outlet
temperatures are locally different, the
fan design is based on the mean temperature
t LM.
Due to physical interrelations temperature
overlaps cannot be prevented by
using larger and/or more powerful fans
alone. They must be prevented by ensuring
a substantially increased cooling
air flow with all the consequences
(e. g. regarding cost, power, and noise
emission).
ϑ 15K
ϑ A
∆ L35K
ϑ ü
T
ϑ4
ϑ 1b
1
s KM
ϑ E = ϑ 1,2
Fig. 8: Simplified schematic representation of temperature behaviour with a cross
feed radiator with temperature overlap.
ϑ 3
ϑ 1a
1…4 = Reference points in cooling circuit
VKM = Volume of cooling medium flow
VL = Volume of cooling air flow
ϑn = Cooling medium temperatures
tLn = Cooling air temperatures
ϑ15K = Average cool-down cooling medium 15 K
ϑ 30K
∆ L40K
ϑ A
T
2
1c
1
s KM
3
1b
13
ϑ1,5 = ϑU 4
6
1a
5,6
4,7
3,8
2,9
12
7
t L2
tL3 tL4 s L
t Lm
2
3
4
V KM, ϑ E
ϑ A
1
1a
VL, tLE 1b
∆L35K = Average warm-up cooling medium 35 K
sKM = Travelling distance of cooling medium
in cooling circuit
sL = Travelling distance of cooling air
through cooling circuit
T = Temperature
ϑÜ = Temperature overlap
11
8
ϑ E = ϑ 9,10
10
13
s L
10
11
tLm 11
12
12
13
ϑ E
V KM
6 5
ϑU Fig. 9: Simplified schematic representation of temperature behaviour with a counterflow
cross feed radiator.
1…13 = Reference points in cooling circuit
VKM = Volume of cooling medium flow
VL = Volume of cooling air flow
ϑE = Inlet temperature of cooling medium
ϑU = Average temperature of cooling medium
at counterflow
ϑA = Outlet temperature of cooling medium
ϑ30K = Average cool-down of cooling medium 30 K
ϑ A
1a
1b
1c
∆L40K = Average warm-up of cooling medium
40 K
sKM = Travelling distance of cooling medium
in cooling circuit
sL = Travelling distance of cooling air
through cooling circuit
T = Temperature
9
8
7
2
3
4
1
V L, t LE
11

3.3 Cooling air demand, flow
resistances, and safety
margins
Various factors are to be observed
when designing the radiator in order to
ensure efficient heat exchange. As far
as radiator type and cooling core selection
are concerned, upper and lower
flow velocity limits apply both for the
inside and outside. The corresponding
flow resistances in the air flow are not
only decisive for fan rating. They also
influence the coolant flow, together with
the flow resistances of the coolant lines
in the cooler group and vehicle.
Cooling cores with fins like those in use
today in vehicle construction require a
low cooling air flow even for high cooling
capacities. They produce a high
temperature increase of the cooling air
and the flow resistance in the fins is
high. However, they are more liable to
dirt accumulation than smooth or rippled
fins, which require a higher cooling
air flow for comparable cooling capacities
and have a lower flow resistance.
Apart from that, smooth or rippled
fins must have a higher block depth
than cut fins to achieve the same cooling
capacity.
In order to guarantee a specified cooling
capacity also when the radiator is
dirty, a possible degree of dirt accumulation
must be defined at the design
stage and a corresponding safety margin
must be provided.
If the safety margin is taken into account
in the form of an increase in heat
load, which is often done also to compensate
for inaccuracies in the calculation
of the exact heat load, errors will
result when determining the temperatures
of mass flows, including the cooling
air flow. The consequences of that
will be a sub-optimal radiator volume
and an incorrrectly chosen operating
point for the fan.
12
Cooling air inlet Safety
heat transfer 1)
∆p
in %
110
101.5
100
93
3
3’
1
2’
2
96 100 108
Clogging,
obstruction
under the roof 5% 10%
Vehicle
front or 5% 10 – 20%
side
Under the floor 10% 15 – 30%
Under the floor 2) 5% 10%
Table 3: Empirical data for dirt accumulation and safety margins of
radiators.
