Developmental Feasibility for a Ceramic Wankel Rotary Combustion Chamber

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DEVELOPMENTAL FEASIBILITY FOR A CERAMIC WANKEL ROTARY

COMBUSTION CHAMBER USING COMPUTER AIDED ANALYSIS

A Thesis f

Presented to the faculty of the Department Of Mechanical Engineering

California State University, Sacramento

Submitted in partial satisfaction of the requirements for the degree of

MASTER OF SCIENCE

Mechanical Engineering

Ian Hellstrom

FALL 2018

DEVELOPMENTAL FEASIBILITY FOR A CERAMIC WANKEL ROTARY

COMBUSTION CHAMBER USING COMPUTER AIDED ANALYSIS

A Thesis

Ian Hellstrom

Approved by:

______, Committee Chair Dr. Rustin Vogt

______, Second Reader Dr. Akihiko Kumagai

______Date

Student: Ian Hellstrom

I certify that this student has met the requirements for format contained in the University format manual, and that this thesis is suitable for shelving in the Library and credit is to be awarded for the thesis.

______, Graduate Coordinator ______Dr. Troy D. Topping Date

Department of Mechanical Engineering

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Abstract

DEVELOPMENTAL FEASIBILITY FOR A CERAMIC WANKEL ROTARY

COMBUSTION CHAMBER USING COMPUTER AIDED ANALYSIS

Ian Hellstrom

The goal to achieve higher fuel-efficient internal combustion engines is a common goal for engine manufacturers. Current engines reject approximately 20 percent of the fuels energy to the engine cooling system. The waste energy is rejected to the atmosphere, in order to maintain the proper operating temperatures of the engine components. Controlling the energy loss by providing material that can operate a higher temperature can be one solution to minimize energy loss. Ceramics allow higher operating temperatures than current alloys that are common in mass production engines. Using ceramics as part of the combustion chamber can lower the energy loss.

An engine with ceramic inserts to control temperature and alloys to provide the load bearing structure where testing using computer aided thermal simulations. Using SolidWork’s Thermal

Simulations, a rotary engine was analyzed for heat losses, using inputs from field-collected data and data available from literature reviews. Results showed that the ceramic inserts provided a higher resistance to heat flow to the coolant systems, thus allowing more energy available in the combustion chamber.

______, Committee Chair Dr. Rustin Vogt

______Date

TABLE OF CONTENTS Page

List of Tables ...... vi

List of Figures ...... vii

Chapter

1. INTRODUCTION ...... 1

2. BACKGROUND OF THE WANKEL ROTARY ENGINE ...... 6

2.1 Operating Principles ...... 8

2.2 Future of the Wankel Rotary Engine ...... 10

3. PROFILE GEOMETRY ...... 12

3.1 Rotor Housing Profile ...... 13

3.2 Rotor Profile ...... 13

4. COMPRESSION ANALYSIS ...... 15

5. INITIAL THERMAL ANALYSIS ...... 18

5.1 Rotor Housing Thermal Analysis ...... 19

5.2 Rotor Thermal Analysis ...... 21

5.3 Thermal Analysis of a Detailed Rotor Housing ...... 24

6. CERAMIC INSERT AND ROTOR HOUSING THERMAL ANALYSIS ...... 29

7. CONCLUSIONS AND FUTURE WORKS ...... 31

References/Work Cited ...... 33

LIST OF TABLES Tables Page

1. Input variables for studies 1-5 for the rotor housing steady state analysis .. …. 20

2. Input variables for studies 6-9 for the rotor housing steady state analysis . …. 25

LIST OF FIGURES Figures Page

1. Pressure casting components using local reinforced ceramic inserts …………. 3

2. Rotor housing inner surface composed of three separate materials ... …………. 4

3. Major Components of a Wankel rotary engine ...... …………………. 9

4. Wankel Rotary Engine Combustion Cycle… ...... …………………………. 10

5. Illustration of the rotor housing and rotor curve …… …………………………. 13

6. Picture with both the rotor profile and rotor housing profile . …………………. 14

7. The 345-T6 rotor housing displacement with 64,000 pounds on front face . ….16

8. Stress on the 345-T6 aluminum rotor housing at 64,000 pounds ……………. 17

9. Two inch thick rotor housing heat transfer profile for studies 1 and 2 .………. 19

10. One inch thick rotor heat transfer profile with the ‘combustion’ profile . ……. 21

11. Steady state ceramic rotor temperature profile …… ... …………………………. 22

12. Rotor heat transfer with the transient profile on lower face . …………………. 23

13. Ceramic on left and cast iron rotor on the right showing maximum temperature

profile after 75 seconds of combustion function at 6000 rpm ... ……………….24

14. Temperature distribution of the inner surface of the rotor housing plotted

for three different operating revolutions per minutes …………………………. 25

15. Study 6, steady state temperature on ceramic rotor housing …………………. 26

16. Study 7, steady state temperature on ceramic rotor housing . ………………….26

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17. Study 8, steady state temperature on ceramic rotor housing with no

