Aa?a?eiey real estate bankruptcy

Aa?a?eiey real estate bankruptcy

Jun 15,  · Bankruptcy in American Airlines Group filed for Chapter 11 bankruptcy on November 29, , in the United States Bankruptcy Court after suffering losses for about four consecutive A provision in a contract to buy real estate stated, "This contract shall be terminated if the buyer doesn't sell his home at 24 Highland Drive within 30 days." This provision is an example of a. contingency. Ted Logan signed a property management agreement with property manager John Burton for a day period. Shortly thereafter Logan died. May 13,  · Source(s): Lots of experience in foreclosure and real estate and bankruptcy fields. 0 0 0. Login to reply the answers Post; Anonymous. 1 decade ago. Depends on how you obtained the property. If there is a paper trail and there usually always is, they can find it (such as you paid for the property by withdrawing a large amount of money from your.

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Challenging your home's assessed value if you live in an area where real estate prices have depreciated. Bankruptcy cases where the value of real estate is required. Tax planning documentation. Needing an accurate estimate of a home's square footage. We can do it quickly and at a reasonable price! Fig, Figure 12 representfi an ordinary wall craae, or warehooae crane. It ia Bupported in a footatcp at the bottom, and ateadicd by a strap at the top. The strain on the tie ia taken as shear- ing strain on the top of the pillar, and transmitted to the strap which holds the top of the pillar.

For some purposes, as in erecting works, and in railway management, moyable cranes are required, and these are arranged on trucks, so as to run on a permanent way. A crane of this sort is shown at Figure In this case it will be seen that the strain on the pillar cannot bo taken by the foundation ; therefore the load on the jibs of the machine is balanced by a weight, C, moving, which may be moved along the tail-piece D at the back of the crane, in order to adjust it to the weights to be raised.

Let the load on the crane be lbs. The tie, it will be observed, is continued past the pillar being bolted to it, however , to the end of the tail-piece. In order to prevent vibration on the springs, the track may be either blocked up off the line, or sbackles may be pat on and the springs compressed on to blocks placed on the axle boxes ; in this case, the crane is very steadily held, being clamped down firmly to the rails.

Thns there is a great strain on the lags to which the tie rods are attached, and they will, in the generality of cases, be of cast iron. Let the tension of a rod be 20, lbs. The resistance of cast iron to tensile force amonnts to only 18, lbs. Let the factor of safety be 6, then the breaking weight of the lags would be lbs. Hence the number of inclies sectional area required will be square inches Which, in practice, would be made 7 sectional square inches.

The general form of hoists is too well understood to require an elaborate de- scription ; hence we shall herein touch only upon some of the details. Hoists are of three kinds : those worked by air, called pneumatic lifts ; by water, called hydrostatic lifts ; and by steam. Air lifts are only suitable for low heights and light weights, and even then, except under peculiar cir- cumstances, there is not much to recommend them.

In principle an air lift is simply a float, and its rising is ex- actly in the same manner as the rising of a gas-holder as it fills. The air being prevented from escaping by the water water-seal around its lower edges, it is evident that the height of the surface of the water above the edge of the vessels rules the amount of pressure which may be main- tained within the lift.

Let us now see how a load is sus- tained by this air pressure. The lowest depth of water above the edge of the lift will be found when it is at the top of its stroke; hence this is all that is to be calculated upon. Let the lowest depth be two feel. Suppose the weight to be lifted be not more than lbs. We may, however, simplify the calculation, and find a rule to give at once the diameter of the tube or cylinder.

The result will be the re- quired diameter. Let the load required to be lifted be lbs. The height of the lift is, of coarse, regulated by the length of the cylinder ; thus, if the height required is 30 feet, the length of the tube must be somewhat more than 30 feet, — say 32 feet to allow for that part which is below the surface of the water when the lift is at the top of its range. We will first, speak of this class of lift, and assume the ram or plunger of the press to represent the whole length of the stroke, 60 that the table of the lift or hoist is placed directly upon the top of the piston.

Let the load to be lifted be lbs. In most cases this rule will be sufficient, as the given quantity will be the greatest available head of water.

