Efficient Volvo Bus Cooling System,Using Electrical Fans

A comparison between hydraulic and electrical fans

RITA BAILAO MARTINS FERNANDES

Master’s Degree ProjectStockholm, Sweden June 2014

TRITA-EE 2014:xxx

AbstractEconomical and environmental factors together with energy policies towards more efficient sys-

tems are the driving force for the development of the vehicle industry. Significant changes have beenmade to fulfill new emissions legislation but the basic internal combustion vehicle architecture hasbeen kept. New emission treatment systems that increase the thermal loading of the cooling systemhad been added within the same package envelope as before, which means less space to place coolingfans and a greater need for airflow. Changes in the cooling system, namely the replacement of thehydraulic fan drive system by electrical fans is one of the energy efficient alternatives for severalcity buses under certain environments, like the ”typical red city buses”, well-known in the UnitedKingdom. In this thesis study, hydraulic fans are compared with electrical fans and a road-mapof the benefits and drawbacks of the two systems is developed, based on real traffic performanceperformance data and the results of existing simulations and tests. In addition, new simulations arepresented in order to find the most efficient design for the cooling system as well as a comparisonof these results with previous ones. This road map will be used later by Volvo-Buses Group as atool to better understand in which circumstances electrical fans can be beneficial, in terms of fuelconsumption, noise production, cooling performance, control of the fans and associated costs.

Keywords : hydraulic fan cooling system, electrical fan drive system, radiator, fan efficiency, fanshroud, static pressure, oil cooler, charge air cooler, built-in-resistance.

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Part I

PrefaceDear reader of this thesis. This master thesis project is the final sign off of the last year of my masterof science. It is based on engineering reports made by Volvo engineers, scientific articles and simulationresults from Volvo GTT and Volvo Buses. The research was carried out from later January 2014 to July2014, and presented in the end to Volvo AB and to KTH (Royal Institute of Technology), in Stockholm.The topic was based on the company’s request to analyze the current options for the fans of the buscooling systems and to better understand electrical fans in terms of cooling performance. The masterthesis project was carried out by Rita Fernandes as an intern, Reza Fakhrai as supervisor at the universityChrister Kjellgren as supervisor at the company. The project was financed by the AB Volvo, namely bythe department of Powertrain Development and conducted by the Cooling System group, with CharlotteEldh as the line manager. The empirical research took place in Volvo Buses offices in Gothenburg, mostlylocated in Arendal. All of the co-workers: Chirster Kjellgren, Erik LindÃľn, Joel SÃűrborn, Eva BjÃűrk,Peter Gullberg, Erik Dahl, Dalibor Cuturic, Jessica LexÃľn, Stephan SchÃűnfeld and the remaininggroup are highly appreciated for their help and availability during the thesis work. Moreover the inputfrom all the interviewees was crucial and without it, there would not be the possibility to get so quicklythe results of this thesis. Also a special thanks to my friend Andrea who helped to feel home in this newcity of Gothemburg and my Portuguese friends who guided me through all the aerodynamics concepts.It has been very nice to get to know you, and I hope to continue to learn more from you. Francisco,thanks for helping me with proper English. Finally, I would like to thank my family and all of my friendsfor their warm and kind support. At the end of the day, you have always been there for me. Thank You.

Confidentiality clause Due to confidentially reasons, all the parts with sensitive information weredeleted form this report. This includes all the tables, figures and graphs with valuable content for VolvoAB. A different and complete report of this thesis was delivered to Volvo AB, as well as a shorter versionwith the main results and recommendations.Gothemburg, July 2014 Rita Fernandes

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Contents

I Preface 2

II Introduction 8

III Objectives 9

IV Volvo Buses and cooling system 10

1 Thermodynamic and Heat Transfer fundamentals 101.1 Conduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 101.2 Convection . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11

2 Bus Cooling System 112.1 Fan Systems . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 122.2 Radiator . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 132.3 Engine water pumps . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 142.4 Charge air cooler . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 152.5 Oil cooler . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 152.6 Fans shroud . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 152.7 Thermostats and other sensors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 152.8 Header Tank and Recovery System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 16

3 Hydraulic Fans 163.1 Coolant Temperature Sensor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 163.2 Hydraulic pump . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 163.3 Pump control plate . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 173.4 Control valve and check valve . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 173.5 Hydraulic fan motor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 17

4 Electrical Fans 174.1 Radial, Axial fans and Diagonal fans . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 194.2 How to select the right fan . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 20

4.2.1 The total cooling requirements . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 214.2.2 Static pressure . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 214.2.3 Total System Resistance / System characteristic curve . . . . . . . . . . . . . . . . 214.2.4 System Operating Point . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 224.2.5 Stall effect and instability regions . . . . . . . . . . . . . . . . . . . . . . . . . . . . 224.2.6 Efficiency of electrical fans . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 22

V Modeling and analyzing electrical fans 23

5 What is limiting electrical fans 235.1 Fan blade types . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 235.2 Fan Laws . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 235.3 Influence of density . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 245.4 Impact of Fan Diameter . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 245.5 Series and Parallel Operation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 25

5.5.1 5 SPAL electrical fans (305 mm) vs 2 SPAL electrical fans (405 mm) . . . . . . . . 285.6 Blade angle . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 285.7 Distance between fan blades and fan ring . . . . . . . . . . . . . . . . . . . . . . . . . . . 28

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5.8 Voltage imposed and consequences in fan curves characteristics . . . . . . . . . . . . . . . 285.9 Possible fan suppliers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 315.10 Pusher Fans vs. Puller Fans . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 31

6 Evaluation of an electrical cooling system: London buses case: B5LH (hybrid) andB9TL 346.1 Exhaust gas recirculation (EGR) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 346.2 Real Data taken from Volvo Data Base (LVD) . . . . . . . . . . . . . . . . . . . . . . . . 346.3 LAT and IMTD . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 346.4 Derating . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 35

VI Simulations and Tests 36

7 City-buses 367.1 Methodology . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 367.2 AMESim . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 367.3 Input Data . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 377.4 Implementation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 377.5 Built-in-resistance . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 377.6 Effect of removing the oil cooler . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 397.7 Cooling performance: LAT and IMTD . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 407.8 Separated CAC and radiator installation . . . . . . . . . . . . . . . . . . . . . . . . . . . . 407.9 Effect of the fan shroud . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 42

8 Coaches cooling performance using electrical fans 448.1 Cooling performance: LAT and IMTD . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 44

VII Conclusions 45

VIII Recommendations and future work 47

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List of Figures1 Scheme of bus cooling system heat exchangers, air flow direction and respective pressure

differences . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 122 Scheme of a turbocharger.Adapted from [39]. . . . . . . . . . . . . . . . . . . . . . . . . . 133 Scheme of a water pump. Adapted from [42]. . . . . . . . . . . . . . . . . . . . . . . . . . 144 Scheme of hydraulic valves . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 175 The effect of multiple fans on system pressure and flow rate. Adapted from [3]. . . . . . . 206 Change of static pressurer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 217 Sketch of fan and system curve. Adapted from [3] . . . . . . . . . . . . . . . . . . . . . . . 228 Stall: unstable zone in the fan curve . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 239 Effect of an increase in fan diameter in air flow and static pressure . . . . . . . . . . . . . 2510 The effect of multiple fans on system pressure and flow rate . . . . . . . . . . . . . . . . . 2611 Lower duct pressure due to fans placed in series . . . . . . . . . . . . . . . . . . . . . . . . 2612 5 fan curves in a parallel configuration, with a diameter of 305 mm and 3750 rpm speed . 2713 2 fan curves in a parallel configuration, with a diameter of 450 mm and 3750 rpm speed . 2714 Different fan curves for the same system resistance . . . . . . . . . . . . . . . . . . . . . . 2815 A comparison between 2 real SPAL fans versus 5 small SPAL fans . . . . . . . . . . . . . 2916 Dependency between pressure drop and blade angle. Adapted from [3]. . . . . . . . . . . . 2917 Fan ring location in the cooling module. Adapted from [3]. . . . . . . . . . . . . . . . . . 3018 Dependency between pressure drop and blade-ring distance. Adapted from [3]. . . . . . . 3019 Effect of voltage in fan performance . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3120 Sound level for different electrical fans . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3321 5 electrical fans model: Heat stack (radiator and CAC) connected to 5 electrical fans . . . 3822 Heat module: CAC, oil cooler and a radiator in the backside . . . . . . . . . . . . . . . . 3923 CFD calculations: case a) without oil cooler and case b) with oil cooler . . . . . . . . . . 4024 LAT change versus fan speed . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4125 IMTD versus fan speed . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4126 CFD model with fans . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4227 CFD model without fans . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4228 Different EBMpapst fan combinations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4429 Radiator performance curves for different coolant flows . . . . . . . . . . . . . . . . . . . . 4530 Different fan curves for coaches and operating points . . . . . . . . . . . . . . . . . . . . . 46

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Nomenclature∆p Static pressure drop (Pa)

m mass flow (kg/s)

A surface area (mˆ2)

AMESim Advanced Modelling Environment for Performing Simulations of Engineering Systems

AV Volumetric flow (m3/s)

B5TL Hybrid bus model with 5L cylinders

B9TL Bus model with 9L cylinders

BiR Built in Resistance (-)

CAC Charge air cooler

CFD Computational Fluid Dynamics

cp Specific heat at constant pressure (J/(g.K))

D Fan diameter mm

E Total energy of the system (J)

e Flow energy (J)

EBMpapst Germany fan supplier

EGR Exhaust gases recirculation

g Gravitational constant (m/s2)

GTT Volvo Group Trucks Technology

H Blades shape factor (dB)

h Heat transfer coefficient ([W/m2K])

ICE Internal combustion engine

IMTD Intake manifold temperature difference (℃)

Lw Sound power level (dB)

LV D Logged Vehicle Data Analysis Tool

P Power (W)

p Total pressure (Pa)

pD Dynamic Pressure (Pa)

PWM Pulse-width modulation

Q Heat transferred (J)

qspecific Specific Heat of the Radiator (kW/K)

RAD LOW Radiator type used for City-Buses

RAD WIDE Radiator type used for Coaches

Re Reynolds number (-)

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rpm Revolutions per minute

SPAL Italian fan supplier

Tf Fluid temperature (℃)

Ts Surface temperature (℃)

TTT Top tank temperature (℃)

u Fan tip speed (m/s)

V Fluid velocity (m/s)

V BC Volvo Bus Corporation

W Sound power (W)

W Work done on the system (J)

W0 Reference sound power

z Height (m)

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Part II

IntroductionThe improvement of the thermal management of a vehicle can contribute significantly to reducing emis-sions and lower fuel consumption. The demand for cooling power varies greatly with the surroundingenvironment. Under certain conditions, wherein the climate is not too adverse and the city’s topographyis smooth enough, the cooling demand is much lower than a situation where the topography is roughand the climate is hot. Electrical fans represent a good solution for climates where the cooling demandis lower since they can work at lower speeds and are almost totally independent from the engine speed.When used on many auxiliary applications, electrical fans can provide additional air flow to preventoverheating, supplement the existing belt-driven fan as a new means to control unacceptable noise levelsand aid in the redesign of the engine compartment by relocating the heat exchangers. This leads to thecentral question of this thesis:

Are electrical fans the best option when compared to hydraulic fans, for the same coolingperformance demand?

The immediate answer seems to be positive since electrical fans consume only 2 kW and hydraulic fansconsume up to 25 kW (for their maximum speed). This represents 12.5 times more power consumption,even idling or operating at lower speeds. This is due to several reasons, and the aim of this thesis is toexplain and compare the two solutions, with a simple 1D model, (in steady state conditions), togetherwith the results of several previous simulations and tests, done during the past 10 years by Volvo Groupengineers. It is important to understand how both systems behave when varying the heat load, fanspeed, air flow, coolant flow, etc, over time and also what happens if one introduces changes in thesystem configuration, namely when changing the fan shroud, the consequence of removal of the oil coolerused by the hydraulic system and also to analyze the changes in static pressure if one splits the radiatorand the charge air cooler. Thus, the scope of this thesis is to better understand the difference in powerconsumption and cooling performance of a cooling system moved by a hydraulic fan and another, movedby electrical fans. To do that, the city-buses and coaches are used as case-studies. For the city-buses,two different bus models were selected: B9TL and B5LH. B9TL is basically a 9L engine, typically usedin London and known as the London red city-buses. This model was recently replaced by a B5LH,(5L engine). Every once in a while, vehicle manufacturers have to make changes regarding emissionsand fuel consumption [34], since there are rules made by the European Union which state the maximumtemperatures allowed in the different parts of the engine system in order to control the emissions of NOx,total hydrocarbons, particulate matter and COx. They are known as the European Emission Standards:Euro 5 and Euro 6. Euro 6 was implemented in 31th December 2013 for buses and trucks. B5LH isan hybrid bus that in Euro 5 use a hydraulic fan and in Euro 6 electrical fans. Although the engine isdifferent for the two emission standards, it is interesting to compare this model.

