MOBILE VEHICLE SHOWER SYSTEM
Joshua Paul Perron
B.S., California State University, Sacramento, 2007
PROJECT
Submitted in partial satisfaction of
the requirements for the degree of
MASTER OF SCIENCE
in
MECHANICAL ENGINEERING
at
CALIFORNIA STATE UNIVERSITY, SACRAMENTO
FALL
2009
MOBILE VEHICLE SHOWER SYSTEM
A Project
by
Joshua Paul Perron
Approved by:
__________________________________, Committee Chair
Dr. Timothy Marbach
____________________________
Date
ii
Student: Joshua Paul Perron
I certify that this student has met the requirements for format contained in the University format
manual, and that this project is suitable for shelving in the Library and credit is to be awarded for
the Project.
__________________________, Department Chair
Dr. Susan Holl
Department of Mechanical Engineering
iii
________________
Date
Abstract
of
MOBILE VEHICLE SHOWER SYSTEM
by
Joshua Paul Perron
Statement of Problem - Many people enjoy outdoor activities, such as riding dirt bikes or
all terrain vehicles, camping, and hunting. Sometimes these excursions occur over
several days, and it is not always very easy to keep clean. Although water may be
available, such as a lake or stream, it is usually rather cold, and definitely not as warm as
most showers taken at home. A heat exchanger was designed in order to take advantage
of the heat produced by an engine in a vehicle by using it to heat fresh water for use.
Sources of Data - A parametric study was conducted using Microsoft Excel. Inputs such as
flow rates, inlet and outlet temperatures, and the size of the heat exchanger were varied.
The parametric study was used to estimate the theoretical amount of heat transfer that can
be expected between the two working fluids before the system was implemented.
iv
Conclusions Reached - After the system was constructed and tested in the field, the
performance of the heat exchanger was not as high as calculated in the parametric study.
The differences can be explained in the implementation of the shower system and the
method by which it had to be connected to the test vehicle. Overall, the greatest
performance was observed while heating fresh water at a temperature of sixty-eight
degrees Fahrenheit flowing at three gallons per minute to seventy-four degrees
Fahrenheit.
_______________________, Committee Chair
Dr. Timothy Marbach
_______________________
Date
v
ACKNOWLEDGMENTS
I would like to thank my advisor Dr. Timothy Marbach for his guidance and assistance during my
graduate education.
I would also like to thank Dr. Susan Holl for her assistance and encouragement during
my entire attendance at California State University, Sacramento.
Finally, I would especially like to thank my entire family for all of their support during
my education endeavors.
vi
TABLE OF CONTENTS
Page
Acknowledgments.................................................................................................................... vi
Table of Contents ................................................................................................................... vii
List of Tables ........................................................................................................................ viii
List of Figures .......................................................................................................................... ix
Chapter
1. INTRODUCTION.......................………………………………………………………… 1
1.1 Previous Designs.................................................................................................... 1
1.2 Proposed General Design ...................................................................................... 3
1.3 Installation on Vehicle ........................................................................................... 3
2. PARAMETRIC STUDY .................................................................................................... 6
2.1 Description ............................................................................................................. 6
2.2 Previous Study ....................................................................................................... 7
2.3 Sample Equations .................................................................................................. 8
2.4 Alternate Method for Calculating Heat Transfer ................................................... 9
2.5 LMTD Sample Equations ...................................................................................... 9
3. EXPERIMENTS ............................................................................................................... 12
3.1 Purpose................................................................................................................. 12
3.2 Set-Up .................................................................................................................. 12
3.3 Trials .................................................................................................................... 15
3.4 Results.................................................................................................................. 15
3.5 Experiment Conclusions ...................................................................................... 16
4. FINDINGS AND INTERPRETATIONS ......................................................................... 20
4.1 Recommendations ................................................................................................ 21
Appendix A. Parametric Study Figures ................................................................................ 27
Bibliography ........................................................................................................................... 44
vii
LIST OF TABLES
Page
1.
Table 3.1 Testing Data………………………….…………………………….. 16
viii
LIST OF FIGURES
Page
1.
Figure 3.1 Heat Exchanger………………...…...………………………………. 12
2.
Figure 3.2 Test Vehicle Set-Up……..….………………………………………. 14
3.
Figure 3.3 Shower System Schematic….………………………………………. 15
ix
1
Chapter 1
INTRODUCTION
Camping, four-wheeling, and fishing trips are just a few examples of the many
activities humans like to take part in outdoors. These excursions generally occur over a
period of a couple days, such as a weekend, and can sometimes last a week or more. It is
usually the case that people are limited in how much gear they may take with them, and
many luxuries are not included. When participating in these activities, especially fourwheeling and camping, people tend to get dirty just from the environment where these
activities take place. One seemingly small luxury would be the ability to take a hot
shower on the trail or at camp. But unless a recreational vehicle (RV) or camping trailer
is present, this is a luxury that many people do not have, especially on some of the more
difficult four wheeling trails, for example, where the purpose of the trip is to drive a
vehicle in places where most vehicles are not capable of traveling.
1.1 Previous Designs
The private sector has developed a few different models of mobile shower
systems. One example is from R&M Specialty Products. This system, which is mounted
to a vehicle, uses a copper heat exchanger measuring roughly two feet long and 2.5
inches in diameter to transfer the heat from the vehicle’s cooling system to the incoming
fresh water. The system has a 1.6 gallon per minute (GPM) pump from an RV that
provides a maximum pressure of 30 pounds per square inch (psi). Running at this
2
pressure is a very nice feature of this pump. Most vehicles’ cooling systems run between
10 and 17 psi. Therefore, if a problem were ever to develop with the system, such as a
leak inside the heat exchanger between the separating chambers, the fresh water will
always flow into the vehicle cooling system, and thus keep coolant from flowing out and
onto the user. If such a situation occurs, fresh water flowing into the coolant system of
the vehicle will not damage the system or the engine, and the RV pump includes an
integrated filter which prevents harmful particles from entering into the cooling system of
the engine.
There are a few reviews of this product given by users of the shower system. One
individual wrote that he used a five gallon bucket as a fresh water tank filled with water
at fifty degree Fahrenheit. When he ran the water through the heat exchanger, the water
was heated to 105 degrees Fahrenheit, a temperature rise of fifty-five degrees. The
problem with this system, however, is that the flow of water is only 1.6 GPM maximum,
and the fifty-five degree temperature rise was reached at a flow rate of roughly 0.8 GPM,
which is a relatively low flow rate for a ‘comfortable’ shower.
Another example of a mobile shower system is one provided from a company
called Bushranger. Their rectangular heat exchanger measures roughly 4.75 inches wide
by a little more than three inches thick, and is slightly longer than one foot. The pump
has a flow rate of about 3 GPM, roughly double the volume than the unit from R&M
Specialty Products.
3
1.2 Proposed General Design
The new shower system being proposed has many similar features as those in the
systems discussed above. This new design, however, will try to incorporate the better
features of the earlier designs all into one complete system. The shower system will use
a similar style marine bilge pump that will bring fresh water from a river or stream (or
even an on-board tank) and pump it through the heat exchanger’s fresh water tubes. All
fluid-cooled vehicles come equipped with an on-board water pump which is used to
pump the coolant through the vehicle’s engine, then through the radiator, and then back
into the engine, creating a closed loop.
