A mechanical power transmission system for an engine or motor driven vehicle is disclosed. The transmission of the present invention includes an input transmission, a compound planetary transmission, and an output transmission. The compound planetary transmission includes two input and two outputs. The input transmission has a single input connected to the engine or motor of the vehicle and two outputs connected to the two inputs to the planetary transmission. The output transmission includes two inputs connected to the two outputs of the planetary transmission and one output connected through suitable gearing to the wheels of the vehicle or any other mechanism to be driven by a motoring device at various speed ratios.

FIELD OF INVENTION
 The present invention relates to a multispeed mechanical transmission with
 an optional hydrostatic attachment, useful in, but not limited to
 agricultural tractors.
 BACKGROUND OF THE INVENTION
 The prior art is replete with various transmissions for agricultural
 tractors and the like. Multispeed transmissions having countershafts are
 widely used in the power train of tractor arrangements because a plurality
 of rotating clutch assemblies and associated gears can be positioned on
 parallel shafts to allow considerable flexibility in adapting them to
 different space requirements and "gear spacing".
 "Gear spacing" is the ratio change between gears which produces the change
 in vehicle speed when the operator shifts to a different gear. The smaller
 this gear spacing the better the optimum engine speed can be matched to
 the optimum ground speed. The more gear selections that are available, the
 finer the gear spacing can be designed. However, the number of clutches
 and gears increases with added gear selections, increasing the cost, etc.
 "Shift quality" is the operator's perception of how smoothly a transmission
 reacts when making a shift. Many factors affect shift quality, such as
 rapid changes in speed of elements with large inertia within the
 transmission, poor timing of the pressure control, large torque
 interruptions at heavy loads, large gear spacing, and most of all, the
 number of clutch "swaps" required from one gear selection to the next. (A
 single clutch swap is defined as the disengagement of one clutch and the
 engagement of another clutch to complete a shift.) All currently
 manufactured powershift transmission have multiclutch swaps during some
 shifts in the operating range. It is difficult, if not impossible, to make
 multiclutch swaps smooth because during a shift, one or more of the
 engaging clutches opposes the direction of the shift. For example, in one
 typical transmission, during a triple clutch swap upshift from 6.sup.th to
 7.sup.th gear, one of the clutches shifts up while the other two clutches
 shift down.
 Any sequence of clutch engagements will cause torque reversals. This effect
 is inherent in all multiswap shifts of current designs. Only single clutch
 swaps can be shifted smoothly by overlapping the engagement of oncoming
 clutch with the disengagement of the outgoing clutch.
 There are special applications for the type of vehicle using the
 transmission described herein where exact speed control is important, such
 as trenching or certain planting and harvesting operations, among others.
 In many of these applications, most of the engine power is used to drive
 the mechanism of the towed attachment through the power take off (PTO)
 with only part of the power used for the forward motion of the vehicle.
 Here, transmission efficiency is of secondary importance, with the primary
 importance being the ability to vary the ground speed at small increments
 independently of the engine. In such cases, an continuously variable
 transmission is desirable. Continuous variability can be achieved, for
 example, through the use of hydrostatic units, in which case, transmission
 efficiency is sacrificed. Other continuously variable transmissions
 include electrical generator-motor sets or variable friction drives.
 Two types of the continuously variable transmission are of interest:
 1. Continuous variability from a certain minimum vehicle speed to the
 maximum vehicle speed with full power transmission capability in one mode
 of operation, and continuous variability from zero to a certain low
 vehicle speed with maximum traction capability in another mode of
 operation. In the full power mode a means (such as a clutch) must be
 available to start the vehicle in motion at full load. As pertains to the
 present invention, this type will be called a partial continuously
 variable transmission (PCVT).
 2. Continuous variability from zero to maximum vehicle speed with full
 power and maximum traction capability within a single mode of operation.
 As pertains to the present invention, this type will be called a full
 continuously variable transmission (FCVT).
 SUMMARY OF THE INVENTION
 It is the object of the present invention to provide the largest known
 number of gear selections with a given number of clutches. For example,
 current transmissions with 9 clutches provide between 16 and 18 speeds
 forward with 4 to 6 reverse speeds. A typical arrangement of the present
 invention provides 23 forward and 8 reverse speeds using 9 clutches. These
 five extra speeds allow a closer gear spacing.
 It is a further object of this invention to provide a variety of
 transmission systems from which to choose for specific applications.
 It is a further object of the present invention to provide single clutch
 swaps for all single step and double step shifts throughout the normal
 operating range. Single step means sequential shifts and double step means
 skipping a gear selection.
 It is a further object of the present invention to maximize the efficiency
 by keeping the number of gear meshes low. Also, the clutch sizes and
 speeds can be minimized for low losses.
 It is a further object of the present invention to provide continuous
 variability of the type PCVT via an optional attachment. In this case, the
 present invention proposes the use of modulated clutches for starting the
 vehicle motion, thus allowing for the use of relatively small hydrostatic
 units.
 It is a further object of the present invention to provide an alternate
 transmission system with continuous variability of the type FCVT. A
 continuously variable transmission from zero to maximum vehicle speed is
 realized by making the hydrostatic units large enough for sufficient
 torque to start the vehicle motion.
 It is a further object of this invention to provide a compact transmission
 package using a unique gear, shaft, clutch and bearing arrangement.
 General Description of the Basic Powershift Transmission
 "Transmission" as used in this document is an arrangement of gears, shafts,
 clutches and bearings located in a housing for the purpose of transmitting
 rotational power, "input" meaning power absorbing, "output" meaning power
 delivering. The transmission system consists of an input transmission, a
 compound planetary system and an output transmission.
 A compound planetary system is a planetary system consisting of at least
 two simple planetaries.(FIGS. 16, 17 and 18) There are two kinds of simple
 planetaries, the single and the double planetary. These are defined as
 follows.
 A single planetary is comprised of a internal ring gear, a sun gear and a
 set of planets (usually three) which are rotatably mounted on sets of
 shafts anchored in a planet carrier. Each planet gear is in mesh with the
 ring gear and with the sun gear thus acting as an idler gear between the
 ring and the sun gear.
 A double planetary is comprised of an internal ring gear, a sun gear and
 two sets of planet gears which are rotatably mounted on two sets of shafts
 each set of shafts being anchored in a planet carrier. One set of planets
 is in mesh with the ring gear and with the other set of planets which in
 turn is in mesh with the sun gear.
