WO2007129039A1

WO2007129039A1 - A turbo-expansion valve

Info

Publication number: WO2007129039A1
Application number: PCT/GB2007/001594
Authority: WO
Inventors: Peter John Bayram
Original assignee: Peter John Bayram
Priority date: 2006-05-02
Filing date: 2007-05-01
Publication date: 2007-11-15
Also published as: GB0609654D0; EP2013548A1; GB2449590A; US20110061412A1; GB0609326D0; GB0816082D0

Abstract

Liquid refrigerant (1) is flashed-off by ultrasonic vibrations and/or diffuser plate (5) into saturated vapour that is static pressure reduced/velocity increased by the converging inlet tract and volute of turbine (4). The turbine wheel expansion cools (***) this saturated vapour - less a turbine efficiency loss heating effect and powers compressor (7). The %age free work of compressor (7) is the product of the efficiencies of turbine (4) and compressor (7). Valve (2) controls pump-down, superheat via sensors (10) & (11). Valve (16) controls hot gas bypass/turbine speed via sonic sensor (9). Compressor (13) can be any type of compressor with reduced compression and motor speed - horsepower reducing 55 to 75% plus. Exact required speed of an automobile A/C compressor is obtained via the belt drive. (***) For the same evaporator capacity: mass flow rate and compressor horsepower reduces 10 to 20% due to expansion cooling increasing refrigerating effect (RE) - similarly further reduced by elimination of TEV losses.

Description

A TURBO-EXPANSION VALVE

This invention concerns the work of expansion and adiabatic expansion cooling generated by the flow of high pressure saturated refrigerant vapour through a turbine, or any other type of reversible expansion engine to a low pressure evaporator.

5

Conventionally high pressure liquid refrigerant is raetered into a refrigeration evaporator via a thermostatic expansion valve (TEV). The TEV being actuated by aetal bellows connected to a

IO thermal bulb at the outlet of the evaporator such that the temperature of vapour leaving the evaporator is maintained at a superheated condition relative to the refrigeration evaporating temperature; or the TEV may be electrically actuated and electronically controlled via the inputs of temperature sensors in S the evaporator and at it* s outlet. A TEV may also be pressure controlled to βaintain a constant evaporating pressure/temperature. In any case, a TEV simply operates as a pressure reducing valve: flashing—off high pressure liquid from a condenser into low pressure saturated vapour that is supplied to an O evaporator such that there is a coaplete loss of the potential static pressure energy difference between that of the condenser and the evaporator. Not only is potential energy lost, but heat is also generated from turbulent and surface shear frictions through the TEV, associated distributor and capillary tubes to 5 the evaporator - such heat as is generated evaporates liquid droplets in suspension in saturated vapour <at a constant temperature), thus reducing the evaporator refrigeration effect (RE). Since there is no change in temperature this effect has, hitherto, been unrealised; but should be quantifiable by means O of computational fluid dynamics (CFD). Since TEVs are substituted with this invention their losses need not necessarily be quantified. Also see page 7 note N.B.I.

5 According to this invention high pressure liquid refrigerant is flashed-off into saturated vapour and subsequently pressure re- reduced and expansion cooled by a radial-flow turbine, or any other type of reversible expansion engine by braking it* s output shaft. Such expansion also adiabat ically cools the satu- O rated vapour, less a turbine efficiency loss heating effect.

Pressure would not be significantly reduced by the turbine unless a brake load was imposed on it's output shaft, causing it to work and generate a brakehorsepower output. 5 The output shaft of the turbine could, most simply and most efficiently, be co-axial Iy connected to the input shaft of a radial-flow centrifugal (1st stage) compressor. A 2nd stage refrigeration compressor and motor, equal in capacity to the efficiency losses of the turbine and 1st stage compressor also O being required. Alternatively, a motor could be mounted on a co-axial shaft Joining the turbine to a (single stage) compressor — with the motor sized to power the compressor, less the power input of the turbine. Similarly, the turbine could be co-axial Iy connected to a generator that may be speed variable.