convection on the cooling passages …… ...... …………………………. 27

18. Study 9, steady state temperature on the combustion section on ceramic

rotor housing, with convection on all surfaces of the cooling passages .. …….27

19. Temperature profile of the adjusted convection coefficient to yield

approximately 175°F…… ...... …………………………. 28

20. Temperature profile ceramic and aluminum housing …………………………. 30

21. Displacement (left) and stress profile of the ceramic and 345-T6 rotor

housing at 64,000 pounds force on front face… ...... …………………………. 30

viii

1. INTRODUCTION

In modern vehicles that utilize internal combustion engines, one of the leading goals is to design engines that are more efficient than the last design iteration. One quest for higher efficiency is through the process of optimizing the engines fuel consumption efficiency. Higher efficiency could be achieved by utilizing as much of the potential energy the fuel possesses, maximizing the conversion from potential energy to actual work. In a four-stroke internal combustion engine, 17-26 percent of input energy from gas is converted to waste energy and rejected through the cooling system using components like the radiator to reject the heat energy [1]. All the energy rejected to the cooling system is wasted energy.

The Wankel rotary engine (RE) when installed and operational has almost all the same external components as an automotive piston gasoline engine. The biggest difference being how the work is converted inside the engine. The current design of the rotary has two cooling sections of the engine, the oil system, which ran through the center of the engine, and doubles as lubrication for the internal moving parts. The water based coolant system; runs parallel to the axial direction of the engine on the peripheral section of the intake, combustion and exhaust chambers through the housings. In the RE, one thought to increase fuel efficiency is fully utilizing the fuel ignited in the internal combustion engine by the means of insulating the combustion chamber. Insulating the combustion chamber in order to reduce; the heat rejected to the cooling system through the rotor housing and heat rejected through the rotor to the oil system. Lowering the heat

2 rejection could allow reduction in the required cooling system size; increasing the overall efficiency of the system, i.e. vehicle, due to the weight reduction.

A feature of the RE, from an operational standpoint, is the air/fuel mixture from the intake system does not cool the combustion chamber. Every combustion cycle in the

RE takes place in the in the same one third of the rotor housing. Thus, the combustion section is operating at elevated temperatures compared to the rest of the engine.

Inhomogeneous cooling (or heating) causes thermal stresses minimized by local reinforcement with ceramic inserts [2]. Analyzing the feasibility to utilize newer methods of combining metals and ceramic to minimize heat rejection during the combustion cycle of a RE was a main goal. In addition, creating thermal analysis studies to maintain a uniform temperature at the metal that would makes up the main structure of the engine.

The emerging manufacturing process is the manufacturing of metal components with ceramic freeze casted inserts called pressure die-casting. The casted ceramic inserts could be utilized in components such as, engine blocks, engine heads, exhaust runners, etc., where the main structure is steel or other structural metal. Figure 1 shows an illustration of the setup of pressure die casting with ceramic inserts.

Figure 1: Pressure casting components using local reinforced ceramic inserts [2].

Current production RE rotor housings are manufactured using three separate materials; chrome, steel and aluminum, see figure 2. These three separate metal layers combine to make a single rotor housing. Using the production layout while substituting in ceramics in place of the chrome and the steel, will be the focus of the analysis. The natural properties of ceramic materials allow operation at higher temperatures than common metal used in the manufacturing of internal combustion engines. One drawback of ceramics is that they are brittle and have low fracture toughness. This makes it difficult to select them as the primary structure in an internal combustion. This application feasibility will investigate the use of bonded ceramic and metal rotor housing of an RE. A ceramic insert installed in the combustion section. The aluminum housing utilized to stabilize the ceramic insert and provide the main structure of the engine. This layout will

4 be similar to the production RE geometry. Analysis was ran on the interaction between aluminum and ceramic to see viable stress in the structure.

Figure 2: Rotor housing inner surface composed of three separate materials [16].