 Another type of compound planetary system uses cluster planet gears. (as
 schematically represented in FIGS. 19, 20 and 21) Usually an equivalent
 planetary system of the first type described above can be constructed. The
 input transmission has one input member connected to the engine or motor
 and two output members each of which is selectively connectable to the
 input member by clutches through various gear ratios in forward and
 reverse. The output members are connected to the two input members of the
 compound planetary system.
 The compound planetary system has two input members connected to the input
 transmission (as stated in the previous paragraph) and two output members
 connected to the output transmission.
 The output transmission has two input members connected to two output
 members of the planetary system (as stated in the previous paragraph) and
 one output member connected to the rear and/or to the front wheel drive of
 the tractor. The output member is selectively connectable by clutches
 through various gear ratios to each of the two input members. Various
 accessory drives can be included wherever convenient to the input and
 output transmission.

Description of One Typical Embodiment of the Basic Powershift Transmission
 In what follows, a specific transmission system (shown in FIG. 1) is
 described. It is understood, however, that the invention is not limited to
 this specific example. The compound planetary system consists of one ring
 gear R, two sun gears, S.sub.1 and S.sub.2, and one planet carrier C on
 which two sets of planets, P.sub.i and P.sub.O, are rotatably mounted by
 two sets of shafts.
 The input transmission of this system has two sets of three clutches. One
 set, denoted F, REV, and B, selectively connects the input shaft to the R
 shaft through various gears or it stops the R shaft. The other set,
 denoted f, rev, and b, selectively connects the input shaft to the S.sub.1
 shaft through various gears or it stops the S.sub.1 shaft.
 The output transmission has two sets of clutches. One set, denoted 1, 3,
 and 5, selectively connects the output shaft to the C shaft through three
 corresponding gear ratios. The other set, denoted 2, 4, and 6, selectively
 connects the output shaft to the S.sub.2 shaft through three corresponding
 gear ratios.
 The compound planetary systems shown in FIGS. 16, 17, 18 and 25 is
 incorporated in FIGS. 1, 3, 4, 5, 6, 8, 9, 11, 12, and 14. It is comprised
 of a single input planetary P.sub.i and a double output planetary P.sub.O.
 The members of the two planetaries are connected as shown: The ring gear R
 of the planetary P.sub.i is connected to the ring gear of the planetary
 P.sub.O, the carrier of the planetary P.sub.i is connected to the carrier
 C of the planetary P.sub.O. The two input connections, R and S.sub.1, are
 attached to the ring gear and to the sun gear of the planetary P.sub.i,
 respectively, and the two output connections, C and S.sub.2, are attached
 to the carrier and the sun gear of the output planetary P.sub.O,
 respectively.
 The compound planetary system has the following characteristics:
 1. The ratio of speeds of any two of the four members (two input and two
 output members) determines the ratio of speeds of all members to each
 other.
 2. At a constant speed of one input member (the ring gear R), as the other
 input member (the sun gear S.sub.1) is varied from a negative speed to
 zero to a positive speed, the speed of one output member (sun gear
 S.sub.2) decreases and the speed of the other output member (planet
 carrier C) increases. (Of course, as the sun gear S.sub.1 is varied from a
 positive speed to zero to a negative speed, the speed of S.sub.2 increases
 and the speed of C decreases.)
 An alternate planetary system, equivalent to the one described above, is
 shown in FIGS. 19, 20, 21 and 26. The first planet set P.sub.i is made up
 of three cluster gears each having a gear on the left and a gear on the
 right. The gears on the left are in mesh with the ring gear R at the
 outside and with the first sun gear S.sub.1 at the inside. The gears on
 the right are in mesh with the gears of the second planet set P.sub.O. The
 second planet set P.sub.o is also in mesh with the second sun gear
 S.sub.2. Both planet sets, P.sub.i and P.sub.O, are rotatably mounted to
 the same carrier C.
 A second alternate planetary system is shown in FIGS. 22, 23, 24 and 27.
 This compound planetary system is comprised of two single planetaries
 P.sub.i and P.sub.O, each having a ring gear, a sun gear and planet gears
 rotatably mounted by sets of shafts in respective planet carriers. The
 members of the two planetaries are interconnected as shown: The ring gear
 of the planetary P.sub.i is connected to the carrier of the planetary
 P.sub.O and the carrier of the planetary P.sub.i is connected to the ring
 gear of the planetary P.sub.O. The two input connections, R and S.sub.1,
 are attached to the ring gear and to the sungear of planetary P.sub.i,
 respectively. The two output connections, C and S.sub.2, are attached to
 the ring gear and to the sun gear of planetary P.sub.O, respectively.
 A third alternate planetary system is shown in FIG. 28. This compound
 planetary system is comprised of two double planetaries P.sub.i and
 P.sub.O, each having a ring gear, a sun gear and planet gears rotatably
 mounted by sets of shafts in respective planet carriers. The members of
 the two planetaries are interconnected as shown: the ring gear of the
 planetary P.sub.i is connected to the ring gear of the planetary P.sub.O
 and the carrier of the planetary P.sub.i is connected to the sun gear of
 the planetary P.sub.O. The two input connections, R and S.sub.1, are
 attached to the ring gear and to the sun gear of planetary P.sub.i,
 respectively. The two output connections, C and S.sub.2, are attached to
 the carrier and to the sun gear of planetary P.sub.O, respectively.
 A fourth alternate planetary system is shown in FIG. 29. This compound
 planetary system is comprised of two single planetaries P.sub.i and
 P.sub.O, each having a ring gear, a sun gear and planet gears rotatably
 mounted by sets of shafts in respective planet carriers. The members of
 the two planetaries are interconnected as shown: The ring gear of the
 planetary P.sub.i is connected to the ring gear of the planetary P.sub.O
 and the sun gear of the planetary P.sub.i is connected to the sun gear of
 the planetary P.sub.O. The two input connections, R and S.sub.1, are
 attached to the carrier and to the sun gear of planetary P.sub.i,
 respectively. The two output connections, C and S.sub.2, are attached to
 the ring gear and to the sun gear of planetary P.sub.O, respectively.