In many hydrostatic hoists the stroke of the press is less than the lift, the difference being adjusted by pulleys and chains ; but in this case the total pressure on the ram of the press must be much greater than the load to be lifted, in the same ratio as the height of lift is greater than the stroke of the press. Let the load to be lifted be 24, lbs. And by applying the previous rule, the necessary diameter of ram to work this load can be found. We will now proceed to consider the questions apper- tahung to the hydrostatic press, regarding, in the first timi dMs irorked by hand.

Square of diameter of plunger lbs. Embodying the two rules now into one general rule, we have EuLB. Then midtiply the distance from fulcrum of pump lever in inches to centre of pump plunger by square of diameter of pump plunger ; and divide the former product by the latter, the quotient will be the total pressure the ram is exerting.

We will assume the same proportions as above : — lbs. The result will show the number of times by which the diameter of the ram must exceed that of the pump in order that the required pressure may be obtained. Let the ayailable pressure on the pump plunger bo 8, lbs. So much for the power of presses. We will now pro- ceed to consider the strain on the materials of which they are constructed. As far as the ram is concerned, we do not require to make any calculation as hydrostatic presses are never worked to a sufficient pressure per square inch to endanger the ram, about 8, lbs.

Hence we have only to consider how the thickness of the metal in the cylinder is to be determined. Let the diameter of the cylinder be 8 inches, and the internal pressure to which it will be subjected lbs. Assume that the working pressure should be but one-fifth of the bursting pressure, we haye lbs. Whence we find the thickness by means of the foregoing rule in the following manner : lbs.

Let us now apply this rule to thick cylinders, which are those subject to heavy internal pressure, and show wherein it fails.

As a rule for thick cylinders of cast iron, we take the under one extracted from a yery popular Engineer's Pocket Book. Without digressing further, let us now apply the thin cylinder rule to the calculation of the thickness of a cylin- der 10 inches in diameter to sustain safely a pressure of lbs.

Using the same factor of safety as before, the bursting pressure would be lbs. Hence the thickness is thus found :— lbs. This, however, would not be sufficient. We can- not compare this with the rule a , as we hare nothing stated about the factor of safety use ; but we will investi- gate it independently, assuming for simplicity the thickness to be taken at 6 inches.

Under strain, metal, in common with other elastic bodies, is temporarily stretched or compressed, according to the intensity of the strain, and in a cylinder under internal pressure, the inner layers must be more strained than the outer, because the inner ones must be distended to press on the outer.

Hence, the inner layer may be strained very much, whilst the outer one is scarcely strained at all ; and in all cases the straining of the fibres in a given cylinder will vary inversely as their distances from the centre of the cylinder; and as action and reaction are equal and op- posite, the resistance of each fibre will correspond to the point to which it is strained. When the breaking disten- sion of the inner layer of fibres is reached, the vessel must give way ; hence, if it be very thick, the mean break- ing strain of the cylinder per square inch is less than the ultimate strength of a square inch bar under tension.

When the cylinder is thin, as in the previous case, the rule holds good, as the difference of distension between the inner and outer layers of fibres is practically inappreciable ; but in the present case it is different. The breaking tension per square inch of sectional area on cast iron may be taken in even numbers at 17,, and to find the tension on the outer layer of fibres, when the inner are at breaking point, we have the fol- lowing : Rule.

In the present case the Inner diameter is 10 inches, and the outer, being the sum of the inner and two thicknesses of metal, is 22 inches. Which may be assumed in round numbers as lbs. We, therefore, find the internal pressure per squaie inch at which the inner layer of fibres is on the breaking point, by reducing the assumed bursting pressure in proportion to the ratio of to , that is to say by multiplying h hj , and dividing by We assumed 20, lbs.

We may eyidontly take a much lower one than 5, as that was adopted as the assumption of all the fibres coming to the breaking stress at once, so probably 8 will be amply suf- ficient; then we hare 3 lbs. It is a matter much to be regretted that there has been no series of experimental trials of thick cylinders upon which some satisfactory formulas might be based.

We may now make a few remarks upon the modes of packing rams of hydrostatic presses. The plan most generally adopted in England has been to use leather collars, into which the water obtained access under pressure, and so pressed the leather tightly against the cylinder on the one side, and against the ram of the press on the other, and thus preyenting the escape of water. There is, however, a great objection to these leathers, inas- much as they require to be renewed every few months ; which, of course, causes some delay and expense.