In order to reduce the complexity of the overall project, the rotor apex seals, side seal and oils control seals are assumed to properly operate at elevated combustion section temperatures. The developments of apex seals that can operate at these extreme temperatures is too complex of an issue to incorporate in this paper. Further development of these items would absolutely be required to operate a RE in these conditions. The research would need to include additional material development to tolerate wear at high temperature and remain close to current longevity, and tolerances. In addition, due to the elevated temperatures in the combustion section there is a possible requirement for a compound to lubricate the apex seals rather than the current oil injection system [3].

The ceramic insert in the rotor housing is used to change the overall thermal resistance of the rotor housing. The higher the thermal resistance value, the higher the temperature gradient from the combustion chamber to the outside of the housing. In addition, the higher the resistance, the higher the temperate gradient, the lower the rejected heat out through the rotor housing. The equations used in the calculated of heat transfer and thermal resistance used in the thermal analysis studies are [1]:

Q = A*h₁*(T₁-T₂), turbulent convection for gases, combustion gas to wall

Q = A*k/t*(T₁-T₂), conduction through rotor housing

Q = A*h₂*(T₁-T₂), turbulent convection to a liquid, rotor housing to coolant

Q = A*h*(T₁-T₂), is the overall heat transfer

1/h = 1/h₁ + t/k + 1/h₂, is the overall thermal resistance

Where, Q is the heat transfer

A, is the area of surface h₁, is the heat transfer coefficient for a gas

h₂, is the heat transfer coefficient for a liquid

h, is the heat transfer coefficient for the system

k, is the thermal conductivity t, is the wall thickness

T₁, is the hot temperature T₂ is the cold temperature

2. BACKGROUND OF THE WANKEL ROTARY ENGINE

The Wankel rotary engine, designed by Felix Wankel in 1929 is the best known and most developed rotary engine. Before about 1960, technologists identified four types of internal-combustion engines: rocket, gas turbine, diesel, and gasoline piston. That year a fifth type, the rotary engine, was in development [13]. The Wankel engine was licensed, between 1958 and 1973, to 26 manufacturers [4]. Mazda signed a licensing agreement with Wankel in 1961. Mazda made the first prototype Wankel automotive engine in approximately 1967 [13]. Developers in the late 1970s, a victim of toughening emissions standards and rising fuel-economy demands shelved the engine [10].

Mazda Motor Company was the only major manufacture that was able to develop this engine and provide the engine to power their production vehicles. The rotary engine has found its way into at least five different Mazda vehicles. The most popular being the

Mazda RX-7. Mazda sold nearly two million vehicles powered by rotary engines before abandoning the design in 1995 [12]. Mazda continued to develop the rotary, relaunching the design in its RX-8 sports coupe in 2002 [11]. Mazda produced their last Wankel- powered vehicle, the RX-8, in 2012, partly because they could not meet European Union emission standards [13].

Advantages of the rotary engine: [4]

- Higher power to weight ratio than a piston engine

- There are no reciprocating moving parts

- It is running with almost no vibration

- Approximately one-third the parts of a piston engine

- Cheaper to mass-produce because it contains fewer parts

- The filling of the combustion chamber made at 270 degrees of the main

shaft compared to 180 degrees in a piston engine

- The rotary produces torque for about two-thirds of the combustion cycle,

compared to a quarter in a piston engine

- The wide range of speeds provides greater adaptability

- It can use various octane fuels

- The rotary does not suffer from the "scale effect" in limiting its size

- Has about a third the size of a piston engine of equivalent power

- The crankcase oil remains uncontaminated by the combustion process

and thus it requires no oil changes

- The main crankshaft oil completely sealed from the combustion process.

Disadvantages of the rotary engine: [4]

- The rotor sealing was poor due to large temperature differences between

sections of the combustion chamber

- The combustion process was slow and bulky due to the combustion

chamber being in motion, causing a thin stream and preventing the flame

in the combustion chamber in reaching the outer rims

- High fuel consumption at frequent changes of RPM, and high emissions

of pollutants

- Low torque due to the high rate of repetition of the Otto cycle;

- Straight sealing line, will never reach the same tightness as a piston rings

in a round cylinder bore [8]

-Rotary apex seals have higher sliding velocities compared to the velocity

piston rings [8]

2.1 Operating Principles

The rotary engine is comprised of four major components. These components are the side housing, the rotor housing, the rotor, and the eccentric shaft. See figure 3, for the photo of each major component. The rotor is made of machined cast iron. The rotor is made of a heavy material but it is always rotating in the same direction. The inertia of the rotor produced less internal stress since it is not starting and stopping like a reciprocating engine. The rotor housing is comprised of three press fitted pieces. The outer section of the housing is cast aluminum and the inner section is steel with a chrome surface. The aluminum is used to minimize weight, the steel liner is used to limit thermal and reduced wear. The chrome surface is to limit apex seal chattering. The side housing is cast iron and services multiple functions; routes intake and exhaust gases into and out of the combustion chamber and provides wear surfaces for the rotor seals. The eccentric shaft is forged steel, and takes the bearing loads that are in perpendicular to the axis of rotation in the rotary engine. The eccentric shaft also orbits the rotors through their planetary motion.