 A fifth alternate planetary system is shown in FIG. 30. This compound
 planetary system is comprised of a double planetaries P.sub.i and a single
 planetary P.sub.O, each having a ring gear, a sun gear and planet gears
 rotatably mounted by sets of shafts in respective planet carriers. The
 members of the two planetaries are interconnected as shown: The ring gear
 of the planetary P.sub.i is connected to the carrier of the planetary
 P.sub.O and the carrier of the planetary P.sub.i is connected to the sun
 gear of the planetary P.sub.O. The two input connections R and S.sub.1 are
 attached to the ring gear and to the sun gear of planetary P.sub.i
 respectively, and the two output connections C and S.sub.2 are attached to
 the ring gear and to the sun gear of planetary P.sub.O respectively
 A sixth alternate planetary system is shown in FIG. 31. This compound
 planetary system is comprised of a single planetaries P.sub.i and a double
 planetary P.sub.O, each having a ring gear, a sun gear and planet gears
 rotatably mounted by sets of shafts in respective planet carriers. The
 members of the two planetaries are interconnected as shown: The ring gear
 of the planetary P.sub.i is connected to the sun gear of the planetary
 P.sub.O and the carrier of the planetary P.sub.i is connected to the ring
 gear of the planetary P.sub.O. The two input connections R and S.sub.1 are
 attached to the carrier and to the sun gear of planetary P.sub.i
 respectively, and the two output connections C and S.sub.2 are attached to
 the carrier and to the sun gear of planetary P.sub.O respectively
 A seventh alternate planetary system is shown in FIG. 32. This compound
 planetary system is comprised of a single planetary P.sub.i and a single
 planetary P.sub.O, each having a ring gear, a sun gear and planet gears
 rotatably mounted by sets of shafts in respective planet carriers. The
 members of the two planetaries are interconnected as shown: The carrier of
 the planetary P.sub.i is connected to the ring gear of the planetary
 P.sub.O and the sun gear of the planetary P.sub.i is connected to the sun
 gear of the planetary P.sub.O. The two input connections R and S.sub.1 are
 attached to the carrier and to the sun gear of planetary P.sub.i
 respectively, and the two output connections C and S.sub.2 are attached to
 the carrier of the planetary P.sub.O and to the ring gear of planetary
 P.sub.i, respectively.
 An eighth alternate planetary system is shown in FIG. 33. This compound
 planetary system is comprised of a single planetaries P.sub.i and a single
 planetary P.sub.O, each having a ring gear, a sun gear and planet gears
 rotatably mounted by sets of shafts in respective planet carriers. The
 members of the two planetaries are interconnected as shown: The carrier of
 the planetary P.sub.i is connected to the carrier of the planetary
 P.sub.O, and the sun gear of the planetary P.sub.i is connected to the sun
 gear of the planetary P.sub.O. The two input connections R and S.sub.1 are
 attached to the ring gear and to the sun gear of planetary P.sub.i
 respectively. The two output connections C and S.sub.2 are attached to the
 carrier and to the ring gear of planetary P.sub.O, respectively
 A ninth alternate planetary system is shown in FIG. 34. This compound
 planetary system is comprised of a single planetaries P.sub.i and a double
 planetary P.sub.O, each having a ring gear, a sun gear and planet gears
 rotatably mounted by sets of shafts in respective planet carriers. The
 members of the two planetaries are interconnected as shown: the carrier of
 the planetary P.sub.i is connected to the carrier of the planetary
 P.sub.O, and the sun gear of the planetary P.sub.i is connected to the
 ring gear of the planetary P.sub.O. The two input connections, R and
 S.sub.1, are attached to the ring gear and to the sun gear of planetary
 P.sub.i, respectively. The two output connections, C and S.sub.2, are
 attached to the carrier and to the sun gear of planetary P.sub.O,
 respectively
 In the following description power flow may be positive or negative.
 The input transmission absorbs engine power through the input shaft,
 transmitting power selectively through two pairs of clutches to the two
 input members, S.sub.1 and R, of the compound planetary. The first pair of
 clutches is denoted f and F, and the second pair is denoted rev and REV.
 The first pair of clutches is mounted on the input shaft to selectively
 drive the ring gear R through clutch F, via a gear set 34F to 42, and to
 selectively drive the sun gear S.sub.1 through clutch f, via a gear set
 23f to 53.
 The second pair of clutches is mounted on the reverse shaft. This shaft is
 driven through a 41 to 41r gear set from the input shaft. The reverse
 shaft is located such that the 34Rev gear is in mesh with the 42 gear and
 the 23rev gear is in mesh with the 53 gear. Thus, the selective engagement
 of the clutches REV and rev will respectively drive the ring gear R and
 the sun gear S.sub.1 in the opposite direction at the same speeds as the F
 and f clutches.
 In addition to these two pairs of clutches, there are two clutches B and b.
 Engagement of B will stop the ring gear R and of b will stop the sun gear
 S.sub.1.
 The output transmission absorbs power selectively through two shafts
 S.sub.2 and C. Power is transmitted through either one shaft or the other
 or through both shafts in certain proportions. The shafts of both S.sub.2
 and C have clutches mounted on them which in combination with the clutches
 mounted on the output shaft provide three selective gear ratios each from
 S.sub.2 to the output shaft and from C to the output shaft. The even
 numbered clutches, 2, 4 and 6, selectively connect the S.sub.2 shaft to
 the output shaft at certain ratios provided by the gear sets 16(2) to
 64(2), 30(4) to 50(4) and 48(6) to 33(6), respectively. The odd numbered
 clutches, 1, 3 and 5, selectively connect the C shaft to the output shaft
 at certain ratios provided by the gear sets 16(1) to 64(1), 30(3) to 50(3)
 and 48(5) to 33(5), respectively.
 To establish a connection from the engine to the output shaft of the output
 transmission either:
 (1) Two input clutches and one output clutch must be engaged, or
 (2) One input clutch and two output clutches must be engaged.
 (1) Two input and one output clutches engaged:
 Two input clutches are engaged, one driving the ring gear R at a certain
 ratio with respect to the engine, and one driving the sun gear S.sub.1 at
 a certain other or same ratio with respect to the engine, will establish a
 ratio of two members of the planetary system; also one input clutch F or
 REV driving the ring gear R with the sun gear S.sub.1 stopped (b engaged),
 or one input clutch f or r driving the sun gear S.sub.1 with the ring gear
 R stopped (B engaged), will establish a ratio of two members R and S.sub.1
 of the planetary system. Then by Item (1) on page 4, the ratios of all
 members of the planetary system are established. Thus, the engagement of
 any one clutch (1, 2, 3, 4, 5, or 6) in the output transmission will
 establish an overall ratio from the input (engine) to the output of the
 transmission system.
