MOBILE VEHICLE SHOWER SYSTEM

(1)

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

(2)

ii

A Project

by

Joshua Paul Perron

Approved by:

__________________________________, Committee Chair Dr. Timothy Marbach

____________________________

(3)

iii

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 Date

(4)

iv

Abstract

of

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.

(5)

v

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

_______________________

(6)

vi

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.

(7)

vii

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

(8)

viii

LIST OF TABLES

Page

(9)

ix

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

(10)

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 four-wheeling 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

(11)

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.

(12)

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

(13)

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

(14)

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.

(15)

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.

(16)

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

(17)

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:

sec 18927 . 0 ) 1000 1 )( 1 1 )( 1 1 )( 1 1000 )( 1 7854 . 3 )( sec 60 min 1 )( min 1 0 . 3 ( 0 . 3 ₃ 3 kg g kg cm g mL cm L mL gal L gal gpm m_in     sec 253 . 1 ) 000151 . 0 1 )( 1 003785 . 0 )( sec 60 min 1 )( min 0 . 3 ( ₂ 3 m m gal m gal v  k c_p  Pr (Prandtl number)   H vD  Re (Reynolds number) 4 . 0 8 . 0 (Pr) (Re) 023 . 0  Nu (Nusselt number) H c D k Nu

h_,₂   (Convective heat transfer coefficient of the engine coolant)

) )( ( 1 1 1 , 1 , 1 L circ h A h R c c 

 , which is the resistance to heat transfer of the cold water

kA L

R₂  , which is the resistance to heat transfer through the heat exchanger wall which separates the two fluids

(18)

) )( ( 1 1 2 , 2 , 3 L circ h A h R c c 

 , which is the resistance to heat transfer of the hot fluid

3 2 1 R R R Rtotal    Total avg R T Q  

, 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

(19)

this new method. However, examples of the new values and formulas needed when using the LMTD method are described below:

out C in H a T T T  _,  _,

 , which is the difference in temperatures between the hot fluid coming into the system and the cold fluid going out of the system.

in C out H b T T T  ,  ,

 , which is the difference in temperatures between the hot fluid going out of the system and the cold fluid coming into the system.

) 1 ( * * * 2 ) ln( * ( ) * ( 1 1 , , 2 , , , 2 tube c Cu tube tube tube o c tube i tube o h L k ID OD SA h SA SA U     

, which is the overall heat transfer

coefficient of the system.

) ln( * * b a b a o T T T T SA U Q     

 , which is the total amount of heat transferred in Watts.

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,

(20)

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.

(21)

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

Barbed Hose End Swivel Copper Union Adapter Hose Fitting/Adapter

(22)

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

(23)

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

Heat Exchanger located underneath truck cab

Fresh H2O Inlet (used standard spigot/hose for ease of testing) Coolant Outlet (Counter-flow Design) Heater Core Inlet Shower Heat Exchanger Coolant Return Shower Heat

Exchanger Coolant Supply

Coolant Supplied From Engine

Heater Core Outlet

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

(24)

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 Vehicle Heater Core Inlet Line

Heat Exchanger Coolant Inlet Coolant Outlet Marine-style Bilge Pump Heated Fresh Water Out Cold Fresh Water In Fresh Water In (pump) Flow Meter Vehicle Heater

(25)

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 Flow Rate (GPM) Corresponding Engine RPM Vehicle Coolant Temp, in (Deg. F) Vehicle Coolant Temp, out (Deg. F) Fresh Water Flow Rate (GPM) Fresh H2O Temp, in (Deg. F) Fresh H2O Temp, out (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.

(26)

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.

(27)

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

(28)

shell is only 1/16”. If the soldering is not performed correctly and with care, the heat exchanger will not work as designed.

(29)

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

(30)

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

(31)

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

(32)

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

(33)

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.

