T/torque converter-direct-173.txt 173

Extremely harsh lock-up clutch engagements may occur or the clutch slip codes may be set. The "Late" 25 spline F family converters have three I. This unit is a single plate Modulating Clutch System. The "Late" 25 spline F family converters with four I. Ink stamp: A Pinto, Bobcat.

Pinto Bobcat. Some impellers will have dimples, some will not. Lincoln Mercury All. All with 4. Stamped: 23, Introduction to General Motors Clutch Type Converters The following pages are devoted to the introduction of clutch type torque converters as used by General Motors in passenger cars and trucks since Certain criteria must be dealt with to insure proper identification of the clutch type converters other than simple "looks like" identification.

Stall speed, clutch type, stall torque ratio, dampening spring rate, and mounting lug type and style must be considered as well as the transmission that will receive the converter.

Do not forget the axle ratio or final-drive. Read the code from the converter I. Every effort has been made to ensure accuracy as well as vehicle coverage combined with an easy to use format. Some applications are not as complete as others but we feel certain that most application questions can be answered.

Please study the following explanation guide before using the application charts. Production Years. K-Factor and determined stall are calculated with an input torque of A change of input torque will increase or decrease determined stall speed. The STR is the ability of the converter to multiply the torque of the engine as delivered to the converter. Read the original code tag on the torque converter or the metal stamp on the converter body to establish stall speed.

Codes: The code sticker will identify the transmission type that the converter fits, the stall speed, the type of clutch and damper assembly as well as spring rate, and the type of mounting lugs on the cover. The metal stamp found between two of the "dimples" on the converter body was the early system used to identify the angle of the impeller blades. Some early units do not have this metal stamp and identification by "dimple" angle and engine application must be used for correct application.

This is the standard identification code system as used by General Motors. Note: Units with Carbon Filled or Woven Graphite clutch material should be replaced with the same code. See opposite page.

B D G 5 F 3rd Digit will not be used unless a specific code should be replaced exactly. This includes converters that are used in Cadillac vehicles that "read" stall speed and clutch release such as codes CF9C etc Some converters will not have a code sticker, but will have a number or letter stamped between the dimples of the impeller.

Number 5 stamped in body or RPM depending on stator. Code B is and Code F is Letter C stamped in body RPM approx. Same as GM code.

The impeller package "dimples" and stamped numbers between the dimples will aid in identification if the I. The 14 blade stator is used with the 5 impeller package to produce a stall speed of RPM. The 19 blade stator is used with all impeller packages to produce the stall speeds as identified on the following pages. TC Orange Label Code, 2. Hub size: 1. Stall torque ratio: O. Stall torque ratio: 1. Ink Code B last letter of 3 or 4 digit code describes the length of pilot, bolt circle center of pilot to center of thread X 2 and height of mounting pads.

Overall height of thick pad converter is 5. Measure the length of the mounting pad from the converter body to the face of the pad. Record the eight digit part number, Ink Stamp, and the four letter code if available.

The bolt circle may be measured from the center of the pilot to the center of the bolt hole and multiplied by 2 on three pad converters. Measure across the pilot on six pad converters, bolt center to bolt center. These torque converters were basic Chrysler torque converters without ring-gears. The TC-6 and TC-8 torque converters followed the Chrysler torque converters as lock-up units and have now followed as A and A units in The through A and A lock-up and non-lock-up torque converters are the same as used in A and A non-lock-up and lock-up transmissions.

To begin your identification, follow these simple guidelines. Count the splines. A non-lock-up torque converters have 27 turbine splines, same as A non-lock-up converters. A lock-up torque converters have 26 turbine splines, same as A lock-ups. A non-lock-up torque converters have 24 turbine splines, same as A non-lock-up converters. A lock-up torque converters have 23 turbine splines, same as A lock-ups. Identify the stall of the converter by the color of the label that is attached to the top of the converter.

Most and later converters are of the high stall variation. Check the mounting pads and look at all units for the 90 degree symbol on the cover. These torque converters will not inter-change with or be inter-changed by earlier units. The bolt pattern is different and does not have an "off-set" mounting pad. The earlier butterfly weight will not fit the flex-plate. The un-balance weight is on the flex-plate. In Chrysler starts the use of a thrust bearing under the impeller hub and on top of the converter stator.