 (2) One input and two output clutches engaged:
 Two output clutches engaged, with one of the clutches 2, 4, or 6 connecting
 the sun gear S.sub.2 to the output shaft of the output transmission at a
 certain ratio, and one of the clutches 1, 3, or 5 connecting the carrier C
 to the output shaft at a certain other or same ratio, will establish a
 ratio between S.sub.2 and C. Again, by Item (1) on page 4, the ratios of
 all members of the planetary system are established. Thus, the engagement
 of any one input clutch, F or Rev, which connects the engine to the ring
 gear R, or any one input clutch f or rev, which connects the engine to the
 sun gear S.sub.1, will establish a ratio from the engine to the output
 shaft of the transmission system.
 Engaging the clutches as in either (1) or (2) above establishes a "gear"
 for the transmission system. The number of possible "gears" is the total
 number of possible combinations in (1) and (2). Thus,
 F and rev combined with 6 output clutches=6 forward "gears",
 REV and rev combined with 6 output clutches=6 reverse,
 F and b combined with 6 output clutches=6 forward,
 REV and b combined with 6 output clutches=6 reverse,
 F and f combined with 6 output clutches=6 forward,
 REV and f combined with 6 output clutches=6 reverse,
 f and B combined with 6 output clutches=6 forward, rev and B combined with
 6 output clutches=6 reverse,
 F at input combined with 9* combinations at output=9 forward,
 REV at input combined with 9* combinations at output=9 reverse,
 f at input combined with 9* combinations at output=9 forward,
 rev at input combined with 9* combinations at output=9 reverse

FNT * The 9 combinations at output are: 1-2, 1-4, 1-6, 2-3, 2-5, 3-4, 3-6, 4-5,
 5-6.
 Thus, there are 42 forward "gears" and 42 reverse "gears" possible in the
 transmission schematically illustrated in FIG. 1.
 In order to provide single clutch swap shifts, a certain sequence of clutch
 combinations for each "gear" is chosen, as shown in column 2 of FIG. 2. By
 selecting certain gear ratios for all gears in the system it is possible
 to provide reasonably equal geometric steps (11+ or -2 % shown as % speed
 change in column 9 in FIG. 2) from one gear selection to the next by
 choosing the sequence shown in FIG. 2. Note that some of the possible
 combinations of clutches are not used because they either provide
 redundant speeds or they will not fall within the single clutch swap
 sequence. Thus, from the 42 possible combinations available only 33 are
 used.
 FIG. 1 schematically illustrates the various clutches engaged in each gear
 selection (column 2) with the corresponding speeds of each shaft in the
 typical transmission described. Note the clutch REV and the clutch B are
 considered optional at added cost. The "standard" transmission has only 27
 speeds forward and 9 speeds reverse. The addition of the clutch REV will
 make these 9 speeds reverse redundant. Therefore, a 27 speed forward and
 27 speed reverse will result by adding the REV clutch. The addition of the
 clutch B will add 6 forward and 6 reverse speeds. These 6 speeds are the
 creep option for applications at very low vehicle speeds.
 General Comments on the Basic Powershift Transmission
 Planetary systems, which exhibit the characteristics as described above,
 may be currently in use or described in existing patents. It is the idea
 of this invention to combine a planetary system having these
 characteristics with an input transmission and an output transmission,
 each having a plurality of selectable gear ratios.
 It is an additional idea to design the compound planetary system with two
 input means and two output means, to design the input transmission with
 one input and two output means and to design the output transmission with
 two input and one output means.
 It is an additional idea to select from a number of planetary systems the
 one that is best for the particular application, depending on the speed
 range to be covered, the number of gear selections, manufacturing
 capabilities, space limitations, etc.
 It is an additional idea of this invention to design the input transmission
 to provide two output means, each of which has a forward-reverse symmetry
 with or without a zero speed (lock) between forward and reverse. Thus,
 there is an equal reverse gear ratio corresponding to every forward gear
 ratio for each of the output means which are selectively connectable to
 two input means of the planetary system. (It is understood, however, that
 this forward-reverse symmetry is not a requirement).
 It is an additional idea of this invention to design the output
 transmission to provide two input means, each of which has several gear
 ratios connectable to the output shaft of the transmission system.
 It is an additional idea to select the gear ratios in the planetary system
 such that the following differential equation is essentially* met:

FNT * Note: Since all gear ratios are ratios of integral numbers the above
 equations may only be approximated.
 ##EQU1##
 with the constraint
 R=constant,
 S.sub.2 r.sub.2 =KCr.sub.1
 at S.sub.1 =0,
 where
 S.sub.2 is the speed of one output member of the planetary system,
 r.sub.2 is the lowest selectable gear ratio between the sun gear S.sub.2
 and the output shaft of the transmission system,
 C is the speed of the other output member of the planetary system,
 r.sub.1 is the lowest selectable gear ratio between the carrier C and the
 output shaft of the transmission system,
 S.sub.1 is the speed of the input member of the planetary system, which is
 selectively connectable to the engine at one or more ratios in forward and
 reverse,
 R is the speed of that input member of the planetary system, which is
 connected to the engine at a constant gear ratio through a clutch, which
 remains engaged during shifting in the normal operating range,
 K is a constant calculated by the following equation:
 ##EQU2##
 Note: exp. means that the number in the second brackets is the exponent to
 the number in the first brackets.
 where
 H is the highest desired rated output speed of the transmission system,
 L is the lowest desired rated output speed to be attained in the normal
 operating range (this is the range in which the clutch F remains engaged),
 N.sub.0 is the number clutches in the output transmission,
 N.sub.s is the number of clutches in the set which selectively connects the
 engine through gear ratios to the input member S.sub.1.
 It is an additional idea of this invention to make the geometric steps from
 one gear selection to the next equal to each other in a certain interval.
 For an odd number N.sub.s of clutches connectable to the sun gear S.sub.1,
 this interval is from the gear in which two output clutches are engaged to
 the higher gear in which the sun gear S.sub.1 is stopped. (As seen in FIG.