1.3 Installation on Vehicle
This design will tap into one of the heater hoses running to the vehicle’s heater
core using a 3-way ‘T’ connector and a valve. The heater core is a smaller radiator in
which coolant runs through when the vehicle’s inside heater is utilized. The smaller
hoses running to this unit provide an easy point for the shower system to tap into, and the
valve will allow the vehicle’s cooling system to operate normally with no fluid running
through the shower system if the user so desires. This is a very important design feature
of this particular system. Although many older vehicles come equipped with brass and
copper radiators, many newer vehicles are now being outfitted with aluminum radiators.
Also, because of the increased cooling capacities and efficiencies of aluminum radiators,
many older off-road vehicles are retrofitted by their owners to accept these newer
aluminum designs. When the copper shower system is connected to the vehicle’s cooling
4
system which is utilizing an aluminum radiator, the dissimilar metals will cause
electrolysis through the cooling system. Electrolysis is the transfer of ions between two
dissimilar metals. This will cause a degradation of the cooling system, most notably the
vehicle’s water pump, which obviously affects a vehicle’s cooling ability. Although
electrolysis will occur while the shower system is actually in use, it will not be a problem
during normal vehicle operations. The amount of time electrolysis will be occurring
while the system is in use is not great, creating a generally acceptable risk to the vehicle’s
cooling system. If the user is concerned about damaging the vehicle’s cooling system,
the electrolysis can be greatly reduced by either grounding the aluminum radiator to the
frame of the vehicle or by using a specially designed radiator cap that incorporates a
sacrificial metal (zinc) nugget. Over time, this nugget is ‘eaten away’ through the
electrolysis process, protecting the aluminum radiator and the vehicle’s water pump.
The heat exchanger unit will be mounted on the vehicle, and will be a modified
counter-flow, single pass, tube-and-shell style heat exchanger. The modification will be
that the inner tube, although it only runs through the shell one time, will be in a spiral,
similar to that of an automotive suspension coil spring. This will provide the fresh water
with more time spent inside the shell, boosting heat transfer between the two liquids. The
spiral design will also increase the velocity of the water through the heat exchanger, as
well as create turbulence within the tube, both of which increase convective heat transfer
coefficients, resulting in higher heat transfer rates.
A counter-flow heat exchanger has the inlets for the two working fluids located at
opposite ends. This design allows the two fluids to maintain a larger difference in
5
temperatures throughout their time flowing through the heat exchanger, also increasing
heat transfer efficiency. Connected to the outlet end of the heat exchanger will be the
coolant return line, which re-introduces the coolant back into the vehicle’s original
cooling system. There will also be an output line that carries the newly heated fresh
water out to an ordinary shower head or spigot. It is at this time that the water is now
ready for use.
6
Chapter 2
PARAMETRIC STUDY
In order to design this system, a parametric study was conducted using Excel as
the modeling software. The purpose of the parametric study is to simulate different
designs of the heat exchanger by changing certain design aspects, such as length of the
heat exchanger tube, the heat exchanger shell, inlet and outlet fluid temperatures, or the
flow rates of the water or ethylene glycol (vehicle coolant). Five specific trials were
performed, each of which had a different size heat exchanger modeled. Each of these
trials were run using all combinations of four different fresh water flow rates and three
different coolant flow rates.
2.1 Description
This particular study was based on a small four-cylinder import engine, very
similar to that found in the proposed test vehicle for this project. Certain inputs were
needed in order to begin the study, such as temperatures and flow rates of the vehicle’s
coolant system, which will be the heat source for the shower system. The rate of flow of
a vehicle’s cooling system is directly related to its temperature. As the temperature rises,
the thermostat on the engine opens wider, allowing a larger flow rate. During times when
the coolant loses energy, the thermostat will begin to close.
7
2.2 Previous Study
Temperatures and corresponding coolant flow rates for a very similar engine as
that in the test vehicle were previously measured in a study entitled, “Thermal Flow
Analysis of Vehicle Engine Cooling System,” by Kyoung Suk Park, from the Department
of Mechanical Engineering at Kyung Hee University, and Jong Phil Won and Hyung
Seok Heo, from the Korea Automotive Technology Institute. Flow rates can be indirectly
controlled by increasing the number of cycles an engine undergoes per minute (RPMs).
As the RPMs increase, so does the vehicle’s temperature, thus opening the thermostat and
allowing for a greater coolant flow rate.
Other inputs needed for this parametric study were the thermodynamic properties
of the fluids used in the experiment. These include density, specific heat capacity,
thermal conductivity, and dynamic viscosity, and whose values were compiled in several
charts within the study described above. In order to perform the theoretical calculations,
the data from these charts was inputted into an excel spreadsheet. Then, in order to find
the values at the corresponding temperatures assumed in this shower system’s theoretical
study, the data was used for interpolation calculations. Utilizing these inputs, along with
simple unit conversions in order to calculate values such as velocities of the fluids at the
corresponding flow rates, it was possible to calculate Prandtl numbers, Reynolds
numbers, and Nusselt numbers for each case. Using these values made it possible to
calculate the heat transfer coefficient. Next, the heat transfer coefficients were used to
calculate resistances to heat transfer. There are three different resistances to heat transfer
that needed to be calculated – the resistance between the cold water and the exchanger
8
wall, the resistance as heat flows through the heat exchanger wall which separates the
two fluids, and the resistance between the hot coolant and the exchanger wall. Finally,
the resistances were used in a formula to find the overall amount of heat transferred.
2.3 Sample Equations
The following are thermodynamic property inputs based on the working fluids as
well as sample formulas used to perform the calculations described above:

min  3.0 gpm  (
3.0 gal 1min 3.7854 L 1000mL 1cm 3 1g
1kg
kg
)(
)(
)(
)(
)(
)(
)  0.18927
3
1min 60 sec
1gal
1L
1mL 1cm 1000 g
sec
3.0 gal 1min 0.003785m 3
1
m
v(
)(
)(
)(
)  1.253
2
min
60 sec
1gal
sec
0.000151m
Pr 
cp
k
(Prandtl number)
vDH 
Re 

(Reynolds number)
Nu  0.023(Re) 0.8 (Pr) 0.4 (Nusselt number)
hc , 2 
R1 
R2 
Nu  k
(Convective heat transfer coefficient of the engine coolant)
DH
1
hc,1 A

1
, which is the resistance to heat transfer of the cold water
hc,1 (circ )( L)
L
, which is the resistance to heat transfer through the heat exchanger wall which
kA
separates the two fluids
9
R3 
1
hc, 2 A

1
, which is the resistance to heat transfer of the hot fluid
hc, 2 (circ )( L)
Rtotal  R1  R2  R3

Q
Tavg
RTotal
, which is the average amount of heat transfer in Watts per unit length.
2.4 Alternate Method for Calculating Heat Transfer
These calculations give us a good idea of what we can expect in terms of
performance from several different heat exchanger designs. However, there is a more
accurate method to determine the amount of heat transferred within a heat exchanger
called the log mean temperature difference method, or LMTD. This method more
accurately calculates heat transfer by taking into account the change in temperature of
both fluids as they travel through the heat exchanger as opposed to simply using
thermodynamic properties at only the inlet and outlet temperatures. The counter-flow
design of this heat exchanger allows for a more constant, greater temperature difference,
which is much better calculated using the LMTD method.