This difficulty may be obviated by packing the rams with hemp in exactly the same way as the piston rod of a steam engine is packed, and the hemp will be found to wear out many leathers. We cannot say how long the hemp packing will last ; but we know of one now that has been in use in an accumulator 18 inches in diameter, for three years, and it is not yet worn out. Before leaving the subject of hydrostatic presses, it may be desirable to explain what an accumulator is, and the part it takes amongst hydraulic machinery.

Let us suppose that an hydrostatic press is required to be used once every ten minutes, and it would take say nine minutes to fill it so that it would make a complete stroke, and further suppose that it is required to act quickly. A slightly larger press is made, and its ram is loaded with a weight corresponding to the pressure re- quired in the press which it is intended to work.

Now if we assume that an engine is engaged constantly in pump- ing water into the accumulator except it stops when the accumulator is full, we shall get instantaneous action of the liydrostatic press ; for, after the engine has worked nine minutes, there will be enough water pumped into the accu- mulator to make one stroke of the ram of the press. Self- acting gear stops the accumulator until the stroke of the press has been made when it b again started and filled so as to be ready for another stroke of the press, and thus the required end is attained.

Now it is evident that if the accumulator be made large enough it can be used for more than one hydrostatic press. And even if all the presses be not required to produce the same gross pressure, this can be attained by varying the diameters of the presses so that they may aW.

We will now proceed to show how to determine the size of the tank or of the weights required for the accu- mulator. Let us assume that a number of hydrostatic presses working at lbs. A cubic foot of cast iron of average quality weighs shont lbs. We will assume the length of the ram including the head to be equal to 15 feet. We will ascertain how many cubic feet it contains. The rest being made up by the weight of the tank and its appurtenances.

We will now pass on to the last kind of hoist of which we purpose treating, namely the steam hoist, not however in ibis place entering into the calculations of the engines themselyes, which will be duly considered in a subsequent chapter.

If the lift is light, say under a ton weight, and not over 20 feet in height, it may be worked by a single-acting en- gine having a chain attached to a fixed pulley and rove through a grooved pulley, carried on the end of the piston rod, and thence taken over guide pulleys to the required position. The end of the chain will then travel twice as far as the piston, the gross pressure on which must of course be twice the weight to be lifted.

Thus, if the lift be 14 feet, and the weight to be lifted 12 cwt. The mode of action is this : the piston is balanced at the top of its stroke by a counter weight attached to the platform of the hoist, which is at the bot- tom of the lift when the piston is at the top of its stroke. The platform being now loaded steam is admitted above the piston, and forcing it down, raises the load. When the load is removed the steam is shut off, and that in the cylinder being allowed to escape, the counter-weight again draws the piston to the top of its stroke, ready for another lift.

The chief distinction between other hoists rests in whe- ther tbey be driven by quick or slow Tuuuiug engines. Before proceeding to the mode of calculating the tangent screw arrangement, it seems desirable to discuss the objec- tions which have been raised to it, and also the advantages claimed for it.

The first objection commonly raised is on account of the amount of friction which is involved in its use, the fact of the matter is, however, not that there is more friction in this arrangement than in the train of spur wheels, but the friction, instead of being distributed through a number of shafts is concentrated in one. Let us assume that gearing is required to reduce the velocity of the primary shaft to one-fiftieth in the final shaft. If we can place the two shafts at right angles, this can be done at once by the tangent-screw and worm-wheel arrangement, but if it be done by spur-wheels, a train of them will be requisite.

Thus, say we have on the primary shaft a pinion 8 inches in diameter gearing into a wheel 40 inches in diameter, on a shaft carrying another 8-inch-pinion gearing into another inch wheel on a shaft carrying a 12 -inch pinion gearing into a inch wheel on the final shaft ; the end will be attained, as the final shaft will make one revolution for every fifty revolutions of the primary shaft Here it will be observed we have four shafts amongst which the friction is distributed, whereas in the other arrangement there are but two, the worm-wheel shaft and the tangent screw shaft.

