water cooling thesis f41
DESCRIPTION
fsaeTRANSCRIPT
DESIGN AND OPTIMZATION OF A FORMULA SAE COOLING
SYSTEM
Neal Persaud
A thesis submitted in partial fulfillment of the
requirements for the degree of
BACHELOR OF APPLIED SCIENCE
Supervisor Professor M. Bussmann.
Department of Mechanical and Industrial Engineering University Of Toronto
March 2007
ABSTRACT This thesis documents the testing, design and analysis performed to determine the optimal design for the 2007 cooling system for the University of Toronto Formula SAE race car. The main focus of this project is the physical testing and analysis of the data collected. The cooling system design has been refined as a result of the testing carried out in this project. The test data has been analyzed to identify heat rejection requirements, optimal engine operating temperature, and other important design parameters. It has been found that the cooling system used in the race car must reject 9500 Watts of heat energy, and should aim to maintain an operating temperature of 85°C. The 2007 cooling system promises to be successful and an improvement over last years system. The new system is 10% lighter than the system it replaces, contributing to the overall improvement of vehicle performance of the 2007 University of Toronto FSAE race car.
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Acknowledgements The author wishes to acknowledge the following people for their much appreciated help and support during the course of this thesis:
o Professor M. Bussmann for graciously agreeing to supervise my thesis and the for his role as the University of Toronto Formula SAE Faculty Advisor.
o Long Dana Manufacturing for providing radiators and continued support of the University of Toronto FSAE team. In particular I would like to thank Martine Banville, Stephanie Sesitito, Nick Kalman, and David Bruce from Long Dana manufacturing for assisting in the radiator procurement as well as providing proprietary testing data that proved to be invaluable.
o Jeremy Koudelka and Jerry Zielinski for assisting me during my first year as cooling system design leader
o Nilufar Damji and James Correia for their help in conducting performing cooling tests and test component manufacture
o Huang Iu and Saquib Siddiqui for their help in developing, troubleshooting, maintaining, and operating the engine dyno
o Andrew Wong for his excellent photography skills
o To past and present members of the Formula SAE team who have helped me to develop into a competent young engineer. The knowledge and practical experience gained by being a part of this team has profoundly affected my life and will continue to do so for many years to come.
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Table of Contents
ABSTRACT .. ……………………………………………………I ACKNOWLEDGEMENT .. ……………………………………II TABLE OF CONTENTS…...…………………………………III GLOSSARY ................................................................................... V 1 INTRODUCTION……………………………………………. 1 1.1 Purpose.................................................................................................... 1 1.2 Formula SAE.......................................................................................... 1 1.2.1 Background........................................................................................ 1
1.2.2 Formula SAE Rules .......................................................................... 3 1.3 Literature Review.................................................................................. 3 1.4 Motivation............................................................................................... 4 1.4.1 History .............................................................................................. 4 2 OBJECTIVES .............................................................................. 6
2.1 Design Objectives................................................................................... 6 2.1.1 Design Constraints............................................................................ 7 2.2 Design Requirements ............................................................................ 7 2.2.1 Benchmarking................................................................................... 8 2.2.2 2005 Cooling System vs. 2006 Cooling System .......................... 10 2.3 UT2007 Initiatives................................................................................. 12 3 Testing And Analysis................................................................. 14 3.1 Track Testing ........................................................................................ 14 3.2 Flow Rate Testing ................................................................................. 19 3.3 Heat Rejection Requirements ............................................................. 21 3.4 Optimal Engine Operating Temperature.......................................... 23 3.5 Radiator Configuration Testing .......................................................... 26 4 Component Selection ................................................................. 30 4.1 Component Selection.............................................................................. 30 4.2 Water Pump Selection............................................................................ 30 4.3 Radiator Selection................................................................................... 31 4.4 Cooling Fan Selection............................................................................. 33 4.5 Thermostat vs. Flow Restrictors........................................................... 35 4.6 Temperature Regulation........................................................................ 36 4.7 Weight Reduction ................................................................................... 36
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5 Component Manufacture.......................................................... 38 5.1 Radiator ................................................................................................... 38 5.2 Cooling Fan ............................................................................................. 38 5.3 Cooling Duct and Shroud Assembly .................................................... 38 5.4 Mechanical Water Pump....................................................................... 39 5.5 Thermostat and Thermostat Housing.................................................. 39 5.6 Swirl Pot................................................................................................... 39 5.7 Hardlines.................................................................................................. 40 5.8 Hoses and Hose Clamps......................................................................... 40 6 Recommendations..................................................................... 41 7 Conclusions ................................................................................. 42 References ...................................................................................... 43 Appendix A Weight Comparison Matrix ...................................... 44 Appendix B Flow Rate vs. RPM Data ………………………….46 Appendix C Water Temperature vs. Torque Plot……………….47 Appendix D Cooling System Design Data……………………...48
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Glossary AFR: Acronym for Air-Fuel ratio. A lean air fuel ratio indicates excess air and results in high combustion temperature. A rich air fuel ratio is desired for maximum engine power output
CC: Acronym for cubic centimeter.
CBR 600 F4i: The model of engine used in the University of Toronto FSAE Race Car.
Dynamometer: A device that is used to measure torque.
ECU: Acronym for Engine Control Unit. Responsible for controlling many electronic functions such as cooling fan relay, spark timing, fuel injector timing, etc.
RPM: Acronym for revolutions per minute, the most common units for measuring engine speed.
SAE: Acronym for Society Of Automotive engineers, the governing body of the Formula SAE competition.
Thermistor: A resistor whose resistance changes with temperature. Used for logging water temperatures.
UT XXXX: Referring to University of Toronto FSAE vehicle, year XXXX (e.g. UT 2005)
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1 Introduction
1.1 Purpose
The purpose of this report is to perform extensive testing and research to establish a set of
guidelines for designing and optimizing a cooling system for a Formula SAE race vehicle.
The design objectives of the cooling system will be identified, and through rigorous testing
that has been completed during the course of this thesis, the fundamental requirements will
be presented. Using the testing data compiled, an in-depth analysis will be conducted to
satisfy all of the design objectives. Justification of design decisions and system components
selection will also be presented.
The scope of this analysis will be to establish a reference for cooling system designs in
future UT Formula SAE vehicles.
1.2 Formula SAE
1.2.1 Background
The Formula SAE competition is a student design competition that is open to
undergraduate and graduate students around the world. The students are to conceive,
design, and manufacture an open wheel formula style vehicle that is suited for a weekend
autocross-racing car. Each team represents a fictional manufacturing company that is to
produce these vehicles at a rate of four vehicles per day, while maintaining a cost below
$25,000 USD. The target audience is the recreational autocrosser who wishes to participate
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in non-professional autocross events. At the competition, the formula cars are judged in
static and dynamic events, with the following point breakdown:
Table 1.1 Formula SAE Competition Points Breakdown
Static Events Max Points Available
Presentation 75
Engineering Design 150
Cost Analysis 100
Dynamic Events 75
Dynamic Events
Acceleration 75
Skid-Pad 50
Autocross 150
Fuel Economy 50
Endurance 350
Total: 1000 Points
It is evident that the majority of the points are contained within the dynamic events, and
thus a larger proportion of time is spent on refining the dynamic performance of the
vehicle. Given that the Endurance event accounts for 35% of the points, much effort is
spent on ensuring reliability of the vehicle as a whole to withstand the tremendous loads
placed on the vehicle during this event.