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Part III

ObjectivesThe objective of this thesis is to find possible fan cooling systems configurations and to better under-stand why electrical fans are advantageous compared to hydraulic fans. The first step is a literaturereview about the bus cooling system, in order to find a definition of cooling performance and thus todiscover what is limiting electrical fans, by studying the different types of fans and respective suppliers,the possible system voltage and configuration. The background analysis consists in the analysis of theLVD, a Volvo database that keeps tracking the vehicles’ performance in real time each time they go toservice. They represent the data in real traffic conditions and several variables can be chosen includingthe fuel consumption, the voltage, cooling capacity, fan clutch type, fan speed and engine speed. To-gether with the real values of such variables, a 1D model using AMESim software package is used. Theconstructed steady-state model should be useful to evaluate different heat transfer parameters of thewhole system. These simulations will then be used for other VBC future projects since they will allow toreach conclusions related to the consequences of switching to electrical fans, namely in what concerns fuelconsumption, the optimized number and configuration of the electrical fans, cooling performance, noise,product cost, powertrain performance and ventilation of engine compartment. Some CFD calculationsare also performed in order to reach conclusions related to the configuration of the system, namely inwhat concerns the disposition of the charge air cooler and the radiator, since due to spacial constraints,they are in a sandwich configuration. The cooling performance gained if they are placed side by sidewas not so far calculated; a measurement of the engine bay temperature was the only study done at thetime this thesis was written, with a gain of 20℃, though with sandwich configuration using hydraulicfans and split configuration using electrical fans. CFD is also used to measure the velocity of the airthrough the radiator’s surface, since the cooling performance depends also on that together with thebuilt-in-resistance.

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Part IV

Volvo Buses and cooling systemThis chapter provides a brief description of Volvo Buses. Volvo Buses offers a broad range of passengertransport solutions to match the diverse customer demands. The product offer can be classified intothree segments: city, commuter and coaches. This thesis will analyze the cooling system of city-busesand coaches.Volvo Buses has a relatively small business and restricted budgets so it cannot afford torun standalone process development efforts. However, the business scenario is rapidly changing with theintroduction of emerging economies in Asia and South America, which have different requirements onproducts than their European counterparts and the strategy of adapting European products to thesecountries is no longer a feasible idea since the environmental conditions are most part of the timesconsiderable different.

1 Thermodynamic and Heat Transfer fundamentalsThe engine cooling system from a perspective of heat transfer, can be summarized in three points.

• Heat transmission from engine to engine coolant

• Engine cooling circuit, that transports the heat from the engine to the heat exchangers (radiatorand charge air cooler)

• Dissipation of heat from the hot coolant into the ambient air through forced convection

According to the second law of thermodynamics, if an isolated system is not in equilibrium, its entropywill tend to increase over time, until it reaches an equilibrium for its maximum value. This means that itsenergy disperses over time, and less energy is available to do useful work. On the other hand, accordingto the first law of thermodynamics, energy cannot be destroyed and it changes from one process toanother. ∫

dE =∮∂Q−

∮∂W (1)

Exergy is the term used to describe the destroyed energy when a process involves a temperature change,which is proportional to the entropy increase of the system and its surroundings. Heat cannot flowfrom a material at lower temperature to a material at higher temperature spontaneously. It is necessaryto provide energy for that to happen, either by work or heat. For a cooling system, heat at a lowertemperature is worth more than heat at a higher temperature. Therefore, the efforts are in the directionof cooling a flow with another flow that is as close in temperature as possible. Whenever there is atemperature difference in a medium or between media, heat transfer must occur, either by conduction,convection or radiation. Conduction can occur between solids or a stationary fluids, convection froma surface to a moving fluid and in radiation heat exchanges between two surfaces. Since the majorityof the materials present in the cooling system are made of aluminum, iron and other materials withlower emissivity coefficient, ε, radiation transfer is neglected in this study, since according to the Stefan-Boltzmann Law, it is dependent on the materials emissivity. (even if the emissivity coefficients for somematerials vary with the temperature.)

1.1 ConductionEnergy is transferred on a molecular scale and no movement of macroscopic matter relative to oneanother. The energy transfer occurs by interaction of particles more energetic to less energetic ones.This net transfer of energy by random molecular motion is called a diffusion of energy. This also occursin liquids but molecules are more closely space. In solids the conduction is due to lattice vibrations.Fourier’s law represents the relation between the time rate of heat transfer through a material and thetemperature gradient of a certain area, through which the heat is flowing. This law can be expressed bythe differential and integral form. The differential form of Fourier’s Law of thermal conduction showsthat the local heat flux, Q, is equal to the product of thermal conductivity, k, and the negative local

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temperature gradient, −∆T . The heat flux density is the amount of energy that flows through a unitarea per unit time.

Q = −k∆T (2)

In Equation 2 Q is the local heat flux (W), k is the material’s conductivity (W Kˆ-1), and ∆T is thetemperature difference (K). The thermal conductivity of a material generally varies with temperature,thus, k is a measure of the rate at which heat flows through a given material. It is measured by theamount of heat, Q, that flows through the material of a thickness, dx, per unit of temperature difference,∆T For many simple applications, Fourier’s law is used in its one-dimensional (1D) form [29]. Forinstances, in the x-direction,

qx = −kdTdx

(3)

unit

1.2 ConvectionConvection occurs due to temperature differences between a fluid and a solid boundary. Due to the fluidmotion redistribution of energy partly happens due to conduction. If the motion is due to an externaldevice, like a fan in the case of an engine cooling system, it is called forced convection. In the otherhand if the convection is due entirely to density gradients in fluid it is called natural convection. Heattransfer through convection is governed by Newton’s Law of Cooling [56].

Convective Heat Flux q = h(Ts − Tf ) (4)

Convective Heat Transfer RateQ = hA(Ts − Tf ) (5)

The area A is the surface exposed to the convective heat transfer, Ts is the surface temperature and Tfis the fluid bulk temperature. The convective heat transfer coefficient h[W/m2K] will vary for differentflow regimes (i.e. laminar or turbulent flow), fluid properties and temperature differences. Forced aircoolers and heaters are examples of equipment that transfer heat primarily by forced convection. Whena fluid flows over a flat plate, a boundary layer forms adjacent to the plate. The velocity of the fluid atthe plate surface is zero and then increases to its maximum value just past the edge of the boundarylayer. This boundary layer formation is important, since the temperature change from plate to the fluid(thermal resistance) is concentrated there. Where the boundary layer is thick, thermal resistance ishigher and the heat transfer coefficient is small. At the edge of the plate, the boundary layer thicknessis theoretical zero, and the heat transfer coefficient is infinite.

2 Bus Cooling SystemThe bus cooling systems plays two roles: the first is to make fuel efficient engines and the second in thiscontext is that the cooling systems constitutes a loss. Internal combustion engines generate mechanicalpower by the energy generated by the heat flows in the engine compartment. As like any other machine,the efficiency of the engines is not unitary, so more heat has to come to the engine than what wouldbe required if the process was isotropic. The outcome difference is waste heat that has to be removedin order to keep the components working at its designed conditions, avoiding that the lubricants andoil burn. This excess heat is also used to run the turbo-compressor. Thus, internal combustion enginesremove heat through the hot exhaust gases, through cool intake air and mainly, by the cooling system.A typical bus cooling system is composed by several components that differ to a system that is driven byone hydraulic fan or multiple electrical fans. The main components that are common to both systemsare the charge air cooler, the radiator, thermostats and thermocouples, valves and pumps. Nowadaysthe cooling system is mounted in the rear of the vehicle due to space requirements. Previous studies alsoconcluded that the cooling performance for the rear mounted installation is favorable compared to thefront mounted cooling package. This was mainly due to the low vehicle speed, the high fan speed and tofewer obstacles around the cooling module resulting in a lower system restriction within the installation.A radiator is basically a set of heat exchangers where the material that cools the engine (the coolant)

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Figure 1: Scheme of bus cooling system heat exchangers, air flow direction and respective pressuredifferences

passes by when its temperature reaches a certain value. This value is measured by a thermostat locatednear the engine. When the coolants arrives into the radiator, with an initial temperature of 90 ℃, itstemperature lowers due to the heat transfer from the outside air that passes firstly by the charge aircooler and finally by the radiator. This air flow is originated by the movement of the fans that pull theambient air. Thus, in summary, the cooling system of a bus is a sandwich of a charge air cooler, an oilcooler, a radiator and fans [59].

The reason why there is a charge air cooler (or intercooler) is because buses have a turbocharger thatneeds also cooling. A turbocharger is a very complex mechanical system that increases the efficiency offuel combustion by increasing intake air charge density through nearly isobaric cooling. A turbochargerhelps to supply air by forcing it into the combustion chamber making the combustion more efficient.Exhaust gases, normally at 220 ℃[7],[60] are channeled through the turbine housing where they increasespeed. The gas then flows through the turbine wheel where it slows down again, releasing energy. Theturbine wheel drives the common shaft that connects it to the compression wheel. The compressingwheel is driven by ambient air into the compressor housing, raising both its pressure and air density,forcing this air to the engine [38].

2.1 Fan SystemsThere are three different fan systems: direct driven, hydraulic driven and electrical driven fans.

• Direct Fan Drive SystemDirect driven fan systems use an old technology: the fan is either directly connected to the enginecrank or driven by a belt mechanism. As there is dependency on the engine mechanical energyoutput, the speed of the fan is proportional to the speed of the engine. Excessive fan noise isa problem, especially during the vehicle’s acceleration through the gears. In addition, due to

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Figure 2: Scheme of a turbocharger.Adapted from [39].

continuous fan rotation there is always energy consumption, even when it is not needed.

• Hydraulic Driven SystemThis type of system uses a fan with an electric motor and the coolant moves through the radiator bymeans of a hydraulic pump. Thus, this system is more complex because it requires the hydromotor,the hydraulic pump, control valves, oil, the coolant, connection hoses and hydraulic tanks. Thesecomponents have high power density and high efficiency when compared to electric driven fans, forsituations/vehicles that require high cooling rates.

• Electrical FansBy definition, electrical fan defines the fan blade and the electrical motor. Since it is electricallydriven, this system is easier to control. The fan is not driven by the engine ‘directly but by analternator.

2.2 RadiatorThe radiator is an aluminum mesh with a high number of fins all attached to the tubes that transportwater and oil. There are two types of heat exchangers: unidirectional and cross flow heat exchangers,being the latter the most common in buses. In this case, the two fluid streams flow at right angles and so,fluid temperatures vary in both the direction of the flow and at right angles to that direction. Within thecross flow heat exchangers, the flow can be mixed or unmixed according to the temperature distributionfrom one row to another. The size of the heat exchanger matrix assumes primordial importance forthe desired outcome of the cooling system to meet the specified heat transfer rate and pressure droprequirements. Bigger surface areas implies more drag and thus more cooling. The ideal radiator hasto be wide, tall, thin and have a large frontal surface area, since the space constraint is a permanentissue. Regarding the fin density, high fin density means an increase in the pressure drop due to thereduction in the air flow, which can be a problem when the vehicle is stopped and when this restricts theair flow, being a problem for the case of the hydraulic fan. Lower fin density has the advantage of beingeasier to clean. In the case of hydraulic fans, the coolant flow rate to the radiator is also important:the higher the flow rate, the higher the energy that can be removed from it. The temperature differencewill be greater since the coolant will be moving at high temperatures as it is traveling through. Theoverall energy removed from the total system is less in the case of lower flow rates, even though theoutlet temperature is lower. This is due to the fact that for higher flow rates, the temperature differencebetween the environment and the coolant is higher [41]. For air conditioning, there is another radiator,also called the air conditioner condenser, which also needs to be cooled. As long as the air conditioningis turned on, the fan keeps running, even when the engine is not at high temperatures.

13

Discharge

Impeller eye

ImpellerImpellerVolute

Figure 3: Scheme of a water pump. Adapted from [42].

2.3 Engine water pumpsIn order to make the coolant flow from the engine to the radiator, a water pump is used. This processoccurs by throttle heating, a mechanism by which the flow of a fluid is managed by obstruction orconstriction. It controls the engine power by restriction of inlet gases. In an internal combustion enginea throttle is a valve that regulates the amount of air entering the engine, indirectly controlling the chargeof air and fuel burned on each cycle. In a water pump, the majority of fluid is forced in a directionnormal to the rotating axis of the water pump shaft. Thus, the rotor or impeller blades tend to be 2dimensional and are more easily manufactured, unlike the mixed flow pumps where the blades take on atwist in the 3rd dimension [15]. As the coolant moves forwards, the diameter increases and so the areaincreases, leading to a consequent decrease of velocity and an increase of pressure.