This location had previously used a hardened thrust washer. Stall speed and K-factors will change along with the use of stamped numbers on the torque converter as a means of identification. Look for the three digit number stamped on the transmission side of the torque converter as an identification number. Flag for inappropriate content. Download now.

Save Save tonyduramax. Original Title: tonyduramax. Related titles. Carousel Previous Carousel Next. Jump to Page. Search inside document. GM Duramax 6. Documents Similar To tonyduramax. Giapy Phuc Tran. Je Rel. Scott Kramer. Gateway Chevrolet Cadillac. Irvin Lopez. Kieran Ryan. Troy D Gasperino. Anonymous DihPNFy. Marcin Baj. Ubaldo Sanchez. Norm Sai. Roberto Mascarenhas. Jamey Basham. David Corrales. Adam T-Man. More From Andres Contreras.

Andres Contreras. Sistema de aire motor ford 6. Popular in Engine Technology. Don Patriot. With the availability of a suitable fluid and a design to reduce forces and stresses, the traction drive becomes competitive enough to warrant a more thorough evaluation. Variable ratio traction transmissions of various types have been studied over the years. The toroidal type has emerged as the best design for highest power density, reasonable life, and good efficiency.

Sundstrand, Lycoming, Rotax, General Motors, English Electric, and others have built and tested the toroidal types; Tractor also is developing a modified torodal type. Lycoming has been the only company to market a toroidal traction transmission but other companies have successfully tested prototype designs. There are many options open in the design of a toroidal traction variable speed drive. One of the basic design considerations is the use of a two row dual toroid design or a single row single toroid design.

A single row toroidal drive derives its name from the fact that there is only one set of traction rollers, typically three rollers per set, that transmit power between the input toroidal disk and the output toroidal disk.

In a two row device, the power flow is split from the center, or input toroid, through two sets of traction rollers. The two sets of rollers, with one set on either side of the input toroid, transmit power to the output toroids which are located at each end of the traction drive.

The benefit from the use of a two row device is that since it is symmetrical, the thrust loads from the two halves of the unit cancel each other and do not have to be taken thru thrust bearings. On the other hand, a single row traction drive requires very large thrust bearings to react to the large axial thrust loads, and also, since there is load sharing between the two halves of a two row device, the required toroid diameter for a two row unit is considerably smaller than for a single row unit.

The relatively large power loss that would be encountered in the large high speed thrust bearings of a single row unit make the two row design much more attractive from an efficiency point of view. Although a single row traction drive would cost less to produce, it is felt that the advantages to be realized from using a two row unit more than offset the cost. The basic scheme of the toroidal type is shown in Figure IV Principal components are the input toroid, rollers, and output toroid.

The rollers are steerable and are "steered" to the necessary angle to provide the input to output speed ratio desired. When the input toroid is rotated, the rollers turn and exert a traction force on the output toroid. The toroids must be held together to insure sufficient traction exists with the rollers to transmit the desired power.

The rollers are steered or tilted to the angle which produces the desired output speed. With the rollers angled as shown in the figure, output speed is lower than input speed. At the opposite angle, output speed is higher than input. Power capacity for a given size and given number of rollers is a function of the clamping force between input and output toroids across the rollers, and the traction coefficient of the fluid being used. The torque producing force at the point of roller contact is the product of the clamping force and the traction coefficient.

Life of the unit is a function of this force and devices have been developed to vary this force in proportion to the load with a resultant increase in life. The traction transmission design for this application must have a disengaging device to permit zero output speed when the engine is running. Three ways of achieving this were considered, and they are listed below along with the effects of each type on the traction drive unit. For narrow speed range engines, where minimum speed is close to maximum speed, the power which must be dissipated in the clutch is high.

This is an important consideration with variable thrust types of traction drive units in that thrust must always be maintained sufficient to prevent gross slip between roller and toroid.

These shock loads could be seen in the drive train under such conditions as accelerating on an ice-patched surface or sudden wheel-lock when braking. The hydraulic ratio control system Pa9e 14 Sundstrand Aviation Without torque limiting, the torque would theoretically go to infinity under conditions of holding the output stalled at wide open throttle.

The traction drive unit would have to be sized to take this limiting torque value, which could be several times greater than maximum engine torque. It could also be arranged to give a "built-in" reverse ratio capability, obviating the need for a reverse gearbox. This is illustrated graphically in Figure IV An input mounted torque converter will up-size the traction drive. The input mounted converter offers lesser protection.