 2, the step from the gear selection F 12 to F 2 f is equal to the step
 from the gear selection F 2 f to F 2 b. In the case of an even number of
 clutches connectable to the sun gear S.sub.1, the interval is from the
 gear in which two output clutches are engaged to the first gear in which
 the direction of the speed of the sun gear S.sub.1 is changed. (As seen in
 FIG. 13, the step from F 12 to F 2 f.sub.2 is equal to the step from F, 2,
 f.sub.2, to F, 2, f.sub.1, and equal to the step from F, 2, f.sub.1, to F,
 2, r.sub.1. Note that in the last step the direction of S.sub.1 is changed
 from f.sub.1, to r.sub.1). This choice of steps will provide the best
 compromise for gear spacing throughout the total range of the transmission
 system.
 It is an additional idea of this invention to design the gear ratios in
 each branch of the output transmission such that a geometric progression
 results with the common ratio equal to K.sup.2, where K is defined above.
 Thus, the consecutive gear ratios in the branch associated with r.sub.1
 are
 r.sub.1, r.sub.1 K.sup.2, r.sub.1 K.sup.4, . . .
 and those ratios in the branch associated with r.sub.2 are
 r.sub.2, r.sub.2 K.sup.2, r.sub.2 K.sup.4, . . .
 It is an additional idea of this invention to utilize the speed ratios
 which are available by engaging two clutches of the output transmission in
 combination with one clutch of the input transmission. This increases the
 number of "gears" available for a given total number of clutches with
 respect to conventional countershaft transmissions. This also allows for a
 shift sequence with single clutch swap shifting.
 It is an additional idea of this invention to make use of the
 forward-reverse symmetry of the input transmission and the feature of
 engaging two clutches of the output transmission simultaneously as
 described above to provide a powershift transmission system with the
 unique feature of single clutch swap shifting for all sequential shifts
 and skip shifts. For a "range shift" (meaning the sequential shift into a
 "gear" in which two output clutches are engaged), the single clutch swap
 shift is accomplished by matching the ratios in the output transmission
 and in the input transmission to the planetary system such that the speed
 ratios provided by the engagement of two output clutches and one input
 clutch (F) falls in between the speed ratios provided by the engagement of
 one of the two output clutches combined with two input clutches and (in
 between) the engagement of the other of the two output clutches combined
 with the same two input clutches. Thus, the engagement of the output
 clutches 1 and 2 combined with the input clutch F provides a speed ratio
 in between the engagement of output clutch 1 combined with input clutches
 F and f, and the engagement of output clutch 2 combined with the same
 input clutches F and f. Similarly, F 23 falls between F 2 r and F 3 r.
 To show the versatility of the concept of this invention, four variations
 of a transmission for different applications are schematically illustrated
 in FIGS. 1, 9, 12 and 14 with their resultant outputs being schematically
 represented in FIGS. 2, 10, 13 and 15 respectively.
 While the present invention has been described in conjunction with a
 specific embodiment, it is understood that many alternatives,
 modifications and variations will be apparent to those skilled in the art
 in light of the foregoing description. Accordingly, this invention is
 intended to embrace all such alternatives, modifications and variations
 which fall within the spirit and scope of this invention.
 General Description of the Continuously Variable (Hydrostatic) Options
 In the description below the continuously variable element is referred to
 as the hydrostatic transmission. It is understood that other continuously
 variable system such as electrical generator-motor sets or variable
 friction drives could be used in place of hydrostatic units.
 Two hydrostatic options are described below:
 (1) Add-on option (PCVT):
 A low cost option requiring relatively small hydrostatic units with minor
 changes on the transmission to install the units provides continuous
 variability over two ranges of operating speeds, one range from a certain
 low speed to maximum speed at ful power capacity and the other from a
 certain low reverse speed through zero to a certain low forward speed at
 full tractive load capacity. The first range requires the gradual
 engagement of a master clutch to start the vehicle motion, if the startup
 load is higher than say 53% of the vehicle weight. (53% of the weight is
 the amount for a specific example shown on Chart 1H line PH 1, 2) After
 the master clutch is fully engaged continuous variability over the full
 hydromechanical range from a certain low speed to maximum speed is
 available at full load or full power capacity. For the second range there
 are gear selections available which provide full load continuously
 variable capacity from a certain low reverse speed through zero to a
 certain low forward speed, but the vehicle must be stopped to shift from
 these low speed gear selections to the gear selection for the first range.
 On FIG. 3 schematically illustrates the gear selection for the first range
 are designated hydromechanical (HM) and for the second range pure
 hydrostatic (PH). Note a shift from the pure hydrostatic range PH 1, 2 to
 the hydromechanical range PM 1 or PM 2 can be made at the S.sub.1 speed of
 1523.82 rpm by engaging the F clutch, which is at synchronism at this
 speed.
 (2) Substitution option (FCVT):
 Larger hydrostatic units can be added which provide full load startup
 capacity by gradually increasing the displacement of the variable
 hydrostatic unit, but due to the space limitations these units can not be
 used as an "add-on" feature to the standard powershift transmission.
 Description of One Typical Embodiment of the PCVT (Add-On Option)
 The transmission system consists of the powershift version as described
 above, with the addition of an continuously variable transmission,
 installed such that the input member S.sub.1 to the planetary system can
 be selectively connected to the engine through this continuously variable
 transmission.
 In the description below a specific transmission system (shown in FIGS. 3,
 4 and 5) is described for simplification. It is understood, however, that
 the invention is not limited to this specific example.
 The system has two modes of operations: The mechanical powershift mode and
 the hydrostatic mode. The mechanical powershift mode is described above
 and will not be described here.
 The hydrostatic mode functions in the following manner. The input
 transmission of this system has two sets of three clutches. One set,
 denoted F, B, and Rev, selectively connects the input shaft to the R shaft
 through various gears, and the other set, denoted f, b, rev, selectively
 connects the input shaft to the S.sub.1 shaft through various gears. This
 set, f, b, rev, is inactive in the hydrostatic mode. In its place a fixed
 displacement hydrostatic unit (M) is connected to the S.sub.1 shaft by
 moving the shift collar (HM) rightward, and a variable displacement
 hydrostatic unit (P) is connected to the reverse shaft by moving the shift
 collar (HP) rightward.
 The compound planetary system is described hereinabove.
 In the following description the flow of power may be positive or negative.
 The input transmission absorbs engine power through the input shaft,
 transmitting power selectively through one of a set of clutches (F or REV)
 to the first input member R of the compound planetary. The clutch (F) is
 mounted on the input shaft to selectively drive the ring gear R, via a
 gear set 32(F) to 42.