At the outlet of the impeller there is a considerable velocity in the direction tangential of the outsidecircumference of the impeller that could be greater than the velocity required in the discharge section,which will lead to energy losses in the discharge section. To overcome this, the diameter of the casingincreases along it, to provide an initial reduction in velocity, creating static pressure head as any addi-tional velocity. In this way, kinetic energy is converted in static pressure head. The diffuser helps theprocess as well, although it is limited by packaging constraints [60].

• Water pump design

When designing a water pump or a fan, the concept of system characteristic is used and is the sum ofengine resistance, radiator resistance and thermostat resistance. The system characteristic is selected ata flow rate that corresponds to maximum engine power, being that the operating point. That correspondsto an operating condition that the water pump must match.

dP = KQ2 (6)

In the equation 6, k is the system resistance in Ns2.m8, dP the overall pressure drop in Nm−2 and Qthe system overall flow rate in m3s1. In a water pump, the power is proportional to the engine speed tothe power of 3 and the water flow rate is proportional to the engine speed. The water pump speed rangeis determined by package and speed limitations. It is driven by a belt from the crankshaft and has adrive ratio (crank to pump) between 1 and 1.3. The lower ratio is usually dictated by the package spaceavailable as this would render the maximum pulley diameter. The design of a water pump uses a toolcalled Cordier diagram. This diagram uses the specific speed and a dimensionless diameter related to

14

the outside diameter of the impeller, predicting its diameter for a given operating point. A common wayof presenting machinery performances as function of the similarity is through correlation of the specificdiameter, D0 (7) and the specific speed, Ns (8).

D0 = DsQ0,5

gdH0,25 (7)

Ns = ωQ0,5

gdH0,75 (8)

The number of blades is also an important parameter in the design of a water pump. High number ofblades implies frictional losses and blade blockage to flow, notwithstanding a better fluid conformity toblade direction. Normally five to ten blades is the most appropriate number. The eye diameter shouldminimize shock losses, that occur when the global direction of the coolant does not flow parallel to thedrive shaft and in doing so can create an effect of collision of separate flow streams. Regarding theimpeller inlet diameter, the higher the radial distance occupied by the blades, the more time the coolanthas to gain the required velocity and the less violence is associated with the fluid’s passage through theimpeller blades, leading to less cavitation problems and to a better control over the stress on the blades.The inlet diameter is normally about 60% of the outlet diameter.

2.4 Charge air coolerThe charge air cooler is a heat exchanger that uses the compressed hot air from the compressor of theturbo-compressor and cools it, since it leaves the compressor at temperatures near 200 ℃with low density,this means that for the same volume there are less air particles and thus, the quantity of injected fuelwill be less. So, decreasing the air temperature allows the introduction of more fuel that is used in amore efficient way since it has more space in the carburator. The temperature and pressure increaseconsequently the torque and the power of the engine. The charge air cooler works differently from theradiator since it only uses ambient air as a coolant. Despite this, there are also charge air coolers thatuse water as the coolant fluid. They are named air-to-water intercoolers and as they use water, theyrequired more components such a pump and a water reservoir. For the time being, air-to-air charge aircoolers are used in buses but some ideas about integrating an air-to-water with the radiator coolant arenow on the table. The air-to-air type is cheaper but less effective and air-to-water has pressure losses[8],[4]. Regarding the design of the charge air cooler the criteria used are: the fin density, height andthickness. In a similar way to the radiator, the fin density can not be too high since it will decrease thevelocity of the air and thus, lead to a bigger pressure drop. On the other hand, changing the height ofthe charge air cooler will not increase much the temperature difference between the ambient inlet airand the compressed air [44], [15].

2.5 Oil coolerThe oil cooler, used in the case of the hydraulic fan is positioned horizontally along the radiator’s loweredge or vertically along it. It is also an heat exchanger with oil as the coolant.

2.6 Fans shroudThe fan shroud is a plastic or metal funnel shaped piece on the back of the radiator (usually) thatcovers the fan(s) completely. It acts as a director for the air used to cool the radiator and to avoid airleakages between the fans and heat exchangers [56]. It is proved that excessive recirculation within thefan caused inefficient cooling but recent design improvements have tended to reduce recirculation andprovided better cooling to underhood vehicle compartments.

2.7 Thermostats and other sensorsThe purpose of the thermostat is

1. To help minimize the warm up time of the coolant;

15

2. To control the coolant temperature during normal engine running conditions once fully warm;

3. Allow the cooling pack to control coolant temperature when maximum cooling is required;It does so by converting thermal energy into mechanical work. It is composed by a thermal motorcontaining wax element, a valve plate, a return spring and the body. By changing the state of thewax material from a solid into liquid, there is a conversion of thermal energy into mechanical work [46].During warm phase the brass element is heated by the coolant and the heat is transmitted via conductionthrough the casing the and the wax. At a prescribed temperature the wax starts to melt and the resultingexpansion pressurizes the rubber sleeve, forcing the piston outwards. The piston then reacts against thethermostat body to open the valve. During the warm up phase, the engine thermostat will be closed andthe heater circuit in most cases will be open allowing coolant to flow through the engine and into theheater sub circuit but not through the radiator. When the thermostat opens and allows coolant to flowinto radiator, the heater sub-circuit opens as well as allowing coolant to return back to the water pumpinlet [60],[46]. Normally, the start-to-open temperature is 90 ℃[59]. A full piston travel of 9 millimetersis achieved at a temperature of 104 ℃, offering an active temperature range of 14 ℃. Working hand inhand with the thermostat, a relief valve not exposed to the coolant acts preventing that potential airpockets appear until the thermostat opens.

2.8 Header Tank and Recovery SystemA typical cooling system needs a header tank and a recovery system to get a simple and convenientcoolant filling point, allowing the expansion of the coolant without any losses. They have to be placedas high as necessary, since the coolant goes to the engine by hydrostatic pressure, due to the change indensity, as the temperature decreases. The header tank (or expansion tank) has a pressure cap on thetop. If the coolant boils and becomes vapor, which can cause serious damage since the heat transferby conduction decreases. With the pressure cap, it is possible to change the temperature inside theheader tank by changing the pressure inside [7],[60],[46]. As the coolant temperature rises, the coolantexpands and pressure rises,above atmospheric pressure. In a recovery system, hot coolant flows out intothe overflow container. As the engine cools, the coolant contracts and pressure in the radiator drops.Atmospheric pressure in the overflow container then opens a second valve, a vacuum vent valve andoverflow coolant flows back into the radiator. And that stops atmospheric pressure from collapsing thehouses of the cooling system [7],[46].

3 Hydraulic FansWhen the cooling system uses a hydraulic fan, the engine’s control unit receives information on theengine’s temperatures from the coolant sensor. The hydraulic pump capacity is controlled by a pulsemodulated signal (PWM) sent from the engine control unit to the solenoid valve on the hydraulic pumpcontrol valve, and the hydraulic pump’s oil pressure drives the wheel motor [15]. There is, in additiona rotational speed induction (or piston) sensor on the coolant fan, which reads the fan’s actual speedand returns this information to the engine control unit. When necessary, the output signal to thehydraulic pump is adjusted so that a correct rotational speed is maintained, in accordance with theengine thermostat and the remaining programmed parameters.

3.1 Coolant Temperature SensorThe coolant temperature sensor measures the temperature of the coolant of an internal combustionengine. The information measured by the sensor is then sent to the engine’s control unit, being usedafter to adjust the fuel injection and time of ignition. The temperature sensor works through changes onresistance, that either increases or decreases according to a positive or negative temperature coefficient,respectively [7], [59].

3.2 Hydraulic pumpThe hydraulic pump is mounted on the top of the engine or on the left side of the engine (rear). Thepump’s pressure is controlled by the control valve (solenoid valve), via the PWM signal from the engine

16

Figure 4: Scheme of hydraulic valves

control unit. This releases requested amounts of hydraulic oil which increases or decreases the controlplate angle. Thus, the pump piston has longer or shorter strokes, which increases or decreases the oilpressure [59].

3.3 Pump control plateWhen the motor starts, the movable units inside the hydraulic pump housing start rotating.

1. When the control plate is vertical, the oil remains stationary in the cylinder barrel. This meansthat no oil is forced away and thus, no displacement.

2. The angle of the control plate means that the piston can move forwards and backwards in thecylinder barrel when the rotating unit turns and so, the pumping starts.

3. The amount of oil can vary, depending on the control plate’s angle. The larger the angle, thegreater the displacement and higher the pump pressure.

3.4 Control valve and check valveThe control valve controls the upper valve. The longitudinal hole and the restriction on the check-valvemake possible to build up the system’s pressure. This pressure acts the pre-tensioned spring. Therestriction increases the system pressure in the control chamber. Depending on the electrical signal tothe solenoid valve, a control pressure is obtained, which moves the main valve to its right.

3.5 Hydraulic fan motorThe hydraulic motor is a gear wheel type motor and is supplied in two different versions, depending onthe bus specification. The oil pressure from the hydraulic pump is forced between the motor’s two gearwheels and starts the hydraulic motor. For most buses at Volvo Buses Corporation, idling speed can be1120 rpm with a maximum rotational speed of 3000 rpm , that is reached at an engine speed of 1200rpm. For other buses models, idling speed is equal to 750 rpm with a maximum rotational speed of 1950rpm. Only when the engine reaches 1200 rpm the fan speed is maintained above 1800 rpm. A rotationalspeed induction sensor is located by the coolant fan and continuously reads the fan’s rotational speed,that is then sent to the engine’s control unit. When necessary, the output signal to the hydraulic pumpis adjusted. The correct fan speed is therefore maintained with respect to the engine thermostat. Thereis also a hydraulic oil filter located inside the engine compartment and its purpose is to filter the oil onthe return side.

4 Electrical FansAn electric cooling fan system can reduce the fuel consumption and thus, the engine load compared toexisting cooling systems that use hydraulic fans. Thus, in order to know how much energy savings it

17

is therefore important to know how is calculated the fan power.The transfer of mechanical energy isusually accomplished by a rotating shaft, and thus mechanical work is often referred to as shaft work; afan receives shaft work (usually from an electric motor) and transfers it to the fluid as mechanical energy(less frictional losses). Laws of conservation must be taken into account when studying machines likefans and pumps.

• Conservation of massdm

dt= min − mout (9)

• Conservation of energydE

dt= dQ− dW + minein − mouteout (10)

where,dE = change of energydQ = rate of heat addeddW = rate of work donem = mass flowe = flow energy

Fans are a special case of fluid machines that consume power, like the compressors, but since theirvelocities are much lower it is assumed that the air flow is incompressible. Fans are used when thepressure differences are small enough that the density is considered constant . The work for the workinggas is transformed on the pressure energy and the kinetic energy A fan set consists of a rotor followedby a stator and the energy equation for a the fan is shown in equation 11.

dQ− dW = minein − mouteout (11)

dQ− dW = min(∆KineticEnergy + ∆InternalEnergy + ∆PotentialEnergy) (12)

Assuming potential energy neglected, the fan energy equation will be function of the kinetic and internalenergy, related with flow velocity and temperature differences. There will be no heat exchange in thefan. The flow in fans is so slow that is always possible to regard it as incompressible. It is however notpossible to say that the density is constant in the system equations, since it can vary with the operatingpoint and with the location of the system. The density will vary with the temperature so, in the caseof a bus cooling system, where two or more heat exchangers are present, the density change from aheadto after but the mass flow is the constant. If it is introduced pressure and assume the density constantthrough the fan unit, assuming air as an ideal gas,

h = u+ p/ρ (13)

that,

−W = (poutρ

+ uout + v2out

2 ) + (pinρ

+ uin + v2in

2 ) = (poutρ− pin

ρ) (14)

The Bernoulli equation is concerned with the conservation of kinetic, potential and flow energies of afluid stream.