The recommended procedure is to account
for dirt accumulation by assuming
a reduced heat transition due to the
presence of a dirt film on the surfaces
of the cooling air side and to additionally
assume that a certain percentage
of the radiator front area is blocked
(table 3). The blocking of part of the
radiator front area influences flow resistance
in the air flow.
In this way, clear operating points are
found for the fan that can be simulated
IN/09V00
V L in %
Fig. 10: Characteristic of fan IN/09V00 (red); characteristic of radiator
under various conditions (blue).
1 – Design point with dirty cooler group
2’ – Design point with clean cooler group
2 – Operating point with clean cooler group
3’ – Design point with cold and clean cooler group
3 – Operating point with cold and clean cooler group (test rig)
1) These figures also
include possible power
reductions due to
production tolerances.
2) Using fine wire mesh
and filters.
in the vehicle both for a clean and for a
dirty cooler group at full throttle conditions,
steady-state conditions, and defined
air inlet temperatures (fig. 10).
The same applies when additional protective
measures in the form of fine wire
mesh or filters at the cooling air inlet
are taken to reduce dirt deposits due
to long maintenance intervals. The resulting
resistances and their behaviour
in the case of clogging must be taken
into account for the fan design.

3.4 Arrangement of the fan
The vehicle type and cooler group design
determine the fan arrangement in
the cooling system. Preferably, the fan
should be arranged behind the radiator,
i. e. it should act as a suction fan.
This arrangement will result in a relatively
even application of air to the radiator
as well as in favourable inflow
conditions for the fan and low transition
losses from radiator to fan. Good
inflow conditions exist in particular
when the ratio of radiator front area to
fan diameter does not exceed 3:2, or
when the largest radiator dimension is
not greater than 1.5 to 1.8 times the
fan diameter, and when the fan is not
located too close to the radiator. The
distance between radiator and fan
should be more or less the hub diameter,
as a minimum, however, 60 % of
the hub diameter.
A disadvantage in the case of free discharge
are the exit losses, which may
be high. With an air outlet velocity of
30 m/s at 70 °C, the exit losses may
reach a level of about 500 Pa (fig. 11).
When the fan is used as a blower fan,
high transition losses between fan and
radiator (fig. 12) and the non-uniform
application of air to the radiator must
be taken into consideration. This has a
negative effect on thermal performance
and must be taken into account when
designing the system. The impact of
these disadvantages can be alleviated
by increasing the distance between fan
and radiator. Choosing a blower arrangement
may be advisable if the cooling
air temperature after the radiator is
too high for the fan drive with electric
motor so that the motor would become
too big or difficult to cool.
In an increasing number of cases, determining
one operating point as described
above is no longer sufficient.
Further system performance curves
have to be prepared to determine other
V2 = 110%
1
V2 = 110%
2
V1 = 100%
Arrangement Pressure curve
Fig. 11: Pressure loss of suction fan
V2 = 110%
2
V1 = 100%
1
V1 = 100%
Arrangement
Fig. 12: Pressure loss of blower fan
V1 = Cold ccoling air volume flow
V2 = Hot ccoling air volume flow
∆p t = Total pressure difference of fan
∆p A = Dynamic outlet pressure loss
operating points before it is possible to
assess the fan design and check it for
limit conditions.
Important differences result if
- calculated over or underpressures
from the airflow around the vehicle
turn out to be higher or lower than
assumed,
– +
∆p t = 125%
– +
∆p A
∆p A
∆p t = 100%
Pressure curve
1 Ventilator
2 Cooler
- the calculated dirt deposit on new or
cleaned radiators does not exist or
becomes greater than assumed for
various reasons,
- air density is higher at low cooling
air temperatures while the system is
operated at maximum fan speed,
- function checks are performed on
the test rig while the radiator is cold
(fig. 10).