Figure 3: Major Components of a Wankel rotary engine (photo from WikiRotary.com)

There is one combustion process for each revolution of the eccentric shaft. The rotors are driven at one third of the eccentric shaft, so three revolutions of the eccentric is one full revolution of the rotor. In figure 4, the photo shows a cut away of the combustion chamber, the light blue indicates the intake stoke, the blue shows compression, red is the ignition & expansion stroke and the yellow shows the exhaust stroke. For every three revolutions of the eccentric shaft, the rotor makes one full revolution. If the engine was running at 6000 revolutions per minute, which correlates to the speed of the eccentric shaft, the rotor is spinning at a rate of 2000 revolution per minute. A combustion cycle is achieve for every revolution of the eccentric shaft, 100 cycles per second, so the spark plugs are firing every 0.01 second at 6000 rpm. These numbers used in the thermal analysis.

Figure 4: Wankel Rotary Engine Combustion Cycle (photo from speedacademy.com)

2.2 Future of the Wankel Rotary Engine

The rotary has multiple developments pursued in various industrial applications.

In 2009, Mazda patented the Mazda 16B engine and the company is testing the engine for use in hydrogen-powered vehicles. Other small displacement rotary engines are under development to run generators for hybrid power system and automobiles. The engines designed to operating at the most efficient speed to minimize the fuel usage and the exhaust emissions produced. The small displacement engine utilizes single rotor design that allow for a lightweight engine package. BMW created a hybrid Mini Cooper that has a single rotor engine installed; the unit generates 15 kW of electric power at 5,000 rpm, can be scaled up to 25 kW at 7,000 rpm. A slightly bigger version of the rotary engine, displacing 0.357L, can produce 36 kW at 7,000 rpm, while a double-rotor engine could deliver 50 kW [10]. In 2020, Mazda is releasing an electric vehicle, which will utilize the rotary engine as a range extender [14]. At this time, specific details of the new Mazda RE are not available, i.e. size, fuel efficiency, structural materials. Assumptions can be made

11 for the use of the RE because of the leveraging of the advantages of scalability, size and weight.

Geramic Rotary Engines Inc. in Texas has developed a rotary engine design using advanced ceramics for the main components. The company states ceramics allow the engine to run hotter (and thus more efficiently), operate on a variety of fuels, and burn whatever fuel it is using more completely to reduce emissions [12]. The goal of the company is to create a near adiabatic combustion cycle with an estimated increase in thermal efficiency by 30%. The increased efficiency will come from the minimization of heat transfer through the engine components to the atmosphere by maintaining the heat energy in the combustion chamber. These engines were developed for use in a small generator, water pumps, and heaters that run on bio-ethanol and other organic type of fuels. One item that the company does not discuss is the effect on ratio of the emission by-products.

3. PROFILE GEOMETRY

This section also set up base line trial and error analysis using SolidWorks

Simulation. The equations that make the unique rotor and rotor housing profiles were found and the equation drawn in SolidWorks. The rotary engine has a unique shape for the rotor and the rotor housing, see figure 5, for picture of both the rotor and the rotor housing. A parallel trochoid curve is the type that makes the rotor profile. A peritrochoid curve is the type that makes the rotor housing profile. These profiles created in

SolidWorks, to run the analysis for this study. The parametric equations for the two curves:

Peritrochoid Curve:

X=e cos (alpha)+R cos (beta)

Y=e sin (alpha)+R sin (beta)

Parallel Trochoid Curve:

X=e cos (alpha)+R cos (alpha/3)+alpha cos (alpha/3+phi)

Y=e sin (alpha)+R sin (alpha/3)+alpha sin (alpha/3+phi)

The constant e, is called the amount of eccentricity, a function of the dimensions of the eccentric shaft, which is the circle that is formed by the eccentric shaft for one revolution of the crankshaft. R is the generating radius, which is the length from the center axis of the rotor to any one of the three apexes. Phi is the angle of oscillation of the apex seal [16].

Figure 5: Illustration of the rotor housing and rotor curve, to show the placement of e and

R, constants in the equations that make up the curves.