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1.2.2 Formula SAE Rules
The SAE sanctioning body has numerous rules that are strictly enforced due to the nature
of the competition; since these formula cars are designed and raced by the students
themselves. Among the many rules in the Formula SAE Rulebook, the most significant
ones are:
o A four stroke piston engine(s) that has a maximum displacement of 610 CC per
cycle.
o The use of a single circular restrictor placed between the throttle and the engine.
The maximum restrictor diameter size for a gasoline engine is 20mm.
o A minimum wheelbase of at least 60 inches.
o The vehicle must have four wheels that are not in a straight line.
Rules pertaining specifically to the cooling system state that the cooling system must use
water as the working fluid. Rust inhibitor is permitted, to a maximum of 10% of the total
system volume.
1.3 Literature Review
Literary resources on the design of a cooling system to be used in a race application are
mediocre at best. Valkenburg8 and Smith9 both cover the cooling system in famous race car
engineering texts but do not provide any sort of theoretical analysis for determining any of
the system operating condition, or system parameters. They both emphasize the importance
of the cooling system and other engine support systems and give a general overview of the
systems desired function. These two texts are helpful when it comes to sharing industry
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experience of what can be easily overlooked in the system design, construction, and
maintenance. The lack of theoretical literature regarding the subject of a race car cooling
system is one of the reasons that have inspired me to undertake this project.
1.4 Motivation
1.4.1 History
Since the inception of the Formula SAE team at the University of Toronto in 1997 until
2004, all of the UT cars suffered from inadequately designed cooling systems. This was
due to the fact that the cooling system was overlooked as a major subsection, and usually
put together as an afterthought. Consequently, it did not get the attention required to ensure
that the cooling system was able to perform its desired function. All of the UT FSAE cars
prior to 2004 experienced overheating problems on hot, dry, sunny days. In some cases the
overheating was mild, in other cases it resulted in catastrophic engine component failures.
For UT 2004, Technical Director Jeremy Koudelka did some testing and analysis to design
the cooling system. Working with manufacturer test data, he was able to size an appropriate
radiator, fan, and water pump combination to adequately meet the cooling requirements.
This system was highly effective but could easily be improved upon as it was very
overbuilt.
For the UT 2005 and UT 2006 race cars, I was responsible for the cooling system design
and manufacture. While completing the analysis to select the cooling system components,
it was noted that the system input variable, the amount of heat rejection required, had been
estimated for our engine (CBR 600 F4i.) Consequently, it was necessary to either validate
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or refine the existing heat rejection requirement estimate. The current estimated maximum
heat rejection requirement of the motor is 10200 Watts, and the entire cooling load is
assumed to be carried solely by the cooling system, i.e. cooling via radiation and
conduction is neglected. To refine this estimate a test was performed to experimentally
identify the heat rejection requirement.
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2 Objectives
2.1 Design Objectives
In order to optimize the system, the design parameters must be defined. It is important to
remember that the design parameters are similar to that of a production car, however their
order of importance is different, given the context of its racing application. The design
parameters listed in descending order of importance are as follows:
o Maintain an optimal engine operating temperature over a wide range of ambient
conditions.
o Maximize the overall system reliability.
o Meet all of the design criterions while minimizing the system mass.
o Minimize mechanical power requirements.
o Minimize electrical power requirements.
o Meet the packaging constraints of the chassis designer.
o Ensure the system does not negatively impact other sub-sections.
o Keep the system mass as central and as close to the ground as practical.
o Ensure the system can be serviced without undue difficulty.
o Minimize the cost of the system.
2.1.1 Design Constraints
The major design constraints in descending order of importance are:
o Timeline - Completed on time to meet running car deadline
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o Packaging - Fit neatly within the chassis and minimize system weight
without comprising system performance or adjacent sub-system
performance.
o Cost – Minimize cost, and maximize system performance and reliability.
In general, the timeline and packaging are the most restrictive design constraints. Our team
philosophy is to complete the car early and get it tested on track. This limits the amount of
time for development before the system must be finalized and construction must begin.
Packaging constraints represent the physical constraint in terms of overall component sizes
and location. This primarily affects the size of the radiator and fan combination, as well as
its placement on the vehicle. The final constraint is cost. Most of the components are either
donated to the FSAE team or sold at a reduced price, so the cost does not have a great
impact on the design. It does, however, require the relationships built between the team and
sponsors to be maintained in a professional manner.
2.2 Design Requirements
The system design cannot begin until the system requirements are identified. Once
identified, design and testing can begin to converge towards the optimal cooling system
design. The data plots needed to aid system design are described as follows in descending
order of importance:
o Compute the average heat rejection needed for the engine during simulated race
use.
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o Create a characteristic curve for the stock cooling system, a 3D graph containing
pressure drop vs. rpm vs. flow rate.
o Analyze the relationship between water temperature, AFR ratio, fuel economy, and
engine output.
o Compute water side pressure drop for double and single pass radiators.
o Compute heat rejection values for double vs. single pass radiators.
o Analyze the effect of added air side (heat exchanger) velocity to evaluate air side
limitation of FSAE race vehicles.
o Compute engine output losses associated with a mechanical pump.
o Compute engine output losses associated with electrical loads.
2.2.1 Benchmarking
Given the limited time, money, and testing facility resources, benchmarking older systems
is a useful tool to compare system performance and identify areas where the most gains can
be achieved. It allows for the system designer to focus their attention in the appropriate
areas, and allows them to make empirical relations between system parameters and system
performance.
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Refer to the Table 2.1 below for discussion:
Table 2.1 Cooling System Comparison 2003-2006
Vehicle Finned
Surface Area
Airflow
Rate
Water flow Rate Designed Max
Temp (30°C air)
Actual Max Temp
UT 2003 0.075 m2 3.2 m/s 18L/min 90°C 115°C
UT 2004 0.075 m2 4.2 m/s 18L/min 90°C 88°C
UT 2005 0.07245 m2 2.6 m/s 30 L/min (avg) 90°C 115°C
UT 2006 0.07245 m2 4.7 m/s 30 L/min (avg) 90°C 85°C
Table 2.1 serves to illustrate a number of points. The most significant point to comment on
is the apparent cyclical nature of the cooling system performance. It can be seen that the
performance of the system has fluctuated over the past four years. This can be directly
attributed to the system designer having insufficient background information and reference
material when starting the design. The intent of this paper is to serve as that reference
material to ensure the system converges to an optimal design. One should note that the
2003 and 2004 vehicles used the same radiator and water pump, but had different fan and
duct designs. Similarly the 2005 and 2006 vehicles also used the same radiator and water
pump, and also had different fan and duct designs. This illustrates the air side limited
nature of FSAE cooling systems. That is to say that the maximum heat rejection is dictated
by airflow as opposed to water flow through the radiator. Given the low speed nature of the
competition, the air velocity through the radiator (after losses) is significantly less than the
flow velocity induced by an axial flow fan. Thus, ducting and proper fan selections are
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absolutely critical to ensure system performance. The 2004 and 2006 cooling systems both
utilized a fully ducted radiator entrance and shroud exit. Both systems also used larger
diameter fans to ensure the correct airflow was maintained through the radiator core.