RECTIFIED SHEET (RULE 91) ISA/EP It should be noted that it is mentioned in the ASHRAE Handbook Fundamentals volume, Chapter 1 - Thermodynamics and Refrigeration - in the section headed 'Actual Basic Vapor Compression Refrigeration Cycle* that in an ideal cycle: "a reversible adi- 5 abatic expansion engine should be substituted for the expansion valve, and work output from this engine would supply part of the work input to the cycle" and, later, that: "The complexity of equipment needed for these improvements precludes its application* *. However, this invention reveals equipment that is a 0 stunningly simple solution to this long known conundrum.

Means for controlling system capacity include! variable compressor inlet guide vanes; variable turbine expel ler guide blades and/or variable turbine outlet nozzles; compressor hot- 5 gas bypass; varying 2nd stage compressor output with a variable speed motor or variable speed CVT drive (gearing) system, and combinations of these methods. The most energy efficient means of capacity control is that of compressor motor speed control. Other systems that reduce flow rate will actually increase O motor amps as the compressor is unloaded. Hot gas bypass control, however, tends to maintain relatively stable flow rates and pressure differentials across both the turbine and compressors to thereby avoid unstable surge operating conditions, even though motor amps would not reduce at low loads, this is 5 more than offset by the much reduced motor power requirement of a turbo-expansion valve system.

Any type of compressor can be operated in conjunction with a co-axial Iy mounted turbine/compressor (turbo-expansion valve) O as a 2nd stage compressor. However, standard available compressors would require their effective compression ratio to be reduced by speed reduction, or by a combination of a lower synchronous speed (and horsepower) motor and reduced compression ratio — but with unchanged mass flow. 5

A single turbine may serve multiple evaporators whereby each evaporator has an inlet motorised control valve; the evapoi— ating temperature being sensed at the outlet of the expander, and superheat being sensed at the common suction inlet to the O refrigeration compressor. Also, multiple evaporators, each with a turbo-expansion valve and coupled compressor, may be connected to a single, common, 2nd stage compressor; as might suit the modulated operation of a large centrifugal compressor unit. 5 As the saturated vapour is expanded it is cooled, such that the enthalpy difference (RE) of the saturated vapour passing thro* the evaporator, per unit of mass flow, is increased: therefore, for the same evaporator cooling output as a conventional TEV system refrigerant mass flow and required compressor power re- O duces. The enthalpy/RE increase for a theoretically 100 % efficient expander would typically be in the order of plus 15 to 20 *, from which should be subtracted the efficiency loss of the expander, which could be as much as 35 % for a small unit to as little as 15 % for a mult i-megaWatt unit. With very small 5 coupled expander/lst stage compressors having efficiencies of say 65 * each, the 2nd stage compressor would still only require to input 57.75 it of the total pressure increase of an

RECTIFIED SHEET (RULE91) ISA/EP approMi oatel y 10 * reduced mass flow, such that the required power input would be reduced by 5O to 55 * compared to a conventional TEV system. With large centrifugal chillers having conpressor/turbine efficiencies in the order of 85 X, the re-r 5 quired power input would, similarly, be reduced by approκ 75 %.

Since a turbo-expansion valve is oilless, and since it is possible to also have a 2nd stage centrifugal compressor with an oilless planetary roller bearing traction drive gearing-up sys-

10 ten - evaporator superheat is also not necessarily required. It is, therefore, feasible to have a refrigeration system that would, in effect, become self-balancing without a superheat control system, e.g. saturated liquid carryover into the 1st stage compressor would have the effect of diminishing the teβ- 5 perature rise through the compressor which, in turn reduces condenser capacity which would reduce the flow of liquid to the evaporator and thus, compensat ingly, reduce carryover from the evaporator — a similar effect would occur if there were carryover of liquid from the flashing—off process — also the reverse O effects would tend to occur if there was too much superheat, or lack of flow into the flashing-off process.