3.1 Rotor Housing Profile

Equations converted for use in SolidWorks, e and R verified measurements from a production Mazda rotor housing:

X-coordinate=0.59375*cos(3*t)+4.15625*cos(t)

Y-coordinate=0.59375*sin(3*t)+4.15625*sin(t)

When drawn in SolidWorks t varied from 0 to pi, which created half of the housing. The housing half mirrored to create the closed profile. Through trial and error it was found that these parametric equations could not run the functions from 0 to 2*pi, the equation would not solve due to being self-intersecting at the start and end point.

3.2 Rotor Profile

Equations converted for use in SolidWorks, e and R verified measurements from a production Mazda rotor:

X-coordinate =

4.15625*cos(2*t)+3*(0.59375^2)/(2*4.15625)*(cos((8*t))-cos((4*t)))+ (0.59375*(1-

(9*0.59375^2)/(4.15625^2))*(sin((3*t)))^2)^0.5)*(cos((5*t))+cos(t))

Y-coordinate=

4.15625*sin((2*t))+3*(0.59375^2)/(2*4.15625)*(sin((8*t))+sin((4*t)))+ (0.59375*(1-

((9*0.59375^2)/(4.15625^2))*(sin((3*t)))^2)^0.5)*(sin((5*t))-sin((t)))

Where t varies from 0.523598776 to 1.570796327, this function creates one third of the rotor profile. Using the circular pattern command the complete profile.

Figure 6: Picture with both the rotor profile (inside profile) and rotor housing profile

(outside profile).

4. COMPRESSION ANALYSIS

The first analysis is to calculate amount of compression on the rotor housing. The highest static loading on the engine assembly is the compression force on the assembled rotary engine (RE). A fully assembled RE utilizes 18 tension bolts that clamp the two end housings, intermediate housing and the two rotor housing. The tension bolts pass perpendicular to the rotor rotation, and run parallel to the eccentric shaft. The tension bolts are all located on the peripheral of the engine. Determining the compression forces on the engine assembly, rotor housing deformation will be calculated. The calculated housing deformation used further in later section in this thesis to minimize clamping forces and associated stresses on ceramic parts, and maintaining the static loading on the aluminum structure of the housing. Referencing the Mazda rotary engine factory service manual the torque specifications on all the tension bolts are 23-29 foot pounds. All the tension bolts have a 12-millimeter shank diameter. Using equation #1, where Tᵢ is the tightening torque, the preload is Fᵢ, d is the diameter of the bolt, Kᵢ is the torque coefficient. Where, the force from one bolt is (equation from Machine Design, 3rd edition, equation 14.18d):

Fᵢ = Tᵢ / [Kᵢ * d] equation #1

Using the maximum torque value from the service manual and converting all the units to inch, pounds, seconds; Tᵢ = 29 ft-lbs, d = 0.0394 ft, and using the reference from Machine

Design, 3rd edition, table 14-11 is Kᵢ = 0.21

Fi = Tᵢ / [Kᵢ * d] = 29 ft-lbs / [0.21 * 0.394 ft] = 3,500 lbs

Total force on the assembled engine, for 18 bolts at 3500 lbs is 63,000 lbs. where if the lower limit of the torque value was used, the force on the housing for all 18 bolts is

50,000 lbs.

Figure 7: The 345-T6 rotor housing displacement with 64,000 pounds on front face.

Utilizing the calculated clamping force, an analysis was ran on the housing with the 64,000 pound force (lb-f) on the assembly to determine the change in width of the rotor housing. The material used for the analysis is 356-T6 aluminum. The analysis ran with a total force of 64,000 lb-f on the rotor housing as seen in the above figure 7. The total displacement was 0.00068 inches. Dimensional tolerances for the aluminum and the ceramic, aluminum will require at least 0.001 in greater tolerance call out, for static

17 assembly. Temperature tolerances and the analyzed change in thickness due to compression, incorporated in later sections.

Figure 8: Stress on the 345-T6 aluminum rotor housing at 64,000 pounds.

A second simulation was ran to determine the static stress on the rotor housing applied by the clamping force from the tension bolts, figure 8. The maximum stress seen in this configuration is 6.4 kilopound per square inch (ksi), approximately one quarter of the yield strength of the 356-T6 aluminum, 22 ksi. These results used as a reference for the parts analyzed made of both ceramic and aluminum, in later sections.