It is also important to note that both the 2003 and 2005 systems had high operating
temperatures, but only the UT 2003 overheated catastrophically. The 2005 system was able
to sustain high engine temperatures due to the fact that it used a mechanical water pump
that increased water flow rate substantially. It also ran a relatively high system pressure at
1.6 bar. The additional water flow rate reduced the chances of localized hot spots within the
engine, and the high system pressure raised the boiling point of the water.
2.2.2 2005 Cooling System vs. 2006 Cooling System
As mentioned above, both UT 2005 and UT 2006 used the same radiator and water pump
combination. The UT 2007 cooling system will continue to use the same radiator and water
pump combination and the analysis for this decision is discussed further in the report. A
benefit to this decision is that it allows for direct comparisons to be made between the
systems since many of the system variables are being held constant. This greatly reduces
the uncertainties and possible sources of error during analysis.
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Table 2.2 is a thermal performance comparison between the 2005 and 2006 cooling system.
Table 2.2 Cooling System Comparisons Thermal Performance vs. Mass
Vehicle Actual Specific Heat
Dissipation
Overall System
Weight
Specific Heat Dissipation per unit
mass
UT 2005 135 W/°C 8.7 kg 15.52 W/°C*kg
UT 2006 186 W/°C 7.0 kg 26.57 W/°C*kg
One should note that the specific heat dissipation refers to the amount of heat that is
rejected per 1°C increase, and is normalized against mass to quantify each system’s
performance to mass ratio. It is evident that the system design of UT 2006 is substantially
improved in all respects when compared to that of UT 2005. It is over 1.5 kg lighter and
can reject 28% more heat. The weight savings between the two systems is primarily from
weight reduction measures in system packaging, as well as the removal of the stock oil
cooler. The stock oil cooler accounts for more than half of the weight savings. Testing
completed in 2005 indicates that the oil cooler can be removed without any detriment to
engine performance or longevity. The additional cooling performance can be attributed to
the use of a duct and shroud for the radiator rather than a side pod. The duct and shroud
assembly are about 0.7 kg lighter than a full side pod, while the performance gain is a result
of the seal generated by the duct and shroud. This seal prevents air from escaping through
low-pressure zones around the radiator, rather than passing through the radiator core. A full
weight breakdown and comparison can be found in Appendix A.
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2.3 UT 2007 Initiatives
The most important initiative for this report is to experimentally quantify the amount of
heat rejected by the motor to the cooling system, since subsequent design calculations use
this value repeatedly. The testing has been completed and will be discussed in further detail
later in this report.
After reviewing the previous year’s cooling system data, there are two main design
initiatives being considered for UT 2007. Despite the fact that the optimal engine operating
temperature has not yet been established, reference material10,11, engine builders, and racing
industry professionals all agree that a safe steady-state operating temperature range for an
engine is between 80°C – 105°C. Temperatures above this will result in overheating of the
engine and other peripheral components, and temperatures below this have negative effects
on engine power output, fuel consumption, and engine wear. The optimal operating
temperature typically lies somewhere between the 80°C – 105°C range, although some
engines show peak performance around 75°C. This optimal temperature will vary from
engine to engine as parasitic losses are generally a function of temperature. The
temperature at which these losses are a minimum is the optimal operating temperature for
the engine. With this in mind, the design initiatives for UT 2007 are intended to reduce the
weight of the cooling system, while satisfying all of the aforementioned design constraints.
The first initiative is to switch from the current “Medium” performance 11” Spal fan to the
“Standard” performance 11” Spal fan, which has the same overall diameter as the current
fan but uses a smaller motor which reduces the mass of the fan. The use of a smaller motor
also reduces the electrical load on the electrical system. Switching to the standard
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performance fan results in a weight savings of 0.36 kg. One may suggest that going to a
smaller fan will result in the system losing some heat rejection ability. This is true; however
analysis of on track data suggests that the UT 2006 system is overbuilt. The justification
and analysis of on-track data will be examined later in this report.
The second design initiative is to remove the water-pump by-pass line and thermostat
assembly to reduce system complexity and reduce the weight of the system. This assembly
is to be replaced by a flow restriction disc. However this requires careful consideration of
the intended function of the thermostat and by-pass lines, and to verify that the flow
restriction disc can be used without adverse effects on the performance of the cooling
system and the engine. If these design initiatives are implemented, the overall system
weight savings over the 2006 system is about 0.7kg, which equates to a 10% weight
savings.
3 Testing and Analysis:
3.1 Track Testing
As noted above, the 2006 cooling system was suspected to have been overbuilt for its
application. However, prior to this discussion and analysis, it is important to understand the
general function of the thermostat (see Fig. 3.1) and the warm up phase of an engine. The
function of the thermostat is to bring the engine to its operating temperature as quickly as
possible. This is due to the fact that engines operating below ~ 70°C10,11 typically exhibit
poor performance, poor fuel economy, and high engine wear. The thermostat is a flow
restriction device that is placed at the water outlet of the engine and remains closed until
the engine is close to the operating temperature. In the closed position, water is not allowed
to enter the radiator, and is recirculated back into the water pump via the bypass line. As
the temperature approaches the engines operating temperature the thermostat valve begins
to open allowing water to flow to the radiator to be cooled.
Figure 3.1: Automotive Thermostat
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In the case of the CBR 600 F4i engine, the thermostat begins to open around 82°C and is
fully open around 90°C12.
In August of 2006, on-track testing was completed to examine the cooling system
performance by varying a number of system parameters. The first test was to complete a
simulated endurance race with the cooling fan running for the duration of the race to see
what temperature the system would operate at for steady state race use. Figure 3.2,
generated by race studio analysis, the data acquisition software used in conjunction with the
data acquisition system, shows a plot of engine temperature as a function of time. The
ambient conditions are at 30°C and the car is undergoing a simulated endurance run.
Figure 3.2: Race Studio Water Temp (°C) vs. Time (sec)
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Figure 3.3 Measures
Figure 3.3 shows the data collected on lap 26 of 31. Each lap is on average 24 seconds
long. The car starts the test at a water temperature of 78°C and the cooling fan is turned on
and remains on during the course of the race simulation. The plot in Fig 3.2 shows that the
engine water temperature averaged about 85-86°C. Given the above operating conditions
and the discussion about the thermostat, closer inspection of the plot suggests that the
thermostat was not in the fully open position and the thermostat was cycling between
opened and closed positions while the engine was running. This explains the apparent noise
in the water temperature signal. One should note that design calculations indicate that the
car should run at 85°C. However, if the thermostat is indeed cycling between the open and
close positions, this suggests that the car can run cooler if the thermostat were to be
removed. Due to time constraints, this test was not actually performed.
Given that the above simulation represents the most extreme conditions, the data suggests
that the system is overbuilt. Since Formula SAE cooling systems are typically airside
limited, the UT2006 system is likely to have more airflow than needed through the radiator
core. Consequently a lighter duty fan is being considered for use in UT 2007.