2.5 LMTD Sample Equations
When using the LMTD method, there are a few alternate calculations that must be
performed in order to obtain the correct inputs. The values found for the different
properties of the working fluids may be carried over from the original method when using
10
this new method. However, examples of the new values and formulas needed when using
the LMTD method are described below:
Ta  TH ,in  TC ,out , which is the difference in temperatures between the hot fluid coming
into the system and the cold fluid going out of the system.
Tb  TH ,out  TC ,in , which is the difference in temperatures between the hot fluid going
out of the system and the cold fluid coming into the system.
U
1
ODtube
SAo ,tube * ln(
)
SAo ,tube
IDtube
1
(
)(
( )
SAi ,tube * hc , 2
2 *  * k Cu 2  * Ltube
hc ,1
, which is the overall heat transfer
coefficient of the system.
Q  U * SAo *
Ta  Tb
, which is the total amount of heat transferred in Watts.
Ta
ln(
)
Tb
No matter which method was chosen, it was still necessary to assume certain
values in order to perform the theoretical calculations. The most difficult assumptions
were the outlet temperatures of both working fluids that might be expected. These values
were needed in order to accurately calculate the different properties of these working
fluids, such as density, specific heat, dynamic viscosity, and thermal conductivity. Once
these values were found, it was possible to perform the intermediate calculations,
including finding the resistances to heat transfer and overall heat transfer coefficients.
11
Finally, these values made it possible to find the theoretical rates of heat transfer from the
cooling system of the vehicle to the shower system.
12
Chapter 3
TRIALS
3.1 Purpose
One component of this research project included conducting three experimental
trials. The main purpose for these trials was to verify that the system will function as
designed. The other reason for conducting the experiments was to compare the
theoretical calculations used in the project to a ‘real-world’ system and examine the
differences between the theoretical and actual systems. The primary values compared
were the rates of heat transfer between the two working fluids. Although it was required
to assume outlet temperatures when performing the theoretical calculations, these values
were physically measured during the experimental trials. Figure 3.1 below details the
heat exchanger for the system.
Figure 3.1 Heat Exchanger
Swivel Copper
Union
Adapter
Barbed Hose
End
Hose
Fitting/Adapter
13
3.2 Set-Up
The physical trials for this project were conducted using a 1981 Toyota SR5
pickup truck with a 144.4 cubic inch gasoline engine. The coolant side of the heat
exchanger was connected to the vehicle’s coolant system at the heater core inlet line by
two ‘T’-style fittings, one for the supply and the other for the return. Both fittings also
incorporated valve assemblies. These features allowed the heater inside the vehicle to be
operated without flowing coolant through the shower system heat exchanger. Hoses were
connected to the valve side of the ‘T’-style fittings and routed down to the heat
exchanger’s respective inlet and outlet fittings. However, in between the heat exchanger
inlet ‘T’-style fitting and the actual unit, a flow meter was mounted in order to monitor
the flow rate of the coolant running through the heat exchanger. Because the actual flow
of the engine cooling system was not initially known, several flow meters were purchased
in order to monitor the system. The final flow meter utilized had an operating
temperature range between 20 degrees below zero Fahrenheit and 240 degrees above zero
Fahrenheit. The meter was able to measure flow rates between 0.2 and 2 GPM.
The heat exchanger was strapped to the vehicle’s frame in between the frame and
the exhaust pipe spanning down the length of the pickup. This was found to be the
optimal location for the heat exchanger because the area is large enough to comfortably
hold the heat exchanger as well as allow for the inlet and outlet lines of both the coolant
and the fresh water to be connected without clearance issues. A bonus feature of this
location is that heat radiating from the exhaust pipe as gasses flow through it helped heat
14
the heat exchanger from the outside. Although the heat exchanger (when connected to
the vehicle for general use) will be insulated, there will still be a small amount of heat
loss to the environment. Locating the shower system’s heat exchanger near the exhaust
system will help reduce this loss at least on one side of the heat exchanger. While this
might not be an ideal location for all vehicles on which this system may be implemented,
most vehicles generally have a location similar to the experimental vehicle’s location
where it is possible to mount the heat exchanger. Figure 3.2 below is the actual test
vehicle and shower system set-up.
Figure 3.2 Test Vehicle Set-Up
Heater Core
Inlet
Heat Exchanger located
underneath truck cab
Coolant Supplied
From Engine
Heater Core
Outlet
Coolant Outlet
(Counter-flow Design)
Shower Heat
Exchanger Coolant Supply
Shower Heat
Exchanger Coolant Return
Fresh H2O Inlet
(used standard
spigot/hose for
ease of testing)
Located near the coolant outlet of the heat exchanger is the fresh water inlet
fitting. The outlet side of the marine-style bilge pump will be connected to this fitting via
an ordinary garden hose; however, for this experiment, a garden hose connected to a
water spigot was utilized not only for ease of testing, but also for the fact that use of this
15
system on the trail will require either an on-board water storage supply or some type of
body of water, such as a creek or lake, which was not readily available at the testing site.
The outlet fitting of the fresh water side is located in close proximity to the coolant inlet
side. For the experiment, water temperature was measured at this point using a K-type
thermal couple connected to an ordinary multi-meter display unit. For actual use, a
flexible shower line and head will be attached at this point in order to splay the fresh
water out for a more useful and enjoyable shower. The schematic below portrays the
basic system in a graphical context:
Figure 3.3 Shower System Schematic
‘T’-Style Fitting w/
Valve Assembly
Vehicle Heater
Core Inlet Line
Flow
Meter
Coolant
Inlet
Vehicle Heater
Core Inlet Line
Vehicle
Heater Core
Inlet Line
Coolant
Outlet
Marinestyle Bilge
Pump
Heat Exchanger
Heated Fresh
Water Out
Cold Fresh
Water In
Fresh
Water In
(pump)
16
3.3 Trials
Three different experiments were performed to test the actual shower system. For
each test, the fresh water flow rate was adjusted in order to gain a broad view of how the
system will perform across the entire design spectrum. Accordingly, the engine
revolutions were raised in order to increase both engine heat and coolant flow. Below is
a tabulated form of the results obtained during these three tests:
Vehicle Coolant
Corresponding
Vehicle Coolant
Vehicle Coolant
Fresh Water
Fresh H2O
Fresh H2O
Flow Rate
Engine
Temp, in
Temp, out
Flow Rate
Temp, in
Temp, out
(GPM)
RPM
(Deg. F)
(Deg. F)
(GPM)
(Deg. F)
(Deg. F)
0.45
900
163
154
1
67
71
0.92
2,000
179
168
2
67
73
1.20
3,000
182
175
3
68
74
Table 3.1 Testing Data
3.4 Results
After conducting the testing experiments, results obtained indicated differences
between the as-tested system and the theoretical system. It can be seen that there was a
transfer of energy (heat) between the two working fluids. The rates of heat transfer found
after conducting the three experiments were found to be roughly 1.1 kW, 2.4 kW, and 3.6
kW, respectively. Although these rates were much lower than expected, the system is
still able to use excess heat from the engine that would otherwise be wasted and radiated
to the environment.