From these considerations it was decided to use an output mounted torque converter as it is the only device which gives any degree of shock load protection to the traction drive unit, and it allows the use of the smallest possible traction unit. The torque converter has the additional advantages of low cost, excellent reliability and virtually zero maintenance. The computer efficiency and consequently the fuel consumption of the two systems were very close.

Attention must be given to the gearing to ensure that noise is reduced to a minimum. Sundstrand is committed to meeting acceptable noise levels with the present production Dual Mode hydromechanical transmission DMT. Experience gained from testing the DMT has been applied in the proposed Tri-Mode hydromechanical transmission design.

It appears at this time that the noise from either transmission can be brought within acceptable limits. Tooling requirements will be similar. More companies are involved in actual testing and evaluation of hydromechanical transmissions than with traction transmissions. Production hydromechanical vehicle transmissions are being offered for sale, while production traction transmissions have been produced only for aircraft constant speed drive applications.

Design and development of a pre-prototype transmission by late or early can be accomplished for either transmission type, although the traction type represents a somewhat greater risk. Either type could be ready for production by The hydromechanical transmission primary development task will be the integration of its controls with the engine. Although the basic control scheme has been mechanized and demonstrated, operation with the specified engines will require additional effort.

The major development task for the traction type transmission is assurance and demonstration of the required life. Since the life capability is highly dependent upon the vehicle load toroid and roller stress , vehicle and engine speed traction ratio and time at each condition, determination of the actual vehicle duty cycle is very important. Since little experience has been obtained with traction transmissions in vehicles this definition of the "real" operational requirements and the mechanical design reflecting these parameters, becomes the major development item.

As a result, it was decided to continue the detailed evaluation of both transmissions through to the completion of the study. Mechanical Operation The following is a discussion of the mechanical operation of the tri-mode hydromechanical transmission with regard to the direction of power flow, component speed and torque relationships, and variable unit displacement. The transmission is shown in simplified schematic form on Figure V The transmission has three distinct modes of operation in forward.

At lower engine speeds the shift points occur at proportionately lower vehicle speeds. During Mode 1, the output from the fixed displacement hydraulic unit is geared directly to the output. In Mode 2 and Mode 3 operations, the fixed unit is geared into the plantetary. Reverse is the same as Mode 1, but in opposite direction and is obtained by stroking the variable displacement hydraulic unit in the reverse direction. Figure V-2 shows schematically the geartrain arrangement.

In Mode 1 the fixed displacement hydraulic unit is geared directly to the output planetary link, so in Mode 1 its speed will be directly proportional to output speed, hence, vehicle speed. When the fixed displacement hydraulic unit speed increases to the point where it is equal to variable displacement hydraulic unit speed, a mode shift from Mode 1 to Mode 2 occurs. In Mode 2, the fixed displacement hydraulic unit is geared to a leg of the planetary which causes power to be transmitted both hydraulically through the hydraulic units and mechanically through the planetary.

The fixed unit speed decreases with increasing vehicle speed until it passes through zero speed and then increases in the opposite direction. When the fixed displacement hydraulic unit speed increases to minus one times the variable displacement hydraulic unit speed, a second mode shift from Mode 2 to Mode 3 occurs. Both Mode 1 and Mode 2 shifts are accomplished when the driving and driven clutch discs are at essentially equal speeds.

In Mode 3, the fixed displacement hydraulic unit is geared to another leg of the planetary, different from that of Mode 2, which again causes power to be transmitted both hydraulically and mechanically. The characteristics of Mode 2 and Mode 3 are very closely related, the only difference being the speed and torque ratios between the various elements.

Increasing vehicle speed further after the Mode 2 to Mode 3 shift results in decreasing fixed displacement hydraulic unit speed from its negative maximum until it passes through zero, and then increases to its positive maximum speed one times the variable displacement hydraulic unit speed at maximum vehicle speed. Figure V-3 shows the hydraulic unit speeds schematically. The speeds of the various links of the compound summer in this case a four element planetary can also be represented on a nomograph, shown on Figure V Thus, when output speed and engine speed are known, fixed and variable displacement hydraulic unit link speeds can be found.

Since the fixed and variable displacement hydraulic units are related directly to their respective planetary links by gear ratios, all the system speeds can be calculated.

Unknown torques may be found by applying the equations of statics to the torque vector-beam analogy of the planetary see the following sketch.