Whateyer rubbing there is on the tooth of the spur wheel is radial, but the rubbing on the teeth of the worm- wheel is parallel to the axis of its shaft. We know by ex- perience that the wear of a well-formed tangent screw and worm-wheel well constructed is by no means great if due attention bo paid to its lubrication, but if this be neglected the worm will of course soon perish, the same as would a journal if dry.

Let the strain on the chain be 22, lbs. If the strain be taken in tons the thrust will of course be in tons. Let the strain on the chain be 24 tons, the radius of the barrel 18 inches, and the radius of the worm wheel 11 inches. This thrust would hare to be taken up by the two plummer blocks supporting the shaft on either side of the tangent screw, and sufficient bolts must be used to fasten the plummer blocks down to the framing to prerent any risk of their being sheared through.

The total sectional area of the bolts will be found by dividing the thrust on the screw shaft by the resistance per square inch of wrought iron to shearing strain. The second example shows an enormous thrust due to the diameter of the chain barrel being so much greater than that of the worm wheel, whereas in the former case the worm wheel was much larger than the barrel in diameter, which is a far better arrangement ; but we have inserted the second case because we have known of such apparatus being put up, although the use of larger chain barrels is altogether contrary to good practice, involving as they do very heavy gearing.

Within the small space allotted to the chapters in the present treatise, it would be impossible to enter minutely into all the details of thermo-motiYe engines, and therefore we shall direct our attention more particularly to those points in which there is most chance of errors arising through either carelessness or ignorance. It behoyes every one who buys an engine, and more especially one of the cheap class, to examine it thoroughly, or have it ex- amined by some competent person ; for, although any engineering firm of position would, for its own credit sake, try and see that no imperfect machinery left their shops, yet experience shows us that there are those who make engines simply to sell, and consequently knock them to- gether as cheaply as possible, without much regard to efficiency.

We have known several cases of engines supplied with the valves so made that they would not, until altered, do one third of the work guaranteed, — and that too, within the last two or three years. Telocity at which the engines were intended to be run. Now this information shoald be sufficient for any practical engineer to detail the engine, so that it should perform its work satisfactorily.

For the pressure of steam being known, and the size of the cylinder, it is a simple matter to determine the size of the steam ports and passages. However, although the makers undertook to guarantee the performance of the machine, when set to work it would only lift four tons, and indeed barely that.

In most re- spects the engines were tolerably right, only the steam passages were certainly too small ; but the great error lay in the yalyes haying a great excess of lap.

The engines were intended to run at a speed of from to revolutions per. This having been done, the engines just did the stipulated work, but with nothing to spare ; but, had the steam passages been properly propor- tioned, there is no doubt that fifteen tons might have readily been picked up.

Haying said thus much about the area of the steam ports, we will now give a rule, deduced from the laws of the flow of gases, by which the proper area may be determined nndcr any given set of cir- cumstances. From this we obtain the proper area of the steam pas- sages for the particular cases that may demand our con- sideration.

For safety take the minimum difference between pressures in cylinder and boiler with maximum pressure in the cylinder. Let the diameter of cylinder be 20 inches, speed of pis- ton feet per minute, absolute pressure in boiler dO lbs.

As a general practice, the exhaust port is made twice the area of the steam port. From the 20 inches area we may determine the diameter of the steam pipe by the ordinary rule. Haying decided upon the amount of lap and lead to be given to the valve, the travel of the valve is at once deter- mined by the following rule.

It is a common practice in proportioning piston rods to make them one-tenth of the diameter of the cylinder, unless the steam pressure is very high when the following rule may be used. So the piston rod should be 3 inches in diameter. The metal in the cylinder of a steam engine has two duties to perform ; one is to resist the actual bursting pressure of the steam within it; the other, to withstand the vibration due to the motion of the machinery ; and in addi- tion to this, allowance must be made for wear.

Including nil, the following rule has been found satisfactory. Let the pressure of steam be 25 lbs. This, which would be made inch, is the minimum thick- ness to be used. The main shaft of the engine upon which is fixed the crank or crank plate, as the case may be, is proportioned to resist the torsion to which it is subjected by the follow- ing rule : — Bulb.