17
3.5 Experiment Conclusions
The most notable differentiating feature was the coolant flow rate of the vehicle
through the shower system’s heat exchanger. The theoretical calculations had been
performed using the entire coolant system’s flow rate based on the coolant pump driven
by the accessory drive system of the vehicle’s engine, while the actual test set-up only
utilized a small portion of the actual coolant. This problem may be remedied by altering
the location where the shower system heat exchanger is tapped into the cooling system of
the engine.
Many various vehicle cooling system designs exist, with different vehicle
manufacturers employing their own designs according to their specific performance
criteria. The cooling system of this particular test vehicle utilizes a 5/8” diameter hose
spliced to the larger engine cooling inlet hose and is then routed up to the firewall of the
vehicle, where it is connected to a fitting that feeds into the heater core inside the cab of
the truck. This heater core is used as a small radiator that heats the interior of the vehicle
when the vehicle’s interior climate control is utilized. As shown in the system schematic
above, the shower system is spliced into the cooling system of the vehicle in between the
heater core inlet and the line routed from the engine inlet hose of the vehicle’s cooling
system. This allows vehicle coolant to flow into the shower system heat exchanger;
however, the difference in flow rates stems from the fact that the shower system was
tapped in at this splice. Because the full coolant flow does not flow through this 5/8”
spliced line, it is impossible for the full amount of coolant to flow through the shower
system heat exchanger.
18
There is another issue with connecting the shower system at the test location on
this particular vehicle. The coolant system of the vehicle is designed so that by this point,
coolant has already flowed through the main radiator in front of the engine. The purpose
of this primary radiator is to remove as much heat energy from the coolant system as
possible, and is usually paired with a fan. The fan increases air flow over the fins of this
air-to-liquid heat exchanger, greatly increasing its performance. While this is needed to
ensure the engine of the vehicle will run at its appropriate operating temperatures, the
loss of energy greatly affects the performance of the shower system. Temperatures
measured at the inlet of the shower system heat exchanger were much lower than
assumed when performing the theoretical energy transfer calculations, thus resulting in a
significantly lower amount of heat transferred inside the system.
The final potential problem that will have to be addressed for this design is the
overall workmanship when assembling the heat exchanger. The 3/8” O.D. tube that is
coiled inside the 2.5” O.D. shell must be twisted rather tightly, causing the soft copper
tube to flatten, which can potentially restrict the fresh water flow rates. Although there
were no problems observed during the experiments performed for this project, it is a real
possibility. The other workmanship-related problems may come about during the
soldering process. There are many different components that must be soldered together
in order to correctly assemble this heat exchanger, and they must be soldered in a specific
sequence. There are a couple fittings that must be soldered directly perpendicular to the
tangent point of the curved wall of the shell. This presents the highest level of difficulty
during the construction process of the heat exchanger because the wall thickness of the
19
shell is only 1/16”. If the soldering is not performed correctly and with care, the heat
exchanger will not work as designed.
20
Chapter 4
FINDINGS AND INTERPRETATIONS
Although the shower system did not perform exactly as predicted by the
parametric study, it did properly function as a tube-and-shell heat exchanger. The
relatively low levels of heat transfer between the two working fluids can be explained by
the cooler-than-expected inlet temperatures of the coolant flowing into the heat
exchanger. However, it was still possible to use the measurements taken during the
experiments and input them into the parametric study equations, thus calculating the
actual rates of heat transfer in the system during the trials.
While performing the parametric study, a very surprising result was produced
regarding the properties of the vehicle’s coolant. Initially, it was believed that the coolant
convective heat transfer coefficient would change in as linear progression, depending on
the flow rate and temperature. At a low temperature, and thus low flow rate, the
convective heat transfer coefficient was almost constant, with only a slightly positive
slope. When at a moderate temperature and flow rate, the convective heat transfer
coefficient varied at a much steeper, still linear, rate. At a high temperature and flow
rate, however, came the most unusual results. On the shorter end of the heat exchanger
tube size range, the heat transfer coefficient changed as expected, dependent on length of
tube. Unexpectedly, at the other end of the heat exchanger tube length range (the longer
end), the coolant convective heat transfer coefficient actually decreased. This clearly
21
was not a predicted result; however, it can possibly be due to the specific properties of
the 50/50 water ethylene mixture.
4.1 Recommendations
After comparing the theoretical calculations to the testing results, there are a
couple recommendations that should be made for future designs. These alternate design
features have the potential to offer rather large increases in both performance and
efficiency in return for relatively small increases in cost and mild design alterations. Any
of these suggested alterations, if attempted, should be tested to verify their effectiveness.
The first issue relates to one of the problems mentioned near the end of the
description of the experiments. The coolant line routed to the heat exchanger from the
vehicle coolant system must be larger in diameter than what was used in the experiments
in order to increase the coolant flow rate through the shower system heat exchanger.
Additionally, this line must be spliced to a larger vehicle coolant line instead of the 5/8”
line that is tapped off of the main line and runs up to the heater core on the inside of the
vehicle as done in the tests. The solution to this would be to locate a splice in between
the main coolant line, which is two inches in diameter. Coolant would be diverted to the
shower system heat exchanger, pass through, and then be routed back into the vehicle’s
cooling system. In order to perform these tasks, larger fittings with incorporated valves
must be used, which is significantly more expensive than the original design. This
solution adds to the cost of the system as well as increases the difficulty with the initial
system installation. Although the testing of the shower system was performed by using a
22
rather simple set-up, the simplicity must be reduced in order to accomplish the desired
results.
Another recommendation for future heat exchanger designs is to alter the size of
the tube coiled inside the shell. The maximum flow rate through the 3/8” O.D. tube,
despite having been misshapen due to the coiling procedure, greatly surpassed the design
criteria for this particular project. It would be possible to use a smaller diameter tube,
such as one measuring 1/4" O.D. This would still allow a sufficient amount of fresh
water to flow through the system while allowing the tube to keep its circular shape.
Using a smaller diameter tube will make it possible to increase the length of tube used
inside the heat exchanger, thus increasing the length of time the fresh water is exposed to
the heating process. Although the surface area per unit length of tube will be reduced
when using a smaller diameter tube, the actual amount of heat transfer may increase due
to the added length of tubing that is made possible to fit inside the shell of the heat
exchanger. However, this solution would require more testing, as any change will affect
the heat transfer rates inside the system.
One other physical design modification that can be applied to a future heat
exchanger was actually utilized for the second attempt of this project’s trial experiments
(the first attempt at construction of the heat exchanger failed due to inadequate materials).
Instead of simply soldering a reducer at the end of the shell, a coupler was soldered in its
place. The reducer was soldered at the other end of this coupling. The addition of the
coupling allowed for a wall thickness at these soldering points double that for which the
system was originally designed. This change provided a much more sufficient amount of
23
material to which the inlet and outlet fittings of the fresh water tube could be soldered.
This greatly improved ease of construction of the shower system heat exchanger.
Another recommendation that would improve the performance of the shower
system would be to wrap the inner tube around the exhaust pipe of the vehicle before it
enters the heat exchanger. Although this would require a significantly additional amount
of copper tubing, it could prove to be an overall benefit to the system. More analysis
would be required in order to validate this modification.