P

ρ+ V 2

2 + gz = cte (15)

Therefore, the kinetic and potential energies of the fluid can be converted to flow energy, as a form ofpressure. This can be seen by multiplying the Bernoulli equation by the density, ρ.

p

+ρV 2

2 + ρgz = cte (16)

where p is the static pressure used to evaluate fan performance, the second term the dynamic pressureand ρgz is the hidrostatic pressure, neglected in this case. This leads to the fan energy equation,

−W = ∆pfanρ

+ ∆plossρ

(17)

18

The rotational energy from the engine is converted to electrical energy through an alternator to drivethe electric motor(s). The motor again converts this electrical energy to rotational energy and theninto flow energy. Some of this energy is useful but other output energy is wasted like the air swirl atthe fan’s outlet. Concluding, the two types of useful energy in a fan are the static pressure (considereda form of energy) and the kinetic energy (due to the air particles movements). The kinetic energy ismeasured as the air flow rate. The fan performance is measured at atmospheric pressure and is a resultof a combination of the two types of energy: kinetic energy and static pressure. The relative percentageof each type of energy will be dependent of the resistance of the system to be cooled by the fan [14]. Anelectrical fan system optimizes the operational efficiency due to the easiness of control the fan accordingto the engine load, regardless of the engine speed. Together with that, they are reliable and simple,even though they may not move as much air flow as a hydraulic fan. Hotter running engines havebetter thermal efficiency, which means that heat losses are reduced and intern more heat is used to makepower. Electrical fans allow to safely increase the operating temperature of the coolant in the engine,which means less heat to rid of in the cooling system, which decreases the system capacity. In the caseswhere the hydraulic fans move more air than needed, electrical fans are beneficial. Other possible reasonsfor the less power required to reach the same cooling performance are:

1. Cooling performance (LAT, IMTD)

2. Oil cooler deleted (Power dissipated, static pressure)

3. nfan at vehicle (diesel) idling (stop-and-go test)

4. Heat rejection of the engine

5. Could be exhaust recirculation gas (EGR) mixed into the cold CAC-flow

6. Fan shroud design

7. Built-in-resistance

8. Derate strategy

9. Sandwich vs. Separated (rad and CAC)

10. TTT difference between the models

Roughly, can be said that the dissipated power is proportional do the mass flow power to 0,7, for higherair flows. For these air flows, the heat has to be dissipated and pressure drop will pay for this dissipatedheat. This pressure drop is proportional to the mass flow to the power of 1,7. The fan can be modeledby its speed and the torque and the work done by the fan is the product between the fan speed and thetorque.

Pdiss by the cooling module[W ] ∝ m0,7 (18)

∆Pfan ∝ m1,7 (19)

The energy changes for fans are usually measured as pressure differences.

4.1 Radial, Axial fans and Diagonal fansFans can be radial or axial. For large volume flows and low pressure increase axial fans are preferred Ifthe total pressure increase is very low, the rotor looks like a propeller. Axial fans are also used when theradial when the available radial space is restricted.

Radial Fans Radial fans (or blowers) are the most common ones used ones. The blades can be arrangedstraight ones in the radial direction, which is said to make the fan useful for gases contaminated withparticles. They use a rotating impeller to move the air stream, increasing its velocity.The speed increasesas it reaches the ends of the blades and is then converted to an increase in pressure. These fans can workunder high temperatures and high pressurized conditions [23], [33].

19

CHAPTER 1. FANS

(a) Axial fan (b) Radial fan

Figure 1.1. Different types of fans

The fan in this thesis is a radial fan. The considered fan delivers the coolingairflow for an electric motor and is placed at the end of the motor and rotates withthe same speed as the rotor. This is called an open self-ventilated motor.

The air flows axially into the inlet and through the channels in the motor andis deflected within the fan in order to leave the motor radial through the outlets.In figure 1.2 the position of the fan is visualized at the back of the motor and theairflow is denoted by arrows.

Figure 1.2. Open self-ventilated electric motor

2

Figure 5: The effect of multiple fans on system pressure and flow rate. Adapted from [3].

Axial Fans Axial fans are normally used for cooling systems, since they allow large volume flows andlow pressure increase. If the total pressure increase is very low, the rotor looks like a propeller. Axialfans are also used when the available radial space is restricted [28]. The blades in axial fans are mostlyformed as wing profiles. To get a good efficiency, they should be skewed in the radial direction Theywork by moving an air stream along the axis of the fan. It can be compared to a propeller on an airplane:the fan blades generate an aerodynamic lift that pressurizes the air [20], [23],[24]. They are inexpensive,compact and light which makes them the best choice for the cooling systems. The highest total efficiencyis obtained with blades curved backward, even though they do not achieve an energy conversion as highas forward blades. However, this type of fan operates stably because the pressure difference providedby the fan drops if the flow rate goes up. If the opposite were true, an increased flow rate would causeincrease fan power, which would be unstable [11]. This gives them the design advantage that makesthem the best choice for the bus cooling systems.

Diagonal Fans Cooling fans for automobile industry have predominatly been axial fans since theirlow cost, thinness and ease of mounting. However, many tests show that due to the flow resistances infront and back of then fan are somehow strong (heat exchangers and front grille), the air flow directiontends to flow diagonally at the fan outlet side. An hybrid solution recently introduced in the market iscalled diagonal fans. These fans take an intermediate position between the axial and radial fans. Theirconfiguration allows an air-flow similar to the axial fans whilst at the same time, achieving a higherstatic pressure, overcoming the problem of the counter pressure of the radial fans [45]. These fans have aconical rotor hub and they suck the air axially. This conical hub avoids vortex formation which reducessignificantly the noise. The hub gas a small cross-section in the inlet area. In the outlet side, the diameterincreases gradually, which provides an higher circumferential speed of the blade tips, meaning a highercentrifugal air acceleration[19]. Regarding the operating point of the diagonal fans, in comparison withaxial fans, lies in a higher pressure range, which means that they can give more pressure with a significanthigher air flow.

4.2 How to select the right fanThe first step is to know the total cooling requirements of the system. This can be translated in theneeded air flow to dissipate heat. This air flow is calculated according to the power consumed by thesystem and the amount of air needed to remove heat from the system in order to do not increase itstemperature.

20

2.2. NUMERICAL SIMULATIONS

Figure 2.3: Trace of static pressure through an engine bay.

small contribution from ram air (1)), so to overcome the pressuredrop in the heat exchangers the fan is fully engaged and the majordriver of airflow (5).

The red curve in Figure 2.3 represents a light load cruise condi-tion. Here, there is more access to ram air, and there is no significantneed of cooling, hence the fan does not need to engage to any greatextent.

The model used as a base in this thesis was proposed by Dav-enport [8] in 1974. Recently, work has been done on this model byCowell [9].

This model has the form of Equation (2.1).

∆pF = ∆pR +∆pSys +1

2ρv2F − 1

2Fρv20 (2.1)

In this equation the notation in Table 2.1 apply.

Term Explanation∆pF Fan Pressure Rise∆pR Radiator or Cooling Package Pressure Drop∆pSys System Restriction12ρv

2F Fan Dynamic Head

12Fρv20 Ram Air Pressure, F is ram air effectiveness

Table 2.1: Local nomenclature to Equation (2.1).

13

Figure 6: Change of static pressurer

4.2.1 The total cooling requirements

It must be known the heat as a temperature difference that must be transferred, the heat transfer powerto offset the specified ∆T and the amount of air flow needed to remove the heat. The volume of airflow required can be thus calculated if one knows the internal heat dissipation and the total rise intemperature allowed.

Q = cP ∗ m ∗∆T (20)

where,∆ q = amount of heat transferredcP = specific heat of air∆T = temperature risem = air mass flow

4.2.2 Static pressure

The total pressure of a system is the sum of the static pressure and the dynamic or velocity pressure.The dynamic pressure is the pressure term of the Bernoulli equation associated with the velocity of theflow, and thus with the motion of the fluid.The static pressure is the pressure of the fluid if it were atrest. Thus it is independent of the air velocity but it is dependent of the air flow. Static pressure isalso dependent on the blade profile, number of blades, pitch, hub space and aerodynamic characteristicsof the fan impeller [17]. The force with which the air molecules move against the fan blades derivesfrom static pressure. In this way, it is an indication of how much restriction a fan can overcome; higherstatic pressure means that a fan can push through a more dense radiator. A good way to understandthis concept is to imagine using the fan to inflate a balloon; a fan with a higher static pressure wouldinflate the balloon to a larger size than one with a lower static pressure. The static pressure thereforecan indicate how strong a fan is, and how good it is at overcoming resistance. [20]. The concept ”systemresistance” is used when referring to the static pressure. The overall system resistance is the sum ofstatic pressure losses, being also dependent of configuration of ducts and houses. The system resistancevaries with the square volume of air flowing through the system.

4.2.3 Total System Resistance / System characteristic curve

To evaluate a specific fan with respect to its ability to transport a certain airflow, fan curves are used.The pressure loss due to the resistance of the components of the cooling system varies with air flow and

21

CHAPTER 1. FANS

shut-off operatingpoint

QFigure 1.5. Sketch of fan and system curve

0 0.5 1 1.50

500

1000

1500

2000

2500

3000

3500

4000

4500

Q in m3/s

dP in

Pa

Figure 1.6. System curve from ABB measurements

The system in our application is the set of channels and other conduits throughthe motor from the inflow from ‘free’ air to the outflow again to ‘free’ air. Thesystem curve is determined by the flow resistances in the flow paths and is given infigure 1.6.

The fan curve can be measured if a prototype is available and is therefore pro-vided by the fan manufacturer. The fan curve can also be computed by using CFDsoftware. This is done in this thesis.

The resulting intersection points are very useful to compare different fan designs

6

Figure 7: Sketch of fan and system curve. Adapted from [3]

is known as system resistance. It describes the amount of flow that will flow through the radiator andthe charge air cooler; as more air is pushed through the radiator/CAC, more pressure is required. Thesystem characteristic curve is given by:

Dp = KQn (21)where K is the system characteristic constant, Q is the air flow, n a turbulence factor (1<n<2, whereas1 is laminar and 2 turbulent flow.)

4.2.4 System Operating Point

A specific fan curve together with a specific system curve (or system impedance curve) of the cooledmotor yield an operating point. This operating point represents the airflow that is delivered by the fanto the system. This means that in this point the fan is operating at the static pressure of that point andgiving the respective flow, which is not the maximum flow rate. Thus, when selecting a fan is is not onlyimportant to look at the extreme values but to the system operating point[51].

4.2.5 Stall effect and instability regions

Stall on axial flow is a status of non uniform flow through the impeller of the fans that is translated in anunstable zone in the fan curves with a peak point as an indication of stall. During this condition, the flowis separated from the blade surface. Flow separation can be explain as the following: fan blades deflectthe air, depending on its orientation relative to the flow direction. So if one change the orientation of thefan blade relative to the flow direction, it is possible to increase or decrease the amount that the air isdeflected.That orientation is named angle of attack. If one increase gradually the angle of attack of thefan blade, it will increase the amount of air deflection. The angle of attack is responsible to the pressuredifferences between the upper and lower parts of the blades that allows the fan to lift and thus to move[12]. But if the angle of attack becomes too high, the air will no longer follow the blade surface in anuniform manner. The amount of deflection and the pressure difference being generated stops increasingand normally will fall off. This is called the stall point. In a fan, the blades are normally rotating atconstant velocity. Therefore, to change the angle of attack, the system to which the fan is attached mustbe changed. Higher flow rates through the inlet increase the attack angle, and lower flow rates decreaseit. Therefore, if a fan is operating in stall, it is because the air flow rate is too low for it. On a givensystem, this is caused by selecting a fan which is too large (making the air velocities too low in the fan)[22],[54],[58]. [14] [1] [2] [9] [61] [18]

4.2.6 Efficiency of electrical fans

The fan efficiency is the ratio between power transferred to the airflow and the power used by the fanshaft [32].

ηairflowfanblade = PoutPin

(22)

22

BSTALL

REGION

A

C

D0

AIR FLOW

STAT

IC PR

ESSU

RE

RECOMMENDED

SELECTION RANGEUnstable region

STALL REGION

AIR FLOW

STAT

IC P

RESS

URE

A

BC

D

Stable regionȠ=100%

Rotational noise Non-rotational noise EF

FIC

IEN

CY

Figure 8: Stall: unstable zone in the fan curve

ηairflowfanblade =∆pfan ∗ Q

ρfan

Pshaft=

∆pfan ∗ Qρfan

Pelec∗ PelecPshaft

(23)

Part V

Modeling and analyzing electrical fans5 What is limiting electrical fans5.1 Fan blade typesThe blades in axial fans are mostly formed as wing profiles. A fan wheel must impart to the air streaman uniform velocity and pressure over its entire area. To get a good efficiency the blades must be skewedin the radial direction, which is said to make the fan useful when the air is contaminated with particles.The fan is simple to manufacturer but will operate with low efficiency. The highest efficiency is obtainedwith blades curved backward. To get the impeller simpler to manufacture without loosing efficiency thebackward swept blades are normally constructed as flat plates. The most advanced fan blades are madewith a thickness distribution on the blade.

5.2 Fan LawsIf a fan curve has been measured for a specific speed, diameter and density, these laws can theoreticallybe used to calculate the fan curve for any other speed, diameter or density [26]. In order to betterunderstand how does the fans work and compare between several configurations, the fan laws are used.However, since these fans are electrical, the only two parameters that are allowed to scale for are thepressure and flow rate as a function of fan rotation rate. Efficiency and power consumption cannot beaccurately scaled since they are a function of the electrical motor efficiency and fan blade efficiency[55].