(34)

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

(35)

case. Changes can now be implemented on the original test set-up, which can then be re-evaluated. It is expected that the modified system will show improvements in

(36)

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

y = 9.0196x - 278.29 y = 7.1478x - 220.76 y = 5.4731x - 169.27 y = 4.0211x - 124.57 y = 2.9874x - 92.641 200.00 300.00 400.00 500.00 600.00 700.00 800.00 900.00 1000.00 1100.00 115 120 125 130 135 140 145

Working Fluid Temperature Difference (Deg. F)

H e a t T ra n s fe rr e d (W ) 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

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

y = 104.27x - 2168.6 y = 83.41x - 1770.9 y = 64.674x - 1410.4 y = 48.259x - 1086.7 y = 36.185x - 829.98 2000.00 4000.00 6000.00 8000.00 10000.00 12000.00 14000.00 115 120 125 130 135 140 145

Temperature Difference, Deg. F

H e a t T ra n s fe rr e d , (W ) 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

(37)

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

y = 131.17x + 593.08 y = 105.48x + 377.3 y = 82.366x + 191.02 y = 62.007x + 47.797 y = 46.74x - 6.6977 5000.00 7500.00 10000.00 12500.00 15000.00 17500.00 20000.00 22500.00 125 130 135 140 145 150 155

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

y = 9.0719x - 282.03 y = 7.188x - 223.65 y = 5.5027x - 171.4 y = 4.0418x - 126.06 y = 3.0023x - 93.713 200 300 400 500 600 700 800 900 1000 1100 115 120 125 130 135 140 145

Temperature Difference (Deg. F)

H e a t T ra n s fe rr e d (W ) 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

(38)

y = 112.88x - 1860.2 y = 90.179x - 1523.3 y = 69.799x - 1217.3 y = 51.969x - 941.35 y = 38.915x - 720.33 2000.00 4000.00 6000.00 8000.00 10000.00 12000.00 14000.00 16000.00 115 120 125 130 135 140 145

y = 147.33x - 364.42 y = 118.21x - 384.21 y = 92.027x - 394.2 y = 69.017x - 383.3 y = 51.905x - 327.14 5000.00 7500.00 10000.00 12500.00 15000.00 17500.00 20000.00 22500.00 25000.00 125 130 135 140 145 150 155

(39)

y = 3.0118x - 94.396 y = 4.055x - 127.01 y = 5.5216x - 172.75 y = 7.2137x - 225.49 y = 9.1051x - 284.41 200.00 400.00 600.00 800.00 1000.00 1200.00 115 120 125 130 135 140 145

y = 118.93x - 2246.6 y = 94.899x - 1826.1 y = 73.335x - 1445.4 y = 54.494x - 1105.3 y = 40.757x - 840.43 2000.00 4000.00 6000.00 8000.00 10000.00 12000.00 14000.00 16000.00 115 120 125 130 135 140 145

H e a t T ra n s fe rr e d (W ) Tube Length = 6.096 m Tube Length = 6.858 m Tube Length = 7.620 Tube Length = 8.382 m Tube Length = 9.144 m

(40)

y = 56.599x - 633.31 y = 75.376x - 795.56 y = 100.77x - 954.98 y = 129.7x - 1115.4 y = 161.9x - 1285.5 5000.00 7500.00 10000.00 12500.00 15000.00 17500.00 20000.00 22500.00 25000.00 125 130 135 140 145 150 155

y = 3.0185x - 94.873 y = 4.0642x - 127.67 y = 5.5347x - 173.7 y = 7.2315x - 226.77 y = 9.1283x - 286.08 200.00 400.00 600.00 800.00 1000.00 1200.00 115 120 125 130 135 140 145

(41)

y = 42.124x - 931.22 y = 56.369x - 1229.5 y = 75.971x - 1618.9 y = 98.425x - 2057.2 y = 123.45x - 2542.3 3500.00 6000.00 8500.00 11000.00 13500.00 16000.00 115 120 125 130 135 140 145

y = 58.551x - 767.57 y = 78.075x - 979.8 y = 104.6x - 1213.2 y = 134.87x - 1459.8 y = 168.56x - 1726.2 5000.00 10000.00 15000.00 20000.00 25000.00 125 130 135 140 145 150 155

(42)

Convective Heat Transfer Coefficient vs. Fresh Water Inlet Temperature H2O Flow Rate = 1.5 GPM

y = 362.6x + 30211 y = 332.39x + 27693 y = 302.17x + 25176 y = 271.95x + 22658 y = 241.73x + 20141 25000.00 30000.00 35000.00 40000.00 45000.00 50000.00 55000.00 35 40 45 50 55 60 65 70