The final recommendation also stems from a problem mentioned at the end of the
description of the experiments. As previously mentioned, the shower system heat
exchanger is tapped into the coolant system of the vehicle at the point immediately before
the coolant flows into the heater core on the interior of the vehicle. The coolant system
of the vehicle is designed so that by this point, coolant has already flowed through the
main radiator in front of the engine. The vast majority of heat energy is dissipated
through this primary radiator, greatly reducing the temperature of the coolant. After heat
exchanger temperatures were recorded from the coolant inlet side, it was found that the
assumed temperatures used in the theoretical calculations were much greater than the
actual temperatures measured. This is a problem because the performance of the entire
shower system is based, among other factors, on the temperature difference between the
two working fluids.
The solution to this issue would be to not only use larger fittings as described
above, but to locate those fittings directly after the thermostat housing mounted at the top
of the engine. A thermostat in a vehicle is a simply a mechanical valve that is operated
24
via a temperature-dependent spring. Its location at the top of the engine is vital to the
performance and efficiency of the engine. It is here where the coolant, which has been
flowing through the engine’s water passages and providing a heat sink for the engine,
exits and flows back into the radiator to be cooled. Thus, it is at this point where the
temperature of the coolant will be at its highest level. Locating the splice point for the
shower system’s heat exchanger directly aft of this point will allow for the coolant at a
much higher temperature to flow through the heat exchanger. This temperature may be
as much as thirty degrees Fahrenheit hotter than the temperatures measured during the
original experiments.
It is obvious that this solution will increase the performance of the shower system.
The temperature difference between the two working fluids will be much larger, thus
greatly increasing the energy exchange inside the heat exchanger. While this outcome
would be the desired result of this recommended design change, another less-obvious
result stems from this alteration. After the coolant is routed through the shower system
heat exchanger, it is then looped back into the vehicle coolant system and continues on to
the primary radiator at the front of the vehicle. If the insulation around the shower
system heat exchanger is removed, allowing heat to be radiated out, the coolant will have
already undergone a substantial decrease in temperature. In this case, the heat exchanger
is operating as a simple fluid-to-air radiator. Even more energy will be pulled from the
vehicle coolant system as it passes through the main radiator. The extra energy pulled
out of the system will help keep the vehicle’s engine temperature lower than it would be
if there was no shower system.
25
This is a very useful feature of the shower system that was originally overlooked,
as it was not one of the desired outcomes. A common problem plaguing drivers and their
vehicles when four-wheeling is an overheating engine. This is caused by the fact that in
most cases, driving on an off-road trail is done at extremely low speeds. The vehicles are
driven over rather large boulders and fallen trees, and this must be done very slowly so as
not to damage the vehicle as well as create a semi-comfortable ride for the driver and any
passengers. These slow speeds cause airflow through the fins of the primary radiator due
to the velocity of the vehicle to become practically negligible.
In most cases, slow vehicle speeds translate to low vehicle engine RPMs. Most
vehicles, especially those that are typically used on off-road excursions, originally come
equipped with mechanically operated fans from the manufacturer, which are driven by
the accessory drive system of the engine. Some people have remedied engine
overheating issues by replacing the stock manufacturer’s mechanically operated fan with
an electrically driven fan. While effective, this is sometimes a rather expensive solution
to an overheating engine in an off-road vehicle. Once the new design features are
incorporated into this shower system, increasing its heat transfer efficiency, it may prove
to be a viable alternative to an electric fan. The added benefit over an electrically
operated fan will be the ability to use the shower as designed in order to obtain heated
water.
The parametric study has proven to be a valuable resource for the design of this
vehicle shower system. Although the desired results based on the parametric study were
not achieved, it highlights the areas where the actual test system is inferior to the ideal
26
case. Changes can now be implemented on the original test set-up, which can then be reevaluated. It is expected that the modified system will show improvements in
efficiencies and effectiveness over the original design.
27
APPENDIX A
Parametric Study Figures
Overall Heat Transferred vs. Temperature Difference between Working Fluids
Coolant Flow = 0.1321 GPM; T,in(Coolant) = 190 Deg. F; H2O Flow Rate = 1.5 GPM
1100.00
y = 9.0196x - 278.29
1000.00
900.00
y = 7.1478x - 220.76
Heat Transferred (W)
800.00
Tube Length = 6.096 m
Tube Length = 6.858 m
Tube Length = 7.620 m
Tube Length = 8.382 m
Tube Length = 9.144 m
700.00
y = 5.4731x - 169.27
600.00
500.00
y = 4.0211x - 124.57
400.00
y = 2.9874x - 92.641
300.00
200.00
115
120
125
130
135
140
145
Working Fluid Temperature Difference (Deg. F)
Overall Heat Transferred vs. Temperature Difference between Working Fluids