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Another on-track test performed was to evaluate the effect of thermostat removal on the
performance of the cooling system. Some consideration was given to removing the
thermostat and by-pass line to reduce the system weight. However, removal of the
thermostat could have detrimental effects if not considered carefully. The primary function
of the thermostat is to regulate the temperature of the engine and facilitate a shorter warm
up period for the engine. The first test conducted was an endurance simulation with the
thermostat installed and the second test was the same endurance simulation with the
thermostat removed. The purpose of the test was to examine the impact of the thermostat’s
presence on the rate at which the water temperature increases to a reasonable operating
water temperature. The test was conducted by starting the car, idling briefly and then full
out race driving to simulate the first few laps seen in an endurance race. Both tests were
conducted on the same track with the same ambient conditions.
Figure 3.4 Endurance Simulation with Thermostat
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Figure 3.5 Endurance Simulation without Thermostat
Referring to the Figures 3.4 and 3.5 generated by Race Studio, it can be seen that there is a
delay in reaching the engine’s operating temperature when the thermostat is removed. It
should be noted that the X-axis scale on these two graphs are different. To quantify the
difference between the two system arrangements, the average rate of water temperature
increase was computed by taking the slope of line drawn from 45°C to 85°C, which is
presented in Table 3.1.
Table 3.1 Rate of Water Temp Rise Rate of Water Temp Rise Thermostat Removed 0.17°C/s Thermostat Installed 0.5°C/s
This indicates that the thermostat allows the car to reach its operating temperature nearly 3
times as quickly when the thermostat is installed. However, no drivability problems were
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noted during the test with the thermostat removed. During competition, the car is warmed
up prior to running the dynamic events, with the exception of the endurance event. The
endurance and fuel economy events are coupled, so unnecessary fuel consumption used to
warm up the engine prior to the endurance race adversely affects the fuel economy score.
Thus, the only event that needs to be considered when comparing the performance impact
of the thermostat is the endurance event. Analysis regarding the impact of a longer warm-
up period is discussed later in greater detail. However, as mentioned earlier, the engine
should be maintained at an operating temperature between 80° and 105°C, and it should
approach those temperatures during the warm-up phase relatively quickly.
3.2 Flow Rate Testing
Determining the water flow rate through the engine was a particularly important design
parameter needed to make many design calculations and design decisions. The primary
objective of the flow rate testing was to determine the flow rate at the average engine speed
during race use. This value is needed to compute the heat rejection requirement of the CBR
engine and, as mentioned earlier, the heat rejection requirement is the most important
parameter to be computed. All of the subsequent system designs are based on the initial
heat rejection requirement.
For a mechanical water pump, a pump characteristic curve consisting of flow rate vs. rpm
vs. pressure drop is desired. However, due to time and manufacturing constraints, only a
flow rate vs. rpm test was conducted.
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Testing was completed on the UT 2006 vehicle using the stock water pump and radiator.
As noted earlier, flow restriction devices were being considered, so testing was completed
using these flow restriction discs. Three discs were prototyped and tested. The three discs
create varying flow restrictions quantified from low to high restriction. The restrictive disc
is simply a small round 0.050” thick aluminium disc with one or more holes to facilitate
water flow. Note that the medium flow restriction disc is sized to provide the same
diameter opening as the thermostat orifice in the fully open position. Thus, it is assumed
water flow rate curve for a medium restriction disc representative of the water flow rate for
the stock thermostat. The testing was completed using a GPI brand water flow meter and
the ignition cut feature on the engine control system to hold the engine at steady state rpm.
From this test, a scatter plot of flow rate vs. rpm has been created and a line of best fit has
been used to represent the flow rate. Refer to Appendix B for this plot.
Table 3.2 Water Flow Rates
Disc Type Ambient
Conditions
Flow Rate @ Idle
(~1600 rpm)
Flow Rate @ 7500rpm
Low Restriction Hot 9L/min 43 L/min
Medium Restriction Warm 6L/min 30 L/min
High Restriction Cold/Raining 4L/min 25L/min
Table 3.2 shows the flow rates at idle and at 7500 rpm for the three different discs. Note
that the flow rate of 7500 rpm is used since it is the average engine speed during a race.
After a great deal of analysis and consideration, it was decided not to use flow restriction
discs for UT 2007, but to use a modified thermostat. An in-depth analysis can be found
later in this report.
3.3 Heat Rejection Requirements
As noted earlier the most important design parameter is the heat rejection required to
maintain optimal operating temperature. One should note that these are two separate
parameters that are decoupled. The first one is the actual heat rejection by the motor to the
cooling system, and the second is the optimal engine operating temperature. The heat
rejection requirement is governed by the equation TmCPQ ∆= . In this case, the mass flow
rate has been calculated using a flow meter as discussed earlier, and the water flow rate for
a fully open thermostat was used. Since the mass flow rate is a function of engine speed,
the flow rate is taken at the average engine speed during race simulation. The temperature
difference is calculated by placing an additional thermistor in the cooling line at the
radiator exit, coupled with the use of the existing stock thermistor at the engine block
outlet. These two positions represent the coolest and hottest spots respectively. The
temperature rise between these two points is caused by the heat rejected from the engine to
the water. Given the nature of the experiment and the value of the data extracted from the
test, a great deal of time was spent calibrating and the verifying calibration on these
sensors.
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Figure 3.6 Water Temperature Difference between engine inlet and outlet
Referring to figure 3.6 it is apparent that the temperature difference between the two values
remains fairly constant during race simulation. Since heat rejection is a function of engine
speed, this observation indicates that the assumption of constant heat rejection is justified
since the heat rejection requirement remains fairly constant during the simulation. The
temperature difference is on average 4.1°C. Using the average engine speed of 7500 rpm,
the water flow rate is approximately 30L/min. Using the calculation for heat rejection we
have:
TmCPQ ∆=
KKgKKJsKgQ 1.4/205.4/5.0 ∗∗= = 8620 Watts
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When compared to the estimated heat rejection value of 10200 Watts it is apparent that the
estimate used in previous years is fairly conservative. The use of this conservative estimate
explains the characteristics of the UT2006 cooling system, which was capable of cooling
more than what the initial design calculations suggested. Using a safety factor of 1.1, the
heat rejection design requirement for the CBR 600F4i motor is calculated to be 9500 Watts.
The designer should bear in mind that engine operating parameters, such as lean air-fuel
ratio and pre-detonation caused by too much ignition advance, will increase combustion
temperatures. This, in turn, will cause the heat rejection requirement to increase. However,
dyno testing indicates that both of these conditions cause reduced power output and
compromise engine longevity due to excessive mechanical and thermal loads.
3.4 Optimal Engine Operating Temperature
Determining the optimal engine operating temperature is a key test to designing the cooling
system. Once the system designer knows how much heat is rejected by the motor, the next
step is to determine the operating temperature for optimal for engine performance. It would
be difficult to determine the optimal engine temperature through analytical methods. The
time and effort needed to develop an in-depth model would encompass the scope of this
thesis. Additionally, it would have to be verified through physical testing. Consequently,
the physical testing portion was completed using the engine dynamometer, which can be
seen in Figure 3.7, developed throughout the past 6 months.