System characteristic The static pressure is the difference between the absolute pressure in a pointof air stream and the absolute pressure of ambient temperature. The total pressure in a point of a fan’s

23

airflow is the algebraic sum of the total pressure in that point and the velocity pressure. The velocitypressure is the pressure required to accelerate the air and is proportional to the kinetic energy. The fantotal pressure is then the difference between the total pressure measured at the outlet and at the inletof the fan. When choosing a fan for a cooling system, the fan static pressure is the parameter used andcannot be measured directly. It is the fan total pressure minus the dynamic pressure corresponding tothe mean air velocity at the fan outlet [49]. The system resistance varies with the square of the air flow.The affinity laws or fan laws establish that the volumetric air flow (AV) is proportional to the fan speed(N), the static pressure (∆p) is proportional to the square fan speed and the power (P) is proportionalto the speed to power 3, if considered a constant impeller diameter (24). Since fan power consumptionis highly sensitive to fan speed, significant energy savings are achieved if the fan can serve the system ata low speed. However, below about 40% of the motor load, fan efficiency begins to decline [50].

AV2

AV1= N2

N1

∆p2

∆p1= N2

2N2

1

P2

P1= N3

2N3

1(24)

5.3 Influence of densityIf the air temperature changes, then the density will be altered.Although the density change due to theflow is small. If a change in air density is considered, keeping the same fan model and the same air flow,the pressure and the fan power will be influenced by the density.

∆p2

∆p1= ρ2ρ1

P2

P1= ρ2ρ1 (25)

For the flow speed to be unchanged in a fan, the same air flow has to be considered. If the velocitiesare the same so the pressure increase over density is unchanged. When doing the combination of fans, itmust be observed that it is the mass flow that is conserved and the pressure increase has to be matched.For example, two fans in series do have the same mass flow but not necessarily the same volume flow,which is used in the fan diagram. It is also assumed that the pressure level in the heat exchangers willnot be exactly the same for the several fans but it is assumed that such difference will not have anymajor impact in the density.

5.4 Impact of Fan DiameterBy increasing the diameter of the cooling fan, the same amount of air can be generated at a lowervelocity. The area, A varies with the square of the diameter and so the velocity. The energy varies withthe square of the velocity, v.

A ∝ D2; v ∝ D2; hp ∝ v2 (26)

hp2 = hp1D1

D2

4(27)

where,hp1 - power draw of the existing fanhp2 - power draw of the desired fanD1 - diameter of the existing fanD2 - diameter of the desired fanIn the same way, flow rate and static pressure will be influenced by a change in diameter.

AV2

AV1=(D2

D1

)3 ∆p2

∆p1=(D2

D1

)2P2

P1=(D2

D1

)5(28)

A 10% larger fan will use less 32% hp to move the same amount of air. Thus, for instances, if one wantsto reduce the fans diameter by 1/2, but still keeping the same air flow, following the fans laws:

Q ∝ N3 (29)

where Q is the air flow, N is the speed and D is diameter.

N1D31 = N2D

32 and D2 = 0.5D1 ⇒ N2 = 8N1 (30)

The fans with half of the diameter will have to spin 8 times faster to provide the same air flow [10].

24

0

200

400

600

800

1000

1200

1400

0 0.5 1 1.5 2 2.5 3 3.5

Stat

ic p

ress

ure

[kP

a]

Air flow [m3/s]

Fan curves with diameter change from 305 mm to 450 mm

Figure 9: Effect of an increase in fan diameter in air flow and static pressure

5.5 Series and Parallel OperationTwo fans mounted in parallel will produce double the free flow of one, whereas in series the shut offpressure increases to twice that of a single device. At intermediate flow points, the flow and pressureincreases, and whether fans should be in series or parallel depends upon the degree of system resistance,i.e. placing two in series, in a low resistance system would produce a negligible increase in flow andpressure. In general, for higher resistances’ systems, two or several fans in series will produce a higherflow rate than in parallel, which means that this mode is advantageous when there is a high impedanceof the system and there is a wish to increase the airflow. The opposite can be seen in lower resistancesystems where the parallel fans are more beneficial [48]. Parallel fan configurations may also be a safetyrequirement in case of a single fan failure.

In the situation of zero static pressure, the maximum air flow flows in free conditions, without anyobjects around. This free condition is explained by the x-axis (maximum kinetic energy). On the otherhand, the y-axis represents the maximum resistance possible to not have air flow, which means thatare objects in front and to the rear of the fan that prevent air circulation (maximum potential energy).This situation can happen in reality but should be avoid since it can damage the fan. The advantages ofmultiple-fans arrangements include lower average pressures (especially in the case of series arrangements),lower noise generation due to lower average system pressure and higher energy savings. Even thoughlarger electrical motors are more efficient than smaller ones, if the fans operate near their best operatingpoint, higher-speed fans close to their best efficiency point can often achieve a net efficiency advantageover a single, low speed fan [33]. The drawbacks of placing electrical fans in parallel remain on the factthat in extreme cases, one fan might force another to operate far away from its best operating point.If one keeps the fan diameter and the speed constant and using SPAL data for one fan, it is possibleto predict the behavior of 5 fans, placed in parallel. According to the fan laws, it can be seen that themaximum air flow increases with an increase of the number of fans (see figure 12). As can be seen in figure13, if one increases the fan diameter from 305 mm to 450 mm, and instead of 5, only 2 fans are used, themaximum volumetric air flow possible in zero pressure conditions is increased about 20%. Nonetheless,this maximum point in the graph has no meaning in real conditions since the system resistance can notbe zero. Besides this, according to the fan laws regarding a change in diameter, the input power for thefan is also a variable. If one increases the diameter, the power needed to move the fan will increase tothe power 5.

P2

P1=(D2

D1

)5(31)

So for the case of a SPAL fan of 305 mm, assuming an approximately constant current of 16 A, for themaximum fan speed of 3750 rpm, with a constant voltage of 26 V, the power input should be the productof the current and voltage, being 416 W per fan (a total of would need 2kW to work). According to

25

[48]

Figure 10: The effect of multiple fans on system pressure and flow rate

Figure 11: Lower duct pressure due to fans placed in series

26

0

100

200

300

400

500

600

700

0 1 2 3 4 5 6

Stat

ic P

ress

ure

[kP

a]

Air flow [m3/s]

Different fan curves in parallel for a maximum fan speed of 3750 rpm, 305 mm diameter

1 fan 2 fans 3 fans 4 fans 5 fans

Figure 12: 5 fan curves in a parallel configuration, with a diameter of 305 mm and 3750 rpm speed

0

200

400

600

800

1000

1200

1400

0 1 2 3 4 5 6 7

Stat

ic p

ress

ure

[kP

a]

Air flow [m3/s]

Two fans curves installed in parallel (3750 rpm and 450 mm diameter)

Figure 13: 2 fan curves in a parallel configuration, with a diameter of 450 mm and 3750 rpm speed

the equation 31, if one increases the diameter 32% (from 305 mm to 450 mm), it will result in 1864 W,which is almost the same value (see figure 14). Figure 14 represents different fan configurations for thesame system. The fan laws were used to scale one fan curve (with data given by the supplier) in otherpossible fan curves, changing the diameter and/or the fan speed. The hydraulic system is representedas well, for a fan speed of 2600 rpm, correspondent to a full displacement and a change in its velocity ismade in order to have the same operating point as the one of the 5 electrical fans. To reach the samepoint, the fan speed should be 2000 rpm, a value still lower than the fan speed of each smaller fan (3750rpm). At this point, however, the power required to drive the fans differ largely. In a hydraulic system,about more than 10 kW is needed to have the same cooling performance. In real conditions, the fanswork most of the time under much lower speeds: around 800 rpm for the hydraulic fan and 600 rpmfor the electrical fans. If we consider a 800 rpm fan speed and compare the power consumption, it willresult in about 640 W pump shaft power for the hydraulic system and 80 W for the electrical fans, adifference of 8 times more power consumed by the hydraulic system, together with the disadvantage ofthe dependency between engine fan and hydraulic fan speed. If the engine is stopped, and still needs

27

0.0

200.0

400.0

600.0

800.0

1000.0

1200.0

1400.0

1600.0

1800.0

2000.0

0.00 1.00 2.00 3.00 4.00 5.00 6.00 7.00

Stat

ic P

ress

ure

[kP

a]

Air flow [m3/s]

Different fan curves and operating points for the system resistance of a B9TL

System Resistance

1 fan 450 mm, 3750 rpm

5 fans, 305 mm, 3750 rpm

1 fan, 450 mm, 2590 rpm

2 fans 450 mm, 2590 rpm

1 fan, 305 mm, 3750 rpm

Hydraulic Fan, 560 mm, 3200rpm

Hydraulic fan, 560 mm, 2000rpm

Figure 14: Different fan curves for the same system resistance

more air flow, hydraulic fans can not fulfill the cooling demand.

5.5.1 5 SPAL electrical fans (305 mm) vs 2 SPAL electrical fans (405 mm)

If the same blade shape is considered, going from a 5 electrical fans to 2 larger fans, will required amaximum fan speed of 3100 rpm for these fans. However, the existing 405 mm diameter SPAL fan, doesnot have the same behavior as a scaled 405 mm diameter SPAL fan. Two existing SPAL 405 mm fanswill imply 3850 rpm to meet the same operating point of 5 fans. However, the maximum fan speed ofsuch fans is 2450 rpm, which means that this solution will not work the way the fans are designed. Thesolution could be ask for a powerful motor for these fans or to wait for a better design.

5.6 Blade angleThe impeller blade outlet angle, B0 is also an important parameter and should vary between20 ◦ and40 ◦. As more as the angle is increased, the higher is the pressure difference. Previous studies haveshown a roughly linear dependency(figure16 between this angle and the pressure difference, which meansa better fan performance. The reason behind this is that an increased angle means more surface tothe blade that is ”attacking” the air, being more air accelerated and thus transported through the fan,yielding to a higher pressure difference [3]. For larger blade angles the coolant is more directed towardsa radial direction and will increase the dynamics head component. This can be converted to the desiredstatic head if there is room to include a diffuser section at the outlet of the pump. If no diffuser can beaccommodated, there is no need in using a large outlet blade angle as this will only increase flow losses.

5.7 Distance between fan blades and fan ringWith increasing distance between the fan blades and ring the pressure drop decreases linearly. Due tospace constraints the fan ring is placed as close to the fan as possible.

5.8 Voltage imposed and consequences in fan curves characteristicsThe specified voltage range for a bus is typical between 24 and 29 V. This range is nowadays fixed andso, the fans have to be chosen accordingly. These values can vary with outdoor temperature about 3V.

28

0.0  

200.0  

400.0  

600.0  

800.0  

1000.0  

1200.0  

1400.0  

1600.0  

1800.0  

2000.0  

0.00   1.00   2.00   3.00   4.00   5.00   6.00  

Sta.

c  Pressure  [P

a]  

Air  flow  [m3/s]  

Different  fan  curves  and  opera.ng  points  for  the  system  resistance  of  a  B5LH  

System  Resistance  Curve  

5  fans,  305  mm,  3750  rpm  

2  fans,  405  mm,  3100  rpm  (calculated)  

2  fans,  405  mm,  3850  rpm  (scaled)  

Figure 15: A comparison between 2 real SPAL fans versus 5 small SPAL fansCHAPTER 4. FAN PERFORMANCE

However, the achieved increase in pressure difference is only interesting in terms ofthe tendency. The actual gain is not that much, only approximately 150Pa over thewhole angle range.

−30 −20 −10 0 10 20 305500

5550

5600

5650

Angle in degrees

dP in

Pa

rpm = 4500; Q = 1.05 m3/s

data from simulationslinear approximation

Figure 4.7. Dependency between pressure difference and blade end angle

4.3.5 Radius for rounding of ring edge

In figure 4.8 the pressure difference for different diameters of the ring edge is shown.The tendency is visualized with a linear fitting.

60

Figure 16: Dependency between pressure drop and blade angle. Adapted from [3].

29

Volvo Buses VBC, 86113, C Kjellgren, General training - Cooling system

8 2010-12-09

Cooling module

Fan shroud

Fan ring

Fan blade

Figure 17: Fan ring location in the cooling module. Adapted from [3].

CHAPTER 4. FAN PERFORMANCE

0 1 2 3 4 5 65000

5200

5400

5600

5800

6000

6200

6400

6600

6800

7000

distance in mm

dP in

Pa

rpm = 4500; Q = 1.05 m3/s

data from simulationslinear approximation

Figure 4.5. Dependency between pressure difference and blade-ring distance

The pressure losses come to a certain extend also from the increase of k inthe space between blade and ring, as described in section 3.4.2. So as smaller thedistance can be made, the better both important properties of the fan, namely theperformance and the turbulent behavior, will become.