Coolant Flow = 4.623 GPM; T,in(coolant) = 200 Deg. F; H2O Flow Rate = 1.5 GPM
14000.00
y = 104.27x - 2168.6
12000.00
y = 83.41x - 1770.9
Heat Transferred, (W)
10000.00
Tube Length = 6.096 m
Tube Length = 6.858 m
Tube Length = 7.620 m
Tube Length = 8.382 m
Tube Length = 9.144 m
y = 64.674x - 1410.4
8000.00
y = 48.259x - 1086.7
6000.00
y = 36.185x - 829.98
4000.00
2000.00
115
120
125
130
135
Temperature Difference, Deg. F
140
145
28
Overall Heat Transferred vs. Temperature Difference between Working Fluids
Coolant Flow = 7.925 GPM; T,in (Coolant) = 210 Deg. F; H2O Flow = 1.5 GPM
22500.00
y = 131.17x + 593.08
20000.00
17500.00
Heat Transferred (W)
y = 105.48x + 377.3
15000.00
Tube Length = 6.096 m
Tube Length = 6.858 m
Tube Length = 7.620 m
Tube Length = 8.382 m
Tube Length = 9.144 m
y = 82.366x + 191.02
12500.00
y = 62.007x + 47.797
10000.00
y = 46.74x - 6.6977
7500.00
5000.00
125
130
135
140
145
150
155
Working Fluid Temperature Difference (Deg. F)
Overall Heat Transferred vs. Temperature Difference between Working Fluids
Coolant Flow Rate = 0.1231; T,in (Coolant) = 190 Deg. F; H2O Flow Rate = 2.0 GPM
1100
y = 9.0719x - 282.03
1000
900
y = 7.188x - 223.65
Heat Transferred (W)
800
Pipe Length = 6.096 m
Pipe Length = 6.858 m
Pipe Length = 7.620 m
Pipe Length = 8.382 m
Pipe Length = 9.144 m
700
y = 5.5027x - 171.4
600
500
y = 4.0418x - 126.06
400
y = 3.0023x - 93.713
300
200
115
120
125
130
135
Temperature Difference (Deg. F)
140
145
29
Overall Heat Transferred vs. Temperature Difference between Working Fluids
Coolant Flow Rate = 4.623; T,in (Coolant) = 200 Deg. F; H2O Flow Rate = 2.0 GPM
16000.00
y = 112.88x - 1860.2
14000.00
12000.00
Heat Transferred (W)
y = 90.179x - 1523.3
Tube Length = 6.096 m
Tube Length = 6.868 m
Tube Length = 7.620 m
Tube Length = 8.382 m
Tube Length = 9.144 m
10000.00
y = 69.799x - 1217.3
8000.00
y = 51.969x - 941.35
6000.00
y = 38.915x - 720.33
4000.00
2000.00
115
120
125
130
135
140
145
Temperature Difference (Deg. F)
Overall Heat Transferred vs. Temperature Difference between Working Fluids
Coolant Flow Rate = 7.925; T,in (Coolant) = 210 Deg. F; H2O Flow Rate = 2.0 GPM
25000.00
y = 147.33x - 364.42
22500.00
Heat Transferred (W)
20000.00
y = 118.21x - 384.21
17500.00
Tube Length = 6.096 m
Tube Length = 6.858 m
Tube Length = 7.620 m
Tube Length = 8.382 m
Tube Length = 9.144 m
15000.00
y = 92.027x - 394.2
12500.00
y = 69.017x - 383.3
10000.00
y = 51.905x - 327.14
7500.00
5000.00
125
130
135
140
145
Temperature Difference (Deg. F)
150
155
30
Overall Heat Transferred vs. Temperature Difference between Working Fluids
Coolant Flow Rate = 0.1321; T,in (Coolant) = 190 Deg. F; H2O Flow Rate = 2.5 GPM
1200.00
y = 9.1051x - 284.41
Heat Transferred (W)
1000.00
y = 7.2137x - 225.49
800.00
Tube Length = 6.096 m
Tube Length = 6.858 m
Tube Length = 7.620 m
Tube Length = 8.382 m
Tube Length = 9.144 m
y = 5.5216x - 172.75
600.00
y = 4.055x - 127.01
400.00
y = 3.0118x - 94.396
200.00
115
120
125
130
135
140
145
Temperature Difference (Deg. F)
Overall Heat Transferred vs. Temperature Difference between Working Fluids
Coolant Flow Rate = 4.623; T,in (Coolant) = 200 Deg. F; H2O Flow Rate = 2.5 GPM
16000.00
y = 118.93x - 2246.6
14000.00
y = 94.899x - 1826.1
Heat Transferred (W)
12000.00
Tube Length = 6.096 m
Tube Length = 6.858 m
Tube Length = 7.620
Tube Length = 8.382 m
Tube Length = 9.144 m
10000.00
y = 73.335x - 1445.4
8000.00
y = 54.494x - 1105.3
6000.00
y = 40.757x - 840.43
4000.00
2000.00
115
120
125
130
135
Temperature Difference (Deg. F)
140
145
31
Overall Heat Transferred vs. Temperature Difference between Working Fluids
Coolant Flow Rate = 7.925; T,in (Coolant) = 210 Deg. F; H2O Flow Rate = 2.5 GPM
25000.00
y = 161.9x - 1285.5
22500.00
20000.00
Heat Transferred (W)
y = 129.7x - 1115.4
17500.00
Tube Length = 6.096 m
Tube Length = 6.858 m
Tube Length = 7.620 m
Tube Length = 8.382 m
Tube Length = 9.144 m
y = 100.77x - 954.98
15000.00
12500.00
y = 75.376x - 795.56
10000.00
y = 56.599x - 633.31
7500.00
5000.00
125
130
135
140
145
150
155
Temperature Difference (Deg. F)
Overall Heat Transferred vs. Temperature Difference between Working Fluids
Coolant Flow Rate = 0.1321; T,in (Coolant) = 190 Deg. F; H2O Flow Rate = 3.0 GPM
1200.00
y = 9.1283x - 286.08
Heat Transferred (W)
1000.00
y = 7.2315x - 226.77
800.00
Tube Length =
Tube Length =
Tube Length =
Tube Length =
Tube Length =
y = 5.5347x - 173.7
600.00
y = 4.0642x - 127.67
400.00
y = 3.0185x - 94.873
200.00
115
120
125
130
135
Temperature Difference (Deg. F)
140
145
6.096 m
6.858 m
7.620 m
8.382 m
9.144 m
32
Overall Heat Transferred vs. Temperature Difference between Working Fluids
Coolant Flow Rate = 4.623; T,in (Coolant) = 200 Deg. F; H2O Flow Rate = 3.0 GPM
16000.00
y = 123.45x - 2542.3
13500.00
Heat Transferred (W)
y = 98.425x - 2057.2
11000.00
Tube Length =
Tube Length =
Tube Length =
Tube Length =
Tube Length =
y = 75.971x - 1618.9
8500.00
6.096 m
6.858 m
7.620 m
8.382 m
9.144 m
y = 56.369x - 1229.5
6000.00
y = 42.124x - 931.22
3500.00
115
120
125
130
135
140
145
Temperature Difference (Deg. F)
Overall Heat Transferred vs. Temperature Difference between Working Fluids
Coolant Flow Rate = 7.925; T,in (Coolant) = 210 Deg. F; H2O Flow Rate = 3.0 GPM
25000.00
y = 168.56x - 1726.2
20000.00
Heat Transferred (W)
y = 134.87x - 1459.8
Tube Length = 6.096 m
Tube Length = 6.858 m
Tube Length = 7.620 m
Tube Length = 8.382 m
Tube Length = 9.144 m
y = 104.6x - 1213.2
15000.00
y = 78.075x - 979.8
10000.00
y = 58.551x - 767.57
5000.00
125
130
135
140
145
Temperature Difference (Deg. F)
150
155
33
Convective Heat Transfer Coefficient vs. Fresh Water Inlet Temperature
H2O Flow Rate = 1.5 GPM
55000.00
y = 362.6x + 30211