Let the horse-power be 40, and let the engine make 25 revolutions per minute. One horse- power is agreed to be eqaiyalent to 33, lbs. Let the mean pressure per square inch on the piston be bs. Let the engine be required to be 12 horse-power, the mean effectiye pressure in the cylinder be 20 lbs.

If the pressure taken be the gross effectiye pressure, then the corresponding horse-power is the gross horse- power, and is in excess of the ayailable power, because a part of it is absorbed in oyercoming the friction of the engine. The amount of power absorbed by friction yaries according to the construction of the eugOiQ. It is usual to speak commercially of engines by their nominal horse-power, which in condensing engines is usually about one-third of the actual horse-power.

Let the diameter of the piston be 25 inches. Let the diameter of the piston be 8 inches. The nominal horse-power is, we may say, merely a term used in the buying and selling of engines. Engines driying machinery, whether single or coupled, are usually fitted with a fly-wheel to equalise the motion ; for when the engine is giving off more work than is being absorbed, the fly-wheel, with a slight increase of Telocity, takes it up as accumulated work, and, at some other part of the reyolution where the work done by the engine is below the average of the fly-wheel, parts with the excess of work which it had previously taken up.

Note : the diameter of fly-wheel should be about four times the stroke. Suppose we have an engine with cylinder 10 inches diameter and 15 inches stroke, the pressure of steam being on the average 30 lbs. Thirteen cwts. From this may be determined the sectional area of the rim of the wheel.

In machinery for hoisting it is always necessary to have two engines acting on one shaft, the cranks being at right angles to each other, so that one or other engine always has a hold npon the load being lifted ; for, with a single engine, there would be no getting oyer the dead point, except through the momentum of a heavy fly-wheel, which is not suitable where stoppages are sudden and frequent.

Now, it is evident that, with two engines so arranged, when one of them is on the dead point, the other is ex- erting its full force ; but the engines must be made so large that one cylinder alone, when the piston is at mid- stroke, shafl be capable of sustaining the load.

To find the force which the engine must exert, we have the follow- ing rule : EuLE. Let the load to be held be 22, lbs. The pressure on the piston must be equal to this. Assume the mean pressure per square inch on piston at 25 Ibft. Say 30 square inches for area of piston necessary to sustain load, and add one sixth for friction.

Let us now see what would be the mean force exerted by the two engines during the revolution, for it must be remembered that as the crank pin travels through the cir- rumference of a circle haying a diameter equal to the stroke of the f iston, the piston itself travels only through twice the diameter of the same circle, making two strokes to each revolution.

In the above case we find, exclading allowance for fric- tion, that lbs. From which it will be seen that we have sufficient power for the work. In respect to air engines we have not much to say, as they are not used to any considerable extent in England at all events , and those that are in use are of small size. We must, however, point to a few matters connected with them of practical importance.

In order to keep the heat away from the working parts, the well-known plunger-shaped appendages are always attached to air engine pistons, except in the case of some few which, like Mr.

Gill's, worked under water, the elastic fluid used being a mixture of heated air and steam. There is in the principle of burning fuel imder pressure an element of economy, as was practically established some years since, although it has not been applied to steam machinery ; probably through the patentees of that method of working furnaces not having sufficient interest to press their inven- tion ; and no doubt what extra economy is exhibited by these air-engines in consumption of fuel may be traced to the close furnace, but these certainly have as much or more claim to be called gas engines as air engines ; for all the gaseous products evolved from the solid fuel must aid materially in the propulsion of the machine.

This, how- ever, is not the place to enter more fully into a question which, further pursued, would become almost entirely cbemical. The most durable packings for the working parts have been found to be those made of leather.

We will now pass on to gas engines properly so called. Lenoir's gas engine, which was one of the first, was ac- tuated by exploding a mixture of air and gas at each end of the cylinder alternately, the ignition being effected by means of an electric spark.

In the Hugon gas engine the mixture of gas and air is exploded by gas jets, placed at the two ends of the cylinder, and the slides are so arranged that there is a communication between the mixture of air and gas in the cylinder and the lighted jet outside at the beginning of every stroke. As to the economy of this engine we can- not speak, having had no opportunity of testing it.