4.3.3 Outer blade radius

The effect, how different radii of the blade wheel affect noise and turbulence isalready discussed in section 3.4.3. Now the influence of the diameter on the per-formance will be discussed. In figure 4.6 the change of pressure difference can beseen.

58

Figure 18: Dependency between pressure drop and blade-ring distance. Adapted from [3].

30

[48]

Figure 19: Effect of voltage in fan performance

Everything is parallel connected to that voltage, with a constant frequency. To control the fans, thecurrent changes and this change in current is managed by the vehicle control unit by PWM signal. Thissignal consists on a series of pulses, with different widths and constant frequency. The speed control isperformed by the fan based on the control signal, which is a square function operating at a frequencyof 25 kHz; the duty cycle determines the fan speed. As the fans work in parallel and the availabletechnology does not allow to control each one independently, they have to be at the same speed.

The fan sped vary proportionally to the voltage, affecting the static pressure and consequently theair flow. According to the fans’ laws, the static pressure varies based on the square of the fan speed andthe airflow is proportional to the fan speed. Thus, as can be seen in the picture 19, if one varies thevoltage 10%, this will vary the maximum static pressure between -19 and +21%, and the maximum airflow 10% [40].Hybrid and full electrical buses use one big alternator with 600 V, together with a DC/DC converterfor equipment that uses 24 V such as fans and lighting. The possibility of, in the future, to use highervoltage fans could be interesting since the current drawn by the fans would be smaller. According toVolvo experts, 600 V fans are as efficient as 24 V fans, so current savings will not by itself be a benefit.Besides that, safety reasons are a drawback for 600 V components. They require special requirementsnamely water protection and can not be managed as a 24 V system, which imply higher investment andmaintenance costs.

5.9 Possible fan suppliersEBMpapst and SPAL are the companies with whom Volvo collaborates. EBMpapst has almost the sameperformance as SPAL fans [55] but they are less resistant to high temperatures; SPAL fans motors startderating at 90 ℃. Recently, EBMpapst is developing diagonal fans and S-force fans, that are knownto have higher efficiency in terms of energy conversion. Grayson and multi-wing could be interestingto look at.Grayson is known by selling the complete cooling system[25], Electrical Fan Engineering is amanufacturer of high performance thermal management components. Cooling products include electricand hydraulic fan assemblies, oil-to-air heat exchangers, charged air coolers and radiators. COMAIRROTRON is another company recognized as a premier global provider of air-moving and thermal solutionswith leading fan technology in the EUA [13]. Tonada is another company specialized in brushless DCcondenser fans and evaporator blower for bus air-conditioner [53].

5.10 Pusher Fans vs. Puller FansIn the history of Volvo buses the fans have always been placed behind the heat exchangers, which couldbe called that the fans are placed in a pulling position. The reason for this is that he fans have been belt

31

driven fan or hydraulic driven which makes it very complicated to go through the heat exchangers withthe belt or oil pipes. It has also been a common solution shared by GTT. But for a cooling perspectiveit could be interesting to have the fans on the cooler side where the air has higher density giving a higherair mass flow for a certain fan speed. When Volvo buses is more and more looking towards electricalfans it must be considered to have pushing fans since the motor in a pulling position is exposed to hotair heated by the radiator and charge air cooler. Electrical fans can also be easily mounted in front ofthe heat exchangers in a pushing position. One drawback is that the total cooling package gets thickerdue to that the fans cannot be placed too close to the grille for safety demands and also that the fanshroud needs an extra frame to be mounted against the charge air cooler. One question rises throughhow the fan shroud should be designed for a pushing fan set from an aerodynamic point of view, whichmeans that its optimum size might not be the same as for a pulling fan. In general, pulling air throughthe radiator works better than pushing it through it. Since fans manufactures clam the same efficiencyfor both configurations, especially if the fan shroud is well designed. Otherwise, it can be said thatpusher fans would be better. As there is always some leakage between the fan and the radiator (orcharge air cooler), the fan does not just flow air through itself straight, since it spins causing turbulence.Centrifugal force throws air outward all along the fan as well, but the intake side is pretty much limitedto the area of the fan. When the fan is in front of the remain cooling system, some air goes thrown outand never makes it through the radiator at all. Thus, if one compare total air moved with a pusher fan,less makes it through the radiator than the same fan as a puller. The shroud effect is important whenaddressing this issue, especially regarding pusher fans.

Other simulations comparing these two configurations were made to analyze the air flow velocityand it was concluded that there is not a big difference between them. This test was made in ambienttemperature, which means that the density change, consequence of change in temperature, was nottook into account. The mass flow is constant and will be the same wherever the fan is located. Thedensity change is given by the temperature change in the hear exchangers. The pressure level in theheat exchanger will not be exactly the same for the two fan locations, but it is assumed the differencewill not cause any change in density, as it is a fan and thus not a pressure difference in the system thatcauses density change. Normally the temperature behind the heat exchangers is around 90 ℃, whichentails to a correspondent 0,972 kg/mˆ3, compared to an air temperature of 20 ℃with a density of 1,204kg/mˆ3. This increase of 24% in density could be translated in 24% increase in mass air flow. This meansaccording to 20, that the heat dissipated by the cooling system could be improved the same percentage,if losses are neglected. Therefore, if the fans operate under this high temperature their lifetime is reducedand the conclusion is that in order to get as low power required as possible the fan shall be placed wherethe density is higher, i.e. at the cold side of the heat exchanger [27].

Fan noiseThe emitted noise by the cooling fan is affected by the three parameters: speed, size and fan load [58].There are two main mechanisms that generate the overall fan noise: rotational and non-rotational noise.Rotational noise is generated by the forces acting in the blades, and is provoked by the interaction ofdistorted flow with the blades, unsteady flow, turbulence and the presence of nearby obstacles[16], [58].On the other hand, non-rotational noise is not related to the fan rotational speed; it is dependent onthe relative velocity of the air flow on the blades. It is caused by several factors being the stall themost important. Stall phenomenon (explained in section 4.2.5) is responsible for an increase in the noiselevel. In that zone, between B and C (see figure 8) , the noise level assumes its higher value and thus,the fan should work to the right of that zone. It is claimed that noise reductions of up to 5 dB can beachieved with resulting efficiency improvements of 15% [37]. The sound power, W is a measure of thesound energy per time unit. It is measured in watts and can be computed as the sound intensity timesthe area. The sound power level Lw, 32, represents the difference between two sound powers and can beexpress in (dB), relative to a reference sound power, W0. In air W0 takes the value of 10−12 watt, thatcorresponds to 0 dB.

Lw = 10 log10(P1/P0) (32)According to [16], an axial fan sound power level can be described as

Lq(dB) = 10 log(kwD2u5

W0) = 10 log(u) + 10 log(D2)2︸ ︷︷ ︸

u˙1

+10 log( u5

W0) (33)

32

Figure 20: Sound level for different electrical fans

where u1 isu1 = k + 20 log(D) (34)

Which means that at a certain operation point increases as the square of the diameter and the fifthpower of the (equation 37).

W ∝ D2u5 (35)

or increases as the seventh power of the diameter and the fifth power of the fan speed.

W ∝ D7N5π5 (36)

sinceutip = πND (37)

Since it is not accurate to define the absolute sound level in theory, due to many variables and thedifferences in the shape of the fans, to compare two fan systems is more interesting to define a way tocalculate the sound difference, assuming the fan blades have the same shape.

As can be seen in figure 20, the difference in sound level is explained by the difference in shapebetween the fans. The four fans presented in the figure have the same diameters, so the difference is onlydue to the blades shape factor, H. In this case, the worst fan in terms of noise is the EBM fan, whichblades have a simple plane geometry[52], [21]. Thus, in the case that one wants to compare differentfans in terms of blade shape regarding noise level, it is better to use test values and take the differencein decibels of such difference.

Lwfan2 − Lwfan1 = H + 20 log(D2

D1

)+ 50 log

(u2

u1

)+ 10 log

(nrfans2

nrfans1

)(38)

Equation 38 gives how much the sound power level increases or decreases if a change in fan diameter,fan tip speed or the number of fans changes between two systems. This equation can also be written asa function of the fan speed:

Lwfan2 − Lwfan1 = H + 70 log(D2

D1

)+ 50 log(N2

N1) + 10 log

(nrfans2

nrfans1

)(39)

33

Study cases6 Evaluation of an electrical cooling system: London buses case:

B5LH (hybrid) and B9TL6.1 Exhaust gas recirculation (EGR)The exhaust gas recirculation is a nitrogen oxide emissions reduction technique that works by sendinga portion of the exhaust gases back to the engine cylinders. If it is a gasoline engine, the inert exhaustdisplaces the amount of air fuel mixture in the cylinder and if it is a diesel engine the exhaust gasesreplace some of the excess oxygen in the pre-mixture, reducing the oxidations of the nitrogen oxidesthat required high temperatures to occur. This leads to lower combustion efficiency so it representsa trade-off between engine performance and emissions. since by this method the temperature of thecombustion chamber is reduced. They are also important regarding the control of emissions, especiallyNOx. Exhaust gases that recirculate through the CAC have been used for long time as a measure ofcontrolling NOx emissions, but off late as the emission legislations are becoming increasingly stringent,a significant quantity of these exhaust gases are used to reduce NOx. Generally when exhaust gas isre-circulated back to the engine, it is cooled. The exhaust gases increase the specific heat capacity of thecylinder contents, which lowers the adiabatic flame temperature. Because of the significant quantities ofexhaust gases involved, the increase in heat rejection requirements due to its cooling has been estimatedto be 20%.

6.2 Real Data taken from Volvo Data Base (LVD)LVD is an on-board monitoring system that saves constantly information in the control unit of the vehiclethat later, when it goes to service, is collected in a huge database. This data base collects informationregarding the engine, electrical system, power transmission, brakes and axles, steering and the hydraulicsystem. Concerning the engine load, information about temperatures in different locations within theengine are provided, as well as the power consumption, engine speed and fuel consumption. LVD allowsto choose a model and several variants such as the type of the fan existent in the cooling system or thefuel consumption, in order to filter the results to a specific case study, i.e. a specific bus model in acertain environment. The bus model B9TL was choose to see what happens for instances to the coolanttemperature and fan speed when it changes from hydraulic fans to electrical fans.

It was observed that the ambient temperature and the average engine speed were similar for the twosystems, which leads to a fair comparison. (Due to confidentially reasons, the information can not bedisplayed.)

On the other hand, the coolant temperature distribution shows unexpected results on a first sightsince it should be expected to have the same coolant temperature. This is due to the different idlingspeed requirements in the control unit. Electrical fans have a lower idling speed than hydraulic fansso since they operate 90% (figure ??) of the time at this minimum speed, the vehicles equipped withelectrical fans will have higher coolant temperature going to the radiator, correspondent to the full openthermostat situation, whereas the hydraulic fans that operate at higher idling speed will lead to lower”unnecessary” coolant temperature. In the case of the B9LT, the thermostat opens linearly with anincrease in temperature.

6.3 LAT and IMTDTwo important parameters to validate a cooling system are the Limiting Air Temperature (LAT) and theIntake Manifold Temperature Difference (IMTD). The LAT is the maximum atmospheric temperatureat which the engine can be operated safely and without failure [36], [57].

LAT = (TTT − Tengine outlet coolant) + Tamb (40)

where, TTT is the maximum allowed coolant temperature and Tamb is the outside ambient temperature.The origin of this formula is the Newton’s law of cooling, or convection law. Assuming that the heat

34

transferred from the coolant to the air is constant, i.e. the temperature difference between the coolantand air is constant at a given moment, this leads to the equation 41.

h(Trad in−amb = h(TTT − LAT ) (41)

This equation states that the heat transfer for a certain coolant temperature, when entering in theradiator is cooled by a given ambient air mass temperature. This temperature difference is the same asthe limit situation, where the coolant would reach its maximum allowed temperature (TTT) and thusthe ambient air temperature would reach its limit value, known as (LAT). Solving for LAT,

LAT = (TTT − Trad in) + Tamb (42)

If the LAT is too low, the engine may overheat during peak ambient temperature conditions. To overdesign by selecting an excessively high LAT can cause significant power, fuel and performance penalties.In general, the higher the LAT, the lower the temperature difference between ambient air and hot airand so, more air flow is needed to cool the system. The cooling goal of the CAC is expressed as theamount of heat transfer required, which means the charge air outlet temperature desired. The IMTD away to express this desired outlet temperature and it is the difference in temperature between ambientair and CAC air and EGR mixture in the inlet manifold. The intercooler will never cool below ambientair temperature because that is what it uses to cool itself, and in turn, cool the air charge. IMTD isone way of representation of how much fuel can be saved. The theory behind this is the following: adecrease in air temperature results in an increase in air density. Since the air density is higher, for thesame volume of space, one can push in a higher mass of air. In other words, there is more oxygen in thecylinder for combustion to get a higher power output. In simple terms the overall volumetric efficiency ofthe engine increases. And also it will lower the combustion temperature leading to lower levels of NOx.[43].