5. INITIAL THERMAL ANALYSIS

The housing of the rotary engine (RE) is the equivalent to the cylinder block and cylinder head of the reciprocating piston engine. The combustion chamber of an RE is composed of rotor housing and side housings. In the piston engine, each stroke: intake, compression, expansion and exhaust takes place in the same area. The heat load on the cylinder head and cylinder block more evenly distributed, than the RE. In the RE the working chamber moves as the stoke proceeds, see in figure 4, section 2. The intakes stoke take places at the coldest section of the engine. The expansion stroke occurs in the hottest part of the engine. This portion of the engine continually exposed to the highest temperatures created by the combustion process [16].

In this section, thermal analysis studies were ran with many variables. Each simulation reviewed to identify and provide improved inputs for the next simulation. The multiple studies completed to provide a foundation of testing procedures for the advance assembly analysis, found in later sections, the rotor housing structure with the ceramic insert. In addition, the studies established a process and procedure for running a steady state and transient combustion analysis. The first issue identified using SolidWorks with education license is the inability to edit and or add a material in the SolidWorks domain.

Only ceramic materials available to use in the thermal test were porcelain and alumina.

The ceramic porcelain’s heat transfer coefficient is 2.0x10⁻⁵ BTU/(sec*in*°F ) which is comparable to the engineering ceramic that have been successfully used in the manufacturing of engine components, such as silicon nitride. The heat transfer coefficient for silicon nitride is 4.0x10⁻⁵ BTU/(sec*in*°F ).

5.1 Rotor Housing Thermal Analysis

The first simulation, study 1, evaluated steady state temperatures on a rotor housing that was 2 inches thick, utilizing ceramic porcelain. See figure 9 for the profile of the heat transfer rotor housing. See table 1 for the input variables and output temperatures for studies 1-5. Internal temperature (combustion temperature) for study 1 was 1700 degrees Fahrenheit [14]. All the studies used one of three convection coefficients for on the outer surface of the housing: 3.4x10⁻⁶ BTU/(sec*in²*°F ), defined as slow forced air over a surface, 3.4x10⁻⁵ BTU/(sec*in²*°F ) defined as moderate forced air over a surface, and 10.2x10⁻⁴ BTU/(sec*in²*°F ) defined as moderate force water in a pipe [15].

Simulations were ran using two-dimensional simplification to use less computing power to run the analysis, since all profiles used were simplified when drawn.

Figure 9: Two inch thick rotor housing heat transfer profile for studies 1 and 2.

STUDY INTERNAL CONVECTION CONVECTION OUTER MATERIAL THICKNESS TEMP (°F) BTU/(sec*in²*°F) BULK TEMP TEMP (INCH) [W/(m²*K)] (°F) (°F) 1 1700 3.4x10⁻⁶ [10] 120 1233 Ceramic 2 2 1700 3.4x10⁻⁵ [100] 120 423 Ceramic 2 3 480 3.4x10⁻⁶ [10] 120 373 Ceramic 1 4 480 3.4x10⁻⁵ [100] 120 189 Ceramic 1 5 480 10.2x10⁻⁴ 120 122 Ceramic 1 [3000] Table 1: Input variables for studies 1-5 for the rotor housing steady state analysis.

The initial assumption was the temperature on the outer surface of the part would be close to, or at, the convection bulk temperature input of 120°F. Studies 1 and 2, the variable adjusted were convection on the outer surface of the part. The desired value was to have the middle of the rotor housing down to a temperature close to 150° F, but not achieved. Reviewing the analysis set up the simulation is not realistic to an actual running engine due to the rotor housing is not a constant 1700° F. The 120°F was a measure value of a under hood temperature of a vehicle engine bay. Additional research revealed data regarding the apex seal temperature with respect to engine rpm and oil injection. Since the apex seal is a thin piece of metal exposed to the fuel combustion and is a friction surface, studies 3-5 used the inner surface of the housing of 480°F [9]. Studies 3, 4 and 5 used similar inputs as the first two studies; the difference was the 2-inch thick housing reduced to 1 inch. See table 1 to for the input and out temperatures.

With the simulations ran for studies 1-5 the realization that using ceramics will help insulate but, not as promising as the initial desired value. Further research determined that the turbulent convection of the boundary layer on the housing surface

21 played a large part in realistic analysis. A transient ‘combustion’ temperature function was created to ramp up from 75 to 1700°F over the half a second. The function was used on half of the rotor housing face. This function used to simulate the surface of the rotor housing that experiences the combustion process. The 3.4x10⁻⁶ BTU/(sec*in²*°F ) convection coefficient was used on the outer section of the housing. The highest temperature on the surface of the housing is 537°F. See figure 10, for the profile of the heat transfer with the transient ‘combustion’ heat transfer.

Figure 10: One inch thick rotor heat transfer profile with the ‘combustion’ profile.