The engines are bulky, as are all gas engines ; one of 3 horse- power beipg as large as a 10 horse-power steam engine. In designing pumps of large dimensions it is of great im- portance properly to detail the pipes and passages, and in all cases bends in the pipes should be avoided as much as possible. Wherever it is practicable, the suction and de- livery pipes should be kept of diameter equal to that of the pump plunger ; but if this cannot be arranged, they should be made as near as possible to that size, in order to reduce the friction of the water passing through the pipes as much as possible, and on the same principle in the designing of all details of water works, attention should always be paid that the friction of the water may be kept at a minimum.

Let the head of water be feet. Hariug shown the relation between head of water and pressure, it will now be necessary to supply rules for deter- mining the motion of water produced by such head or pressure. The following rule serres to determine the size of orifice under a given head of water to discharge a giyen quantity per minute : Rule.

The square root of 16 is 4. Square root of head 4 gallons per minute yii square feet. Let the required discharge be 31, gallons per minutOi Ihe head of water 10 feet, and the length of pipe 70 feot.

Adopting the latter plan, we find the fifth root of 32 is 2 ; therefore the diameter of the pipe must be not less than 2 feet. It is, on the other hand, useless to have the suction pipe larger than the pump barrel, because the water would not come in any easier, as its ingress would then be limited by the area of the pump barrel itself.

It does not in the least signify of what material the pipes are made, so long as they are tolerably small for the water adhering to, or wetting, the sides of the pipe. Velocity is the great element in the friction of liquids, pressure in no way affecting it ; but the friction varies as the cube of the velocity. To find the delivery of single-acting pumps we have the following rule : — BuLE. This rule allows about five per cent, for leakage.

Let the diameter of the pump plunger be 20 inches, the stroke 8 feet. Suppose, for instance, a population of 40, required to he supplied with water hy a pumping engine in a manufacturing town.

Allowing SO gallons per head per day, the total quantity to he pumped in 12 hours not working the engine at night will he 30 12 1,, gallons per 12 hours 60 , gallons per hour gallons per minute. Let the engine he single-acting and make 8 strokes per minute : No. We will now show how the amount of this preponderating weight is to be calculated.

Let the water being pumped up be required to be raised feet above the bottom of the pump plunger when at mid-stroke, the pressure per square inch corresponding to this weight is found by a previous rule. The pressure of the water cannot again force the plunger upwards, as the closing of the deliyery yalve prerents its return. It is a matter of yital importance that the pump valves should work accurately and quickly, as otherwise there will be a great loss of water. In some of the early pumps they used the clack or butterfly yalves, but they expose so large a surface to the water in proportion to their weight, that they are yery slow in closing, and two evils result there- from; the first is great loss of water, amounting to as much as 16 or even 20 per cent, of the water; and the whole column of water descending with and upon the valve causes a very great shock, destructive to the machinery.

We will take an example to illustrate the force of such a blow in a large pumping engine. This weight of 86, lbs. In the engine alluded to the yalyes are of excellent construc- tion, so that any one standing close by can just feel them beat and no more ; but if we assume that a flat yalve had been used, at every stroke of the engine a blow equal to 88 tons falling through one foot would hare been inflicted on the machmery.

The double-seated valves have been very much used for large pumps under high lifts, and give good results, and recently india-rubber valves have been employed with great success. We were first shown them in action by Mr. Morris, of the Kent Water Works, Lewisham, and we believe that gentleman was the first to introduce them.

The valve consists in form of a series of cylinders diminish- ing upwards from the base pyramidically, and perforated with vertical slots, so that the valve altogether forms a series of cylindrical gratings, round each of which an india-rubber band is placed. The pressure of the water expands these belts, and passes them, after which they close noiselessly, the superincumbent pressure of the water keeping them perfectly tight.

He later sold a controlling stake in the property, which remains in operation. Trump relinquished the majority control to bondholders and gave up his title of chief executive officer, according to The Press of Atlantic City newspaper.

Atlantic City casinos were also hurting, according to published reports, because of new competition from across the state line in Pennsylvania, where slot machines had come online and were drawing gamblers. The holding company emerged from bankruptcy in February and became a subsidiary of investor Carl Icahn's Icahn Enterprises. Icahn took over the Taj Mahal then sold it in to Hard Rock International, which renovated, rebranded, and reopened the property in Share Flipboard Email.

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