The invention is now described with reference to the Figure 1 schematic flow diagram drawing of a typical example of it. 5

In Figure 1, the flow rate of liquid refrigerant 1 is controlled by valve S, Fitting 3 transforms from the liquid pipe size to the larger size of the inlet tract of the turbine 4 that accommodates the volume increase of liquid flashed—off in— O to saturated vapour after pressure reduction and diffusion through the perforated mesh, or plate 5. The static pressure of the saturated vapour is reduced as it' s velocity, and thereby it' s velocity pressure, is increased by the converging inlet tract and volute of turbine 4. The velocity pressure of the 5 saturated vapour discharged from the volute exerts a force on the vanes of the turbine 4 wheel that transmits this power via rotating shaft & to the centrifugal compressor 7. The pressure reduction of the saturated vapour through turbine 4 down to the evaporator β evaporating pressure, and work done by the turbine O 4 wheel, expands the saturated vapour such that it is expansion cooled thereby increasing the refrigeration effect (RE) of the evaporator β such that, for the same evaporator θ cooling capacity as a conventional TEV system there would be a reduced refrigerant mass flow rate. The flow rate through control valve 5 2 is directly controlled from the inputs of superheat sensors IO & 11, which may be overridden by sonic sensor 9 to maintain turbine 4 speed within manufacturer¹ s recommemded limits by controlling hot gas bypass 15 control valve 1&. Also, system capacity may be varied by controlling hot gas bypass 15 control O valve 1&.

Diffusion flashing-off of liquid refrigerant may be signifi- icant Iy enhanced by means of ultrasonically vibrating diffuser 5, or by ultrasonically vibrating liquid in fitting 3, or 5 liquid droplets immediately downstream of diffuser 5.

Due to the efficiency losses of turbine 4 and compressor 7 the

RECTIFIED SHEET(RULE91) ISA/EP pressure of the discharge gas 12 requires to be further increased by 2nd stage compressor 13 and motor 14. The required capacity of coapressor 13 and aotor IA being equal to the sun of the efficiency losses of turbine A and compressor 7. 5

Particularly where there may be little, or no sub-cooling of the liquid 1, and therefore little pressure drop thro' valve 2, fitting 3 should have the internal sidewalls airfoil shaped with a refleκ at the trailing edge, as per Patent No: GB2399552 O A, to ensure adequate static pressure regain through this fitting (in a practically short length) to inhibit unstable flash- ing-off of liquid prior to the perforated diffuser 5. Where the liquid nay be significantly sub-cooled and/or the condenser is located at a significant elevation above the turbo-expansion 5 valve such that there would be a significant pressure drop thro* valve 2_t such that liquid would spray diffuse out of it, then the diffuser plate and static pressure regain function of fitting 3 may be obviated. O Particularly where the condenser is significantly elevated above turbine A then valve 2 could be substituted with a liquid (refrigeration sub—cooled) turbine powering a variable speed generator, or the 2nd stage compressor - with superheat and turbine speed controls controlling the speed of the generator 5 and valve 2 controlling pump—down.

It should be noted that there is precedent in the design of 'Japanese* spl it/suit i-spl it/MRV systems for an evaporator to be supplied with saturated vapour, rather than with liquid via O a TEU, a distributor and distributor capillary tubes.

The turbo—expansion valve should be hermetically sealed within a casing to obviate the problem of moisture migration via the shaft seals into the refrigerant. 5

Motor IA is better cooled by the lower temperature suction gas 17, when hermetically sealed with compressor IA, by locating it upstream of compressor 7, as shown in Figure 2. O The Figure 1 configuration is suited to that of a condensing unit with a remote evaporator, whereas the Figure 2 arrangement is suited to that of a chiller or single package O/C unit.