The basic methods of solving for torque and speed in the transmission were defined previously. The magnitude of the horsepower in any link is the torque in that link times the speed of that link divided by the appropriate dimensional constant. The direction of horsepower flow, on the other hand, must be determined from the direction of link rotation and the direction of applied torque.

Sign conventions were established for the planetary speed nomograph Figure V-4 such that any speed above the nomograph absissia is positive, and any speed below is negative.

In the planetary torque balance beam sketch , any vector pointing up is positive and any vector pointing down is negative. Hardware Description The following is a brief description of the various components which make up the tri-mode hydromechanical transmission. Reference should be made to the cross section drawing, A-L5, shown in Appendix V-2 for indication of component arrangement and relative size.

Figure V-5 shows a schematic cross section of a typical hydraulic unit of this configuration. The hydraulic units are identical in construction to hydraulic units presently being manufactured by Sundstrand for hydromechanical transmission applications where they have proven their reliability, low cost, and good efficiency. One unit is variable displacement, the other is fixed displacement. The units are designed for psi nominal, psi overloads, and psi proof pressure. Page 25 Sundstrand Aviation..

Mounting the units in this manner provides for a shorter transmission length and allows for better noise reduction techniques to be utilized. These clutches are of the conventional mu I tip late disc type common to automotive applications. They are simple to control, inexpensive, and have high torque capability.

At the shift, the shaft speeds are essentially synchronized, thereby allowing the use of light duty clutches and are thus sized on torque capability and not energy dissipation. Clutch design follows standard automotive practice. Steel separator plates are used with organic linings and the drums are ductile cast iron. The piston and the back-up ring are aluminum.

A centrifugal operated pressure sensitive check valve is incorporated within each clutch to preclude centrifugal pressure from actuating the clutch.

Rotating seals between concentric shafts are of the cast iron piston ring type common with standard automotive practice. The gears are all designed to permit use of economical mass production techniques. It has been sized to provide for main hydraulic unit charging, control operation, clutch application and cooling, gear and bearing lubrication, and flow to the transmission cooler.

Bearings of this type are widely used in automotive applications as they are inexpensive, reliable, and have minimum lubrication requirements. Tapered roller bearings are used in the hydraulic units as needle bearings are not suitable at these locations.

The hydraulic unit assembly and its drive gears, along with the charge pump, are mounted entirely on the intermediate plate which is mounted to the main housing.

This type of construction allows for very easy assembly, maintenance, and gives the best possible noise isolation. The valve bodies are cast iron, the spools are hardened steel and, where applicable, steel sleeves are used. The control linkages from the driver will be of similar type and construction to those presently used in automotive applications.

Speed sensing governors are of the rotating flyweight type and act directly on a valve stem. As this transmission does not vary speed ratio by dissipating energy, such as the torque converter, the cooling capacity would be less than required for a conventional automatic transmission while the transmission fluid flow rate will be about the same. Size and Weight The tri-mode hydromechanical transmission is designed to fit within the requirements stated in paragraph 6 of the "Prototype Vehicle Performance Specification" see Appendix In brief, the transmission tunnel is not widened so as to decrease clearance between the accelerator pedal and the tunnel; the tunnel height does not affect full fore and aft movement of the front seat; it does not violate the ground clearance lines; it does not violate the space allocated for wheel jounce and steering clearances; and it does not degrade the handling characteristics of the vehicle.

The input or mounting flange is not a standard to fit the conventional internal combustion engine. However, as a reduction gearbox is required at the Rankine or Brayton cycle engine output, the mounting flange and output shaft location may be located to suit the proposed transmission.

The weight of the tri-mode transmission is 92 pounds dry. A weight breakdown is shown in Table V Design Analysis By far, the majority of components in an automotive transmission are sized by considerations other than material stress such as economy of manufacture, or requirements of fitting over or around some other component.

When weight is not a major consideration, components are often oversized to "keep out of trouble," and no heed is taken or calculations made of the exact margin of safety. Page 28 Sundstrand Aviation d,.. Appendix V-5 gives the summary of sizing this class of component along with a schematic which shows torques and speeds.

The hydraulic units are sized by proprietary Sundstrand methods to meet their rated speeds and pressures. In a study of this type where basic concept and feasibility are of prime importance, it is not appropriate to go into extensive sizing detail analysis. There are no areas in the transmission that are so critical that any increase in component size, that may be required after a detailed design study, would precipitate any significant cost performance or weight penalty.

Mechanical Operation This subsection is a discussion of the mechanical operation of the traction drive-torque converter transmission.