Figure 3.7 Engine Dyno Testing Rig
The testing procedure was to hold the engine at constant speeds while under full load. This
procedure was repeated at various operating temperatures and plotted against output torque
at the brake. The test was conducted on three separate occasions. On each occasion the test
was completed by sweeping in both ascending and descending temperature to verify the
repeatability of the results. The dyno was calibrated prior to each testing session and the
data trends (as opposed to absolute values) were compared to each other. This method of
comparison eliminated any uncertainty due to changing ambient conditions, which would
also impact the engine torque output. Each testing session showed the same trends. The
data from a single session can be seen in Figure 3.8 for an engine speed of 10000 rpm.
Refer to Appendix C for the same plot at 6000 rpm.
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Torque vs. Water Temp at 10000 Rpm
65
67
69
71
73
75
77
79
81
83
85
60 65 70 75 80 85 90 95 100 Water Temp (°C)
Torq
ue (l
b-ft)
Torque @ 70°CTorque @ 80°CTorque @ 95°C
Figure 3.8 Torque vs. Water Temperature
The scatter plots for both 6000 and 10000 rpm indicate that the optimal operating engine
temperature is approximately 70°C. Both plots have similar trends that indicate that the
torque output is at a maximum near 70°C and the output decreases and temperature
increases. Both graphs indicate a 3.2% power output difference between 70°C and 95°C.
This suggests that the function of engine operating temperature and torque output is not
related to engine speed. Given the packaging and weight restraints, designing a cooling
system to operate at 70°C would require a system that is considerably larger and heavier
than the existing system. Also, noting that the torque output difference between 70°C and
85°C is about 1%, a desired operating temperature of 85°C was selected as it is a tradeoff
between optimal torque output and maintaining low system weight. Additionally, one
should note that it is possible to maintain temperatures closer to 70°C for short dynamic
events, such as the acceleration or autocross event.
25
3.5 Radiator Configuration Testing
The radiator configuration refers to the number of times the water must pass through the
radiator core. A single pass, or standard cross flow radiator, is a radiator configuration that
allows the water to pass through the core once. A serpentine cross flow, or multipass
radiator, allows the water to pass through the core multiple times. Fig 3.9 shows a single
Figure 3.9 Radiator Configuration Schematics
and a multipass radiator configuration to illustrate this point. The crossflow radiator is the
traditional radiator configuration, since it is typically cheaper to manufacture. One should
note that this type of heat exchange must transfer heat from the heat source (engine) to the
water and then to atmospheric air. If one of the heat transfer rates to one of the fluids is
smaller than the other, then one side of the system will limit the overall system
performance. For example, a system in which the heat exchanger has a small air flow rate,
but an infinite water flow rate, will be limited by the maximum heat transfer of the water to
the air. A simple analogy to this would be a fluidic pipe system where one pipe is much
smaller than all the others. This would result in a bottleneck and limit the flow.
26
A multipass radiator is often considered for cooling systems that are air side limited. This is
because the water must remain in the core for a longer period of time and will transfer heat
to the air for a greater period of time. However, the pressure drop in the system increases
for a multipass radiator. Consequently, for a fixed water pump, a multipass radiator would
have a lower water flow rate than a single pass radiator. Based on manufacturer test data
and experience with previous cooling systems, I believe that the cooling system of an
Fig 3.10 Radiator Single Pass Configuration Testing Set-up
FSAE vehicle is very much airside limited. This hypothesis is supported by the dramatic
performance increases that are seen by using higher flow rate fans. However, in the past,
the decision to use a double pass radiator was speculative and definitive testing in a
controlled manner has never been completed. Thus a test was developed to compare the
performance of a single and double pass radiator in a Formula SAE application. Note that a
27
triple pass radiator was not considered due to time constraints, and the fact that a triple pass
radiator has a considerably larger pressure drop. Using a single pass radiator as a baseline
reference, a double pass radiator increases the pressure drop by 16 times and a triple pass
radiator increases the pressure drop by 64 times15. The test was completed using a single
(Fig 3.10) and double pass (Fig 3.11) configuration of the same radiator, and placed on the
UT2006 vehicle. Testing was completed with the same ambient conditions. The vehicle
was stationary and the motor was started and the engine speed was held constant at 5000
rpm. Once the water temperature reached 100°C, the cooling fan was turned on and air was
blown across the core. Once the coolant temperature reached a steady state temperature
recordings were taken. The same test was repeated with a double pass configuration.
Figure 3.11 Double Pass Radiator Configuration
28
29
The test results found in Table 3.3 indicate that the double pass radiator has a significant
advantage in comparison to the single pass radiator in this particular application. There was
some small variability in test conditions, but a 33.5°C temperature drop is a clear indication
that the double pass radiator performs much better this application.
Table 3.3 Radiator Configuration Testing Results
Single Pass Double Pass Delta
Ambient Air Temp -4°C -4.5°C 0.5°C
Starting Water Temp 100°C 101°C 1°C
Steady State Water Temp 98°C 64.5°C 33.5°C
It is important to note that this experiment is specific to this system and one cannot assume
that a double pass will perform better than a single pass in every application. In this case of
a Formula SAE vehicle, the air side limitation causes the double pass radiator to perform
better than a single pass.
30
4 Component Selection
4.1 Component Selection
After compiling all of the necessary data collected during testing, cooling system design
can proceed. The first step is to size a water pump, cooling fan, and radiator combination
that will maintain an optimal operating temperature of 85°C while rejecting 9500 Watts of
heat with an ambient air temperature of 30°C. Using manufacturer supplied test data, found
in Appendix D, we can calculate system parameters that will allow the engine to operate at
85°C during the worst case scenario. As mentioned above the worst case scenario is
approximated as a hot day with an ambient temperature of 30°C. In the case that the car
should be exposed to a hotter ambient temperature, the water temperature rise is directly
proportional to the rise in ambient temperature, i.e. a 40°C ambient temperature would
result in a water temperature of 95°C.
4.2 Water Pump Selection
Due to time constraints, the analysis of parasitic draw for mechanical and electrical pumps
was not completed. However, existing test data from various automotive and pump
manufacturers suggests that electrical pumps create less parasitic load on the motor than
mechanical pump. However, past experience with electrical pumps on the UT FSAE cars
has resulted in pump failure and consequent overheating. The electrical pumps cannot
withstand the harsh operating conditions endured in a Formula SAE vehicle. Additionally
the pumps that are suited for use on a Formula SAE car have significantly lower flow rates
than a mechanical pump. The water flow rate for an electrical pump is constant, whereas
31
the flow rate for a mechanical pump is proportional to engine speed. This is beneficial since
high engine speeds have greater cooling requirements. Despite the fact that a mechanical
pump likely requires a greater power input than an electric pump, a mechanical pump was
selected due to its superior reliability and greater flow rates at high rpms.