 The clutch (REV) is mounted on the reverse shaft. This shaft is driven
 through a 46 to 46r gear set from the input shaft. The reverse shaft is
 located such that the 32(Rev) gear is in mesh with the 42 gear. Thus the
 selective engagement of the REV clutch will drive the ring gear R
 correspondingly in the opposite direction at the same speed as the F
 clutch. In addition to the two clutches (F and REV), there is the clutch
 B. Engagement of B will stop the ring gear R from rotation.
 The sun gear S.sub.1 is driven by the hydrostatic transmission at a
 variable speed from a certain maximum reverse speed to a certain maximum
 forward speed.
 The output transmission absorbs power selectively through two shafts
 S.sub.2 and C. The clutches mounted on the C-shaft and on the output
 shaft, provide two selective gear ratios from S.sub.2 to the output shaft
 and three selective gear ratios from C to the output shaft. The even
 numbered clutches 2 and 4 are selectively connecting the S.sub.2 shaft to
 the output shaft at ratios of 19(2) to 67(2), and 36(4) to 46(4),
 respectively. The odd numbered clutches 1, 3 and 5 are selectively
 connecting the C shaft to the output shaft at ratios of 19(1) to 67(1),
 36(3) to 46(3) and 56(5) to 26(5), respectively.
 To establish a connection from the engine to the output shaft of the output
 transmission either, one of the input clutches (F, B, or REV) plus the
 hydrostatic transmission in addition to one of the output clutches must be
 connected, or, two output clutches must be engaged plus either the
 hydrostatic transmission or one of the input clutches (F or REV).
 Thus, there are four modes of operation:
 1. The low speed, pure hydrostatic (PH) mode with the clutch B engaged has
 five ranges, each providing continuous variability from a certain maximum
 reverse speed through zero speed to the same maximum forward speed as the
 hydrostatic transmission is varied from a maximum reverse speed through
 zero speed to the same maximum forward speed. To initially start the
 vehicle motion in this mode the clutch B and the output clutch for the
 particular maximum end speed are engaged after the hydrostatic
 transmission has been set to neutral (zero speed). Now the hydrostatic
 transmission can be moved to forward or reverse at any desired rate.
 2. The normal hydromechanical (HM) forward speed mode with the clutch F
 engaged has five ranges, each providing continuous variability from a
 certain minimum forward speed to a certain maximum forward speed as the
 hydrostatic transmission ratio is varied from a maximum speed in one
 direction through zero speed to a maximum speed in the other direction. To
 provide continuous variability over the total normal forward speed range a
 shift in the range transmission is made at the point of synchronism of two
 adjacent clutches. (The point of synchronism is the hydrostatic speed at
 which either of two output clutches provide the same system output speed.)
 Thus, starting with clutch 1 engaged, the hydrostatic transmission is
 varied from a maximum reverse speed through zero speed to a certain
 forward speed at which the relative speed in the friction elements of the
 clutch 2 is zero. At this point the shift from clutch 1 to clutch 2 is
 made. To further increase the vehicle speed the hydrostatic transmission
 is varied from the shiftpoint-speed back through zero to a certain reverse
 speed at which synchronous speed in clutch 3 is reached. At this point the
 shift from clutch 2 to clutch 3 is made. To further increase the vehicle
 speed the hydrostatic transmission is varied from the shiftpoint-speed
 back through zero to a certain forward speed, at which another range shift
 can be made at synchronism. From this progression it follows that the
 total vehicle speed range is continuously variable. It is assumed that
 sensors linked to a computer will sense the shiftpoint and perform the
 shift automatically. To initially start the vehicle motion a certain
 vehicle speed up to a certain maximum is preselected by the operator.
 Based on this selection the computer will set the hydrostatic unit, engage
 the F clutch and select the applicable output clutch (1 or 2). The output
 clutch pressure will be automatically modulated at a certain rate to
 provide a smooth transition from standstill to the preselected speed.
 Alternately the operator can override the automatic rate of modulation by
 pushing the clutch pedal to control the rate of clutch engagement at will.
 3. The normal HM reverse speed mode with the clutch REV engaged in place of
 the clutch F mirrors the normal forward speed mode.
 4. Two output clutches engaged with the hydrostatic transmission driving
 the input in the PH mode: There are 6 combinations of two output clutches
 engaged providing 6 continuously variable ranges from a maximum vehicle
 reverse speed through zero to a maximum vehicle forward speed. To
 initially start the vehicle motion in this mode both clutches for the
 particular maximum end speed are engaged after the hydrostatic
 transmission has been set to neutral (zero speed). Now the hydrostatic
 transmission can be moved into forward or reverse at any desired rate.
 FIG. 4 schematically illustrates the various speeds within the transmission
 system at various input conditions.
 Column 1 of FIG. 4 shows the mode of operation, either PH or HM. The PH
 modes provide an continuously variable range from the maximum vehicle
 forward speed, indicated in column 8, through zero to the same maximum
 vehicle reverse speed. The HM modes provide continuous variability from
 the lowest speed (2.93 kph) to the highest speed (40.95 kph), however,
 range shifts are required at the shift points.
 Column 2 of FIG. 4 indicates these shift points at the speeds wherever one
 or another output clutch is shown on the same line. In order to provide
 reverse speeds in the hydromechanical mode the reverse clutch REV must be
 engaged in place of the forward clutch F.
 Column 3 of FIG. 4 shows the speeds of the S.sub.1 shaft (fixed hydrostatic
 unit) required to provide the output speeds shown in column 7. The speed
 of the fixed unit M is varied by changing the output flow of the variable
 unit P. In the specific example the variable hydrostatic unit P connected
 to the reverse shaft has the same maximum displacement per revolution as
 the fixed hydrostatic unit M. Since the variable unit runs at engine
 speed, the fixed unit M runs at about 90% of engine speed at maximum
 hydrostatic pressure (10% of the speed is lost due to leakage). Thus the
 maximum speeds shown in column 3 are at 1800 rpm. Note: the shiftpoint
 speeds are at about 1500 rpm. It is assumed that speed sensors linked to a
 computer will perform the shift, thus the operator will not be aware of
 the change from one range to the next.
 Columns 4 through 6 of FIG. 4 show the speeds of the members in the
 planetary system.