Fresh Water Inlet Temperature, Deg. F

C o n ve ct ive H ea t T ran sf er C o ef fi ci en t (W/ m ^ 2-K ) 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 = 304.29x + 25353 y = 342.33x + 28522 y = 380.36x + 31691 y = 418.4x + 34860 y = 456.44x + 38029 35000.00 40000.00 45000.00 50000.00 55000.00 60000.00 65000.00 70000.00 35 40 45 50 55 60 65 70

Fresh Water Inlet Temperature (Deg. F)

C o n v e c ti v e H e a t T ra n s fe r C o e ffi c ie n t (W /K ) 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

(43)

y = 363.76x + 30307 y = 409.23x + 34096 y = 454.7x + 37884 y = 500.17x + 41673 y = 545.64x + 45461 40000.00 47500.00 55000.00 62500.00 70000.00 77500.00 85000.00 35 40 45 50 55 60 65 70

y = 631.33x + 52600 y = 578.72x + 48217 y = 526.11x + 43833 y = 473.5x + 39450 y = 420.88x + 35067 50000.00 60000.00 70000.00 80000.00 90000.00 100000.00 35 40 45 50 55 60 65 70

C o n v e c ti v e H e a t T ra n s fe r C o e ffi c ie n t (W /K ) Tube Length = 6.096 m Tube Length = 6.858 Tube Length = 7.620 m Tube Length = 8.382 m Tube Length = 9.144 m

(44)

Coolant Convective H.T. Coefficient vs. Coolant Outlet Temperature y = -0.9171x2 + 367.7x - 34695 y = 4.1988x + 560.14 y = 0.1737x + 44.851 0.00 500.00 1000.00 1500.00 2000.00 2500.00 135 145 155 165 175 185 195 205 215 T,out (Deg. F) C o n v e c ti v e H e a t T ra n s fe r C o e ffi c ie n t (W /m ^ 2 - K )

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

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

y = 9.0196x - 278.29 y = 7.1478x - 220.76 y = 5.4731x - 169.27 y = 4.0211x - 124.57 y = 2.9874x - 92.641 200.00 300.00 400.00 500.00 600.00 700.00 800.00 900.00 1000.00 1100.00 115 120 125 130 135 140 145

(45)

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

y = 104.27x - 2168.6 y = 83.41x - 1770.9 y = 64.674x - 1410.4 y = 48.259x - 1086.7 y = 36.185x - 829.98 2000.00 4000.00 6000.00 8000.00 10000.00 12000.00 14000.00 115 120 125 130 135 140 145

Temperature Difference, Deg. F

H e a t T ra n s fe rr e d , (W ) 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

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

y = 131.17x + 593.08 y = 105.48x + 377.3 y = 82.366x + 191.02 y = 62.007x + 47.797 y = 46.74x - 6.6977 5000.00 7500.00 10000.00 12500.00 15000.00 17500.00 20000.00 22500.00 125 130 135 140 145 150 155

(46)

y = 9.0719x - 282.03 y = 7.188x - 223.65 y = 5.5027x - 171.4 y = 4.0418x - 126.06 y = 3.0023x - 93.713 200 300 400 500 600 700 800 900 1000 1100 115 120 125 130 135 140 145

H e a t T ra n s fe rr e d (W ) 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

y = 112.88x - 1860.2 y = 90.179x - 1523.3 y = 69.799x - 1217.3 y = 51.969x - 941.35 y = 38.915x - 720.33 2000.00 4000.00 6000.00 8000.00 10000.00 12000.00 14000.00 16000.00 115 120 125 130 135 140 145

(47)

y = 147.33x - 364.42 y = 118.21x - 384.21 y = 92.027x - 394.2 y = 69.017x - 383.3 y = 51.905x - 327.14 5000.00 7500.00 10000.00 12500.00 15000.00 17500.00 20000.00 22500.00 25000.00 125 130 135 140 145 150 155

y = 3.0118x - 94.396 y = 4.055x - 127.01 y = 5.5216x - 172.75 y = 7.2137x - 225.49 y = 9.1051x - 284.41 200.00 400.00 600.00 800.00 1000.00 1200.00 115 120 125 130 135 140 145