4.3 Radiator Selection
Long Dana Manufacturing, a manufacturer of radiators for small ATV’s and personal
recreational vehicles, sponsors the UT FSAE team and supplies prototype radiators. For
UT2007 the radiator selection was limited to the available resources rather than specifying
a specific radiator design. This would be rather cumbersome, and testing and validation of a
radiator design would be difficult. The resources needed to complete such testing would be
more than what the UT Formula SAE team can support. As such the radiator selected is the
same radiator that has been used in UT 2005 and UT 2006. This radiator was selected for a
number of reasons. It packages well within the chassis, it was the smallest radiator that
Long Dana has made available to Formula SAE teams, it has proven to be adequately sized
on UT2006, it has an existing set of molds for a duct and shroud, and its geometry lends
itself well to a 10” fan or larger. The fact that it can accommodate a relatively large fan is
beneficial and will be explained in greater detail in the subsequent section. The molds used
to manufacture the duct and shroud are extremely time consuming to fabricate, so using
existing molds save a great deal of time.
The stock radiator, which can be seen in Figure 4.1 in its factory configuration is a single
pass crossflow design that uses ¾” cooling bungs.
Figure 4.1 Unmodified UT 2007 Radiator
The radiator is modified heavily to make it more efficient. It is double passed by welding a
plate to block flow through the header tank to force the water to flow through the core
twice. The stock bungs are removed and replaced with 1” diameter bungs to increase flow
and reduce head loss. Figure 4.2 shows the modified radiator in process.
32
Figure 4.2 UT 2007 Radiator in Process
4.4 Cooling Fan Selection
Since the mechanical pump was selected, the water flow rate value is a fixed design
parameter. Working under the assumption that the thermostat will be used, average water
speed is about 30L/min. Also noting that the same radiator used in 2006 was chosen, the
only design variable remaining is the cooling fan selection. It is important to note that
Spal’s fan product line comes in three performance levels: standard, medium and high
performance. Each performance level uses a standard motor for each respective
performance level. For a given performance level the motors are the same for each fan
33
34
diameter. The only difference is the housing and fan blades. With that being said, the only
difference between a 10” and a 16” fan of the same performance level is the plastic fan
blade and housing. The two fans share the same motor. The difference between a 10”
standard performance fan and a 10” medium performance fan is the motor output, i.e. the
10” medium performance fan has a larger, heavier, and more powerful motor. Therefore to
maximize performance and minimize weight, it is desirable to choose the largest diameter
fan that can be packaged in the lowest performance level.
It can be seen from heat rejection graph found in Appendix D, that the desired output value
of 176W/°C requires an air flow velocity of approximately 4.0 m/s through the core. There
are three fans that are in the vicinity of 4.0 m/s. They are the 10” and 11” inch Spal
medium performance, and the 11” Spal regular performance fans. Given the
aforementioned description of Spal fans and the fact that UT2006 used a Spal 11” medium
performance fan, an 11” standard performance fan was chosen for 2007. This choice was
made for a number of reasons; the cooling system on UT2006 was suspected of being
overbuilt, and since the other two system components remained the same as last year, this
is the most suitable way of optimizing the 2007 cooling system. The 11” standard
performance fan can only pull air through the core at a velocity of 3.1m/s, which is 0.9m/s
short of the 4.0m/s needed to achieve the desired operating temperature for the worst case
scenario. However, one must take into consideration the following: First, the duct assembly
will aid to channel air through the core of the radiator. This will help increase the actual
core velocity to be greater than 3.1 m/s. On track testing will help to determine if this fan
combination is capable of meeting the design objectives. Second one must also consider
35
that the worst case scenario is unlikely to be encountered during the FSAE competition, as
it is held in mid May, where peak ambient temperatures are usually between 15-23°C.
Additionally the 11” standard fan has the same bolt pattern as the 11” medium performance
fan. This feature ensures that if the system is inadequate to maintain a temperature of 84°C,
the medium performance fan can be swapped in to increase cooling capacity.
4.5 T-Stat vs. Flow Restrictors
Post competition testing indicated that with the thermostat removed it takes about 3 times
as long to bring the engine up to operating temperature. As mentioned earlier, the idea
behind removing the thermostat was to reduce the weight of the cooling system, by
approximately 0.5kg. However, the extended warm up period is undesirable, and a device
is needed to accelerate the warm-up phase. The proposed solution was to use a flow
restriction disc in place of the thermostat that serves a two-fold purpose. First, it serves to
allow water to flow through the thermostat housing, such that the by-pass lines can be
removed without pressurizing the system and risking damage to the water pump. Second, it
restricts the water flow rate to aid in the warm-up process. Given the low dependence of
flow rate, and high dependence of air speed, it was thought that the system would be able to
increase its water temperature when the vehicle is stationary, without sacrificing
performance when running on track. However, varying ambient conditions would require
frequent changing of the flow restriction disc to ensure a reasonably fast warm-up phase.
Consequently, it was decided to use a modified thermostat. The stock thermostat has been
modified by drilling three 1/8” holes on the periphery to allow a small amount of coolant
36
flow while the thermostat is in the closed position. This feature will allow the thermostat to
greatly reduce water flow at low temperatures to aid warm-up. Once the water temperature
approaches 82°C the thermostat will begin to open and function as it would normally. This
allows the thermostat to remain in the car for all ambient conditions without greatly
increasing the warm-up period, or compromising flow at high water temperatures. It also
facilitates the removal of the by-pass line since water flow is permitted (at a reduced flow
rate) during the closed position thermostat operation. This ensures that the block inlet side
is not highly pressurized, which may damage the water pump impeller. This design has yet
to be tested on the track, but data analysis suggests that this is the best solution to achieve
the goals of weight reduction without compromising component longevity or greatly
increasing engine warm up phase.
4.6 Temperature Regulation The cooling temperature is thermostatically controlled via the feedback loop between the
water temperature sensor, the cooling fan relay, and the ECU. Once the water temperature
exceeds 88°C the ECU powers up the cooling fan relay and cooling fan is turned on. Once
the water temperature cools down to 83°C the cooling fan shuts off. The set temperature for
the cooling fan for turning the fan on and off the fan is adjustable via the ECU.
4.7 Weight Reduction
The majority of the weight reduction comes from the removal of the water pump by-pass
line and the switch from a medium performance fan to a standard performance fan. The
37
removal of the by-pass line accounts for 0.2 kg savings, and the fan switch accounts for
0.35 kg. Thus, the total weight savings over UT 2006 is conservatively estimated at 0.55kg.
However, a concerted effort was also placed on reduction of water volume by minimizing
cooling line lengths. Since the final placement has not yet been decided, it is unclear
whether this initiative will yield any additional weight savings.
38
5 Components and Manufacture
5.1 Radiator As mentioned earlier the radiator is donated by Long Dana Manufacturing. The radiator is
then modified by cutting off the old cooling bungs and welding larger lower restriction
bungs. The cooling bungs are custom machined from 1-1/8” 6061T-6 aluminium rod. The
radiator also has a bleeder screw assembly welded to it to facilitate the removal of air
pockets that may accumulate near the top of the radiator. The radiator is mounted off a
chassis tube and the carbon fibre side panel. Each mount is a rubber sandwich mount that
isolates the aluminium radiator from vibration. In this application vibration isolation is
important since fatigue failure of the radiator mounts has occurred in the past.