Convective Heat Transfer Coefficient (W/m^2-K)
50000.00
y = 332.39x + 27693
y = 302.17x + 25176
45000.00
Tube Length = 6.096 m
Tube Length = 6.858 m
Tube Length = 7.620 m
Tube Length = 8.382 m
Tube Length = 9.144 m
y = 271.95x + 22658
40000.00
y = 241.73x + 20141
35000.00
30000.00
25000.00
35
40
45
50
55
60
65
70
Fresh Water Inlet Temperature, Deg. F
Convective Heat Transfer Coefficient vs. Fresh Water Inlet Temperature
H2O Flow Rate = 2.0 GPM
70000.00
y = 456.44x + 38029
Convective Heat Transfer Coefficient (W/K)
65000.00
y = 418.4x + 34860
60000.00
y = 380.36x + 31691
Tube Length = 6.096 m
Tube Length = 6.858 m
Tube Length = 7.620 m
Tube Length = 8.382 m
Tube Length = 9.144 m
55000.00
y = 342.33x + 28522
50000.00
y = 304.29x + 25353
45000.00
40000.00
35000.00
35
40
45
50
55
Fresh Water Inlet Temperature (Deg. F)
60
65
70
34
Convective Heat Transfer Coefficient vs. Fresh Water Inlet Temperature
H2O Flow Rate = 2.5 GPM
85000.00
y = 545.64x + 45461
Convective Heat Transfer Coefficient (W/K)
77500.00
y = 500.17x + 41673
70000.00
y = 454.7x + 37884
62500.00
Tube Length = 6.096 m
Tube Length = 6.858 m
Tube Length = 7.620 m
Tube Length = 8.382 m
Tube Length = 9.144 m
y = 409.23x + 34096
55000.00
y = 363.76x + 30307
47500.00
40000.00
35
40
50
45
60
55
65
70
Fresh Water Inlet Temperature (Deg. F)
Convective Heat Transfer Coefficient vs. Fresh Water Inlet Temperature
H2O Flow Rate = 3.0 GPM
100000.00
Convective Heat Transfer Coefficient (W/K)
y = 631.33x + 52600
90000.00
y = 578.72x + 48217
80000.00
Tube Length = 6.096 m
Tube Length = 6.858
Tube Length = 7.620 m
Tube Length = 8.382 m
Tube Length = 9.144 m
y = 526.11x + 43833
y = 473.5x + 39450
70000.00
y = 420.88x + 35067
60000.00
50000.00
35
40
45
50
55
60
Fresh Water Inlet Temperature (Deg. F)
65
70
35
Coolant Convective H.T. Coefficient vs. Coolant Outlet Temperature
2500.00
Convective Heat Transfer Coefficient (W/m^2 - K)
y = -0.9171x2 + 367.7x - 34695
2000.00
1500.00
y = 4.1988x + 560.14
Flow = 0.1321 GPM, T,in = 190 Deg. F
Flow = 4.623 GPM, T,in = 200 Deg. F
Flow = 7.925 GPM, T,in = 210 Deg. F
1000.00
500.00
y = 0.1737x + 44.851
0.00
135
145
155
165
175
185
195
205
215
T,out (Deg. F)
Overall Heat Transferred vs. Temperature Difference between Working Fluids
Coolant Flow = 0.1321 GPM; T,in(Coolant) = 190 Deg. F; H2O Flow Rate = 1.5 GPM
1100.00
y = 9.0196x - 278.29
1000.00
900.00
y = 7.1478x - 220.76
Heat Transferred (W)
800.00
Tube Length = 6.096 m
Tube Length = 6.858 m
Tube Length = 7.620 m
Tube Length = 8.382 m
Tube Length = 9.144 m
700.00
y = 5.4731x - 169.27
600.00
500.00
y = 4.0211x - 124.57
400.00
y = 2.9874x - 92.641
300.00
200.00
115
120
125
130
135
Working Fluid Temperature Difference (Deg. F)
140
145
36
Overall Heat Transferred vs. Temperature Difference between Working Fluids
Coolant Flow = 4.623 GPM; T,in(coolant) = 200 Deg. F; H2O Flow Rate = 1.5 GPM
14000.00
y = 104.27x - 2168.6
12000.00
y = 83.41x - 1770.9
Heat Transferred, (W)
10000.00
Tube Length = 6.096 m
Tube Length = 6.858 m
Tube Length = 7.620 m
Tube Length = 8.382 m
Tube Length = 9.144 m
y = 64.674x - 1410.4
8000.00
y = 48.259x - 1086.7
6000.00
y = 36.185x - 829.98
4000.00
2000.00
115
120
125
130
135
140
145
Temperature Difference, Deg. F
Overall Heat Transferred vs. Temperature Difference between Working Fluids
Coolant Flow = 7.925 GPM; T,in (Coolant) = 210 Deg. F; H2O Flow = 1.5 GPM
22500.00
y = 131.17x + 593.08
20000.00
17500.00
Heat Transferred (W)
y = 105.48x + 377.3
15000.00
Tube Length = 6.096 m
Tube Length = 6.858 m
Tube Length = 7.620 m
Tube Length = 8.382 m
Tube Length = 9.144 m
y = 82.366x + 191.02
12500.00
y = 62.007x + 47.797
10000.00
y = 46.74x - 6.6977
7500.00
5000.00
125
130
135
140
145
Working Fluid Temperature Difference (Deg. F)
150
155
37
Overall Heat Transferred vs. Temperature Difference between Working Fluids
Coolant Flow Rate = 0.1231; T,in (Coolant) = 190 Deg. F; H2O Flow Rate = 2.0 GPM
1100
y = 9.0719x - 282.03
1000
900
y = 7.188x - 223.65
Heat Transferred (W)
800
Pipe Length = 6.096 m
Pipe Length = 6.858 m
Pipe Length = 7.620 m
Pipe Length = 8.382 m
Pipe Length = 9.144 m
700
y = 5.5027x - 171.4
600
500
y = 4.0418x - 126.06
400
y = 3.0023x - 93.713
300
200
115
120
125
130
135
140
145
Temperature Difference (Deg. F)
Overall Heat Transferred vs. Temperature Difference between Working Fluids
Coolant Flow Rate = 4.623; T,in (Coolant) = 200 Deg. F; H2O Flow Rate = 2.0 GPM
16000.00
y = 112.88x - 1860.2
14000.00
12000.00
Heat Transferred (W)
y = 90.179x - 1523.3
Tube Length = 6.096 m
Tube Length = 6.868 m
Tube Length = 7.620 m
Tube Length = 8.382 m
Tube Length = 9.144 m
10000.00
y = 69.799x - 1217.3
8000.00
y = 51.969x - 941.35
6000.00
y = 38.915x - 720.33
4000.00
2000.00
115
120
125
130
135
Temperature Difference (Deg. F)
140
145
38
Overall Heat Transferred vs. Temperature Difference between Working Fluids
Coolant Flow Rate = 7.925; T,in (Coolant) = 210 Deg. F; H2O Flow Rate = 2.0 GPM
25000.00
y = 147.33x - 364.42
22500.00
Heat Transferred (W)
20000.00
y = 118.21x - 384.21
17500.00
Tube Length = 6.096 m
Tube Length = 6.858 m
Tube Length = 7.620 m
Tube Length = 8.382 m
Tube Length = 9.144 m
15000.00
y = 92.027x - 394.2
12500.00
y = 69.017x - 383.3
10000.00
y = 51.905x - 327.14
7500.00
5000.00
125
130
135
140
145
150
155
Temperature Difference (Deg. F)
Overall Heat Transferred vs. Temperature Difference between Working Fluids
Coolant Flow Rate = 0.1321; T,in (Coolant) = 190 Deg. F; H2O Flow Rate = 2.5 GPM
1200.00
y = 9.1051x - 284.41
Heat Transferred (W)
1000.00
y = 7.2137x - 225.49
800.00
Tube Length = 6.096 m
Tube Length = 6.858 m
Tube Length = 7.620 m
Tube Length = 8.382 m
Tube Length = 9.144 m
y = 5.5216x - 172.75
600.00
y = 4.055x - 127.01
400.00
y = 3.0118x - 94.396