5.2 Rotor Thermal Analysis

The next set of simulations, which completed the rotor profile, evaluating steady state temperatures on the rotor, using the porcelain ceramic, see figure 11. The main

22 change in the study was the convection coefficient was set to 10x10⁻⁴ BTU/(sec*in²*°F ) the assumption is that this is equivalent to the oil that would run through the internal section of the rotor. The outer temperature was set to 480°F.

Figure 11: Steady state ceramic rotor temperature profile.

Improving on the previous studies, the combustion set of simulations were ran with transient temperature profiles to simulate the combustion temperature on the face of the rotor. The initial temperature of the rotor was at 75 F. The first varying load simulation on the rotor was for a period of 15 seconds on only one face. The transient temperature function, cycled as if the rotor was seeing the combustion temperatures as if the engine was running at 6000 revolutions per minute. Time values for combustion, discussed in section 2.1, the face of the rotor is subjected to a 1700°F for 0.03seconds

(combustion), and 0.06 seconds (intake and compression) it is subjected to 75°F, see

23 figure 12 for this profile. The profile improved with a high temperature gradient compared to the initial profile as figure 9 and 10.

Figure 12: Rotor heat transfer with the transient profile on lower face.

The next simulation run, the same set up as the profile seen in figure 12, with the addition that all three face received the transient function. Data from a spreadsheet was used to create a function curve for time verse temperature. The maximum data points which can be entered into the function curve is 5000, so the simulation was limited to 75 second run. The maximum temperature profile seen in figure 13 (left). The last rotor simulation was ran identical to the simulation ran in figure 13 (right) but the material of the rotor was changed to cast iron. This simulation was to evaluate the actual material used in the production engine. Evaluating the temperature gradients of the two profiles, the ceramic internal temperature was 157.7°F and the cast iron internal temperature was

389.3°F, both have the surface temperature of 1700°F.

Figure 13: Ceramic on left and cast iron rotor on the right showing maximum temperature

profile after 75 seconds of combustion function at 6000 rpm.

5.3 Thermal Analysis of Detailed Rotor Housing

In prior simulations, additional features left off the rotor housing profile to allow the simulation to compute as quickly as possible. This section of simulations included the cooling passages found on the production rotor housing. The cooling passages transport coolant through the engine and draw waste heat out of the engine. The rotor housing is sectioned to incorporate only the combustion section to allow the fast simulation calculations as possible. Initial studies utilize the convection coefficients outlined in section 5.1. In Kenichi Yamamoto’s paper “Rotary Engine, he provides a temperature profile of the surface temperature on the inner face of a rotor housing. The hottest section, the combustion section, subjected to surface temperatures of 374°F (190°C), see figure 1, for the temperature profile. In addition, the convection bulk temperatures increased to 175°F, due to temperature measurements taken on an operational RE at

25 normal operating conditions. Table 2 shows the input variables for studies 6-10 for the rotor housing steady state analysis. All simulations use ceramic as the study material.

STUDY INTERNAL COOLANT COOLANT OUTER SURFACE OUTER OUTER TEMP (°F) CONVECTION CONVECTION CONVECTION SURFACE TEMP BTU/(sec*in²*°F) BULK TEMP BTU/(sec*in²*°F) CONVECTION (°F) [W/(m²*K)] (°F) [W/(m²*K)] BULK TEMP (°F) 6 374 10.2x10⁻⁴ 175 3.4x10⁻⁵ [100] 175 188 [3000] 7 1700 10.2x10⁻⁴ 175 3.4x10⁻⁵ [100] 175 250 [3000] 8 1700 N/A N/A 3.4x10⁻⁵ [100] 175 312 9 374 10.2x10⁻⁴ 175 3.4x10⁻⁵ [100] 175 220 [3000] Table 2: Input variables for studies 6-9 for the rotor housing steady state analysis.

Figure 14: Temperature distribution of the inner surface of the rotor housing plotted for

three different operating revolutions per minutes [16].

Figure 15: Study 6, steady state temperature on ceramic rotor housing.

Figure 16: Study 7, steady state temperature on ceramic rotor housing.

Figure 17: Study 8, steady state temperature on ceramic rotor housing with no convection

on the cooling passages.

Figure 18: Study 9, steady state temperature on the combustion section on ceramic rotor

housing, with convection on all surfaces of the cooling passages.

The final of simulations for this section was to run studies while adjusting the convection coefficient of the coolant passages, using the aluminum housing. The inner surface temperature was the temperate from the plot seen in figure 16. The convection coefficient was increased an additiona1 6.7x10⁻⁴ BTU/(sec*in²*°F ) in each simulation.