IMTD = TCAC outlet − Tamb (43)

where TCAC outlet is the temperature at the charge air cooler outlet. Engineering tests performed in thebus model B9TL, the typical red London bus concluded that the LAT is increased around 10 ℃. For themaximum fan speed of 5000 rpm some important parameters were simulated for the B9TL with electricalfans for different vehicle average speeds.

Can be observed that the power required to drive the electrical fans do not vary with the engine horsepower, even though the LAT decreases a few degrees.

6.4 DeratingPower output of an ICE is proportional to mass of air and quantity of fuel burnt during combustion.When ambient temperature increases, the density of air and fuel are reduced and hence, the mass. Dueto this increase in ambient temperature, the power output reduces by 0,45% per degree increase [60].Derating is the operation of an engine at less then its rated power in order to prolong its life. Thisimplies a reduction of current, since the voltage is constant.

35

Part VI

Simulations and Tests7 City-buses7.1 MethodologyThe first Volvo bus hybrid project to be launched in 2014 for series production will be the case study forthe simulations for the city-buses. This hybrid bus includes a 120 kW electrical machine parallel withthe diesel engine and a 600 V traction battery. Besides that battery, there is another electrical systemwith 24 V to other lower voltage components such the electrical fans, lighting, etc. The diesel engine hasa maximum power of 215 hp and a maximum torque of 820 Nm. The electrical engine has a maximumpower of 120 kW and a maximum torque of 800 Nm. A ”bullet list” was thought to be the basis of thesimulations and intents to ask this thesis question: why has been accepted a system that consumes more10kW, ((when the fully engaged), than a system with electrical fans?

• Cooling performance (LAT, IMTD)

• Oil cooler deleted (Power dissipated, static pressure)

• nfan at vehicle (diesel) idling (stop-and-go test)

• Heat rejection of the engine

• EGR-gases mixed into the cold CAC-flow

• Fan shroud design

• Built-in-resistance

• Derate strategy

• Sandwich vs. Separated (rad and CAC)

• TTT difference between the models

This list of reasons that would justify the difference in power consumed and consequent difference incooling performance will be analyzed in AMESim.

7.2 AMESimAMESim (Advanced Modeling Environment for performing Simulations of engineering systems) is a 1Dsimulation software for analyzing multi-domain system for example to perform simulations related toautomotive, aviation, aerospac and energy industries. In this thesis, the vehicle thermal managementanalysis has been performed and hence all the libraries associated to thermal management have beenused to carry out the analysis. AMESim is a virtual tool to design, analyze, and optimize the fluidflow characteristic of a system, complementing the physical tests and CFD calculations [43]. All thephysical elements in AMESim have been designed specifically for modeling the effects of heat transferin both steady state and transient conditions but for the purpose of this thesis work, only steady statewill be considered. All the elements in AMESim are representations of several nonlinear time-dependentanalytical equations that represent the hydraulic, pneumatic, thermal, electric or mechanical systembehavior [30]. These elements are then connected to form a circuit, by linking the inputs of one icon tothe outputs of other icon, which in turn represents an actual system, for example, the cooling systemcircuit of a vehicle [43],[30].

36

7.3 Input DataData from several reports regarding individual components were used for the simulation. For steadystate simulations, the model has been run at a constant speed and loading condition. The model hasbeen run at full load condition, and all the parameters have been evaluated in this particular situation.City-buses use a 788x980x56 mm radiator, (RAD LOW type), whose geometry was used in the model,as well as the charge air cooler. (Due to confidentially reasons, the information can not be displayed.)

7.4 ImplementationAll the necessary input data relevant for all the components like the pressure drop data, heat load data,engine performance, fan performance, charge air flow, coolant flow and so on, were collected from severalengineering reports and engine specification files. These data were modified according to the input formatof the AMESim elements with proper units. As far as possible, the latest test data have been used, butfor certain components, the latest data were not available.

Cooling performance: LAT and IMTD Changing LAT demand +4 ℃and -4 ℃, changing theIMTD demand +2 ℃and -2 ℃, changing from steady state full load test cycle to stop-and-go cycle(same LAT demand and ignoring IMTD).

Effect of removing the oil cooler If the oil cooler disappears from the cooling system, an increaseof 14% cooling area is achieved. The objective of this step is to evaluate the consequent pressure dropand the possible decrease in air temperature and the impact of this in the fan power. CFD with andwithout oil cooler to see the effect on air temperature into radiator, air flow and air speed distributionover the radiator surface.

Fan speed at vehicle stand-still Study what happens in the stop-and-go cycle if the fan speed needsto go down at bus stop.

Effect on the cooling performance of a separated CAC and radiator compared with sandwichconfiguration In terms of cooling performance, removing the oil cooler decreases the IMTD in 1,55℃and reduced the recirculation air temperature in about 0,5 ℃.

7.5 Built-in-resistanceThe built-in-resistance is the net restriction offered for fresh air passing through the cooling circuit.To evaluate the differences in the built-in-resistance of the system for electrical fans and an hydraulicfan, AMESim is used, through a simple model with a radiator, a CAC, a oil cooler (when applicable)and a set of fans. This model receives as input the air flow, coolant flow and fan characteristics. Allthe components are defined for a set of properties, either mechanical, thermodynamic and chemical.The simulations use data from a hybrid double-decker, that operates in London, for the situations ofmaximum power and maximum engine torque, with electrical fans. To calibrate the model with thereal values, the built-in-resistance is adjusted in order to have similar values of air flow and coolant flowfor real and simulated values. The adjust in the built-in-resistance is done by changing the dynamicpressure, given by equation 44 in order to get similar test and real values.

pD = 12 ρ v

2, (44)

, where pD is the dynamic pressure (Pa), ρ the fluid density in (kg/mˆ3) and v is the fluid velocity in(m/s). The way used to change the dynamic pressure and thus the built-in-resistance is by using thedefinition of pressure loss coefficient, ζ = f(Re).

ζ = ∆pρV 2

2(45)

Together with the built-in-resistance, the calibration of the model includes also the coolant flow andair flow through the radiator and CAC. At a first analysis, it was thought to build two models, one

37

Figu

re21

:5

elec

tric

alfa

nsm

odel

:H

eat

stac

k(r

adia

tor

and

CA

C)

conn

ecte

dto

5el

ectr

ical

fans

38

Figure 22: Heat module: CAC, oil cooler and a radiator in the backside

for the hydraulic-fan driven cooling system and another one for the electrical fan cooling system. Thiswould imply to use the same bus model (same engine and cooling performance requirements) in order tonormalize both models, which was left behind since the normalization was incompatible with the enginedata. The same bus B5LH use different engines when it goes from Euro 5 to Euro 6 policy, which wouldlead to larger errors if one proceed with the normalization. Thus, just an electrical fan model will beused for this study and changes in its configuration will be discussed further. Even though, just usinga model with electrical fans will lead to about 10% error. The hydraulic fan model that would be usedas a comparison should include a hydraulic fan with 560 mm diameter that, when at the maximumdisplacement works at 3200 rpm and consumes around 15 kW. This fan is modeled according to itspump and motor efficiency, being that data the input as well as the fan curve (static pressure versusairflow). The electrical fan model uses the fan curve for one fan, in this case with 305 mm diameter andworking at its maximum power, 3750 rpm, consuming 400 W each, which means 2 kW. After that, airflow and coolant temperatures are checked in several points of the radiator and CAC, as well as air flowand coolant flow. in order to calibrate the model. The calibration of the model is made to have the sameinput and output values as the reports or tests made and the engine specifications.

7.6 Effect of removing the oil coolerCFD and AMESim simulations were made to see what happens if the oil cooler is removed 23. CFDresults were made to verify and calibrate the AMESim model. CFD model used 38 ℃as ambient temper-ature and it was assumed that the vehicle was running at 30 km/h, which means that there is ram air.0.4 ℃due to air recirculation between the outside air and the CAC was taken into account. A constantfan speed of 3400 rpm was used to find the differences between two cases:

Case a): 5 electrical fans with a maximum fan speed of 3400 rpm, without oil cooler.

Case b): everything the same as case a), but with the oil cooler (pressure drop and 6 kW heat). Theradiator dissipates 100 kW and the charge air cooler 28 kW of heat. The coolant flow used was 3.34 kg/swith an average temperature of 90 ℃. The charge air flow (from the turbocompressor) to the CAC was0.2 kg/s with a temperature of 195 ℃. The oil cooler reduces by 10% the air flow pulled by the fans.It was assumed a 20% loss between the fans and the radiator, leading to the difference seen in tables?? and ??. The oil cooler increases by 6 ℃the air temperature that passes through the radiator and

39

Figure 23: CFD calculations: case a) without oil cooler and case b) with oil cooler

12 ℃the air temperature at the radiator outlet. The effect on air flow of changing the cooling systemconfiguration was calculated for three situations: a reference case, without grill and without oil cooler.(Due to confidentially reasons, the information can not be displayed.) The AMESim model had to becalibrated together with the reference values taken from the engine specifications data and the CFDresults, in order to find the required fan speed for the two cases. A reference case was used setting acertain fan speed, for the 5 electrical fans in a situation without oil cooler. If the oil cooler is mountedin the cooling system the required fan speed to keep the same cooling performance, (i.e. to keep thesame coolant temperature at the outlet of the radiator and the desired charge air out from the CAC)increases 8.8% in fan speed and, according to the fan laws, a 640 W fan shaft power increase. (Due toconfidentially reasons, the information can not be displayed.)

7.7 Cooling performance: LAT and IMTDAs mentioned before, the cooling performance is dependent on LAT and IMTD. To evaluate the coolingperformance a change in LAT or IMTD is performed in order to see how does it affect the fans electricalpower consumption and also, to see how much it cost in the required fan speed. Reminding the definitionsof LAT and IMTD, (see equations 40 and 43) the LAT is related to the radiator cooling performancewhereas the IMTD related to the CAC performance. For the IMTD, a proportional-integral-derivativecontroller (PID controller) was added to the model, after the CAC output in order to input the requiredtemperature difference in the IMTD and thus to get the result consequence in the fan speed as an outputin a faster way. In the case of LAT, it varies with fan speed as a negative second degree polynomialwhereas the IMTD varies as a positive second degree polynomial, which confirms that the larger theLAT, the more fan speed is required.

7.8 Separated CAC and radiator installationIf the radiator is placed somewhere else in the bus and not in a sandwich configuration, the result interms of fan speed is considerable. This situation was simulated in the model, assuming the same BiRfor both cases, with 5 electrical fans. The separated configuration was made putting the radiator side byside with the CAC, by changing the coordinates of the heat exchangers and assuming the same externalair flow for both cases, where each heat exchanger was fed by 5 electrical fans. The reference situation

40

0  500  1000  1500  2000  2500  3000  3500  

00  

10  

20  

30  

40  

50  

2706   2906   3106   3306   3506   3706   3906   4106   4306  

Fan  Po

wer  [W

]  

LAT  [°C]  

Fan  Speed  [rpm]  

Limi7ng  air  temperature  vs  fan  speed  and  fan  power  [°C]  

LAT  [°C]   Fan  power  [W]  

Figure 24: LAT change versus fan speed

0  

2000  

4000  

6000  

8000  

00  

05  

10  

15  

20  

25  

30  

2011   2511   3011   3511   4011   4511   5011   5511   6011  

Fan  Po

wer  [W

]  

IMTD

 [°C]  

Fan  Speed  [rpm]  

Intake  manifold  temperature  difference  vs  fan  speed  and  fan  power  

IMTD  [°C]   Fan  power  [W]  

Figure 25: IMTD versus fan speed

41

Schematic picture of the two fan concepts

P(el,out)

Q = 0.25 - 2.5 kg/s T = 90 °C P = Atm

BIR/radiator (porous media)

Q = 0.25 - 2.5 kg/s T = 90 °C P = Atm

P(hyd,out)

Electrical fans + fan shroud

Hydraulic fan + fan shroud

Figure 26: CFD model with fansAnalysis of the fan shrouds only

P(el,out)

Q = 0.25 - 2.5 kg/s T = 90 °C P = Atm

BIR/radiator (porous media)

Q = 0.25 - 2.5 kg/s T = 90 °C P = Atm

P(hyd,out)

Figure 27: CFD model without fans

was using a fan speed of 3600 rpm which was reduced to 2740 rpm in the case of the radiator and 2150rpm in the case of the CAC, to meet the same cooling performance. This represents a significant 20% difference. As this system represents a new resistance (system resistance), depending on where theheat exchangers are placed in the bus, it is hard to say the optimized number of fans and their requiredspecifications. For instances, if the radiator is placed on the roof of the bus, the ram air will have animportant role and thus, less extra air is needed. Most likely, a lower number of fans would be used forthis configuration.