13

3.5 Fan operating points
Decisive for fan design is the operating
point for the specified service data. The
service data are usually the maximum
conditions for the heat flow, temperatures,
safety margins against dirt, and
driving conditions. A suitable fan characteristic
and fan type are selected for
the service data, e. g. for a volume flow
of 25 m 3 /s with a total pressure loss of
1400 Pa (fig. 13). A prerequisite is an
orderly inflow to the fan. This fan selection
procedure is still adequate today
whenever the operating point lies in the
steep section of the characteristic and
in the favourable area of the fan’s characteristic
diagram. Moreover, there
must be no operating conditions that
14
Speed n = [1/min]
D = [mm]
∆ p t in Pa
1800
1600
1400
1200
1000
800
600
1460
3000
2850
2800
2400
B
2000
1800
1500
ηt
D = 1000
0,8
1400
1200
D = 1120
400
10 15 20 25 30 40
V L in m 3 /s
D = 1250
Fig. 13: Excerpt from fan performance graph IN/09V-BF1.
B = Reference point
L = Performance point
ηt = Overall efficiency
∆pt = Total pressure difference = total pressure increase
L
0,85
deviate very much from the design conditions.
The operating point of the fan is determined
taking into account all the components
exposed to the cooling air, inlet
and outlet losses, and all external
influences such as over and underpressures
on the outside of the vehicle.
These are caused by the speeddependent
flow around the vehicle and
by the installation characteristics of the
cooling system (fig. 14). Secondary air
or return flows of hot cooling air may
have an influence on air density at the
cooling air inlet and thus on flow resistance.
This is the case, for instance,
when the air outlet is toward the track
bed.
D = 1400
0,9
IN/09V94
D = 1600
IN/09V00
0,93
IN/09V06
D = 1800
IN/09V12
∆p
in %
100
93,5
88,5
80
70
70
The resistances of wire mesh, filters,
and radiators is easy to calculate. The
calculation of resistance from air-conducting
housing elements, transition
losses from radiator to fan or vice-versa
is possible based on test results or requires
idealized models, e. g. according
to Idelchik [8]. Resistance values
based on experience are also available
for obstacles inside the cooler group,
as in pipes or various other components,
or obstacles downstream of the
cooler group. Where this is not so, fixed
values must be applied. When detailed
calculations are made, the corresponding
inaccuracies will be negligible.
2’
1
3’
100
2
3
108,5 112,5
IN/09V00
V L in %
Fig. 14: Changes of design and performance points with different
systems and operating conditions.
1 – Design point for dirty cooler group with under/overpressure
2’ – Design point for dirty cooler group without under/overpressure
2 – Operating point for dirty cooler group without under/overpressure
3’ – Design point for clean cooler group without under/overpressure
3 – Operating point for clean cooler group without under/overpressure

3.6 Noise emission and sound
insulation
Reducing noise emission is of paramount
importance today. The limits vehicle
customers specify for noise emission
are getting lower and lower and
non-compliance with these limits is penalized.
One reason for this is an increasing
demand for comfort, another
one is more stringent safety at work and
environmental legislation.
So far it has only been possible to predict
noise emission for individual
sources in the rail vehicle but not for
complete systems comprising several
components. Determining and quantifying
the corresponding influences is
certainly not an easy task. They can
only be estimated based on experience
with similar installations.
The purpose of fans and blowers is to
deliver a volume flow against a resistance.
The resulting noise may be unwelcome,
but it is basically inevitable.
For this reason, the development of
fans has always been linked to investigations
concerning noise emission, the
causes of the noise and possibilities of
reducing it. The immense number of
examinations and research projects
existing in this field shows how complex
interaction between fluidics and
accoustics is [1, 9, 10].
Apart from factors associated with the
fan itself and the system layout, noise
emission is influenced mainly by the
volume flow, pressure rise, efficiency,
speed, and diameter. The total sound
power level of a fan will be about
2 dB(A) lower, for instance, if the required
volume flow can be reduced by
10 %. A reduction in fan speed by 20 %
will reduce the total sound power level
by another 4 dB(A). A reduction in this
level of 3 dB(A) corresponds to a 50 %
reduction in sound pressure so that the
noise only appears half as lound. This
is a substantial improvement, therefore.