200.00
115
120
125
130
135
Temperature Difference (Deg. F)
140
145
39
Overall Heat Transferred vs. Temperature Difference between Working Fluids
Coolant Flow Rate = 4.623; T,in (Coolant) = 200 Deg. F; H2O Flow Rate = 2.5 GPM
16000.00
y = 118.93x - 2246.6
14000.00
y = 94.899x - 1826.1
Heat Transferred (W)
12000.00
Tube Length = 6.096 m
Tube Length = 6.858 m
Tube Length = 7.620
Tube Length = 8.382 m
Tube Length = 9.144 m
10000.00
y = 73.335x - 1445.4
8000.00
y = 54.494x - 1105.3
6000.00
y = 40.757x - 840.43
4000.00
2000.00
115
120
125
130
135
140
145
Temperature Difference (Deg. F)
Overall Heat Transferred vs. Temperature Difference between Working Fluids
Coolant Flow Rate = 7.925; T,in (Coolant) = 210 Deg. F; H2O Flow Rate = 2.5 GPM
25000.00
y = 161.9x - 1285.5
22500.00
20000.00
Heat Transferred (W)
y = 129.7x - 1115.4
17500.00
Tube Length = 6.096 m
Tube Length = 6.858 m
Tube Length = 7.620 m
Tube Length = 8.382 m
Tube Length = 9.144 m
y = 100.77x - 954.98
15000.00
12500.00
y = 75.376x - 795.56
10000.00
y = 56.599x - 633.31
7500.00
5000.00
125
130
135
140
145
Temperature Difference (Deg. F)
150
155
40
Overall Heat Transferred vs. Temperature Difference between Working Fluids
Coolant Flow Rate = 0.1321; T,in (Coolant) = 190 Deg. F; H2O Flow Rate = 3.0 GPM
1200.00
y = 9.1283x - 286.08
Heat Transferred (W)
1000.00
y = 7.2315x - 226.77
800.00
Tube Length = 6.096 m
Tube Length = 6.858 m
Tube Length = 7.620 m
Tube Length = 8.382 m
Tube Length = 9.144 m
y = 5.5347x - 173.7
600.00
y = 4.0642x - 127.67
400.00
y = 3.0185x - 94.873
200.00
115
120
125
130
135
140
145
Temperature Difference (Deg. F)
Overall Heat Transferred vs. Temperature Difference between Working Fluids
Coolant Flow Rate = 4.623; T,in (Coolant) = 200 Deg. F; H2O Flow Rate = 3.0 GPM
16000.00
y = 123.45x - 2542.3
13500.00
Heat Transferred (W)
y = 98.425x - 2057.2
11000.00
Tube Length = 6.096 m
Tube Length = 6.858 m
Tube Length = 7.620 m
Tube Length = 8.382 m
Tube Length = 9.144 m
y = 75.971x - 1618.9
8500.00
y = 56.369x - 1229.5
6000.00
y = 42.124x - 931.22
3500.00
115
120
125
130
135
Temperature Difference (Deg. F)
140
145
41
Overall Heat Transferred vs. Temperature Difference between Working Fluids
Coolant Flow Rate = 7.925; T,in (Coolant) = 210 Deg. F; H2O Flow Rate = 3.0 GPM
25000.00
y = 168.56x - 1726.2
20000.00
Heat Transferred (W)
y = 134.87x - 1459.8
Tube Length = 6.096 m
Tube Length = 6.858 m
Tube Length = 7.620 m
Tube Length = 8.382 m
Tube Length = 9.144 m
y = 104.6x - 1213.2
15000.00
y = 78.075x - 979.8
10000.00
y = 58.551x - 767.57
5000.00
125
130
135
140
145
150
155
Temperature Difference (Deg. F)
Convective Heat Transfer Coefficient vs. Fresh Water Inlet Temperature
H2O Flow Rate = 1.5 GPM
55000.00
y = 362.6x + 30211
Convective Heat Transfer Coefficient (W/m^2-K)
50000.00
y = 332.39x + 27693
y = 302.17x + 25176
45000.00
Tube Length = 6.096 m
Tube Length = 6.858 m
Tube Length = 7.620 m
Tube Length = 8.382 m
Tube Length = 9.144 m
y = 271.95x + 22658
40000.00
y = 241.73x + 20141
35000.00
30000.00
25000.00
35
40
45
50
55
Fresh Water Inlet Temperature, Deg. F
60
65
70
42
Convective Heat Transfer Coefficient vs. Fresh Water Inlet Temperature
H2O Flow Rate = 2.0 GPM
70000.00
y = 456.44x + 38029
Convective Heat Transfer Coefficient (W/K)
65000.00
y = 418.4x + 34860
60000.00
y = 380.36x + 31691
Tube Length = 6.096 m
Tube Length = 6.858 m
Tube Length = 7.620 m
Tube Length = 8.382 m
Tube Length = 9.144 m
55000.00
y = 342.33x + 28522
50000.00
y = 304.29x + 25353
45000.00
40000.00
35000.00
35
40
45
50
55
60
65
70
Fresh Water Inlet Temperature (Deg. F)
Convective Heat Transfer Coefficient vs. Fresh Water Inlet Temperature
H2O Flow Rate = 2.5 GPM
85000.00
y = 545.64x + 45461
Convective Heat Transfer Coefficient (W/K)
77500.00
y = 500.17x + 41673
70000.00
y = 454.7x + 37884
62500.00
Tube Length = 6.096 m
Tube Length = 6.858 m
Tube Length = 7.620 m
Tube Length = 8.382 m
Tube Length = 9.144 m
y = 409.23x + 34096
55000.00
y = 363.76x + 30307
47500.00
40000.00
35
40
45
50
55
Fresh Water Inlet Temperature (Deg. F)
60
65
70
43
Convective Heat Transfer Coefficient vs. Fresh Water Inlet Temperature
H2O Flow Rate = 3.0 GPM
100000.00
Convective Heat Transfer Coefficient (W/K)
y = 631.33x + 52600
90000.00
y = 578.72x + 48217
80000.00
Tube Length = 6.096 m
Tube Length = 6.858
Tube Length = 7.620 m
Tube Length = 8.382 m
Tube Length = 9.144 m
y = 526.11x + 43833
y = 473.5x + 39450
70000.00
y = 420.88x + 35067
60000.00
50000.00
35
40
45
50
55
60
65
70
Fresh Water Inlet Temperature (Deg. F)
Coolant Convective H.T. Coefficient vs. Coolant Outlet Temperature
2500.00
Convective Heat Transfer Coefficient (W/m^2 - K)
y = -0.9171x2 + 367.7x - 34695
2000.00
1500.00
y = 4.1988x + 560.14
Flow = 0.1321 GPM, T,in = 190 Deg. F
Flow = 4.623 GPM, T,in = 200 Deg. F
Flow = 7.925 GPM, T,in = 210 Deg. F
1000.00
500.00
y = 0.1737x + 44.851
0.00
135
145
155
165
175
T,out (Deg. F)
185
195
205
215
44
BIBLIOGRAPHY
Carley, Larry. Underhood Service, April 1999. “Radiator Overheating Causes and
Cures”
<http://www.arrowheadradiator.com/overheating_causes_and_cures.htm>
<http://www.chevyhiperformance.com/index.html>
“Cooling Systems.” © Grape Ape Racing
<http://www.grapeaperacing.com/tech/coolingsystems.pdf>
“Electrolysis: The Silent Killer”
<http://www.sancarlosradiator.com/electrolysis.htm>
Ethylene Glycol Product Guide. © The MEGlobal Group of Companies.
Heo, Hyung Seok; Park, Kyoung Suk; Won, Jong Phil. “Thermal Flow Analysis of
Vehicle Engine Cooling System.” © 2003 NuriMedia Co., Ltd.
Ofria, Charles. “Automotive Cooling Systems”
<http://www.familycar.com/Classroom/CoolingSystem.htm>
“Radiator Repair and Replacement.” © AA1Car
<http://www.aa1car.com/library/radiator_repair.htm>