 Column 9 shows the ratio of the available drawbar pull to the traction
 limit of the machine with sufficiently large hydrostatic units to transmit
 full engine power in the HM mode. The PH modes, however, show relatively
 low ratios. Only at maximum vehicle speeds of less than 3.14 kph is the
 pull to traction ratio greater than 1. This implies that a synchronous
 shift from the PH mode to the HM mode can not be made at maximum tractive
 load since this shift must be made from the clutch selection 1, 2. As
 evident from a review of Column 2 of FIG. 2, a shift made from the clutch
 selection 1, 2 only provides a pull of 53% (column 9) of traction.
 Therefore a high energy clutch must be gradually engaged to start the
 vehicle motion in a preselected hydromechanical range whenever the ratio
 of pull to traction is greater than 0.53. Thus, if there is no need to
 pick up the load hydrostatically from standstill of the vehicle, then a
 small hydrostatic transmission is sufficient. Since large clutches with
 modulation capability are required for the powershift version of the
 transmission system they can also be utilized as energy absorbing clutches
 for the hydrostatic version to start the vehicle motion by modulation.
 Column 10 of FIG. 4 shows the ratio of the available drawbar pull to the
 traction limit with sufficiently large hydrostatic units to provide a
 ratio of 1.30 in the pure hydrostatic selection 1,2. From this selection a
 synchronous shift can be made at 1523.81 rpm by engaging the clutch F and
 then disengaging the clutch 1. Thus a continuously variable transmission
 (CVT) results.
 FIG. 5 schematically represents a realistic gear, shaft, bearing and
 housing arrangement for the transmission illustrated in FIG. 1.
 Description of One Typical Embodiment of the FCVT (Substitution option)
 In the following description a specific CVT (shown in FIG. 6, 7 and 8) is
 described for simplification. It is understood, however, that the
 invention is not limited to this specific example.
 In this system the f, rev, b, and the B clutches with their corresponding
 gears have been removed from the input transmission of the powershift
 version. In their place a gear set (22, 34) to drive the hydrostatic unit
 M has been added and the gear ratio (36, 55) to drive the reverse shaft
 has been added, plus the REV gear 22REV has been added. In effect the
 displacement of the hydrostatic units has been increased due to their
 higher operating speed with respect to the units mounted as shown on FIG.
 1. Additionally the displacement per revolution must be increased to
 provide sufficient hydrostatic torque for the traction limit of the
 vehicle.
 FIG. 7 schematically represents that the CVT can operate at any speed from
 0 to 40.95 kph. There are 4 shift points as indicated in column 2 wherever
 one or another output clutch are shown on the same line.
 At these shift points a shift between the two clutches shown on the
 particular line can be made at synchronism. In other words either clutch
 will provide the same vehicle speed at the speed of the sun gear S.sub.1
 indicated on the line (column 3 of FIG. 7).
 Since the shifts can be made at synchronism all friction clutches could be
 replaced by jaw clutches. The choice becomes a matter of economics
 including the control system which must be more accurate to sense the
 speeds of synchronism. The friction clutches are more forgiving than the
 jaw clutches, however, they generate higher power losses thus affecting
 the efficiency adversely. The friction clutches can be made with fewer
 discs at higher clamping forces and lower cooling flow, since they need
 not absorb high energy and therefore the losses can be minimized.
 Column 10 on FIG. 7 indicates the effect of the larger hydrostatic units in
 the pure hydrostatic (PM) modes with respect to column 9 (smaller units).
 In column 10 the ratio of pull available to traction is greater than 1
 (1.30), when the clutches 1 and 2 are engaged, thus the vehicle can reach
 the shiftpoint under the highest normal load condition. If for some reason
 the traction is unusually high, say the tires are frozen to the ground,
 the pure hydrostatic selection 14 can be made which will more than double
 the available pull (2.16), however, no synchronous shift can be made from
 this selection. The vehicle must be stopped, and 1, 2 must be selected for
 continuous variability from zero speed.
 FIG. 8 schematically represents a realistic gear and bearing arrangement to
 provide for higher speeds for the hydrostatic units.
 General Comments on Hydrostatic Options
 It is an additional idea of this invention to design the input transmission
 such that the hydrostatic branch can be added as an option to a powershift
 transmission with either the hydrostatic mode or the powershift mode being
 selectable by the operator as described above. It is an additional idea to
 use a modulated clutch which is already present in the powershift version
 to start the vehicle motion from standstill rather than the hydrostatic
 transmission whenever the transmission system is started in the
 hydromechanical mode at high startup loads. Thus the hydrostatic units can
 be kept small since the high startup torque is not required making it
 physically possible to provide selectable modes of operation as described
 in the paragraph above.
 It is an additional idea to provide a full continuously variable
 transmission (FCVT) from maximum reverse through zero to maximum forward
 speed by using sufficiently large hydrostatic units without changes to the
 housings, planetary system and output transmission. This is possible
 because there exist synchronous speeds for the F clutch while operating in
 the pure hydrostatic mode at a point where the sun gear speed S.sub.1 is
 equal to the ring gear speed.
 Features in the Arrangement of Clutches, Gears, Shafts and Bearings
 System for a Vehicle with Rear Mounted Transmission
 FIG. 5 schematically represents a cross section of the transmission system
 for a vehicle design in which the transmission system is mounted to the
 rear frame with a relatively large space available at the input side of
 the transmission, thus the axial space available for the input
 transmission allows the use of pairs of clutches and the mounting of
 optional hydrostatic units at the input side of the transmission system.
 Further more this type of vehicle design requires the Power Take Off drive
 to be located across the top of the transmission with a pump drive to be
 located at the side of the transmission.
 The input transmission has two identical clutch-gear assemblies for the
 forward and reverse shafts providing a cost advantage. No reverse idler
 gear is required, again a cost advantage. The clutches are arranged such
 that the high feedback speed on the 22f and 22rev gears does not affect
 the clutch actuation, thus no dump valves are required which are normally
 used to deal with high clutch speeds. The hydrostatic transmission is
 easily attachable to two shafts, requiring no extra gears, shafts or
 bearings.
 The planetary system is arranged such that the input sun gear S.sub.1 and
 the carrier C are floating, with the ring gear semifloating off one
 bearing support BF on the left side. This floatation provides good load
 distribution over all the planets. Only the output sun gear 32 is radially
 fixed on a bearing.