5.2 Cooling Fan The cooling fan is purchased from Spal Fan. The fan remains mostly stock. The only
modifications are the removal of casting flash from the fan blades to reduce turbulence, and
the replacement of the fasteners that clamp the motor to the fan hub. The plastic screws are
replaced with socket head cap screws and locking nuts that will not back off due to
vibration.
5.3 Cooling Duct and Shroud Assembly The duct and shroud assembly are made out of fibreglass material. The assembly is
completed using a wet lay up on a hand made mold. Once the finished parts are removed
from the mold, the shroud and duct assembly are fitted to the radiator and then bonded
using epoxy.
39
5.4 Mechanical Water Pump The water pump is modified by machining the back face of the water pump to mate with
the custom adaptor plate that mates the dry sump to the water pump. The stock water pump
impeller is pressed onto the dry sump pump shaft, which replaces the water pump shaft
inside the motor. The cooling bungs on the water pump used for the stock oil cooler and the
by-pass line are also cut off and then welded shut.
5.5 Thermostat and Thermostat Housing
The thermostat is modified by drilling three 1/8” holes to allow low flow rates during
closed thermostat operation. The thermostat housing has the by-pass line bung cut off and
welded shut.
5.6 Swirl Pot
The swirl pot is made from 3.5” O.D. 0.065” wall thickness aluminium tubing and is 2.5
inches long. It serves to deaerate the cooling system and is placed at the highest point in the
system to facilitate bleeding of the air in the cooling system. The swirl pot also contains the
filler neck assembly, which is welded to the swirl pot cap.
40
5.7 Hardlines
The hardlines are made from 1” O.D. 0.065” wall thickness 6061T6 aluminium tubing. The
lines are filled with sand and then the ends are crimped. The sand prevents the collapse of
the hardline during bending. The line is then heated and bent using solid metal stock to
achieve the desired bend radius. Once the lines are completed, the ends of the line are
beaded using aluminium filler. The bead is added to increase the interference fit with the
cooling hoses.
5.8 Hoses and hose clamps
Silicon hoses are used for the cooling system due to their high temperature resistance and
high burst pressure rating. The hoses clamps used are double lined hose clamps with
radiused edges. It is important to use these particular hose clamps as they do not cut the
cooling hose. Cheaper perforated hose clamps are notorious for puncturing cooling hoses
due to the sharp edges.
41
6 Recommendations
During the course of this thesis, a great deal of testing and development has occurred. The
subsequent cooling system designers have an abundant source of data that will aid in
making sound design decisions. However, there are a number of tests that could not be
completed due to time and manufacturing constraints. Thus, it is my recommendation that
future system designers complete pressure drop testing mentioned earlier in this report. In
this thesis it did not hinder development since all of the testing was done on the same
radiator, fan, and water pump combination that is being used for the current vehicle.
However, this data would be useful for comparing system components that have different
pressure drops across them.
I also recommend that future testing involve load testing for engine output losses generated
by electrical pump loads and mechanical pump loads. It is my opinion that electrical water
pumps are not suited for use on an FSAE car, but load testing of this sort would provide
additional insight for a designer who must select one of these two pump options.
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7 Conclusions
This thesis has been successful in identifying the cooling requirements of the CBR 600 F4i
motor used in the UT FSAE vehicle. Substantial testing was completed to determine that
the cooling system must be sized to reject 9500 Watts of energy, and must also be able to
maintain a steady state operating water temperature of 85°C for an ambient air temperature
of 30°C. The UT2007 cooling system is set to be the most optimized cooling system to date
that is designed to maximize performance and minimize weight, without compromising
system reliability. The new system is approximately 10% lighter than last year, thereby
improving the overall performance of the vehicle. During the testing season the design will
be validated through on track testing and extensive data logging and data analysis.
43
References [1] “Formula SAE Rules”
[2] Smith, Carroll., “Carroll Smith’s Nuts, Bolts, Fasteners and Plumbing Handbook”,
MBI, 1990.
[3] Çengel, Yunus A., “Heat and Mass Transfer: a practical approach.”, McGraw-Hill,
2007.
[4] Çengel, Yunus A., “Thermodynamics: an engineering approach”, McGraw-Hill,
2006.
[5] Ngy-Srun AP., “A Simple Engine Cooling System Simulation Model.” SAE
International. Technical paper no 1999-01-0237.
[6] Smith, Jeffrey D., “Engine Cooling Fan and Shroud Optimization for Blow
Through Operation.” SAE International. Technical paper no. 860764.
[7] McKenzie, A.B. “Axial flow fans and compressors: aerodynamic design and
performance.” Ashgate, 2007.
[8] Valkenburgh, Paul Van., “Race Car Engineering & Mechanics.” Published by
Author, 2000.
[9] Smith, Carroll., “Engineer to win.” MBI, 1990.
[10] Bosch, Robert., “Automotive Handbook.” Robert Bosch, 2004.
[11] Stockel, Martin T., “Automotive Mechanics Fundamentals” Goodheart-Wilcox,
1990.
[12] Coombs, M. “Honda CBR 600 F4i Service and Repair Manual”, Haynes, 1998.