With the same sound pressure, high frequencies
are considered louder and
more unpleasant than low ones. The dominating
frequency of the noise radiated
by the fan is calculated by multiplying fan
speed by blade number. These two factors
should therefore be kept small.
Having a small number of blades is
basically possible and also preferable
for aerodynamic reasons. However, it
will result in a higher depth of the
impeller as the dimensions of the blade
profiles will increase.
Reducing fan speed has its limits. The
angle of attack of the blades can be a-
•
dapted to the specified volume flow V, the
corresponding velocity vectors, and the
assigned energy Y. There is a fundamental
relation between peripheral velocity at
the outer diameter of the impeller and the
total impeller cross-section. For the pressure
coefficient, the peripheral velocity u
appears in the denominator squared
ψ =
2 • Y
(1)
u 2
whereas for the flow coefficient it appears
in the denominator to the power
of one together with the total cross-section
of the fan
ϕ = (2)
π • D
This means that special blade profiles
are needed to realize low fan speeds.
2 •
4 • VL • u
Fans with irregular blade arrangements
and unusual blade shapes have proved
unsuitable for cooling systems of rail
vehicles.
When the fan is installed in the centre of
the system, it is somewhat shielded by
the radiator, filter, deflectors and housing
so that sound propagation is restrained.
A further reduction in noise
emission is possible by using suitable
insulating materials. Sound reflection inside
the cooler group and radiation from
the housing are reduced by absorption.
Freely sucking and freely discharging
fans are a problem. They emit sound unobstructed.
Any measures to reduce
noise emission would be too costly in this
case, mainly because of the high flow
rates involved that would require excessively
large silencers. Moreover, such silencers
would not be efficient because
they would cause considerable flow loss.
Some of the silencing measures already
mentioned, in contrast, can be
adopted easily and at low cost. Using
them in a clever and systematic way
can yield considerable results, as can
be seen from the following examples.
On a commuter train with underfloormounted
cooling system it was found
that sound radiation of the freely sucking
fan was about 3 dB(A) lower when
the air does not enter the system via
grids in the side skirts as usual. A cover
plate coated with sound-absorbing
material will therefore be installed on
the vehicle at a suitable distance in front
of the fan. This plate serves at the same
time as a dirt deflector against the track
bed underneath it.
For a tilting train set with underfloormounted
cooling system the sound
pressure originating from a freely discharging
fan measured beside the vehicle
during development research
could be reduced by about 3 dB(A) by
modifying the housing to provide an
indirect cooling air outlet. Fitting the unit
with sound-absorbing material reduced
the noise by another 4 dB(A).
Efficiency influences noise emission,
too. It is determined by the location of
the operating point in the characteristic
diagram and there is always a
greater or smaller distance from the
point of optimum efficiency. The greater
this distance, the higher the noise level.
A ratio is derived from this distance that
influences the noise. The closer the
operating point to the efficiency optimum,
the lower the noise emission.
15

4. Modern cooling systems and their design
Great demands are made on rail vehicle
components today, combined with
a high power density and reduced
safety margins.
New developments were fostered by
new generations of vehicles for highspeed
and commuter service. Availability
targets and the competitive situation
and the resulting cost pressure on
manufacturers and operators played a
major role in this process.
Today’s cooling systems must offer the
following features:
• competitive and economical,
• compact design with a high power
density,
• light and robust at the same time,
• technically innovative,
•low and precisely controlled auxiliary
machine power (fans, pumps)
in order to save energy,
• long service life,
• high availability, i. e. low maintenance,
high reliability, trouble-free,
• recyclable and environmentally
compatible with regard to materials
and consumables used and
productionprocesses employed,
•low noise emission [9, 10].
These requirements – some of which
contradict each other – are detailed in
tender specifications for new vehicles.
Both vehicle manufacturers and end
users expect a competent cooling system
manufacturer to contribute his
know-how to the vehicle design and
offer solutions to possible problems.