Figure V-6 shows the general schematic. Transmission input speed is proportional to engine speed. Therefore, the speed of the input toric disk is proportional to engine speed since it is driven by a gear on the transmission input shaft. The speed of the output toric disks relative to the speed of the input toric disk is a function of the inclination of the traction rollers.

The speed ratio across the traction drive is the same as the ratio of the radius of rolling contact on the output toric disk to the radius of rolling contact on the input toric disk with respect to the axis of the traction drive. Transmission ratio changes are effected by changing the "tilt angle" of the roller axis which varies the radius of the two points of contact with the toroids. The "tilt angle" of the rollers can be changed by either of 2 methods: a Application of an external force to the roller mounting yoke and physically forcing the axis of rotation of the roller to the required angle.

An explanation of how this is achieved is as follows: Figure V-7a shows a cross-section of the traction unit with the roller in the ratio position.

Figure V-7b shows the roller in an end view of the traction unit. The velocity vector of the roller at point of contact with the toroid is shown by vector "V". With the roller positioned as shown with zero slip between the roller and the toroid, vector "V" also represent the velocity vector of the toroid. Figure V-7c shows the axis of rotation of the roller displaced an amount "X" to the left of a parallel center line ggoing through the axis of rotation of the toroid.

The velocity vectors at the point of contact between the roller Vp and the toroid Bj bonger coincide. This difference causes a relative slip between the two members, represented by "Vg". This slip vector will be "down" at the point of contact between the roller and the toroid shown, and "up" as the other point of contact because the other toroid is rotating in the opposite direction. These two equal and opposite speed vectors produce equal and opposite forces on the roller which, if unrestrained at the roller bearing support will cause a turning movement on the roller, at right angles to its axis of rotation, which will "steer" the roller to some new angle to achieve equilibrium.

The angle to which the roller axis will tilt to achieve equilibrium is a function of the distance "X". The proposed design utilizes the "steering force" approach. Translation of the roller axis of rotation is accomplished by applying a force to them through the hydraulic suspension and control cylinders.

When the desired change in ratio is achieved, the steering forces will be cancelled and the unit will operate at the new ratio until the next ratio change is requested by the control system. The traction rollers are hydraulically interconnected in, such a way that their tangential loads, rather than their absolute positions, must correspond. Therefore, load sharing is positively assured. Mechanization of this approach utilized in the design is defined by an English patent by McGill.

This tangential load is sensed by the hydraulic control-suspension system, and the hydraulic pressure thus generated is applied to the hydraulic thruster which produces the axial clamping force across the rollers. Thus, the normal force necessary to allow a torque producing tangential force to develop at the roller contacting points is directly propcrtional to the torque being transmitted.

The normal force then is only as large as it must be to prevent traction roller skidding, and unit life which is inversely proportional to the cube of the normal force is greatly extended. Initial pre-load is provided by a belville washer which develops sufficient initial force to allow charge pressure build-up. This force is negated when charge pressure is applied moving the piston out of contact with the toroid See Appendix V The one-way clutch is provided between the output of the traction drive unit and the transmission housing.

This clutch prevents the output of the traction drive unit from rotating backwards such as would happen if the vehicle were allowed to roll backwards while engaged in forward drive. This reverse rotation of the traction drive could cause the traction rollers to "steer" themselves out of position. The output of the traction drive is connected directly to the torque converter input member, the impeller. The speed of the torque converter output member, the turbine, is a function of vehicle speed and the ratio of the transmission output gears.

It serves to bring the relatively high torque converter speed down to a more favorable transmission output speed, and it also provides the capability for reverse vehicle operation. Hardware Description The following is a brief description of the principal components of the traction drive torque converter transmission. Reference may be made to the cross section drawing, A-L4, shown in Appendix V-4 for indication of component arrangement and relative size. The elements are typical automotive pressed steel construction scaled down in size from a standard automotive converter.

The maximum diameter of the oil path is 6. There are three 2. The rollers rotate at a radius of 1. At the forward side of the first toroid disk is the variable thrust device. A Belleville type spring imposes a pound thrust preload on the toroids and rollers. This preload is held constant as the control pressure builds up sufficiently to overcome the constant spring force. From then on the clamping force is directly proportional to the control pressure.

Maximum control pressure is psi. The traction roller steering and suspension mechanism is all hydraulic and is based on an existing design which ensures the accurate load sharing described earlier.