 The output transmission is arranged such that the two high energy clutches
 (master clutches 1 and 2) with a large friction surface area and dual
 cooling flow are located on the output shaft where the torque is high. The
 low energy, low torque clutches 3 and 5 are located on one of the input
 shafts (C). The remaining clutch 4 uses the otherwise "dead" space on the
 output shaft below the bearing mounting of the S.sub.2 shaft. The three
 clutches 1, 2 and 4, located on the output shaft, are designed such that
 the actuating piston is rotating with the clutch gear to eliminate the
 effect of the high speed of the output shaft in the higher gear
 selections. This clutch design is described in the provisional application
 Ser. No. 60/071661 entitled Clutch Assembly For Transmission on Jan. 16,
 1998.
 The input shaft (S.sub.2 shaft) of the output transmission is mounted on
 two bearings. The gears 19(2) and 36(4) are cantilevered on the right side
 and the sun gear S.sub.2 is cantilevered on the left side of the shaft
 (S.sub.2 shaft). An extension of the one input shaft (S.sub.2 shaft) acts
 as the bearing mount for the other input shaft (C shaft). This eliminates
 the need for a bearing mount attached to the housing, which would require
 additional axial space and cause an assembly problem. The 38a accessory
 drive gear is coupled to the input shaft to rotate at engine speed. It is
 in mesh with the 38i idler gear, which drives the PTOin shaft and which
 also is in mesh with the 29p pump gear (not shown on the Layout but
 indicated on FIG. 1H). The PTOin shaft drives a gear 33 located at the
 rear of the transmission assembly where it is in mesh with the PTO output
 gear 30 which in turn is connected to the PTOout shaft. Note, this
 arrangement makes it possible to bring the engine speed physically around
 the range transmission back to the center location of the assembly at the
 rear of the transmission in line with a PTO clutch and gear reduction (not
 shown).
 The park brake is located on the output shaft below the pressure control
 manifold for the clutches 3 and 5. Again this axial space would otherwise
 not be utilized.
 The mechanical front drive clutch (MFD) is located as far to the rear as
 possible to allow for maximum ground clearance at the center of the
 tractor.
 System for a Vehicle with a Front Mounted Transmission
 FIGS. 9, 10 and 11 schematically represents a transmission system for a
 vehicle design in which the transmission is mounted towards the engine
 side allowing relatively small axial space for the input transmission. In
 this design it is possible to arrange the gear system such that one single
 clutch per shaft is used for the two reverse clutches and one forward
 clutch. Also the optional hydrostatic units are moved towards the output
 side of the transmission. This design requires the Power Take Off to be
 located at the bottom of the transmission system, therefore, a geartrain
 from the input shaft to the bottom is required, which fits in perfectly
 with the arrangement of gears for the input transmission.
 The input transmission is comprised of an input shaft on which the clutch f
 is mounted said clutch f selectively connecting the shaft S.sub.1 to the
 input shaft through gears 22f and 55, is further comprised of one reverse
 shaft (rev shaft) on which the rev clutch is mounted said rev clutch
 selectively connecting the S.sub.1 shaft to the input shaft through gears
 32I to 32r to 22r to 55, is further comprised of one other reverse shaft
 on which the REV clutch is mounted, said REV clutch selectively connecting
 the R shaft to the input shaft through gears 32I to 32Ri to 32Ro to 43R,
 is further comprised of one hollow R shaft concentric to the S.sub.1 shaft
 on which the F clutch is mounted said F clutch selectively connecting the
 input shaft to the R shaft through gears 32I to 43F, is further comprised
 of one b clutch which selectively stops the S.sub.1 shaft from rotation
 and is further comprised of one B clutch which selectively stops the R
 shaft from rotation. The S.sub.1 shaft is connected to the sungear of the
 planetary system and the R shaft is connected to the ring gear of the
 planetary system.
 The transmission is further comprised of a Front Drive Shaft on which the
 Front Wheel drive Clutch, the Park Brake and the idler cluster gear 28-38
 is mounted. The Front Wheel drive Clutch selectively connects the front
 wheels to the output shaft, the Park Brake selectively stops the output
 shaft from rotation and the idler cluster gear connects the input shaft to
 the PTO shaft through gears 32I to 43F to 38 to 28 to 51.
 The compound planetary system and the output transmission are arranged in
 the same manner as described above under System for vehicle with rear
 mounted transmission.
 The optional hydrostatic transmission is mounted at the rear of the
 transmission such that the fixed displacement unit M is selectively
 connectable to the H shaft which is located concentrically inside the C
 shaft and attached to the sungear 18, and such that the variable
 displacement unit P is selectively connectable to a shaft driven by a gear
 from the PTO shaft. The hydrostatic units M and P are hydraulically
 interconnected by fluid conduits.
 FIGS. 10 and 11 schematically represents a realistic gear, shaft, bearing
 and housing arrangement for the front mounted transmission.
 General Comments on Features in the Arrangement of Clutches, Gears, Shafts
 and Bearings
 From the above description associated with FIGS. 5 and 11, it is apparent
 that several ideas are unique, including but not limited to the following:
 1. The arrangement of clutches, gears and shafts in the input transmission
 which makes it possible to use identical parts for the forward and reverse
 clutch assemblies, and also makes it possible to attach the hydrostatic
 units optionally, without the addition of gears, bearings and other major
 parts. (FIG. 5)
 2. The arrangement of clutches gears and shafts in the input transmission
 which make it possible to minimize the axial space required for the input
 transmission and to provide for a Power Take Off. (FIG. 11)
 3. The mounting of the planetary system, providing flotation of the sun
 gear S.sub.1 and the carrier C in addition to providing stability through
 the radially fixed sun gear S.sub.2 in combination with the
 semifloating-floating ring gear R, which is supported on one side by the
 bearing BF.
 4. The mounting of the S.sub.2 shaft on two bearings with cantilevered
 mounting of gears on each side in addition to provide a bearing mount for
 the C shaft through a cantilevered extension of the S.sub.2 shaft.
 5. The arrangement of gears, clutches and shafts in the output transmission
 providing for a maximum utilization of axial space for essentially a five
 speed package plus a park brake on two centers.
 6. The arrangement of the pump drive and the PTO which utilizes the
 necessary idler gear 38i for two purposes: first to bring the pump drive
 gear 27p far enough out to provide clearance for the pump and second to
 provide a connection for the PTOin shaft. (Note that the direction of
 rotation of the PTOout shaft requires the idler gear.)
 While the present invention has been described in conjunction with a
 specific embodiment, it is understood that many alternatives,
 modifications and variations will be apparent to those skilled in the art
 in light of the foregoing description. Accordingly, this invention is
 intended to embrace all such alternatives, modifications and variations
 which fall within the spirit and scope of this invention.