16
Consequently, the cooling system
manufacturer must be consulted already
when drafting the tender specification
at the project planning stage,
especially with regard to the arrangement
of components and the cooler
group in the vehicle. The arrangement
and properties of components often
have a negative influence on their cooling,
e. g.
• insufficient space for cooling,
• unfavourable routing of cooling air
(outside and inside the cooler
group),
• problematic air intake and discharge
conditions (dirt, leak air),
• speed-dependent external pressure
conditions due to the air flow around
the vehicle, and
• high auxiliary machine power.
To achieve low noise emission levels,
low costs, and high efficiency in the
system, all parties concerned must examine
and assess all possible options
and influences. The consequences of
changes and alternative solutions must
be presented and the best solution for
a given project must be found in cooperation.
An example will show that this
way of proceeding is both appropriate
and successful.
When designing a series of electric locomotives
rated at 4 to 8 MW using
standardized modules, various concepts
and variants derived from them
were calculated completely, both with
regard to technical parameters and
cost. Experts in the field of electrical
driveline components (traction inverter,
transformer, traction motors), vehicle
construction, aerodynamics, accoustics,
and cooling were involved.
The following design variants were considered,
among other things:
- central cooling air intake with and
without air cleaning for all consumers
(inverter/transformer cooling,
traction motor cooling, air-conditioning
units, auxiliary equipment,
machine compartment ventilation),
- individual intake with air cleaning at
the consumers that need clean
cooling air,
- combinations of different cooling
devices vis-à-vis single or multiple
arrangement with different types of
redundancy and drive power.
In order to achieve redundancy in air
supply, additional measures had to be
taken for the combination solutions, e.
g. for cooling air supply by the sister
unit. This required a high drive power,
among other things due to high flow
losses in air ducts and to the required
flow restriction when matching the cooling
air flows to the various consumers.
Also assessed were the weight, space
requirements, ease of installation, accessibility,
noise emission, and damping
requirements as well as the energy
consumption of auxiliary machine systems.
The assessment revealed advantages
of combination solutions in the categories
of space requirements and initial
cost. However, these solutions later
proved to be impracticable due to high
operating costs for energy supply and
high costs for redundancy. Partial combinations
like those successfully used
for inverter and transformer cooling, in
contrast, came out well. This applies in
conjunction with central air intake – with
air cleaning and silencing facilities if
necessary – and the possibility to have
different segments for the various consumers.
The examinations also showed that optimization
of the total system should be
given preference to cost reductions for
individual components only. This applies
even more to new vehicles as the demands
made on them continue to rise.

5. Economical considerations
Development costs money. It can only
be financed today if there is a financing
concept that is suitable for small lot
production as found in the rail vehicle
sector. Taking into account market appreciation
of high-quality components
would be desirable, but there are no
signs that it might ever be possible under
the current competitive conditions.
There is a wide variety of fan types and
designs available. Cheap offers are
based on standard modules for different
applications and will rarely be suitable
for rail vehicles. Just comparing
prices is not the way to find a suitable
fan. An analysis of the price/benefit ratio
and of life cycle costs is indispensable.
The price difference of a few thousand
Euro between a fan with a high overall
efficiency of up to 90 % - such efficiency
levels are possible by using slim blade
profiles with optimum twist - and one
of 50 to 60 % efficiency is often compensated
by fuel savings already after
two to three years.
An example of a railcar with 480 kW
engine output illustrates this. With a
cooling requirement of 2 x 5 m 3 /s and a
total pressure of about 1000 Pa, the use
of high-efficiency fans permits fuel savings
of about 2400 l/yr. (tables 4
and 5).
Standard system
Voith railway system
Efficiency 55% 84%
Power consumption ca. 9.0 kW ca. 6.0 kW
Hydraulic pump * ca. 24 kW ca. 16 kW
Table 4: Comparison of batch-manufactured standard cooling systems
vis-à-vis special rail vehicle cooling systems.