All seals are typical of those found in standard automatic transmissions. Maximum flow is approximately 6 GPM at maximum engine speed. The speed sensing governor is the rotating flyweight type acting directly on a valve stem.

Control linkages may be of similar type as presently used with automatic transmissions. The capacity required is equivalent to the present automotive automatic type transmission cooler. Size and Weight The traction transmission is designed to fit within the requirements stated in paragraph 6 of the "Prototype Vehicle Performance Specification" see Appendix However, as a reduction gearbox is required at the Rankine or Brayton cycle engine output, the mounting flange and output shaft location may be made suitable for the proposed transmission.

The weight of the traction transmission is 77 pounds dry. Design Analysis The double row traction transmission is basically designed from a stress-cycle curve and previous experience in designing and testing traction drives.

The maximum input power requirement was calculated from the acceleration and grade velocity requirements. These input powers, the specified life of hours, a transmission speed range of five to one, and appropriate toroid geometry ratios with a particular traction coefficient provide the basis for the traction transmission design.

The toroid geometry ratios involved in the design are 1 the toroid pitch diameter to roller diameter ratio, and 2 the conformity ratio, which is defined as the ratio of roller crown radius to roller pitch radius. The first ratio 1 defines the size of the machine and the amount of rolling to twisting contact that the rollers experience with the toroids. Since large toroid pitch radius to roller radius ratios approach more nearly pure rolling, the traction coefficient increases and the speed range decreases with this ratio.

Therefore, in order to accommodate a 5 to 1 speed range and stay within package size limits, a ratio of 1. The second ratio 2 affects the shape and size of the footprint as well as the normal stresses. Higher conformity ratios for the same load result in higher stresses. Experience dictates a circular footprint or one that has its major axis in the direction of rolling. The traction coefficient also affects the overall transmission size and decreases as rolling contact velocity increases.

For this design, at an input speed of RPM, the rolling contact velocity is inches per second. A traction coefficient of 0. The maximum stresses calculated for this design are ksi at maximum input power of HP and ksi at the mean input power. Mean input power is weighted average power over the EPA combined driving cycle as defined by the duty cycle. Using an assumed stress cycle curve and scaling from 1,, cycles at ksi for M50 tool steel, it was determined that the hour life requirement was satisfied.

Appendix V-6 shows a schematic of the transmission with individual component speeds and torques. The actual component sizing summary is the same as for the hydromechanical sizing shown in Appendix V This is especially true when the design is being made by personnel with many years of transmission experience.

There are no areas in the transmission that are so critical that any increase in component size, that may be required after a detailed design study, would precipitate any significant cost, performance, or weight penalty. Traction ratio and torque converter optimization An optimization study was carried out to determine the required system gear ratios, traction drive ratio range, and torque converter type and size.

Many different torque converters were simulated and studied to gain a better understanding of the effects of converter characteristics on vehicle performance. Another important factor was torque converter diameter. Making the converter diameter larger makes it "tighter;" that is, it slips less, and is therefore more efficient.

However the power absorbed at engine idle by a torque converter also increases with diameter, and must be considered. A study was also made replacing the torque converter with a friction clutch. It should be noted that in realizing these gains, the advantages of having a torque converter as discussed in Section IV are lost. It would appear that these advantages outweigh the efficiency disadvantage.

However, it is not completely evident that a clutch could not be used. A more detailed study of this would be made prior to a hardware design commitment. Studies were also made using a torque converter lock-up clutch, and an input clutch in the system.

The result of these studies, and some of the other optimization studies, are summarized in Table V Figures are for the Aerojet Rankine engine. The parameters chosen for the final transmission design were a torque converter ratio of 3 to 1 and a transmission speed ratio of 5 to 1. These result in only 3 HP absorbed at idle speed and A trade-off study of all of the studies and computer runs involving complexity, cost, and overall economics resulted in the choice of the above parameters.

Maintainability It is expected that either the tri-mode or the traction transmission should provide no greater maintainability problems than present automotive automatic transmissions. Ratio Converter Stall Torq. Converter Stall Torq. Ratio 2. NOISE The transmission noise whether air-borne or structure borne is an important consideration for any automotive transmission.

It is of particuliar concern because the vehicle levels required are relatively low. The hydromechanical transmission is inherently a higher noise generation source than the traction transmission. Hydromechanical The primary potential noise source is the hydraulic units and the secondary source is the gears.