Figure 3 shows an additional pressure reducing turbine lθ powering electrical generator 19 <that may be variable speed) - as nay be necessary to accommodate high pressure drops. Figure A shows an alternative arrangement whereby the outputs of tu»— bine lβ and motor 7 power compressor 13. If three (3) stages of pressure reduction may be required, the turbine lβ and gener— ator 19 of Figure 3 may be added upstream of the turbine lβ of Figure A. Alternatively, although much less efficiently, a 2nd stage of turbine pressure reduction could be obviated simply by absorbing the excess pressure that a single stage turbine cannot handle by absorbing this pressure thro* valve 2.

Figure 5A shows an automobile air-conditioning system whereby a multipiston swash-plate compressor 20 is belt and pulley 21

RECTIFIED SHEET(RULE 91) ISA/EP driven froa the engine crankshaft 22. A variation would be to substitute the turbo-expansion valve with a mult ipi ston swash- plate expander directly driving a oultipiston swash-plate compressor, or to have the co-axial shaft joining the expander and 5 coBpressor to be belt and pulley driven from the engine crankshaft such that there would not require to be a 2nd stage compressor.

Figure 5B shows a further variation whereby compressor 13 is O driven by turbine 23 from the engine exhaust. Turbine 23 and compressor 13 are shown as being Joined with insulated connectors 25. Bypass valve 24 modulates to the closed position as the evaporating temperature increases, and vice versa. Ceramic shaft bearing seals will likely be required. S

With auto aii—conditioning liquid refrigerant could be pumped down during periods of high output and stored for release during engine off conditions - with bypass valve 17 open. Bypass valve 17 can also be opened to regulate the output of the O oultipiston compressor 2O,

The refrigerant inlet arrangement of cooling coil β requires to be different than for a conventional TEV system. As shown in Figure 1, there would require to be one, or more, inlet coil 5 header plenums, or tubes, preferably with bell mouthed inlets to the coil tubes. Also, the inlet to each coil tube should, ideally, be fitted with inlet guide vanes such that droplets of liquid refrigerant in suspension are centrifuged onto the tube walls such that the rate of heat transfer is increased, or O other means be employed for imparting rotation to the saturated vapour, e.g. spiral inner fins, or 'rifling*.

The system may also be reverse—cycled to operate as a heat pump using two (2) conventional reversing valves 2β. The heating 5 cycle C. O. P. would similarly be increased in the order of 2OO to AOO)C, as it is for cooling. Figures 6A A 6B show schematic flow^' diagrams for, respectively, the cooling and heating cycles of such a system; with coil 26 being the indoor coil and coil 27 the outdoor coil. O

Calculations for typical Turbo-Expansion Valve System No: 1

For smaller air-conditioning systems the efficiencies of the turbo-expander and 1st stage centrifugal compressor can be ex- 5 pected to be in the order of SO to 75%. On this basis, with R.134a evaporating θ 1.5 C. and condensing <? 4O C, the following applies:

Refrigeraton effect <RE) of refrigerant flowing through the 0 evaporator of a conventional TEV system:

= 155 Nj/kg

Expansion RE increase via a 1OO* efficient turbine expander is equal to (but opposite to) that of a 1OO* efficient compressor¹ s enthalpy increase effect - follows a line of constant entropy from the evaporator outlet condition to the condensing

RECTIFIED SHEET (RULE91) ISA/EP pressure condition. By definition: a turbo-expansion valve - having a hypothetical 100* efficient expander co-axial Iy connected to a 1OO* efficient compressor - the ' work* of expansion will exactly equal the 'work' of compression, since the re- frigerant mass flow rate and pressure change is the same for both. a 26.9 kj/kg (obtained from entropy tables) At 7OX efficiency, actual RE increase:

= lθ.β kj/kg

Therefore, for the same evaporator cooling capacity as for a conventional TEV system, the mass flow rate reduction:

= lθ.β / 155 x lOO* = 11.9*

With the turbo-expander & 1st stage compressor both being 70* efficient the pressure increase of this 1st stage compressors

_* 0.7 x 0.7 x 100* = 49* _of total required

Therefore, the power required for a 2nd stage conventional compressor/condensing unit compared to that required for the single compressor of a conventional TEV system, when also adjusted for reduced mass flow rate:

= UOO - 11.9) K U. OO - O.49>* * 45*

Therefore, power saved by the turbo-expansion valve:

This results in the COP being increased by 2.2 times.