Fan design:
Standard cooling system – plastics, aluminium
(without fan shroud)
Rail vehicle cooling system – aluminium (with fan shroud)
* with 2 fans η Hy = 0.75
Assumption:
• max. power of standard fan 24 kW
• max. power of Voith fan 16 kW
• max. savings 8 kW
• Annual average savings about 30% 2.6 kW
• 4 000 operating hours per year
Energy savings: 2.6 kW * 4 000 h/year = 10 400 kWh/year
Fuel savings: 10 400 kWh * 190 g/kWh ≈ 2 000 kg/year
Table 5: Comparison standard fan / Voith fan
2 000 kg / 0.83 kg/l ≈ 2 400 liter/year
This example shows that the fan as a
dynamic element and aerodynamics
are at least as important as the radiator
element when optimizing the cooling
system as a whole. Optimally tuned
complete systems guarantee high efficiency
and thus also maximum
economy. Future vehicle concepts will
have to take this into account.
17

6. Summary
Energy consumption and noise emission
of auxiliary machines in rail vehicles
can be substantially improved by
selecting suitable fans and choosing a
favourable aerodynamic design for the
cooling system.
When procuring new fans, the entire life
cycle of the cooling system should be
assessed rather than focussing on the
initial cost alone. Paying a higher price
for a technically advanced high-quality
solution may well be economical since
the price difference will easily be compensated
by protection of the environment
and energy savings.
18
References
[1] Marcinowski, H.: Einfluß des Laufradspaltes
und der Luftführung eines
Kühlgebläses axialer Bauart.
(1953) MTZ 14/9, S. 259 - 269
[2] Marcinowski, H.: Der Einfluß des
Laufradspaltes bei leitradlosen, frei
ausblasenden Ventilatoren.
Voith-Forschung und Konstruktion 3
(1958) 8, S. 2.1-2.3.2
[3] Marcinowski, H.: Experimentelle
Untersuchungen in der Lufttechnischen
Abteilung.
Voith-Forschung und Konstruktion 4
(1958) 11, S. 15.1-15.16.
[4] Goepel, W.: Regelventilator mit mechanischem
Antrieb.
MTZ 19 (1958) 8, S. 277-279 und
Voith-Druck G 596, S. 4.68.
[5] Goepel, W.: Entwicklungsstand der Voith-
Kühlanlagen für Diesellokomotiven.
Der Eisenbahningenieur, 2 (1968) und
Voith-Druck G 595.
[6] Schmid, R.: Neue energiesparende
Hilfsmaschinenantriebe für dieselgetriebene
Schienenfahrzeuge.
Voith-Druck G 1513 (1997) H.10
[7] Joas, K. u. Schmid, R.: Regeleinrichtungen
für einen hydrodynamischen
Antrieb.
Patentschrift DE 43 04 403 C2, 1994
[8] Idelchik; J.E.: Handbook of Hydraulic
Resistance, Springer-Verlag,
Berlin 1986
[9] von Hofe, R.: Geräuscharme Kühler-
Lüfter-Systeme. Forschungsberichte
Verbrennungskraftmaschinen.
(1980) H.R 276
[10] Kameier, F.u.W. Neise: Spaltmodifikation
verbessert Ventilatoren.
Heizung, Lüftung, Haustechnik.
45 (1994) 8, S. 395-400.
[11] Voith Turbo GmbH (Hrsg.): Kühlsysteme
für Lokomotiven und Triebwagen.
Voith-Druck G 935d (2000) H.8.
[12] Voith Turbo GmbH (Hrsg.), Kühlsysteme
für Schienenfahrzeuge – Referenzen.
Voith-Druck G 1968d (2001) H.3.

20
Voith Turbo GmbH & Co. KG
Division Rail
Postfach 2030
D-89510 Heidenheim
Tel. (07321) 37-4485
Fax (07321) 37-7096
Email cooling-systems@voith.com
www.voithturbo.com
G 1671 e 3.2002, 1000 Klö/MP Printed in Germany.
Subject to modification. The information given herein is of a descriptive nature only and does not constitute any assurances
as to the performance or the characteristics of any equipment supplied.