Solution to the latter is represented by fairly well known techniques utilized and demonstrated in millions of automotive type transmission. Such techniques will be utilized in the recommended configuration to minimize noise. Of primary importance will be the gear tooth profile and speeds which will be similiar to present automotive transmissions. Considerable effort has been expended in the last few years to understand and reduce hydraulic unit noise. The cause is fairly well known and techniques have been developed to minimize it.

However it must be recognized that because of the large number of variables involved, the only positive assurance of meeting the required noise levels comes through actual hardware demonstrations. The hydraulic units represent the major noise source.

This source is primarily related to the rate of generation of high pressure from low pressure and vice versa, the level of maximum pressure and the porting rate rotational speed. Considerable experience has been gained in the last few years in minimizing porting noise. This is accomplished by modifying the ports between the cylinder block and port plate to prevent large, abrupt pressure transients.

Another means of minimizing the noise is to limit the maximum working pressure within the unit. In the recommended configuration, the working pressure is limited to psi, which would only occur with "floored accelerator" below about 20 MPH.

Hydraulic unit operational speeds can be selected to insure the best noise characteristics. Therefore the variables involved are pressure level, rate of pressure increase or decrease in the individual pistons during parting, porting modifications, hydraulic unit speed and to some degree the stroke or displacement of the hydraulic units. Attenuation of the generated noise to the outside of the transmission is very important. The attenuation itself is very important but it is also important to insure that component natural frequencies are such that no resonants occur.

Minimizing resonances will simplify energy attenuation techniques. Also noise frequencies should be kept as high as possible as attenuation is much easier at higher frequencies. Air-borne noise within the transmission to the main housing has been suppressed by a deep-drawn sound shield made from a special laminated sandwich around the hydraulic unit rotating components.

The oil pan is formed from the same material to prevent the air to fluid-borne noise from being transferred outside. The laminated sandwich is a composite of two metal plates with 2 layers of viscoelastic material between them seperated by a steel screen.

This isolation material has a high crush force and good attenuation above Hz. This double barrier should be very effective in minimizing noise propogation. In addition to this, a similar type of isolation is provided between the main input and output transmission bearings and the main housing, thus eliminating any "hard path" between the noise producing dynamic components and the main housing. As indicated previously noise tends to be in the category of "black art".

Extensive testing and evaluation has defined design techniques which will minimizing noise. Although it is impossible to know at this time what the final noise level will be, it is anticipated that the noise requirements will be met. Traction The primary source of noise in the traction transmission are the gears. The output gears are constant mesh and of the helical type similar to present automotive practice. The input gear mesh to the input toroid is constant mesh and is shown as a spur gear.

The input toroid cannot tolerate any external thrust. To ensure lowest possible noise generation, these input gears will be fine pitch, low pressure angle, and with a modified involute profile. Should it become necessary to further reduce the noise from the input mesh, helical gears with thrust runners directly between the gears to cancel the resultant thrust will be used.

In addition to reducing gear noise to a minimum at its source, laminated sound insulating bearing sleeves are used to isolate the noise from the- main housing. The only normal maintenance required will be to check the transmission oil level as is now done. Repair or overhaul of the transmission should not require any additional complication. The only "new to the business" component would be the hydraulic units or the toroids and rollers. It would be expected that these assemblies would be provided to the garage or overhaul shop as reworked assemblies similar to present torque converter assemblies.

This efficiency is reflected as fuel consumption over the defined driving cycle. Acceleration data is also presented in several forms. This also represents total vehicle system performance. Ground Rules and Transmission Parameter Summary The following ground rules for all performance calculations were either specified or mutually agreed to by the Environmental Protection Agency.

Although differences in air temperature do make a difference in air drag forces, their inclusion is somewhat meaningless without corresponding data on variation in engine performance with temperature, which was unavailable. Accessory power requirements Prototype Vehicle Performance Specification. NOTE: Performance specification accessory losses representative of engine speed range.

For narrower speed range engines, the idle accessory power requirements were assumed unchanged, but the accessory power requirements at maximum engine speed were reduced proportionately. See Appendices and The driving cycle used to calculate fuel consumption was the Combined Duty Cycle. Acceleration and fuel economy performance for the referenced typical 3-speed automatic transmission!