7.9 Effect of the fan shroudThe goal of this section is to compare the fan shroud resistance for a electrical fan installation with ahydraulic fan, using CFD simulations. The reason that lead to this study is because the fan shroudhas a big influence on the average air velocity[5] and consequently in the fan performance [32],[6]. Forthis analysis, the fan shroud of the B5TL bus model was used for the case of the electrical fans andthe fan shroud of the B9TL bus model was used for the case of hydraulic fan. The pressure drops ofboth systems are tuned so that the flow with the electrical fans reaches 1.85 kg/s when the fan speedis 3400 rpm. To simplify the simulation, two channels (inlet and outlet) are used with sufficient length[32] and the air is assumed to be at 90 ℃, since the fans are pulling air through the heat exchangers[5]. 10 steps calculations for mass air flow varying from 0.25 to 2.5 kg/s were performed to tune the fanspeed for both the electrical fans and the hydraulic fan. System curves were also calculated to know howthe flow behaves with an increase of pressure drop. Another complementary simulation was performed,but without any fans and setting the air mass flow at the inlet starting again on 0.25 to 2.5 kg/s. Asystem resistance curve was again calculated. (Due to confidentially reasons, the information can not bedisplayed.) The pressure drop is measured between heat exchangers (CAC and radiator) inlet and fanshroud outlet(s). The pressure drop is also mass averaged. The test made with the hydraulic fan shroudand with the fans placed, have shown that hydraulic fan needs to create a much higher static pressure toget the same airflow. So in that way, electrical fans are advantageous and this explain why they consumeless energy, explained by relation between a change in static pressure and the power consumption, by

42

using the fan laws (equation 46).

AV2

AV1= N2

N1

∆p2

∆p1= N2

2N2

1

P2

P1= N3

2N3

1

P2

P1= ∆p2

∆p1

N2

N1(46)

43

8 Coaches cooling performance using electrical fansBy the moment, only one test was performed using electrical fans in coach buses. This test compared 5cooling systems: one with a big hydraulic fan, one smaller hydraulic fan together with two electrical EMPfans (with and without engine derating), 8 EMP electrical fans in pulling position and 8 EMP electricalfans in pushing position. The cooling performance was evaluated by a weighted LAT for different enginespeeds corresponding to maximum engine torque, maximum engine power and two more points evenlydistributed. The conclusion of that report was that the LAT demand of 40 ℃is only achieved for theStop-and-Go test, which is not enough to fulfill the required cooling performance. The full load testhave showed that 8 electrical fans, at the maximum speed of 5000 rpm only lead to a LAT of 44 ℃(Dueto confidentially reasons, the information can not be displayed.). This simulation intents to providepossible solutions and find several operating points for different fan configurations. To achieve so, 8EMP fans mounted in parallel was the reference case which is completed with the SPAL fans, alreadyused for city-buses. EBMpapst fans are also analyzed for three fan configurations, since a request wasmade to that fan supplier to obtain the performance curves and thus find more possible operating points(figure 28). However, those results will not be totally included in this thesis due to lack of time. It is

0 200 400 600 800 10000

100

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600

700

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1000EBMpapst Diameter 350 mm

Radiator width [mm]

Rad

iato

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ght [

mm

]

COMBINATION 2

0 200 400 600 800 10000

100

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EBMpapst Diameter 350 mm

Radiator width [mm]

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COMBINATION 3

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EBMpapst Diameter 500 mm

Radiator width [mm]

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[mm

]

COMBINATION 6

Figure 28: Different EBMpapst fan combinations

important no not that by the time the report was made, back in 2012, EMP fans were the most powerfulfans available in the market although compromising the noise levels. Nowadays, more modern designsachieve high performance fans with significantly lower noise, so it is interesting to explore if there is apossible solution available in the market that makes possible the use of electrical fans in coaches. Theair flow required for coaches is almost double of the one for city-buses (Due to confidentially reasons, theinformation can not be displayed.). In communication with EBMpapst, it was concluded that to meethigh air flow for coaches, in the order of 7 kg/s (due to confidentially reasons, more information can notbe displayed.), diagonal or radial fans could be the step forward electrical fans. With the axial fan usedfor city-buses, it is not possible to meet an acceptable cooling performance with so low flow rates.

8.1 Cooling performance: LAT and IMTDCooling performance is evaluated by LAT and IMTD for a certain demand. As a basis for the calculationof the cooling performance, measured temperatures from a test report made were used, for four differentengine speeds and consequent fan speeds: A, B, C and D. A represents the engine speed for the maximumengine torque and minimum engine power, D represents the engine speed for minimum engine torqueand maximum engine power and B and C are two evenly distributed points between A and D. In thepresent report test, LAT decreases with the engine speed so it is calculated by equation 47.

LATcoolingperformance,w = 0.4LATD + 0.2(LATA + LATB + LATC) (47)

[35] [47]The specific heat across the radiator is given by equation 48.

qspec = Radiator Heat Rejection

∆T (48)

44

00  

01  

02  

03  

04  

05  

06  

07  

08  

2.5   3.5   4.5   5.5   6.5   7.5   8.5   9.5   10.5   11.5   12.5  

Specific  He

at    [kW

/K]  

Air  Flow  [kg/s]  

Radiator  Wide  Cooling  Performance:  Specific  Heat  vs.  Air  Flow  

Coolant  flow  4kg/s  

Coolant  flow  6kg/s  

Coolant  flow  7kg/s  

Coolant  flow  8kg/s  

Coolant  flow  9kg/s  

Coolant  flow  10kg/s  

Coolant  flow  5kg/s  

Figure 29: Radiator performance curves for different coolant flows

where ∆T is,∆T = Tcoolant inlet − Tambient air inlet (49)

With the specific heat, it is possible to know the required air flow for the radiator that along with the fancurves, give the operating points represented in figure 30. In communication with researchers working atVolvo Trucks it was concluded that to get a system resistance curve that is closer to a real life situation,the theoretical curve obtained with these calculations must be corrected to take into account non-uniformflow an other effects. The value of this correction must be, at least a factor of 2. This topic is furtherdiscussed in [26].

Part VII

ConclusionsIn the start of this thesis the author had a limited knowledge about the Volvo Buses itself and the coolingsystem, which means that a lot of time was spent in order to explore and gather the information neededto start understanding how to reach the targets of the thesis. The purpose was to explore the existentbackground for the configuration and functionality of the buses cooling system, to identify the challengesand drivers for change to a more electrified and efficient cooling systems, meaning lower fuel consumptionand less emissions. The research was based on scientific articles, books, Volvo internal material, Volvodata base, LVD, previous master thesis and PhD. thesis, results of CFD simulations and mainly meetingsto Volvo employees. It was concluded that the electrical fans are beneficial in environments where theaverage ambient temperature is around 15 ℃. For the case of London, the hydraulic and electrical fansidle about 90% of the time which means that the cooling system is over-sized and its design leads toenergy losses. The selection of the fans is limited since the suppliers selected by Volvo do not presentas many options as desirable. An expressive case is the coaches cooling system that could take benefitfrom electrical fans but due to a poor offer in the marker for larger fans. The majority of the supplierscontacted just have fans with an average maximum diameter of 350 mm, which is the size being usedby Volvo, with a combination of 5 fans. The option of less fans with bigger diameters could lead toenergy savings, as well as monetary savings regarding the fans’ electrical motors. The only test done forcoaches with electrical fans, used 8 EMP fans which by that time, did not have the best design leadingto poor performance in terms of noise and cooling performance. Therefore, that test should be repeated

45

-­‐500  

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0   1   2   3   4   5   6   7   8   9   10  

Sta$

c  pressure  [P

a]  

Air  flow  [m3/s]  

Hydraulic  fan  versus  8x  EMP  fans  -­‐  Opera$ng  points  and  calculated  system  resistance  curve  

8x  EMP  fans   Hydraulic  fan   Opera@ng  points  8x  EMP  Fan   Opera@ng  Points  Hydraulic  Fan   System  resistance  curve  

Figure 30: Different fan curves for coaches and operating points

with different fan models. SPAL and EBMpapst fans could work better in terms of noise level. Pulland pushing air configurations results have a 20% error since do not take into consideration air densitydifference between the two configurations. Fan noise is affected by stall and fan design should take thisinto account when choosing the right fan for the cooling system. Nevertheless, the formula reachedduring this thesis regarding noise levels makes explicit that the number of fans does not affect the noiselevel as much as its speed. The contribution of the fan speed to the noise level difference between thetwo systems is 5 times greater than that of the number of fans. The remaining listed parameters tocompare the two cooling systems, like the separated CAC and radiator, brings considerable differencesin terms of required fan speed: about 20% less to meet the same cooling performance. The oil cooler,needed for an hydraulic system reduces by 10% the air flow, increasing the air temperature by 6 ℃.This represents an increase an increase of 8.8% of the fan speed and 640 W power increase. In termsof cooling performance, removing the oil cooler reduces the IMTD by 1.55 ℃and the recirculation airtemperature by 0.5 ℃. Concerning LAT and IMTD, a relation between both concepts and fan speed wasreached: a second degree polynomials, negative for LAT and poisitve for IMTD. It was also discoveredthat hydraulic fan needs to create a much higher static pressure to get the same air flow, when comparedto the static pressure correspondent to the electrical fans, explaining the difference in power consumption.Even though the fan shroud is not the same for the two systems, as it can be seen by the test withoutfans, it seems that they do not affect the static pressure required and the air flow. Finally, the analysisof the coaches was not succeed due to the lack of tests and reports. Nevertheless, some ideas for differentfans configurations were took into consideration, maybe for future work. EBMpapst was contacted withdifferent fans configurations and three of them (figure 28) were analyzed by their engineers. Togetherwith that, coaches cooling performance and 8 EMP fans operating points were analyzed leading to theconclusion that those fan curves does not meet the LAT demands and thus, alternative solutions areimportant to implement electrical fans in coaches. The general conclusion of this work is that electricalfans have a lot of advantages and lead to energy savings since they are more efficient and easy to control,even though their implementation is dependent on the available solutions in the market. These energysavings reflect on the fuel consumption, noise reduction and will lead to the satisfaction of the customer.

46

Part VIII

Recommendations and future workPossible voltage The main electrical bus of the future will be 48 V, and it will be buffered by a 36 Vbattery. As many devices and electronic control units require voltages of 24 V, the conversion from the42 V bus to these other voltages will be necessary. This could pass by electro-mechanical engine valves,which will demand both conversion and sophisticated control at power levels in the 2 kW to 10 kW range[31].

Bigger electrical fans The possibility of using bigger electrical fans with powerful engines should beinvestigated further since money can be saved and higher air flow with consequent lower noise levels canbe achieved. SPAL and EBMpapst are working in bigger fans but it will take some time to see themin the market. Another interesting option could be working together with GTT and find a common fanblade design that could suit both parts.

Radial fans Radial fans could be an interesting option to be explored for the coaches since they aremore efficient at higher air flows for lower static pressures, when compared with axial fans. These fanscan work under high temperatures and high pressurized conditions [23], [33] than axial fans which is anadvantage when the working air is at high temperatures.

Auxiliary fan installation For environments with harsh climate conditions such as Singapore andsouth Europe, an additional electrical fan placed in order to increase the ram air could be interesting toinvestigate [8].

Test diagonal fans and new S-force EBMpapst fans In communication with EBMpapst, it wasconcluded that to meet high air flow for coaches, in the order of 7 kg/s, diagonal or radial fans could bethe step forward electrical fans.

Fans pushing the air If this configuration is used, more possibilities of fans emerge since for instancesEBMpapst fans start to derate at low temperatures. If the fan is place in such a position, it will operateat considerable lower ambient air temperatures.

Build the theoretical background to find fan operating point Special for the coaches usingelectrical fans, the system resistance curve was one of the main difficulties to pursue the study. Thisshould be modeled for each bus in order to choose the best fan. This would lead to a time and costreductions related to physical tests, which is the example of the test done with the 8 EMP fans.

More advanced control options for the fan The discussed option of joining the control fan systemwith GPS in order to predict the cooling needs according to the topography. By doing this, the controlof the coolant temperature would be easier.

Increase the contact with fan suppliers Volvo Buses should be more aware of the new developmentsby fan suppliers. EBMpapst has been investing in new fan designs and improving the noise levels andgeneral efficiency. They should be aware of the willing to try fans with bigger diameters.

47

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