[13] Spal Fans. www.spalusa.com
[14] Bosch Electronics. www.boschautoparts.com
[15] Stewart Racing Components. http://www.stewartcomponents.com/
44
APPENDIX A: Weight Comparison Matrix
2005 Cooling System Weight Analysis
Radiator: overall 240 X 405 mm core area: 315 X 230 mm
Cooling Component Description Mass (g) Length (in) Line ID (in) Volume (ml) Water Weight (g)Rad to Water Pump w/ 2 hoses and 4 clamps 269.00 g 14.875 0.86 141.6193875 141.62 g
Rad - S-pot - T-stat w/ 2 hoses + 4 clamps Line Length S-pot dims: 3.5" long X 2.5"Ø
500.00 g 27.875 0.86 265.3875918 265.39 g
Water Pump to block inlet w/ 2 hoses and 4 clamps 357.00 g 18.5 0.86 176.1316752 176.13 g
Upper T-Stat Line 150.00 g 10.625 0.86 101.1567054 101.16 g
Swirl pot - block inlet bleed line + two clamps 50.00 g 13 0 Negligible
Swirl Pot water Volume 0.00 g 0 0 80.45 80.45 g
Oil Cooler Line - Rad to cooler + two hose clamps 132.00 g 19 0.41 41.11403955 41.11 g
Oil Cooler Line - Cooler - pump + two hose clamps 90.00 g 12.875 0.41 27.86017154 27.86 g
Rubber hose from S-pot to Overflow tank 107.00 g 15.625 Negligible
Oil Cooler 764.00 g Negligible Radiator + Fan 3624.00 g 560.07 560.00 g Thermostat 150.00 g 0.00 g
Bypass Line with hoses + clamps 151.00 g 18.25 0.54 68.50451933 68.50 g
Body Work (50% of one full side pod) 908.00 g 0.00 g Totals 7252.00 g 1462.22 g
Volume of Rad Header tank (cubic cm) 280.035
Total volume of Header tanks (cubic cm) 560.07 Cooling System Component Weights 7252.00 g
Cooling System Water Weight 1462.22 g
Total Cooling System Mass 8714.22 g 19.21 lbs
45
APPENDIX A: Weight Comparison Matrix
2006 Cooling System Weight Analysis
Radiator: overall 240 X 405 mm core area: 315 X 230 mm
Cooling Component Description Mass (g) Length (in) Line ID Volume (ml) Water Weight (g) Rad to Water Pump w/ 2 hoses and 4 clamps 279.50 g 10 0.86 95.20631094 95.21 g
Rad - S-pot - T-stat w/ 2 hoses + 4 clamps Line Length S-pot dims: 3.5" long X 2.5"Ø 662.00 g 31.75 0.86 302.2800372 302.28 g Water Pump to block inlet w/ 2 hoses and 4 clamps 310.00 g 17 0.86 161.8507286 161.85 g Swirl Pot water Volume 0.00 g 0 0 80.45 80.45 g Rubber hose from S-pot to Overflow tank 107.00 g 15.625 0.00 g Radiator + Fan 4100.00 g 560.07 560.07 g Thermostat 150.00 g 0.00 g Bypass Line with hoses + clamps 135.00 g 17 0.54 63.81242896 63.81 g Totals 5743.50 g 1263.67 g Volume of Rad Header tank (cubic cm) 280.035 Total volume of Header tanks (cubic cm) 560.07 Cooling System Component Weights 5743.50 g Cooling System Water Weight 1263.67 g Total Cooling System 7007.17 g 15.45 lbs
APPENDIX B: Flow Rate vs. RPM Data
Cooling System Flow Rate Testing Test Completed Nov 06, 2006
High Restriction Disc Medium Restriction Disc Low Restriction Disc
RPM (x 1000)
Flow Rate
Flow Rate
RPM (x 1000)
Flow Rate
Flow Rate
RPM (x 1000)
Flow Rate
Flow Rate
GPM LPM GPM LPM GPM LPM 1.5 1 3.785 1.5 1.5 5.678 1.5 2.25 8.517 2.8 2.46 9.312 3 3.09 11.697 3 4.56 17.261 5 4.77 18.056 4.5 4.77 18.056 4.5 7.01 26.536 6 5.67 21.463 5.7 5.995 22.694 5.7 8.76 33.160 7 7.09 26.839 7.5 8.055 30.491 7.5 11.51 43.570
8.6 7.665 29.015 9 9.465 35.829 9 13.605 51.501 9.8 7.73 29.261
46
APPENDIX C: Water Temperature vs. Torque Plot
Torque vs Water Temp at 6000 Rpm
48
50
52
54
56
58
60
62
64
66
68
60 65 70 75 80 85 90 95 10
Water Temp (°C)
Torq
ue (l
b-ft)
0
Torque @ 70°CTorque @ 80°CTorque @ 95°C
47
48
APPENDIX D: Design Data
(Radiator, Fan, and Water Pump data) These Calculations do not consider radiation as a method of heat loss.
2003 Car
Operating Parameters Unit Radiator Spec 400 Series - 12 FPI Radiator Area 0.075 sq. m Water Pump Davies Craig 900 L/hr
Fan Spal Std. 10" System Operating Water Flow Rate 18 L/min
System Operating Air Flow Rate 3.2 m/s Thermal Performance 120 W/C
Max Temp 115 C Ambient Temp 30 C
Max Heat Transfer Required 10200 W
2004 Car
Radiator Spec 600 Series - 18 FPI Radiator Area 0.075 sq. m Water Pump Davies Craig 900 L/hr
Fan Spal Std. Mid Performance 12" System Operating Water Flow Rate 18 L/min
System Operating Air Flow Rate 3.6 m/s Thermal Performance 175 W/C
Max Temp 88 C Ambient Temp 30 C
Max Heat Transfer Required 10200 W
2005 Car Heat Transfer 10200 W
Max Temperature Desired 90 C Ambient Temp Designed For 30 C
Adjusted Thermal Performance 170.0 W/C
Radiator Spec 18 FPI, 16 core Radiator Area 0.07245 sq. m Water Pump Stock CBR 600 F4I Mechanical Pump
Fan Zirgo 10" System Operating Water Flow Rate ~ 30L/min @ 7500 rpm L/min
System Operating Air Flow Rate unknown (est. at 2.6) m/s Thermal Performance 135 W/C
Max Temp 110 C Ambient Temp 30 C
49
2006 Car Heat Transfer 10200 W
Max Temperature Desired 90 C Ambient Temp Designed For 30 C
Adjusted Thermal Performance 170.0 W/C
Radiator Spec 18 FPI, 16 core Radiator Area 0.07245 sq. m Water Pump Stock CBR 600 F4I Mechanical Pump
Fan 11" Spal Med Performance System Operating Water Flow Rate ~ 30L/min @ 7500 rpm L/min
System Operating Air Flow Rate 4.72 m/s Thermal Performance 186 W/C
Max Temp 85 C Ambient Temp 30 C
2007 Car
Heat Transfer 9500 W Max Temperature Desired 85 C
Ambient Temp Designed For 30 C Adjusted Thermal Performance 172.7 W/C
Radiator Spec 18 FPI, 16 core Radiator Area 0.07245 sq. m Water Pump Stock CBR 600 F4I Mechanical Pump
Fan 11" Spal Med Performance System Operating Water Flow Rate ~ 30L/min @ 7500 rpm L/min
System Operating Air Flow Rate 3.1 m/s Thermal Performance 154 W/C
Max Temp TBD C Ambient Temp 30 C
Test No. A1166 Ref. No. H1416-1 Rad Fluid Ethylene Glycol 50/50 mm in Header Distance 250 9.84 Core Width 300 11.81 Core Thickness 41 1.61 Fin Type 18 fpi louvered fins
Water Temperature Air Flow Air Temperature Heat Transfer Water Flow Inlet Outlet Inlet Outlet Water Air
L/min oC oC m/s oC oC W/oC W/oC
Percentage Difference
7.6 84.7 72.0 1.8 25.1 59.2 94.1 94.8 0.8 22.7 84.7 80.0 1.8 25.0 61.5 103.4 101.2 -2.1 30.3 85.4 81.8 1.8 25.1 63.4 104.5 104.6 0.1 45.4 85.6 83.2 1.8 25.1 63.4 107.0 104.5 -2.3 7.6 84.4 66.4 3.6 24.9 48.9 133.8 132.9 -0.6
22.7 84.8 77.5 3.6 25.0 55.1 160.7 164.7 2.5 30.3 84.7 79.2 3.6 24.9 54.6 162.9 162.3 -0.4 45.4 85.1 81.3 3.6 25.7 55.6 169.2 164.5 -2.8 7.6 84.8 62.3 6.7 26.0 42.7 169.2 173.7 2.6
22.7 85.1 75.6 6.7 26.0 46.6 213.7 213.4 -0.1 30.3 84.9 77.4 6.7 26.1 47.7 224.1 226.3 1.0 45.4 85.0 79.7 6.7 26.6 50.2 242.9 247.9 2.0
Water Pressure Drop Air Pressure Drop Water Flow Water dP Air Flow Air dP
L/min kPa m/s mmH2O 7.6 2.8 1.8 4.3
22.7 9 3.6 12.7 30.3 14.5 6.7 34.5
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