(48)

y = 118.93x - 2246.6 y = 94.899x - 1826.1 y = 73.335x - 1445.4 y = 54.494x - 1105.3 y = 40.757x - 840.43 2000.00 4000.00 6000.00 8000.00 10000.00 12000.00 14000.00 16000.00 115 120 125 130 135 140 145

H e a t T ra n s fe rr e d (W ) Tube Length = 6.096 m Tube Length = 6.858 m Tube Length = 7.620 Tube Length = 8.382 m Tube Length = 9.144 m

y = 56.599x - 633.31 y = 75.376x - 795.56 y = 100.77x - 954.98 y = 129.7x - 1115.4 y = 161.9x - 1285.5 5000.00 7500.00 10000.00 12500.00 15000.00 17500.00 20000.00 22500.00 25000.00 125 130 135 140 145 150 155

(49)

y = 3.0185x - 94.873 y = 4.0642x - 127.67 y = 5.5347x - 173.7 y = 7.2315x - 226.77 y = 9.1283x - 286.08 200.00 400.00 600.00 800.00 1000.00 1200.00 115 120 125 130 135 140 145

y = 42.124x - 931.22 y = 56.369x - 1229.5 y = 75.971x - 1618.9 y = 98.425x - 2057.2 y = 123.45x - 2542.3 3500.00 6000.00 8500.00 11000.00 13500.00 16000.00 115 120 125 130 135 140 145

(50)

y = 58.551x - 767.57 y = 78.075x - 979.8 y = 104.6x - 1213.2 y = 134.87x - 1459.8 y = 168.56x - 1726.2 5000.00 10000.00 15000.00 20000.00 25000.00 125 130 135 140 145 150 155

y = 362.6x + 30211 y = 332.39x + 27693 y = 302.17x + 25176 y = 271.95x + 22658 y = 241.73x + 20141 25000.00 30000.00 35000.00 40000.00 45000.00 50000.00 55000.00 35 40 45 50 55 60 65 70

Fresh Water Inlet Temperature, Deg. F

C o n ve ct ive H ea t T ran sf er C o ef fi ci en t (W/ m ^ 2-K ) 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

(51)

y = 304.29x + 25353 y = 342.33x + 28522 y = 380.36x + 31691 y = 418.4x + 34860 y = 456.44x + 38029 35000.00 40000.00 45000.00 50000.00 55000.00 60000.00 65000.00 70000.00 35 40 45 50 55 60 65 70

y = 363.76x + 30307 y = 409.23x + 34096 y = 454.7x + 37884 y = 500.17x + 41673 y = 545.64x + 45461 40000.00 47500.00 55000.00 62500.00 70000.00 77500.00 85000.00 35 40 45 50 55 60 65 70

(52)

y = 631.33x + 52600 y = 578.72x + 48217 y = 526.11x + 43833 y = 473.5x + 39450 y = 420.88x + 35067 50000.00 60000.00 70000.00 80000.00 90000.00 100000.00 35 40 45 50 55 60 65 70

C o n v e c ti v e H e a t T ra n s fe r C o e ffi c ie n t (W /K ) Tube Length = 6.096 m Tube Length = 6.858 Tube Length = 7.620 m Tube Length = 8.382 m Tube Length = 9.144 m

Coolant Convective H.T. Coefficient vs. Coolant Outlet Temperature

y = -0.9171x2 + 367.7x - 34695 y = 4.1988x + 560.14 y = 0.1737x + 44.851 0.00 500.00 1000.00 1500.00 2000.00 2500.00 135 145 155 165 175 185 195 205 215 T,out (Deg. F) C o n v e c ti v e H e a t T ra n s fe r C o e ffi c ie n t (W /m ^ 2 - K )

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

(53)

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>

<http://www.grapeaperacing.com/tech/coolingsystems.pdf> “Electrolysis: The Silent Killer”

<http://www.sancarlosradiator.com/electrolysis.htm>

Ofria, Charles. “Automotive Cooling Systems”