ALSO SEE THE FIGURE 7 to 11 PRESSURE - ENTHALPY R.22 DIAGRAMS. This particular diagram is used since it has both metric and Imperial scales on the same diagram.

Calculations for typical Turbo-Expansion Valve System No: 2

For the same evaporating and condensing temperatures as System No: 1, but as applied to a large, i.e. mult i—megaWatt, centri- fugal water chiller having a full load compressor efficiency of up to 87*, the following applies:

RE of refrigerant flowing through the evaporator of a conventional TEV system:

= 155 kj/kg

RE increase via 1OO* efficient expanders a £6.9 kj/kg

At 85* efficiency, actual RE increase:

RECTIFIED SHEET (RULE 91) ISA/EP = 22. 9 kj /kg

Therefore, for the same evaporator cooling capacity as for a conventional system, the mass flow rate reduces by: 5

* 22.9 / 155 x 100 % = IA.8*

With the turbo expander & 1st stage compressor both Θ5% efficient the pressure increase of the 1st stage conpressor: O

- O. β52 x O.852 x 1OO% - 72.6* of total

Therefore, the power required for the 2nd stage centrifugal conpressor, compared to that required for a conventional large 5 centrifugal chiller unit, when also adjusted for reduced mass mass flow rates

* <1OO - 14. θ> x (1.0OO - 0.726)% = 23.35% O Therefore, power saved by a turbo-expansion turbines

- 76.65%

This results in the COP being increased by 4.28 times. 5

N. B. 1

According to the ASHRAE Fundamentals Handbook, Chapter 1, under the heading 'Reversed Carnot Cycle*: "In actual cycles.... al 1 O processes of the refrigerant involve friction (converts into heat)....and cause degradation of available (cooling) energy" (words in brackets added). Most certainly there is turbulent and surface shear flow frictions involved in the flow through a TEV + distributor + capillary tubes that converts into heat 5 that, hitherto, has not been accounted for — probably because this heat evaporates soae small amount of liquid/saturated vapour droplets without any teβperature change and, therefore, is not easily measured without resource to CFD. However, a TEV is unlikely to be any aore efficient than a turbo—expander. There— O fore, since the TEV losses are eliminated when replaced by a turbo-expander the actual increase in performance of a turbo- expanson valve system compared to that of a conventional TEV system is estimated to be in the order of a further IO to 25%, according to application. THIS EFFECT SHOULD BE ALLOWED FOR 5 WHEN CONSIDERING THESE EXAMPLE CALCULATIONS. It is thought that in actual practice TEV losses are "hidden* in actual evaporator performance data,

N. B. 2

For maximum turbo-expansion valve energy recovery and expansion cooling condenser sub—cooling should be minimised.

N. B. 3

For c l ari t y : sub-coo l i ng, superheat and syst em pressure drop effect s have not been included for.

RECTIFIED SHEET (RULE 91) ISA/EP

Claims

C L O I M S

1) A refrigeration turbo-expansion valve comprises of an βdia- batic expansion turbine having an output shaft that inputs power to a motor and/or refrigeration compressor, means for diffusion flashing-off of liquid refrigerant upstream of the volute of the turbine, a control system receiving capacity demand input that controls means for varying the flow rate of refrigerant.

2) A refrigeration turbo-expansion valve as claimed in claim 1 wherein the control system receives input from evaporator superheat and/or turbine operating condition sensors.

3) ft refrigeration turbo-expansion valve as claimed in claims

1, or 2 wherein the output shaft of the turbine inputs power to an electric power generator that may be speed controllable.

4) A refrigeration turbo—expansion valve as claimed in any preceding claim wherein the turbine may be substituted with any other type of reversible adiabatic expansion engine.

RECTIFIED SHEET(RULE 91) ISA/EP