Rankine 7. Transmission Efficiency Transmission efficiency has been calculated for both the tri-mode hydromechanical transmission and the traction drive transmission for both the Aerojet Rankine engine and the AiResearch Brayton engine. Conditions of output speed and load for which transmission efficiency tabulations and graphs have been calculated include: 1 The Federal Driving Cycle, 2 The Simplified Suburban Route, 3 The Simplified Country Route, 4 The Combined Driving Cycle a combination of 1, 2, and 3 , 5 Constant Vehicle Speed cruise , and 6 Part Load tractive effort at , 75, 50, 25, and 10 percent of maximum acceleration tractive effort.

The instantaneous transmission efficiency for each point in the driving cycle was calculated. Also, an accumulative efficiency, that is, an average efficiency, for the driving cycles was calculated and is presented as part of this report see Table VI-1 and Figure VI-1 through VI This average efficiency represents the quiotent of the accumulative power utilized over the given driving cycle and the accumulative power supplied.

Two computer programs, one for systems using hydromechanical transmissions and the other for systems using traction drive transmissions were used to simulate the vehicle, the engine, the transmissions, and the required duty cycles to generate the efficiency data.

In the two programs, every effort was made to simulate the system's realistically. Therefore, the absolute values of efficiency presented in this report should be representative of actual hardware. It should also be emphasized that since the two programs were developed together, the relative efficiencies of the systems considered are also quite meaningful.

Transmission efficiency as used in this report is defined as the total power out of the transmission output divided by the total power into the transmission input. The primary or engine gear reduction has been assumed by Sundstrand to be part of the engine gearbox and is therefore not reflected in the transmission efficiency data presented here. Calculations for the power losses contributed by gears and bearings, planetaries, open clutch spinning, charge pumps, and torque converters are well known and accepted.

The following paragraphs describe the background used in calculating hydraulic unit and traction unit efficiencies or losses. This efficiency is markedly reduced below certain levels of working pressure and displacement or wobbler angle. The hydraulic unit design, and the predicted operating efficiencies used in this study are based on the testing and field experience of the past 30 years. The present axial piston-hydrostatic bearing design has evolved from past experience with many hydraulic unit configurations including radial piston units and anti-friction thrust bearino units, and has proved to be the best design in terms of cost, size, efficiency, and reliability.

See Appendix I. Consequently, there is a power loss due to roller spin. Also rolling resistance is encountered between the toroids and the rollers, as well as in the rolling element bearings.

The overall efficiency of the traction unit is the product of the speed efficiency and the torque efficiency. The speed efficiency is a measure of slip. The torque efficiency, in general, is a measure of spin loss, rolling resistance, and windage. The efficiency of the traction unit was calculated using experience gained from the development and testing of a Sundstrand toroidal type variable input speed constant output speed traction drive for aircraft applications.

The Sundstrand efficiency data correlates well with data published by General Motors and Tracor on the efficiency of rolling contacts. Grade and Acceleration Performance Grade and acceleration performance was calculated for both transmissions and engines.

The grade performance is a function of engine power and transmission efficiency. Acceleration performance is a function of engine power, transmission characteristics, drive line efficiency, tire adhesion, the engine time lag in going from idle to the maximum power condition, and the ratio of engine power going into accelerating the engine, to that which is accelerating the vehicle during the time lag period.

Because of the many variables involved several assumptions were made: 1 Maximum acceleration can be achieved by allowing the engine to accelerate to the maximum power condition unloaded and then applying maximum power to the wheels. For the required mph acceleration time and the distance traveled in 10 seconds, an engine acceleration time of 0. For the mph and mph acceleration times, an engine acceleration time of 0.

In practice, these time lags would probably be unacceptable from the "driver feel" point of view, and to overcome this, the engine power during this engine acceleration period would be split, some going to accelerate the engine, and some to accelerate the vehicle. The exact ratio of this power split would depend very much on "driver feel" and would be determined experimentally. Regardless of the split, it has been assumed that the mph and sec. The actual inertia is not only small but is reduced by the square of the gear ratio between the two.

The power available from the Aerojet Rankine cycle engine, HP at zero vehicle speed increasing to HP at 85 mph, appears to be adequate to meet all performance requirement limits.

No problem was encountered in meeting the gradeability requirements. The maximum achievable vehicle speed along with the corresponding engine power requirements are tabulated in Table VI Table VI-2 also lists the actual acceleration performance of the various systems, taking into account engine lag. Also tabulated are the performance requirement limits. The plots are based on a start from maximum power condition as can be achieved by locking the brakes.