This was ran until the outer surface was approximately 175°F, the temperature obtained by field measurements. It was determined that a convection coefficient of 37.4.2x10⁻⁴

BTU/(sec*in²*°F ) produced results comparable to field measurements of the outer surface of the rotor housing, see figure 21 for the profile. This new coefficient used in section 6 simulations.

Figure 19: Temperature profile of the adjusted convection coefficient to yield

approximately 175°F.

6. CERAMIC INSERT AND ROTOR HOUSING THERMAL SIMULATION

The following study used the following parameters: the ceramic insert is porcelain ceramic, the housing is 345-T6 aluminum, the convection in the cooling passages is

37.4.2x10⁻⁴ BTU/(sec*in²*°F), the convection on the outside of the housing is 3.4x10⁻⁵

BTU/(sec*in²*°F ). The internal surface temperature was 1700°F. The temperature was selected from reviewing data from Yamamoto’s paper. The profile provided in his paper showed the surface temperature profiles of both; a cast iron housing and an aluminum housing. The thermal conductivity of aluminum is two times greater than the iron and the profile shows the surface temperature of the cast iron was twice as much as the aluminum. The surface temperature of the housing directly correlated to the thermal conductivity, since the highest temperature that was field verified immediately out of the exhaust was 1700°F, it is used as the maximum operating temperature. The results evaluated in this section were the temperature profile of the housing and the compression stress on the housing.

Simulations studied used the combination of the ceramic insert and the aluminum housing. The plain ceramic insert is the same contour as the peritrochoid with a thickness of 3/16 of an inch. The results for the thermal study showed that the 3/16 of an inch ceramic insert controlled limited the temperature profile from1700°F to 243°F, at the location of the bond between the ceramic and the aluminum. The convection for the coolant passages was the same as the coolant convection of the aluminum housing.

Interpreted the cooling system with the ceramic insert is much larger than required. The

30 compression test was ran on the assembly using the same inputs as section 4. Results from the compression test were negligibly different from the initial test ran.

Figure 20: Temperature profile ceramic and aluminum housing.

Figure 21: Displacement (left) and stress profile of the ceramic and 345-T6 rotor housing

at 64,000 pounds force on front face.

7. CONCLUSIONS AND FUTURE WORKS

The initial steady states thermal simulations in section 5 were not promising due to the high temperatures that where seen on the opposite side of the part being analyzed.

Using the transient function to simulate the combustion process provided results that were much closer to the initial desired results. Adjustment of the input variable for convection and surface temperature allowed for better results, compared to temperature reading taken on an operational RE. With the transient function used the results from figure 13 shows that a 0.25 inch thick ceramic insert reduces the temperature 1300 degree

Fahrenheit.

The limitation for this project was the inability to add and modify the material library in the SolidWorks domain. The best material to run the analysis was ceramic porcelain, for it low thermal conductivity value.

The use of the ceramic insert in all the studies show the temperature gradient was much higher than the aluminum. This showed that less energy was rejected through the housing to the coolant system. This does not necessarily mean the engine is more efficient, it means that the energy put into the engine retained in the combustion chamber.

Additional research will be need to determine if the actual combustion process is more efficient or that more waste energy is expelled through the exhaust system. If the energy is waste out of the exhaust, items such as turbochargers and heat generators to convert that waste energy to useful energy.

The three major variables ignored for this project were flame propagation, apex seals, and combustion chamber lubricating oil. The flame front propagation would play

32 and increasing complex role in the calculation of the rotor housing surface temperature.

Hence, this is why the flame propagation was ignored. The RE has a long and narrow combustion chamber. The temperature difference between the walls and the expanding gases and flame front, has a high differential where cooler combustion chamber walls extinguish the flame front. This leads to incomplete combustion and high hydrocarbon in the exhaust, due to incomplete combustion. Further research and testing is needed to measure the change in flame front propagation characteristics, from a lower temperature differential, due to higher surface temperatures. This information would provide an indication if the fuel efficiency was increase for the system by the engine to complete the combustion of the fuel entered in to the engine. The second ignored variable was the rotor seals, which seal the combustion section of the rotor. The increase heat on the surface of the rotor housing would can the temperature operating conditions of the apex seals. The increased temperature would require additional research into whether the production apex seal would still operate efficiently, or the pursuit of better materials would be required. Also, the increased temperature might affect the seal lubricating film, which is the engine oil on production vehicles, by burning the oil. The last ignored variable was the wear of the ceramic surface. Additional research is need to identify a ceramic material that would operate in the conditions studies about but a material will also need to tolerate wear.

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